Experimental Investigation of R404A Indirect Refrigeration System Applied Internal Heat Exchanger: Part 2—Exergy Characteristics

Abstract

Although the R404A indirect refrigeration system (IRS) with an internal heat exchanger (IHX) and R744 as the secondary fluid has potential applications in supermarkets and hypermarkets, the exergy characteristics of this IRS have not been extensively investigated. In this study, the factors affecting the R744 exergy characteristics (degree of subcooling (DSB) and degree of superheating (DSP) of the R404A cycle, DSP of the R744 cycle, condensation temperature (CT) and cascade evaporation temperature (CET), and IHX efficiency) were experimentally evaluated to obtain basic data for the design of R404A IRS with R744 as the optimal secondary fluid. The main results can be summarized as follows: (1) Under given conditions, the smallest change in the system exergy destruction rate (EDR) according to the change in each parameter is the DSP of the R744 cycle (0.3–1%), followed by the DSB of the R404A cycle (6.1–8.8%), the IHX efficiency of the R404A cycle (3.8–14.3%), the DSP of the R404A cycle (11.7–15.9%), the CET (29.4–41.9%), and the CT (35–47%). (2) Also, in terms of the exergy efficiency of system (EES), the largest value was obtained for the DSP of the R404A cycle (2.4–12.7%), followed by the IHX efficiency of the R404A cycle (3–10.2%), the CET (2.2–8.7%), the CT (4–6.9%), the DSB of the R404A cycle (2.7–6.2%), and the DSP of the R744 cycle (0.04–1.2%). (3) In order to lower the system EDR, DSP, DSB, and IHX efficiency of the R404A cycle, the CET must be increased to the maximum, and to lower the DSP of the R744 cycle, the CT must be reduced to the minimum.

1. Introduction

The cascade refrigeration system (CRS) [1,2,3,4] and indirect refrigeration system (IRS) [5,6] are common refrigeration systems that are used in numerous supermarkets and hypermarkets worldwide. The CRS is primarily used for low-temperature (−45 to −25 °C) refrigeration, and the IRS is primarily used for medium-temperature (−25 to 5 °C) refrigeration. Single-phase brines, such as ethylene glycol, propylene glycol, ethyl alcohol, methyl alcohol, glycerol, potassium carbonate, calcium chloride, magnesium chloride, and sodium chloride, are commonly used in the IRS. The viscosity of single-phase brines increases rapidly with a decrease in the cooling temperature, while the power consumption of the pump (PCP) significantly increases. R744 (CO₂; carbon dioxide), which is not a single-phase brine, has attracted attention as a secondary fluid or refrigerant that can help address the limitations of single-phase brines. Accordingly, the use of R744 as a secondary fluid for the IRS has been investigated.

The risk of refrigerant leaks from supermarket refrigeration systems has been highlighted in numerous studies [7,8,9,10,11,12]. The annual refrigerant leakage from supermarkets in different countries is listed in

Table 1. Leakage rates of refrigerants from supermarket refrigeration systems [13,14,15].

Country	Year(s)	Annual Refrigerant Loss	References
The Netherlands	1999	3.2%	Hoogen et al., 2002 [10]
Germany	2000–2000	5–10%	Birndt et al., 2000 [7]; Haaf and Heinbokel, 2002 [9]
Denmark	2003	10%	Pedersen, 2003 [11]
Norway	2002–2003	14%	Bivens and Gage, 2004 [8]
Sweden	1993 1998 2001	14% 12.5% 10.4%	Bivens and Gage, 2004 [8]
United Kingdom	1998	14.4%	Radford, 1998 [12]
USA	2000–2002	13% and 18%, 19% and 22%	Bivens and Gage, 2004 [8]
USA	2020	25%	[14]
USA	2019–2021	40%, 59%, and 46%	[15]

[13,14,15]. R404A is classified as a group A1 refrigerant in the ASHRAE 34 safety group [16], and, hence, it is considered safe for supermarkets and hypermarkets [17]. R404A continues to be used in developing countries despite its high global warming potential.

Previous studies have investigated refrigeration systems with secondary refrigerant circulation. For example, Pearson (1995) [18] patented R744 as a secondary refrigerant that can easily evaporate, including a general hot gas defrosting system. Pearson (2005) [19] also used R744 as an easy-to-evaporate secondary refrigerant in Swedish supermarket refrigeration systems.

Pachai [20] examined the secondary refrigerant loop system installed in Helsingborg, Sweden, where a mixture of R290 and R170 (which are hydrocarbon-based refrigerants) was used as the primary refrigerant; R744 and propylene glycol were used as the secondary refrigerants for low and medium temperatures, respectively. The system has been functioning smoothly since 1997 and has been installed in more than 50 stores. Nilsson et al. [21] reported a 15 kW ice rink refrigeration system, with R744 as the secondary refrigerant. The primary refrigerant loop comprised a brine cooler, an expansion valve (EV), a cascade condenser (CC), and a compressor. The secondary refrigerant loop comprised a tank (liquid receiver) for R744 and a pump. The R744 (liquid receiver) tank was cooled by the cascade condenser of the refrigeration system that could adjust the brine temperature to −11 °C or lower.

