Comparative Analysis of Transcritical CO₂ Heat Pump Systems With and Without Ejector: Performance, Exergy, and Economic Perspective

Abstract

To promote renewable energy utilization and enhance the environmental friendliness of refrigerants, this study presents a novel experimental investigation on a transcritical CO₂ double-evaporator heat pump water heater integrating both air and water sources, designed for high-temperature hot water production. A key innovation of this work lies in the integration of an ejector into the dual-source system, aiming to improve system performance and energy efficiency. This study systematically compares the conventional circulation mode and the proposed ejector-assisted circulation mode in terms of system performance, exergy efficiency, and the economic payback period. Experimental results reveal that the ejector-assisted mode not only achieves a higher water outlet temperature and reduces compressor power consumption but also improves the system’s exergy efficiency by 6.6% under the condition of the maximum outlet water temperature. Although the addition of the ejector increases initial manufacturing and maintenance costs, the payback periods of the two modes remain nearly the same. These findings confirm the feasibility and advantage of incorporating an ejector into a transcritical CO₂ compression/ejection heat pump system with integrated air and water sources, offering a promising solution for efficient and environmentally friendly high-temperature water heating applications.

1. Introduction

Heat pump systems are widely recognized as an effective solution for utilizing renewable energy and reducing carbon emissions in residential and commercial space heating and domestic hot water applications [1]. However, growing environmental concerns regarding the use of synthetic refrigerants in conventional heat pump systems have prompted an urgent need to adopt environmentally friendly alternatives [2,3]. In this context, carbon dioxide (CO₂) has emerged as a promising natural refrigerant, offering advantages in terms of environmental sustainability, safety, and thermodynamic performance [4].

Transcritical CO₂ heat pump systems can harness the favorable thermophysical properties of CO₂ in both subcritical and supercritical states and have thus become a focus of intensive research in recent years. These systems exhibit several notable performance advantages: (1) High water outlet temperatures can be achieved due to the low critical temperature of CO₂ (31.1 °C) [5]. In supercritical conditions, the CO₂ temperature at the compressor outlet can exceed 100 °C [6], allowing for efficient heat transfer across a wide temperature range, which is well-matched to the temperature glide of water heating [7,8]. (2) Stable operation in low ambient temperatures is possible, as subcritical CO₂ possesses high latent heat, enabling effective heat absorption under cold conditions. Experimental studies have shown that transcritical CO₂ heat pumps can operate reliably at temperatures as low as –15 °C [9,10]. (3) A compact system design is facilitated by the high operating pressure and pressure ratio (typically above 3.5) of CO₂ in transcritical cycles, which supports the miniaturization of key components [11].

Despite these advantages, one of the main limitations of transcritical CO₂ cycles is the substantial pressure difference across the expansion valve, which traditionally results in energy loss through throttling [12]. To recover this pressure energy, ejectors have been introduced into transcritical CO₂ systems as an effective means of improving energy efficiency [5,13]. Prior studies have demonstrated the potential benefits of this approach.

Zhu et al. [14] experimentally investigated an air-source transcritical CO₂ heat pump system with an ejector and observed a 10.3% performance improvement and an increase in outlet water temperature. Chen et al. [15] reported a 22% increase in COP by integrating an ejector into a transcritical refrigeration cycle. Taleghani et al. [16] further found that optimizing the heat transfer area ratio significantly enhanced system COP and heating capacity under fixed ejector configurations. These studies collectively confirm that ejector integration can enhance system performance in conventional single-source CO₂ heat pump systems.

However, as the demand for multi-source renewable energy utilization grows, conventional single-source systems struggle to meet increasingly complex heating requirements. Integrating multiple heat sources—such as air, water, or solar—with ejector technology has therefore become a new direction in heat pump system research. For instance, Qin et al. [17] proposed a novel compression/ejection transcritical CO₂ heat pump system that utilizes both air and water sources to support simultaneous heating and cooling. Their simulation and experimental results showed that the dual-evaporator configuration achieved the highest total COP, despite slightly reduced heating capacity compared to conventional systems. Erdinc et al. [18] introduced a solar-assisted ejector-augmented heat pump system, demonstrating through simulation that the solar input enhances the ejector’s mainstream inlet enthalpy, thereby improving ejector effectiveness. Similarly, Liu et al. [19] developed a jet compression air-source heat pump system with dual heat sources and conducted a comprehensive 4E analysis (energy, exergy, economic, and environmental). Their findings indicated significant performance improvements over traditional systems.

These advancements underline the potential of combining multiple renewable heat sources with ejector technology to improve the overall performance and sustainability of heat pump systems. However, comparative experimental studies on multi-source transcritical CO₂ compression/ejection systems remain limited in the current literature. In particular, there is a lack of in-depth analysis of how ejector integration affects energy efficiency, exergy performance, and economic feasibility in such configurations.

To address this gap, the present study experimentally investigates a novel transcritical CO₂ double-evaporator heat pump water heater that integrates air and water heat sources, with and without an ejector. By comparing system performance, exergy efficiency, and the payback period under both circulation modes, this study aims to clarify the role of the ejector in improving system efficiency and practical viability. The results will provide valuable insights into the design and optimization of advanced CO₂-based heat pump systems for renewable energy applications.

2. Experimental System and Calculation Model

2.1. Experimental System

2.1.1. System Circulation Mode

The conventional transcritical CO₂ heat pump system typically consists of a compressor, a gas cooler, an expansion valve, an internal heat exchanger, and a single evaporator. To reduce compressor power consumption and enhance the utilization grade of heat sources, a novel transcritical CO₂ double-evaporator compression/ejection heat pump system is proposed. In this advanced configuration, an ejector is incorporated to elevate the CO₂ inlet temperature to the compressor by recovering expansion work, thereby improving system efficiency [20]. Additionally, the system integrates both water-source and air-source evaporators, enabling the utilization of multiple renewable heat sources and enhancing operational flexibility under varying environmental conditions.

The system circulation model is presented in Figure 1, and the experimental platform is illustrated in Figure 2.

To evaluate the impact of the ejector on system performance, three stop valves are incorporated into the proposed system, enabling flexible switching between different circulation modes. By adjusting the valve positions, the system can operate either with or without the ejector. Specifically, (1) the ejector mode (TCPE, Figure 3a) is activated by opening stop valves A and C and closing stop valve B, allowing the ejector to function within the cycle. (2) The conventional mode (TCP, Figure 3b) is realized by closing stop valves A and C and opening stop valve B, resulting in a direct series connection between the two evaporators, bypassing the ejector.

The corresponding pressure–enthalpy (P–h) diagrams for the two circulation modes are illustrated in Figure 3, clearly depicting the thermodynamic differences between the configurations.

A comparison of the theoretical cycles in Figure 3a,b indicates that the integration of the ejector can significantly raise the compressor inlet pressure and increase the enthalpy difference across the evaporator, which are critical factors for improving cycle efficiency. To validate these theoretical insights and quantify the actual performance enhancement, this study proceeds with a detailed experimental investigation of the two circulation modes.

2.1.2. Component-Specific Parameters

The configuration of the main system components and the corresponding measurement equipment are presented in Figure 1 and

Table 1. Detailed parameters of measurement equipment.

Name	Medium	Measurement Span	Number	Accuracy
Mass flowmeter	CO2	0 m3∙h−1 to 6 m3∙h−1	2	0.15%
Temperature sensor	CO2; water	−50 °C to 200 °C	13	0.15%
Pressure sensor	CO2	0 MPa to 25 MPa	9	0.2%
Volumetric flowmeter	Water	0 m3∙h−1 to 15 m3∙h−1	2	0.5%

, respectively.

2.1.3. Experimental Method

The circulation mode of the system can be switched by adjusting the positions of the three stop valves A, B, and C. Within each mode, the compressor discharge pressure is regulated using the expansion valves, allowing for variations in discharge pressure from 8700 kPa to 9960 kPa. Once the system reaches a steady-state condition, operational data are collected via a programmable logic controller (PLC). During the experiments, the inlet temperatures of the gas cooler and the low-temperature evaporator are maintained at 30 °C and 10 °C, respectively. The ambient temperature, which is also the air inlet temperature for the high-temperature evaporator, is controlled at 30 °C.