Jahn [22] examined 300 kW air-conditioning systems used to protect computer servers from overloading at a bank. In that case, cold water condensed R744 at 6 °C. R744 was directed to the server cabinet using a centrifugal pump wherein it evaporated, and subsequently, R744 returned to the condenser. Hinde et al. [5] examined nine low-temperature refrigeration systems using R744 in the USA and Canada in early 2008. In the study, R404A or R507 were used as the primary refrigerants; propylene glycol was used as the medium-temperature secondary refrigerant; and R744 was used as the low-temperature secondary refrigerant. The systems were installed at various locations, ranging from small stores to large supermarkets. Their refrigeration capacity was in the range of 22–160 kW. They investigated the secondary refrigerant system installed in the largest store in North America (total floor area: 11,000 m²). This system could reduce R507 refrigerant charging by up to 90%; that is, the refrigerant charge in the large store could be reduced from 1633 to 159 kg. Propylene glycol and R744 were used as the secondary refrigerants for the medium-temperature and low-temperature systems. The carbon emissions were reduced by 7%.

Kaga et al. [23] developed a compact refrigeration system with an inverter compressor that used R600a as the primary refrigerant and R744 as the secondary refrigerant; the refrigerant was circulated via the thermosiphon effect. The coefficient of performance (COP) of the system was in the range of 2.1–2.6. Its refrigeration capacity was in the range of 170–230 kW depending on the evaporating temperature (ET), and the energy consumption was less than 5% of that of a direct expansion system with R134a. Sawalha and Palm [24] examined direct expansion refrigeration systems with R404A, ammonia, and propane refrigerants (R404A-DX) and an R404A/R744 indirect expansion refrigeration system (hereafter referred to as R404A/R744-IX when the R744 refrigerant is used as the brine); they conducted theoretical and experimental investigations of the effects of direct and indirect expansion on the pressure and temperature drop in the suction tube and compared the results. They reported that the gas phase, which is a single phase, flowed in the suction tube of R404A-DX. The gas-phase pressure drop for R404 or R502 was 3 to 3.5 times higher than that for R744 because the gas-phase pressure drop is proportional to the boiling point and molecular weight of the refrigerant. They also examined the pressure drop and temperature distribution in the air cooler, suction pipe, and evaporator for R404A with respect to R404A-DX and R404A/R744-IX as well as the power consumption of the compressor (PCC). The power consumption of R404A/R744-IX was lower than that of R404A-DX.

Despite numerous studies on the IRS (

Table 2. Studies on the IRS with R744 as the secondary fluid.

Authors	Year	Primary Fluid	Secondary Fluid	Analysis Method
Sawalha and Palm [24]	2003	R404A	R744	Experiment and Simulation
Pachai [20]	2004	R290 + R170	R744, PG	Experiment
Pearson [19]	2005	-	R744	Review
Nilsson et al. [21]	2006		R744	Experiment
Jahn [22]	2008	-	R744	Experiment
Hinde et al. [5]	2008	R404A, R507	R744, PG	Experiment
Kaga et al. [23]	2008	R600a	R744	Experiment

), the IRS with R404A as the primary refrigerant and R744 as the secondary fluid has not been extensively investigated. In addition, in the few cases where an internal heat exchanger (IHX) has been applied, simple comparisons have been made between systems with and without an IHX, and no extensive studies have been conducted for different efficiencies of the IHX, which improves the performance of refrigeration systems. Moreover, to the best of our knowledge, analyses of exergy on the refrigeration system with an IHX are lacking. Therefore, in this study, exergy was analyzed according to the degree of subcooling (DSB) and degree of superheating (DSP) of the R404A cycle, DSP of the R744 cycle, condensing temperature (CT), cascade evaporating temperature (CET), and IHX efficiency under constant conditions to identify the exergy characteristics of a low-temperature (−35 to 5 °C) R404A IRS using an IHX and R744 as the secondary fluid. This study aims to provide basic data for the optimal design of a low-temperature (−35 to 5 °C) R404A IRS using an IHX and R744 as the secondary fluid.

2. Test Apparatus and Data Analysis

An R404A IRS with an IHX that uses R744 as the secondary fluid was constructed to achieve a low temperature of −35 to 5 °C. The factors (the CET (${\mathrm{T}}_{\mathrm{E},\mathrm{CAS}}$), DSB ($\Delta {\mathrm{T}}_{\mathrm{SUB},\mathrm{R}404\mathrm{A}}$), DSP ($\Delta {\mathrm{T}}_{\mathrm{SUP},\mathrm{R}404\mathrm{A}}$, $\Delta {\mathrm{T}}_{\mathrm{SUP},\mathrm{R}744}$), R404A IHX efficiency (${\mathsf{\eta }}_{\mathrm{IHX},\mathrm{R}404\mathrm{A}}$), and CT of the R404A cycle (${\mathrm{T}}_{\mathrm{C}}$)) affecting the R404A IRS were identified via experiments to analyze exergy.

2.1. Test Apparatus

Figure 1 shows a photograph of the constructed IRS using R744 as the secondary fluid. A schematic of the test apparatus [25], used to analyze the exergy characteristics of the R404A IRS, is shown in Figure 2. The refrigerant temperature, pressure, mass flow rate (MFR) at the inlet and outlet of each IRS component, and PCC were measured for exergy analysis. The test apparatus, conditions (summarized in

Table 3. Test conditions: the R404A IRS using R744 as the secondary fluid [25].

Cycle	Component	Range	Unit
R404A cycle	CT	20, 30, 40 *, 50	°C
	IHX efficiency	0 *, 1, 2, 3, 4	stage
	DSB	0 *, 5, 10, 15, 20	°C
	DSP	10, 20 *, 30, 40	°C
	CET	−40, −30, −25 *, −20, −10, 0	°C
	Temperature difference of the cascade heat exchanger	5 *	°C
R744 cycle
	ET and cascade CT	−35, −25, −20 *, −15, −5, 5	°C
	DSB	1 *	°C
	DSP	10 *, 20, 30, 40	°C

* Standard conditions.

), and procedure are identical to those reported in a previous study [25].

Table 4. Main components of the test apparatus: the R404A IRS, using R744 as the secondary fluid [25].

Component	Characteristics
R404A compressor	Bock; model: HGX34P/380-4S. Displacement with 1450 min−1:33.1 m3 h−1. Number of cylinders: 4. Weight: 96 kg. Max. power consumption: 11.1 kW
R744 magnetic drive gear pump	Micropump; model: Series 5000. Flow range: 0–13.5 L/min. Maximum system pressure: 103 bar (1500 psi). Temperature range: −46 to 121 °C
Evaporator	Custom-built; type: horizontal double tube. Material: copper tube. Internal diameter of the inner tube: 11.46 mm. Internal diameter of the outer tube: 33.27 mm. Length of the evaporator: 8000 mm
Condenser	Alfa Laval; model: ACH-70X-50H-F. Heat exchanged: 38.44 kW. Heat transfer area: 2.45 m2
Cascade heat exchanger	Alfa Laval; model: ACH-70X-50H-F. Heat exchanged: 10.86 kW. Heat transfer area: 2.45 m2

summarizes the characteristics of the main components of the test apparatus.

2.2. Data Reduction

The thermal properties of R404A and R744 were calculated using REFPROP (version 8.01), which is a software program developed by the National Institute of Standards and Technology (NIST), Gaithersburg, MD, USA to examine refrigerant characteristics. The obtained thermal properties were used to investigate the COP of the system (${\mathrm{COP}}_{\mathrm{SYS}}$), MFR, and exergy characteristics of the R404A IRS with an IHX that uses R744 as the secondary fluid. The test data comprising ${\mathrm{COP}}_{\mathrm{SYS}}$ and exergy were analyzed in a manner identical to that reported in previous studies [25,26]. The exergy of the R404A IRS with an IHX was analyzed by calculating the exergy destruction rate (${\mathrm{Ex}}_{\mathrm{D}}$, kW) of each component using the formulae listed in

Table 5. Entropy and exergy balance formula for each component in the R404A IRS, using R744 as the secondary fluid [26,27,28,29].

Cycle	Component	${\dot{\mathbf{E}\mathbf{x}}}_{\mathbf{F},\mathbf{k}}$, kW	${\dot{\mathbf{E}\mathbf{x}}}_{\mathbf{P},\mathbf{k}}$, kW	${\dot{\mathbf{E}\mathbf{x}}}_{\mathbf{D},\mathbf{k}}$, kW
R404A cycle	Compressor (1→2)	${\mathrm{W}}_{\mathrm{COM},\mathrm{R}404\mathrm{A}}$	${\dot{\mathrm{Ex}}}_2-{\dot{\mathrm{Ex}}}_1$	${\mathrm{W}}_{\mathrm{COM},\mathrm{R}404\mathrm{A}}-$ (${\dot{\mathrm{Ex}}}_2$-${\dot{\mathrm{Ex}}}_1)$
	Condenser (2→4)	${\dot{\mathrm{Ex}}}_2-{\dot{\mathrm{Ex}}}_4$	${\mathrm{Q}}_{\mathrm{c}}\left[1-{\displaystyle \frac{{\mathrm{T}}_0}{{\mathrm{T}}_{\mathrm{C}}}}\right]$	(${\stackrel{`}{\mathrm{Ex}}}_2-{\stackrel{`}{\mathrm{Ex}}}_4$)$-{\mathrm{Q}}_{\mathrm{c}}\left[1-{\displaystyle \frac{{\mathrm{T}}_0}{{\mathrm{T}}_{\mathrm{C}}}}\right]$
	IHX (4→5, 8→1)	${\dot{\mathrm{Ex}}}_4-{\dot{\mathrm{Ex}}}_5$	${\dot{\mathrm{Ex}}}_8-{\dot{\mathrm{Ex}}}_1$	(${\dot{\mathrm{Ex}}}_4-{\dot{\mathrm{Ex}}}_5$) $-{\dot{\mathrm{Ex}}}_8-{\dot{\mathrm{Ex}}}_1$
	EV (5→6)	${{\dot{\mathrm{Ex}}}_5}^{\mathrm{M}}-{{\dot{\mathrm{Ex}}}_6}^{\mathrm{M}}$	${{\dot{\mathrm{Ex}}}_6}^{\mathrm{T}}-{{\dot{\mathrm{Ex}}}_5}^{\mathrm{T}}$	(${{\dot{\mathrm{Ex}}}_5}^{\mathrm{M}}-{{\dot{\mathrm{Ex}}}_6}^{\mathrm{M}}$)$-$(${{\dot{\mathrm{Ex}}}_6}^{\mathrm{T}}-{{\dot{\mathrm{Ex}}}_5}^{\mathrm{T}}$)
	Cascade heat exchanger (6→8, 12→14)	${\dot{\mathrm{Ex}}}_6-{\dot{\mathrm{Ex}}}_8$	${\dot{\mathrm{Ex}}}_{14}-{\dot{\mathrm{Ex}}}_{12}$	(${\dot{\mathrm{Ex}}}_6-{\dot{\mathrm{Ex}}}_8$) $-$ (${\dot{\mathrm{Ex}}}_{14}-{\dot{\mathrm{Ex}}}_{12}$)
R744 cycle
	Pump (14→15)	${\mathrm{W}}_{\mathrm{PUM},\mathrm{R}744}$	${\dot{\mathrm{Ex}}}_{15}-{\dot{\mathrm{Ex}}}_{14}$	${\mathrm{W}}_{\mathrm{PUM},\mathrm{R}744}-$(${\dot{\mathrm{Ex}}}_{15}-{\dot{\mathrm{Ex}}}_{14}$)
	Evaporator (15→12)	${\dot{\mathrm{Ex}}}_{15}-{\dot{\mathrm{Ex}}}_{12}$	${\mathrm{Q}}_{\mathrm{E}}\left[{\displaystyle \frac{{\mathrm{T}}_0}{\left({\mathrm{T}}_{\mathrm{E}}+\Delta {\mathrm{T}}_{\mathrm{E}}\right)}}-1\right]$	(${\dot{\mathrm{Ex}}}_{15}-{\dot{\mathrm{Ex}}}_{12}$)$-{\mathrm{Q}}_{\mathrm{E}}\left[{\displaystyle \frac{{\mathrm{T}}_0}{\left({\mathrm{T}}_{\mathrm{E}}+\Delta {\mathrm{T}}_{\mathrm{E}}\right)}}-1\right]$