2.2. Calculation Model

2.2.1. Coefficient of Performance

The coefficient of performance (COP) is determined by calculating the heat output on the water side of the gas cooler and the compressor’s power consumption [21]. The specific method for calculating the water-side heating capacity is as follows:

$$\dot{Q_{GC,w}}={\dot{V}}_w·{\rho }_w·{Cp}_w·\left(T_{w,out}-T_{w,in}\right)$$

(1)

where $\dot{Q_{GC,w}}$ is the water heat input, $V_w$ is the volume flow of water, ${\rho }_w$ is the water density, ${Cp}_w$ is the specific heat capacity of water at a constant pressure, and $T_{w,out}$ and $T_{w,in}$ are the water outlet and inlet temperatures.

COP can be calculated as [22]

$$COP={\displaystyle \frac{\dot{Q_{GC,w}}}{\dot{W_{CO}}}}$$

(2)

where $\dot{W_{CO}}$ is the power consumption of the compressor.

2.2.2. Ejector Entrainment Ratio

The entrainment ratio is one of the most critical parameters influencing the performance of the ejector. It is defined as the ratio of the mass flow rate of the secondary stream to that of the primary stream and is expressed as follows:

$$\omega ={\displaystyle \frac{{\dot{m}}_2}{{\dot{m}}_1}}$$

(3)

where $\omega$ is the entrainment ratio, ${\dot{m}}_2$ is the mass flow rate of the secondary flow, and ${\dot{m}}_1$ is the mass flow rate of the primary flow.

2.2.3. Ejector Pressure Rise Ratio

The pressure rise ratio characterizes the influence of the ejector outlet pressure on the inlet pressure of the secondary flow. It is defined as follows:

$$r={\displaystyle \frac{P_3}{P_2}}$$

(4)

where $r$ is the pressure rise ratio, $P_3$ is the ejector outlet pressure, and $P_2$ is the secondary flow inlet pressure.

2.2.4. Exergy Analysis Model

The exergy analysis in this study includes the evaluation of exergy output, exergy destruction, and exergy efficiency. The specific calculation method is detailed as follows [17]:

$$E_P=\left(h_{GC,in}-h_{GC,out}\right)·\left(1-{\displaystyle \frac{T_0·\left(s_{w,out}-s_{w,in}\right)}{h_{w,out}-h_{w,in}}}\right)$$

(5)

where $E_P$ is the exergy production, $h_{GC,in}$ and $h_{GC,out}$ are the CO₂ inlet and outlet enthalpy of the gas cooler, $T_0$ is the reference temperature, $s_{w,in}$ and $s_{w,out}$ are the entropy of the water inlet and outlet, and $h_{w,in}$ and $h_{w,out}$ are the enthalpy of the water inlet and outlet. Since the system investigated functions as a water heater, the exergy output is calculated solely based on the hot water generated by the gas cooler.

The calculation process of exergy destruction of each component and the compressor exergy destruction is as follows:

$$E_{d,CO}=T_0·\left(s_{CO,out}-s_{CO,in}\right)$$

(6)

where $E_{d,CO}$ is the compressor exergy destruction and $s_{CO,in}$ and $s_{CO,out}$ are the entropy of the compressor’s inlet and outlet.

The gas cooler exergy destruction is calculated as follows:

$$E_{d,GC}=\left(h_{GC,in}-h_{GC,out}\right)-T_0·\left(s_{GC,in}-s_{GC,out}\right)-E_P$$

(7)

where $E_{d,GC}$ is the gas cooler exergy destruction and $s_{GC,in}$ and $s_{GC,out}$ are the entropy of the gas cooler’s inlet and outlet.

The internal heat exchanger exergy destruction is as follows:

$$E_{d,IH}=T_0·\left[\left(s_{IH,LP,in}-s_{IH,LP,out}\right)+\left(s_{IH,HP,out}-s_{IH,HP,in}\right)\right]-\left[\left(h_{IH,LP,in}-h_{IH,LP,out}\right)+\left(h_{IH,HP,out}-h_{IH,HP,in}\right)\right]$$

(8)

where $E_{d,IH}$ is the internal heat exchanger exergy destruction and $LP$ and $HP$ represent the low-pressure side and high-pressure side of the internal heat exchanger, respectively. The inlet and outlet of the high-pressure side are the outlet of the gas cooler and the inlet of the expansion valve, respectively, and the inlet and outlet of the low-pressure side are the outlet of the high-temperature evaporator and the inlet of the compressor, respectively.