, and the total COP (COP_SYS) of the system was calculated using Formula (1):

$${\mathrm{COP}}_{\mathrm{SYS}}={\displaystyle \frac{{\mathrm{Q}}_{\mathrm{E}}}{{\mathrm{W}}_{\mathrm{COM},\mathrm{R}404\mathrm{A}}+{\mathrm{W}}_{\mathrm{PUM},\mathrm{R}744}}}$$

(1)

The mass flow ratio (${\dot{\mathrm{m}}}_{\mathrm{RATIO}}$) of the CRS was calculated using Formula (2):

$${\dot{\mathrm{m}}}_{\mathrm{RATIO}}={\displaystyle \frac{{\dot{\mathrm{m}}}_{\mathrm{R}404\mathrm{A}}}{{\dot{\mathrm{m}}}_{\mathrm{R}744}}}$$

(2)

Exergy can be obtained using the following formula under given temperature and pressure conditions [26].

$${\dot{\mathrm{Ex}}}_{\mathrm{k}}=\dot{\mathrm{m}}\left[\left({\mathrm{h}}_{\mathrm{k}}-{\mathrm{h}}_0\right)-{\mathrm{T}}_0\left({\mathrm{s}}_{\mathrm{k}}-{\mathrm{s}}_0\right)\right]$$

(3)

$${\dot{\mathrm{Ex}}}_{\mathrm{k}}-{\dot{\mathrm{Ex}}}_{\mathrm{m}}=\dot{\mathrm{m}}\left[\left({\mathrm{h}}_{\mathrm{k}}-{\mathrm{h}}_{\mathrm{m}}\right)-{\mathrm{T}}_0\left({\mathrm{s}}_{\mathrm{k}}-{\mathrm{s}}_{\mathrm{m}}\right)\right]$$

(4)

The exergy efficiency of the system was calculated using Formula (5) with reference to Sun et al. [26].

$${\mathsf{\eta }}_{\mathrm{Ex},\mathrm{SYS}}={\displaystyle \frac{{\dot{\mathrm{Ex}}}_{\mathrm{F},\mathrm{SYS}}-{\dot{\mathrm{Ex}}}_{\mathrm{D},\mathrm{SYS}}}{{\dot{\mathrm{Ex}}}_{\mathrm{F},\mathrm{SYS}}}}$$

(5)

Here, ${\dot{\mathrm{Ex}}}_{\mathrm{F},\mathrm{SYS}}$ is the sum of the power consumption of the compressor and the pump.

The formulae presented in previous papers [25,26] and Table 5 were used to compute the exergy destruction rate (EDR) of each component, EDR of the system (${\mathrm{Ex}}_{\mathrm{D},\mathrm{SYS}}$), and exergy efficiency of the system (${\mathsf{\eta }}_{\mathrm{Ex},\mathrm{SYS}}$) according to the CT, CET, IHX efficiency, DSP, and DSB of the R404A cycle and ET, DSP, and MFR of the R744 cycle. These factors affect the COP and exergy of the R404A IRS with R744. Figure 3 shows a conceptual diagram of the R404A IRS with an IHX that uses R744 as the secondary fluid.

2.3. Uncertainties

The test data used in engineering interpretation or design are not accurate. The uncertainties in the test results were estimated using the formulae proposed by Kline and McClintock [30] and Moffat [31] and are listed in

Table 6. Parameters and estimated uncertainties.

Parameter	Unit	Uncertainty
MFR	[kg/min]	$\pm$0.0100
PCC	[kW]	$\pm$0.0350
PCP	[W]	$\pm$0.2350
COP of the total IRS	[/]	$\pm$0.0135
EDR of the system	[kW]	$\pm$0.0175
Temperature	[°C]	$\pm$0.2000
$\Delta {\mathrm{T}}_{\mathrm{CAS}}$	[°C]	$\pm$0.4000
Pressure	[kPa]	$\pm$5.2700
$\Delta \mathrm{P}$(Pressure drop)	[kPa]	$\pm$0.0100
MFR of coolant	[kg/h]	$\pm$7.5300

.

3. Results

The exergy characteristics of the IRS with an IHX that uses R744 as the secondary refrigerant were analyzed to obtain basic design data. The EDR of each component, EDR of the system, and exergy efficiency of the system (EES) were examined according to the CT, R404A IHX efficiency, CET, DSB of the R404A cycle, and DSP of the R404A and R744 cycles.

3.1. Effect of the DSP

3.1.1. Effect of the DSP of the R404A Cycle

The changes in the EDR of each component, system EDR, and EES with an increase in the DSP of the R404A cycle were experimentally evaluated under constant conditions (Q_E = 6.73–6.82 kW; T_E = −21.5–−21.3 °C; T_CAS,E = −25.4–−24.9 °C; T_C = 39.8–40.1 °C; ΔT_SUB,R404A = 1.1–1.4 °C; ΔT_SUP,R744 = 9.5–10.6 °C; ΔT_SUB,R744 = 1.3–1.9 °C; ΔT_CAS = 3.5–3.9 °C; and η_IHX,R404A = 0). The DSP was increased from 9.9 to 40.5 °C in approximately 10 °C increments. The observed trends are shown in Figure 4.

As the DSP of the R404A cycle in the IRS increased from 9.9 to 40.5 °C (in approximately 10 °C increments), the EDR of the system decreased by 11.7–15.9% and the EES increased by 2.4–12.7%. The variation in the COP of the system was similar to the variation in the EES. The changes in the EDRs of the pump (3.8–5.2% increase) and evaporator (0.1–2% decrease) of the R744 cycle were negligible. The EDR of the condenser of the R404A cycle increased by 4.1–12.6%, whereas those of the compressor, EV, and cascade heat exchanger (CHX) of the cycle decreased by 20.3–23.8%, 6.3–9.1%, and 22.3–47.5%, respectively.

The change in the EDR of the R744 cycle with an approximately 10 °C increase in the DSP of the R404A cycle was negligible because there was no change in the MFR or all the conditions in Table 5. In contrast, the EDRs of the compressor, EV, and CHX decreased in the R404A cycle because the MFR decreased by 6.3–7.7%. However, the EDR of the condenser (Ex_D,C) increased. The EDR of the condenser was calculated by adding the EDR of the inlet and outlet refrigerant (Ex_F,C: +) and the EDR due to condensation heat (Ex_P,C: −), as shown in Table 5. The EDR of the refrigerant (Ex_F,C) as well as the EDR due to condensation heat (Ex_P,C) decreased with a decrease in the refrigerant flow rate. However, the inlet entropy (IE) of the condenser (s₂) increased despite minimal changes in the outlet entropy (OE) of the condenser (s₄). Consequently, the reduction in the EDR by condensation heat (Ex_P,C) is significant due to the reduction in condensation heat, while the reduction in the EDR of the condenser inlet and outlet refrigerant (Ex_F,C) is insignificant. Hence, the EDR of the condenser (Ex_D,C) increased. Therefore, it can be concluded that the EDR of the system (Ex_D,SYS) decreases and the EES and COP of the system increase with an increase in the DSP of the R404A cycle. These results are in good agreement with the results reported by Yılmaz et al. [32].

3.1.2. Effect of the DSP of the R744 Cycle

The variations in the EDR of each component, exergy efficiency, and EDR of the system with an increase in the DSP of the R404A cycle were experimentally evaluated under constant conditions (Q_E = 6.85–6.89 kW; T_E = −21.3–−21.0 °C; T_CAS,E = −25.4–−24.9 °C; T_C = 39.8–40.1 °C; ΔT_SUP,R404A = 19.7–20.1 °C; ΔT_SUB,R404A = 1.1–1.3 °C; ΔT_SUB,R744 = 1.1–1.6 °C; ΔT_CAS = 3.4–3.9 °C; and η_IHX,R404A = 0). The DSP was increased from 9.7 to 39.5 °C in approximately 10 °C increments. The observed trends are shown in Figure 5.

As the DSP of the R744 cycle in the IRS increased from 9.7 to 39.5 °C, the EDR of the system increased by 0.3–1% and the EES decreased and then increased by 0.04–1.17%. The EDRs of the pump and the evaporator in the R744 cycle decreased by 1.1–3.2% and 6.8–19.4%, respectively. The EDRs of the compressor, condenser, and EV in the R404A cycle increased or decreased by 0.4–1.4%, 0.9–2%, and 0.3–2.9%, respectively, whereas that of the CHX increased by 8.1–27.5%. As the DSP of the R744 cycle increased by approximately 10 °C, the EDRs of the pump and evaporator in the R744 cycle decreased because the MFR in the cycle decreased by 3–3.7%; in contrast, the EDRs of the condenser, compressor, and EV in the R404A cycle remained largely constant because the MFR, DSP, and DSB were constant. The EDR of the CHX (Ex_D,CAS) was calculated as the sum of the EDRs of the cascade evaporator (CE) (Ex_D,E,CAS: +) and CC (Ex_D,C,CAS: −). The change in the EDR of the CE in the R404A cycle was negligible. In contrast, the EDR of the CC in the R744 cycle decreased as the R744 MFR decreased. The EDR of the CHX increased by 8.1–27.5%. The EDRs of the condenser, compressor, and EV in the R404A cycle did not change. The sum of the EDRs of the pump and evaporator in the R744 cycle decreased by 5.1–14.9%. Consequently, the EDR of the system increased.

In other similar studies [33] and theory, it was found that the EES decreases slightly as the DSP of the R744 cycle increases. However, the reason for the slight decrease and increase in the EES in the present experimental results is that the temperature and pressure were kept as constant as possible in the experiments, and the small differences in temperature and pressure are thought to be the cause of the error in the EES. The effect of increasing the DSP in the R744 cycle can be considered insignificant for the R404A IRS.

3.2. Effect of the DSB

Effect of the DSB of the R404A Cycle

The changes in the EDR of each component, EDR of the system, and EES with an increase in the DSB of the R404A cycle were experimentally evaluated under constant conditions (Q_E = 6.87–6.95 kW; T_E = −21.8–−21.4 °C; T_CAS,E = −25.5–−24.7 °C; T_C = 39.8–40.3 °C; ΔT_SUP,R404A = 19.6–20.4 °C; ΔT_SUP,R744 = 9.7–10.2 °C; ΔT_SUB,R744 = 1.1–1.5 °C; ΔT_CAS = 3.1–3.5 °C; and η_IHX,R404A = 0). The DSP was increased from 1.4 to 19.5 °C in approximately 4.5 °C increments. The observed trends are shown in Figure 6.