The expansion valve exergy destruction is calculated as follows:

$$E_{d,EX}=\left(h_{EX,in}-h_{EX,out}\right)-T_0·\left(s_{EX,in}-s_{EX,out}\right)$$

(9)

where $E_{d,EX}$ is the expansion valve exergy destruction.

The ejector exergy destruction is calculated as follows:

$$E_{d,EJ}=\left[\left({\displaystyle \frac{1}{\omega +1}}\right)·h_{EJ,1}+\left({\displaystyle \frac{\omega }{\omega +1}}\right)·h_{EJ,2}-h_{EJ,3}\right]-T_0·\left[\left({\displaystyle \frac{1}{\omega +1}}\right)·s_{EJ,1}+\left({\displaystyle \frac{\omega }{\omega +1}}\right)·s_{EJ,2}-s_{EJ,3}\right]$$

(10)

where $E_{d,EJ}$ is the ejector exergy destruction, $\omega$ is the ejector entrainment ratio, and 1, 2, and 3 are, respectively, the inlet of the primary flow, the inlet of the second flow, and the outlet of the ejector. The inlet of the primary flow is connected to the outlet of the high-pressure side in the internal heat exchanger. The inlet of the second flow is connected to the outlet of the low-temperature evaporator. The outlet of the ejector is connected to the inlet of the high-temperature evaporator.

The evaporator exergy destruction is calculated as follows:

$$E_{d,EV}=T_0·\left(s_{EV,out}-s_{EV,in}\right)-\left(h_{EV,out}-h_{EV,in}\right)$$

(11)

where $E_{d,EV}$ is the evaporator exergy destruction.

The exergy efficiency is calculated as follows [23]:

$${\eta }_E={\displaystyle \frac{E_P}{E_d}}$$

(12)

where ${\eta }_E$ is the system exergy efficiency and $E_d$ is the sum of exergy destruction of all components.

2.2.5. Economic Analysis Model

In this study, the economic analysis is used to compare the cost and the product and obtain the payback period of the system. Among them, the cost in this study includes the system manufacturing cost, the operating cost, the maintenance cost, and the production value. Because the system is only used for water heating, the product is only the heat capacity of the gas cooler.

The operating cost of this study is only related to the power consumption of the compressor, and the calculation method is as follows [24]:

$$OC=\dot{W_{CO}}·N·Z_{el}$$

(13)

where $OC$ is the operating cost; $N$ is the compressor operating time, which is assumed to be 6000 h; and $Z_{el}$ is obtained according to the stepped electricity price of Jiangsu Province (China), which is 0.61 CNY⋅kW⁻¹⋅h⁻¹.

The maintenance cost can be obtained by multiplying the maintenance cost by the maintenance ratio. The maintenance ratio in this study is 1%.

The product value is calculated as follow [25]:

$$PV=N·\dot{Q_{GC,w}}·Z_h$$

(14)

where $PV$ is the production value and $Z_h$ is the heating price, which is assumed as 0.373 CNY ⋅ kW⁻¹⋅h⁻¹ in this analysis.

The specific calculation method of the payback period is [26]

$$PP={\displaystyle \frac{SMC}{PV-OC-MC}}$$

(15)

where $PP$ is the payback period, $SMC$ is the system manufacturing cost, and $MC$ is the maintenance cost.

2.2.6. Uncertainty Analysis

The uncertainty analysis by Arat et al. was used [27,28].

The average measurement value ($\overline{X}$) is

$$\overline{X}={\displaystyle \frac{\sum X_m}{n}}$$

(16)

where $n$ is the number of measurements and $X_m$ is the measured value. The specific values are shown in Table 1.

The total uncertainty (${\delta }_R$) is

$${\delta }_R=\sqrt{{\displaystyle \frac{{\sum }_{m=1}^n{\left(X_m-\overline{X}\right)}^2}{\left(n-1\right)}}}$$

(17)

The total uncertainty range for different operating conditions and parameters ranges from 0.13% to 0.52%.