As the DSB of the R404A cycle in the IRS increased from 1.4 to 19.5 °C (in approximately 4.5 °C increments), the EDR of the system decreased by 6.1–8.8% and the EES increased by 2.7–6.2%. The increase in the EES was accompanied by an increase in the COP of the system. In addition, as the DSB of the R404A cycle increased from 1.4 to 19.5 °C, the EV, evaporator, and MFR of the R744 cycle, as well as the EDR of the CHX, hardly changed; in contrast, the EDRs of the condenser, compressor, and EV in the R404A cycle decreased by 3–7.2%, 1.3–14.1%, and 16.4–21.7%, respectively. The EDRs of the pump and evaporator in the R744 cycle did not change with an increase in the DSB of the R404A cycle because the MFR and other conditions in the R744 cycle did not change based on the formulae in Table 5. Moreover, the EDR of the CHX did not change because the increase in the ethalpy difference in the inlet and outlet of the CE (s₈ − s₆) due to the decrease in the IE of the CE (s₆) caused by the increase in the DSB was offset by the reduction in the MFR caused by energy equilibrium even though there was no change in the EDR of the condenser in the R744 cycle and the evaporation heat capacity (EHC) in the R404A cycle. The EDR of the compressor in the R404A cycle decreased because the MFR decreased by 5.5–6%. The EDR of the condenser was calculated as the sum of the refrigerant EDR at the condenser inlet and outlet (Ex_F,C: +) and the EDR due to the condensation heat capacity (Ex_P,C: −), as shown in Table 5. With an increase in the DSB, the IE of the condenser (s₂) remained constant, whereas the OE (s₄) decreased, thereby increasing the entropy difference (s₂ − s₄) between the inlet and outlet of the condenser. However, the refrigerant EDR and the EDR due to condensation heat capacity decreased. The EDR of the condenser decreased because the decrease in the refrigerant EDR was large. The EDR of the EV was calculated as the sum of the EDR due to pressure drop (Ex_F,EXP,R404A or Ex_D,EXP,R404A^M) and the EDR due to temperature drop (Ex_P,EXP,R404A or Ex_D,EXP,R404A^T), as shown in Table 5. The EDR of the EV decreased because the EDR by pressure drop decreased under the dominant influence of the MFR reduction as the CT increased while the EDR by temperature drop increased due to the large temperature difference.

Thus, the EDR of the system decreases and the EES and COP of the system increase as the DSB of the R404A cycle increases. These results are in good agreement with the results reported by Yılmaz et al. [32].

3.3. Effect of CT

The changes in the EDR of each component, EDR of the system, and EES with an increase in the IRS CT were experimentally evaluated under constant conditions (Q_E = 6.81–6.88 kW; T_E = −21.6–−21.3 °C; T_CAS,E = −25.3–−24.4 °C; ΔT_SUP,R404A = 19.5–20.5 °C; ΔT_SUB,R404A = 1.0–1.5 °C; ΔT_SUP,R744 = 9.8–10.7 °C; ΔT_SUB,R744 = 1.2–1.4 °C; ΔT_CAS = 2.8–3.5 °C; and η_IHX,R404A = 0). The CT was increased from 20.1 to 49.7 °C in approximately 10 °C increments. The observed trends are shown in Figure 7.

As the CT increased from 20.1 to 49.7 °C (in approximately 10 °C increments), the EDR of the system increased by 35–47%, whereas the EES decreased by 4–6.9%. The variation in the COP of the system was similar to that of the EES. The changes in the EDRs of the pump and evaporator in the R744 cycle were negligible because the MFR, ET, CT, DSP, and DSB of the cycle remained constant. The EDRs of the condenser, compressor, and EV in the R404A cycle increased by 39.8–47.8%, 23.9–56.3%, and 57.3–71.1%, respectively.

With an increase in the CT, the IE (s₁) of the compressor in the R404A cycle remained constant, whereas the OE (s₂) increased, thereby increasing the difference in the compressor IE and OE (s₂ − s₁). The EDR of the compressor was calculated as the sum of the PCC (Ex_F,COM,R404A: +) and the refrigerant EDR at the compressor inlet and outlet (Ex_P,COM,R404A: −). The PCC and the refrigerant EDR increased with increasing CT. The EDR of the compressor increased because the magnitude of the increase in the PCC (a positive value) was larger than the magnitude of the increase in the refrigerant EDR. The EDR of the condenser was calculated as the sum of the refrigerant EDR (Ex_F,C: +) at the condenser inlet and outlet and the EDR (Ex_P,C: −) due to the condensation heat capacity. Both the IE and OE of the condenser (s₂ and s₄) increased with increasing CT. The differences (s₂ − s₄) in the IE and OE of the condenser decreased because the magnitude of the increase in the OE (s₄) of the condenser was larger than that at the IE (s₂). The refrigerant EDR and the EDR due to the condensation heat capacity increased because the effect of the increase in the MFR was dominant. The EDR of the condenser increased because the magnitude of the increase in the refrigerant EDR was larger than the magnitude of the increase in the EDR due to the condensation heat capacity. The IE and OE of the EV (s₄ and s₆) increased with increasing CT; the differences in the IE and OE (s₆ − s₄) increased because the magnitude of the increase in the OE was larger. The EDR of the EV increased because the EDR increased due to a pressure drop and the EDR decreased due to a temperature drop. The EDR of the CHX was constant because the reduction in the enthapy difference in the inlet and outlet (s₈ − s₆) of the CE due to the increase in the CE inlet entropy (s₆) and constant outlet entropy (s₈) were offset by the rise in the MFR of the R404A cycle caused by an energy equilibrium even though the condensation heat capacity and MFR of the R744 cycle and the EHC of the R404A cycle were constant. Thus, the EDR of the system increases as the CT of the R404A cycle increases, thereby reducing the COP of the system and EES. These results are in good agreement with results reported by Hendri et al. [34], Parekh and Tailor [35], and Kilicarslan and Hosoz [36].

3.4. Effect of the CET

The variations in the EDR of each component, EDR of the system, and EES with an increase in the CET were experimentally evaluated under constant conditions (Q_E = 13.37–13.45 kW; T_E = −35.5–3.5 °C; T_C = 39.8–40.2 °C; ΔT_SUP,R404A = 19.7–20.4 °C; ΔT_SUB,R404A = 1.4–1.6 °C; ΔT_SUP,R744 = 19.8–20.3 °C; ΔT_SUB,R744 = 1.0–1.6 °C; ΔT_CAS = 3.1–3.5 °C; and η_IHX,R404A = 0). The ET was increased from −39.2 to 0 °C in ~10 °C increments. The observed trends are shown in Figure 8.

As shown in Figure 8, the EDR of the system increased by 29.4–41.9% and the EES decreased by 2.2–8.7% as the CET increased from −39.2 to 0 °C. The variation in the COP of the system was similar to the variation in the EES. In the R744 cycle, the EDR of the pump decreased by 4.4–15%, whereas that of the evaporator increased by 6.3–18.1%. In the R404A cycle, the EDRs of the compressor, condenser, EV, and CE increased by 23.8–52.3%, 26.6–63.2%, 35.6–86.6%, and 1.9–3.9%, respectively. In the R744 cycle, only the EDR of the pump decreased as the CET increased. This can be attributed to the decrease in the MFR and PCC of the R744 cycle, with no change in the refrigerant EDR as the CET increased, based on the formulae in Table 5.

The EDR of the evaporator increased despite the reduction in the MFR of the R744 cycle. This is because the OE of the evaporator (s₁₂) increased and the IE (s₁₅) decreased, thereby leading to an increase in the difference in the evaporator IE and OE (s₁₂ − s₁₅). The EDR of the evaporator in the R744 cycle (Ex_D,E) was calculated as the sum of the refrigerant EDR at the evaporator inlet and outlet (Ex_F,E: +) and the EDR due to the EHC (Ex_P,E: −). Because the influence of the entropy difference is more dominant despite the reduction in the MFR, the refrigerant EDR and the EDR due to the EHC of the R744 cycle increased. The magnitude of the increase in the refrigerant EDR was larger than the magnitude of the increase in the EDR due to the EHC; hence, the EDR of the evaporator in the R744 cycle increased. The increases in the EDRs of the condenser, compressor, EV, and CE in the R404A cycle were attributed to the dominant effect of the increase in the MFR.

Thus, the EDR of the system decreases as the CET increases, thereby increasing the COP of the system and EES. These results are in good agreement with the results reported by Shilliday et al. [37] and Yi [38].

3.5. Effect of IHX Efficiency

The variations in the EDR of each component, EDR of the system, and EES with an increase in the number of IHX stages in the R404A cycle were experimentally evaluated under constant conditions (Q_E = 6.84–6.86 kW; T_E = −6.5–−6.3 °C; T_CAS,E = −10.3–−9.6 °C; T_C = 39.8–40.3 °C; ΔT_SUP,R404A = 19.4–20.3 °C; ΔT_SUB,R404A = 1.1–1.5 °C; ΔT_SUP,R744 = 9.8–10.5 °C; ΔT_SUB,R744 = 1.0–1.5 °C; ΔT_CAS = 3.1–3.6 °C; and η_IHX,R404A = 0). The number of IHX stages was increased from zero to four. The observed trends are shown in Figure 9.

As the number of IHX stages in the R404A cycle increased from zero to four, the IHX efficiency increased from zero to 26.4, 35.2, 42.8, and 49.8. The EDR of the system decreased by 3.8–14.3% whereas the EES increased by 3–10.2%. The EES increased with an increase in the COP of the system. The changes in the EDRs of the pump, evaporator in the R744 cycle, and CE were minimal. In the R404A cycle, the EDRs of the compressor and EV decreased by 3.5–19.3% and 0.97–25.2%, respectively, whereas those of the condenser and IHX increased by 0.8–13.8% and 5.5–14.6%, respectively. The change from stage zero to stage one was omitted because the efficiency of stage zero was zero.

As the number of IHX stages in the R404A cycle increased, the EDRs of the pump, evaporator in the R744 cycle, and CE remained constant because the MFR, ET, CT, DSP, and DSB of the R744 cycle were constant. The IE and OE (s₁ and s₂) at the compressor in the R404A cycle increased with an increase in the number of IHX stages. The entropy difference (s₂ − s₁) between the inlet and outlet decreased because the magnitude of the increase in the IE (s₁) was larger than the magnitude of the increase in the OE (s₂). Accordingly, both the PCC and the refrigerant EDR at the inlet and outlet of the compressor decreased. The decrease in the EDR of the compressor was attributed to the larger decrease in the PCC and the decrease in the MFR. The entropy difference (s₈ − s₆) between the inlet and outlet of the EV increased because the IE (s₆) decreased and the OE (s₈) remained constant. Hence, the EDR of the EV decreased with an increase in the number of IHX stages. Consequently, the EDR decreased due to a pressure drop, and the EDR increased due to a temperature drop.