3. Results and Discussions

3.1. Basic Performance Discussion

3.1.1. Water Outlet Temperature

As shown in Figure 4, the outlet water temperature ($T_{\mathrm{w}}$) of both circulation modes increases initially with discharge pressure and then begins to decline. This trend is consistent with findings in previous studies. The transcritical CO₂ system with an ejector (TCPE) exhibits a clear peak $T_{\mathrm{w}}$ at a discharge pressure of 9800 kPa, reaching a maximum of 86.1 °C. In contrast, the conventional system (TCP) shows a continuously rising trend within the experimental range (up to 9960 kPa) without reaching a peak. This suggests that the incorporation of an ejector shifts the optimal discharge pressure for heat transfer to a lower value, enhancing system efficiency under moderate-pressure conditions. Notably, the temperature difference between the two systems at the same discharge pressure reaches as much as 7.7 °C. This significant improvement highlights the role of the ejector in enhancing the heat pump’s heating performance.

From a thermodynamic perspective, the key factor determining the outlet water temperature is the heating capacity of the gas cooler, which is directly related to the enthalpy difference between the inlet and outlet streams. The ejector improves this enthalpy difference by enhancing the heat absorption of the low-temperature evaporator. Specifically, the ejector increases the suction capacity of the low-temperature evaporator, raising its outlet enthalpy. This elevated enthalpy is then transferred to the gas cooler as higher inlet enthalpy, effectively broadening the enthalpy difference across the gas cooler. As a result, more heat is transferred to the water stream, increasing $T_{\mathrm{w}}$. The integration of an ejector in the transcritical CO₂ system not only enhances heat absorption at the evaporator level but also improves the overall heat transfer capacity of the gas cooler. This leads to a substantial increase in water outlet temperature, even under lower optimal discharge pressures.

3.1.2. Compressor Power Consumption

The compressor power consumption of both the TCP and TCPE systems increases steadily with rising discharge pressure in the range of 8700 kPa to 9960 kPa, as shown in Figure 5. This is consistent with thermodynamic expectations: higher discharge pressure requires more work input for gas compression. However, the magnitude and growth trend of compressor power consumption differ notably between the two systems. At a discharge pressure of 8700 kPa, the compressor power consumption in the TCP system is approximately 0.86 kW higher than that in the TCPE system. This difference gradually diminishes as the discharge pressure increases, reaching only 0.24 kW at 9960 kPa. This trend highlights the beneficial role of the ejector in reducing the load on the compressor, particularly at lower discharge pressures.

The fundamental mechanism lies in the ability of the ejector to pre-compress the suction gas before it enters the compressor. By increasing the compressor inlet pressure, the ejector effectively reduces the compression ratio between the inlet and outlet of the compressor. Since compressor power consumption is strongly dependent on this pressure ratio, a lower ratio results in decreased energy input for the same mass flow. However, as discharge pressure increases, the contribution of the ejector to raising the suction pressure becomes proportionally smaller. The pressure gain from the ejector remains relatively constant or saturates, while the required discharge pressure continues to rise. Consequently, the relative impact of the ejector weakens, leading to a narrowing of the power consumption gap between the two systems. This analysis suggests that the energy-saving effect of the ejector is most significant under moderate or low discharge pressures, where its impact on reducing the compressor pressure ratio is maximized. This characteristic makes the TCPE system particularly suitable for applications where high performance is desired under mid-range operating pressures.

3.1.3. System COP

Figure 6 illustrates the variation in the COP for both the TCP and TCPE systems across a discharge pressure range of 8700 kPa to 9960 kPa. As observed, regardless of the discharge pressure, the COP of the TCPE system consistently outperforms that of the TCP system. The maximum COP difference, 0.34, occurs at a discharge pressure of 8700 kPa, while the minimum difference of 0.17 appears near 9960 kPa. This clearly reflects the efficiency advantage introduced by the ejector under lower-pressure conditions.

Despite this relative advantage, the absolute value of COP for both systems decreases monotonically as discharge pressure increases. This trend correlates well with the rising compressor power consumption previously discussed in Figure 5. Since COP is calculated as the ratio of heating capacity to compressor power input, even though the heating capacity increases slightly with discharge pressure (due to enhanced gas cooler performance), the more pronounced increase in compressor power leads to an overall decline in system COP.

The ejector contributes to higher COP in the TCPE system primarily through two mechanisms: (1) Reduction in the compression ratio: By elevating the compressor inlet pressure, the ejector decreases the pressure ratio across the compressor, thereby reducing its power demand; (2) Enhanced evaporator enthalpy gain: The ejector facilitates the better utilization of the low-temperature evaporator, increasing the inlet enthalpy to the gas cooler and, consequently, improving heat transfer effectiveness.

However, as discharge pressure increases, the ability of the ejector to further reduce the compression ratio or enhance preheating becomes limited. This explains the gradually narrowing COP difference between the two systems at higher pressures. From a design and operation perspective, this analysis suggests that the TCPE system demonstrates a more significant performance advantage at moderate discharge pressures, making it especially suitable for scenarios where energy efficiency is prioritized under variable or low-to-mid-pressure conditions.

These findings align with the conclusions drawn by Qin et al. [4] for conventional transcritical CO₂ heat pump systems, where the compressor power input dominates COP behavior despite simultaneous increases in heating capacity. Thus, optimizing compressor energy consumption remains a key strategy for improving overall system performance.

3.1.4. Ejector Performance

Figure 7 illustrates the variation in the entrainment ratio and pressure lift ratio with discharge pressure, both of which are key indicators for evaluating ejector performance. As observed, the entrainment ratio increases monotonically with discharge pressure, while the pressure lift ratio exhibits a decreasing trend. This opposing behavior highlights a performance trade-off: higher discharge pressures enhance the ability of the ejector to entrain more secondary flow (from the low-temperature evaporator) but simultaneously reduce its capacity to elevate the pressure of the secondary flow.

It is noteworthy that the peak coefficient of performance (COP) of the system occurs at a discharge pressure of approximately 8700 kPa, coinciding with the peak in the pressure lift ratio. This correlation suggests that an increased pressure lift ratio can effectively raise the compressor suction pressure, lower the pressure ratio, and reduce compressor power consumption, thereby improving overall system performance. Furthermore, a combined analysis of the entrainment and pressure lift ratios reveals their synergistic effect on enhancing the outlet water temperature. An increased entrainment ratio implies greater heat absorption in the low-temperature evaporator, which raises the enthalpy of the flow entering the gas cooler and ultimately increases the water outlet temperature. Meanwhile, the pressure lift ratio contributes directly by elevating the compressor inlet pressure, optimizing the thermodynamic path of the refrigerant. Although both factors positively impact system efficiency, the effect of the entrainment ratio appears slightly more pronounced, primarily due to its stronger influence on the evaporative heat absorption and subsequent enthalpy gain.

In summary, optimizing ejector performance requires a balance between entrainment capacity and pressure lift capability. The experimental results indicate that a moderate discharge pressure (around 8700 kPa) yields optimal system performance, suggesting a well-coordinated interaction between these two performance parameters at that point.

3.2. Thermal Economy Analysis

3.2.1. Exergy Destruction and Exergy Efficiency

Figure 8 summarizes the variation in energy and exergy parameters with discharge pressure, using the analytical framework proposed by Arslan et al. [29,30]. The analysis focuses on two critical operating conditions for the system: the condition corresponding to the maximum coefficient of performance (COP) and the condition yielding the highest water outlet temperature. These two points represent key performance extremes for the evaluation of thermodynamic behavior. As discharge pressure increases, both the compressor power consumption and the heat input on the water side exhibit a steady and nearly linear upward trend. This increase reflects the growing energy demand required to overcome the higher pressure differential across the compressor. Correspondingly, the exergy destruction and exergy production in the TCPE system both display distinct peaks near the discharge pressure that yields the maximum outlet water temperature. This suggests that although more energy is being consumed, the system is simultaneously able to generate a higher amount of useful work (exergy output) and utilizes energy more effectively under these specific conditions.