The EDR of the condenser (Ex_D,C) was calculated as the sum of the EDR at the condenser inlet and outlet (Ex_F,C: +) and the EDR due to the condensation heat capacity (Ex_P,C: −). The differences in the IE and OE of the condenser (s₂ − s₄) increased with an increase in the number of IHX stages because the IE (s₂) increased and the OE (s₄) remained constant. Moreover, the MFR decreased. Consequently, both Ex_F,C and Ex_P,C decreased. However, Ex_D,C increased because the magnitude of the decrease in Ex_P,C was higher than the magnitude of the decrease in Ex_F,C.

The EDR of the IHX (Ex_D,IHX,R404A) was calculated as the sum of the EDR at the inlet (condenser outlet) and outlet (EV inlet) of the IHX on the high-pressure side (Ex_F,IHX,R404A: +) and the EDR at the inlet (evaporator outlet) and outlet (compressor inlet) of the IHX on the low-pressure side (Ex_P,IHX,R404A: −). Ex_F,IHX,R404A and Ex_P,IHX,R404A increased with an increase in the number of IHX stages in the R404A cycle. Ex_D,IHX,R404A increased because the rate of increase in Ex_P,IHX,R404A was lower than that of Ex_F,IHX,R404A. Ex_D,IHX,R404A increased despite a decrease in the MFR because the effect of the increase in Ex_P,IHX,R404A was dominant.

Therefore, high IHX efficiency in the R404A IRS can increase the COP and exergy of the system by decreasing the EDR of the system.

Figure 4, Figure 5, Figure 6, Figure 7, Figure 8 and Figure 9 show the relationships among the EDR, COP of the system, and EES of the IRS using R744 under different test conditions. The EDR of the system was inversely proportional to the COP, and the behaviors of the EES and COP of the system were similar. Therefore, the EDR of each component must be decreased to increase the COP of the system.

3.6. Comparison of Experimental and Performance Analysis Data

Figure 4 shows the experimental data on the EDR of each component, EDR of the system, EES, and COP of the system according to the DSP of the R404A cycle in the IRS. Figure 10 shows the results obtained via performance analysis.

The analysis conditions used to obtain the results shown in Figure 10 were as follows: Q_E = 6.71 kW, T_E = −21 °C, T_CAS,E = −24.9 °C, T_C = 40.0 °C, ΔT_SUB,R404A = 1.3 °C, ΔT_SUP,R744 = 10 °C, ΔT_SUB,R744 = 1.6 °C, ΔT_CAS = 3.7 °C, η_IHX,R404A = 0, and η_COM,R404A = 0.555.

The analysis conditions were matched with the test conditions to the best possible extent. Despite marginal variations in the conditions, the trends observed in the experiments were identical to the trends obtained via performance analysis. In addition, the experimental and performance analysis results were in good agreement for the EDR of each component (Ex_D), EDR of the system (Ex_D,SYS), EES ($\eta$_Ex,SYS), and COP of the system (COP_SYS) according to the DSB of the R404A cycle, DSP of the R744 cycle, CT, CET, and IHX efficiency.

4. Conclusions

In this study, exergy characteristics were theoretically identified and factors affecting a low-temperature (−35 to 5 °C) R404A IRS using R744 as the secondary fluid were experimentally analyzed. This study aimed to provide basic data for the optimal design of the R404A IRS with an IHX that uses R744 as the secondary fluid. The results obtained in this study can help ensure energy efficiency and economic feasibility by optimizing the total exergy of the IRS and thereby promote eco-friendly IRS operation and maintenance. The conclusions of this study are as follows:

• In order to lower the system EDR, the DSP, DSB, IHX efficiency and CET of the R404A cycle must be increased to the maximum, and the DSP of the R744 cycle and CT must be reduced to the minimum.
• Under the given conditions, the smallest change in the system EDR according to the change in each parameter is the DSP of the R744 cycle (0.3–1%), followed by the DSB of the R404A cycle (6.1–8.8%), IHX efficiency of the R404A cycle (3.8–14.3%), DSP of the R404A cycle (11.7–15.9%), CET (29.4–41.9%), and CT (35–47%).
• Also, in terms of the EES, the largest value obtained was the DSP of the R404A cycle (2.4–12.7%), followed by the IHX efficiency of the R404A cycle (3–10.2%), CET (2.2–8.7%), CT (4–6.9%), DSB of the R404A cycle (2.7–6.2%), and DSP of the R744 cycle (0.04–1.2%).
• The DSP of the R744 cycle does not need to be considered when designing the R404A IRS using R744 as a secondary fluid because the changes in the EDR of the IRS and EES are insignificant, with ranges of 0.3–1% and 0.04–1.2%, respectively.
• Among the EDRs of the components of the IRS, the largest one is the compressor, followed by the expansion valve, condenser, evaporator, CHX, pump, and IHX. Therefore, from energy and environmental aspects, great efforts must be made to reduce the EDR in the compressor and expansion valve when designing the R404A IRS using R744 as a secondary fluid.
• The trend of the system EDR shows the exact opposite of that of the system exergy efficiency and COP. The EES and COP of the system increase with a decrease in the system EDR.
• The EES can replace the COP of the system.
• The position and size of energy inefficiency, which cannot be identified via COP analysis, can be identified by evaluating the EDR of each component, thereby making it possible to maximize energy efficiency.
• Energy efficiency can be maximized via detailed, quantitative, and advanced exergy analysis. This may be considered in future work.
• The R404A refrigerant is regulated in developed countries, including Europe. Therefore, novel alternatives such as R448A and R449A must be examined.