In contrast, the variation in exergy metrics for the conventional TCP system is relatively minor across the pressure range, indicating a less dynamic response to changes in discharge pressure. This relative insensitivity implies the limited capability of the TCP configuration to enhance performance under varying conditions, reinforcing the advantage of the ejector-assisted TCPE system. A particularly notable observation is the exergy efficiency trend. In the TCP system, exergy efficiency peaks at the discharge pressure that coincides with the highest water outlet temperature, suggesting a narrow optimal operating window. On the other hand, the exergy efficiency of the TCPE system increases consistently with discharge pressure, ranging from 35.8% to 46.1%. This continuous improvement highlights the superior adaptability and performance potential of the TCPE configuration, which is attributed to the ejector’s ability to improve thermodynamic matching between components and recover expansion work.

The integration of the ejector in the TCPE system enhances both energy utilization and exergy performance, especially under high-load conditions. This underlines its effectiveness in reducing irreversibility, boosting system efficiency, and achieving better thermodynamic performance compared to the conventional configuration.

Figure 9 illustrates the exergy destruction distribution across individual components under the operating condition corresponding to the maximum outlet water temperature for both the TCP and TCPE circulatory modes. With the exception of the expansion valve, all components in the TCP system exhibit lower exergy destruction compared to their counterparts in the TCPE system. This is primarily due to the additional thermodynamic interactions introduced by the ejector and the dual-evaporator configuration in TCPE. The ejector—exclusive to the TCPE configuration—accounts for the highest exergy destruction among all components, reaching 15.8 kJ∙kg⁻¹. This indicates that while the ejector contributes to performance enhancement by recovering expansion work and improving system thermodynamics, it also introduces significant irreversibility. Therefore, optimizing the ejector design and operating conditions could yield substantial efficiency improvements. Furthermore, the compressor in both configurations shows considerable exergy destruction, recorded at 15.5 kJ∙kg⁻¹ for TCP and 14.0 kJ∙kg⁻¹ for TCPE. These values underscore the compressor’s dominant role in overall exergy losses, especially under high-load conditions. The slight reduction in compressor exergy destruction in TCPE suggests that the ejector also helps mitigate some of the compression-related losses by increasing the inlet pressure and reducing the pressure ratio.

The exergy analysis reveals that the compressor and ejector are the two most significant sources of irreversibility in the high-temperature operation of the TCPE system. Therefore, future system optimization efforts should focus on these components to further enhance the thermodynamic performance and energy efficiency of transcritical CO₂ heat pump systems.

Another key reference condition for system evaluation is the state at which the maximum coefficient of performance (COP) is achieved, as depicted in Figure 10. Under this condition, the exergy destruction of all components, except the ejector, is consistently lower in the TCPE system compared to TCP. Most notably, the largest disparity is observed in the compressor, where the exergy destruction difference reaches 6.0 kJ∙kg⁻¹. This observation mirrors the COP variation trend shown in Figure 5, where the performance advantage of TCPE over TCP is most pronounced at lower discharge pressures.

As discharge pressure increases, the COP difference between the two systems gradually narrows. A similar trend is observed in exergy destruction: the difference in exergy destruction between TCP and TCPE decreases with rising discharge pressure. This parallel suggests a strong correlation between system COP and total exergy destruction. It can thus be inferred that COP is the dominant factor influencing exergy performance in transcritical CO₂ systems.

A comparison between Figure 9 and Figure 10 reveals that the total exergy destruction under the maximum COP condition is consistently lower than that under the maximum outlet water temperature condition. This confirms that the system operates in a more thermodynamically favorable state at maximum COP, with minimized irreversibility and improved energy utilization. Consequently, the maximum COP state represents the optimal operating point under the current experimental conditions, offering the best balance between heating output and energy efficiency.

As illustrated in Figure 11, the exergy efficiency of the system is determined based on the relationship between exergy production and exergy destruction. According to the analysis in Section 3.1.1, the maximum outlet water temperature achieved by the TCPE system is higher than that of the TCP system. Consequently, the exergy production of TCPE at this condition is also greater, as higher water temperatures correspond to higher thermodynamic potential for useful work. A comparison of the two operating conditions, the maximum water outlet temperature and maximum COP, shows that the exergy production is consistently higher under the former. This is attributed to the greater thermal output delivered by the gas cooler in the high-temperature state, despite the associated increase in exergy destruction.

However, under the maximum COP condition, the exergy production of the TCP system surpasses that of TCPE. This seemingly counterintuitive result highlights a key finding: the system operating at maximum COP minimizes compressor work relative to heating output, thereby reducing overall exergy destruction. The data in Figure 11 further support the conclusion that exergy destruction, particularly in the compressor, is closely linked to COP. As COP increases, the efficiency of energy conversion improves, leading to reduced entropy generation and hence lower exergy destruction. This analysis underscores the trade-off between maximizing heat output and optimizing energy efficiency and reinforces the role of the compressor as a critical component influencing the system’s exergy performance. The specific values of the four exergy efficiencies of two circulatory modes are shown in

Table 2. Exergy efficiency under different states and circulatory modes.

System State Circulatory Mode	Tw,MAX	COPMAX
TCP	46.1%	42.1%
TCPE	39.5%	42.6%

.

3.2.2. Payback Period

The calculation of the payback period in this study incorporates several key economic parameters: the system manufacturing cost (SMC), the operating cost (OC), the maintenance cost (MC), and the production value (PV). A comparative summary of these parameters for the two circulatory modes, TCP and TCPE, is presented in Figure 12.

Since the ejector is not part of the TCP system configuration, the only difference in manufacturing cost between the two systems is attributed to the inclusion of the ejector in the TCPE system. Accordingly, the total manufacturing costs for the TCP and TCPE systems are CNY 127,800 and CNY 129,800, respectively. A detailed breakdown of component-specific costs is provided in

Table 3. Component manufacturing cost.

Component	Cost (CNY)
Compressor	75,000
Gas cooler	8000
Internal heat exchanger	2000
Low-temperature evaporator	4500
High-temperature evaporator	15,000
Expansion valve	300
Ejector	2000
Pipe	3000
Installation	20,000

, based on the actual purchase prices of the equipment used in this study.

The operating cost is directly influenced by the power consumption of the compressor. Therefore, under the operating condition corresponding to the maximum water outlet temperature, both systems exhibit higher operating costs compared to those under the maximum COP condition, reflecting the increased energy demand. Conversely, the production value is associated with the thermal output, which is primarily determined by the water outlet temperature. Thus, the production value is greater in the high-temperature state than under the condition of maximum COP.

The payback period, which is calculated using the aforementioned parameters, is summarized in

Table 4. Payback period under different states and circulatory modes.

System State Circulatory Mode	Tw,MAX	COPMAX
TCP	1.6 years	1.6 years
TCPE	1.5 years	1.4 years

. The analysis reveals that although TCPE entails a slightly higher initial investment due to the addition of the ejector, its improved thermal performance—particularly at higher outlet water temperatures—contributes to greater production value and potentially a more favorable return on investment.

As shown in Table 4, the payback period for all operating states and both circulatory modes remains consistently around one and a half years. This observation indicates that, within the context of this study, the system manufacturing cost and production value exert a relatively minor influence on the overall payback period. It can therefore be inferred that while the addition of auxiliary components may enhance the system’s thermal performance and increase the production value, it simultaneously raises the initial manufacturing cost. As a result, the net impact on the ratio of production value to investment is constrained, leading to only marginal changes in economic performance indicators like the payback period.

In conclusion, the implementation of the ejector in the transcritical CO₂ heat pump system effectively increases the outlet water temperature and improves the COP. However, the enhancement in exergy efficiency and economic viability, represented by the payback period, is limited. This suggests that while thermodynamic advantages can be achieved through system innovation, their translation into economic benefits may be less pronounced due to cost–performance trade-offs.

3.3. Comprehensive Evaluation of Ejector Utilization

The integration of an ejector into the TCPE system exhibits both notable advantages and inherent limitations compared to the conventional TCP system. A radar chart, as illustrated in Figure 13, presents a comprehensive comparison of the two systems across multiple performance dimensions, thereby visualizing the overall effectiveness of ejector utilization.

From a performance perspective, the addition of the ejector enhances the system’s heat absorption capability at the evaporator side and raises the CO₂ inlet pressure of the compressor. This leads to a higher water outlet temperature (with a peak difference of 7.7 °C compared to TCP) and lower compressor power consumption at low to medium discharge pressures. Furthermore, the COP of TCPE consistently outperforms that of TCP, with the largest observed difference being 0.34 at a lower discharge pressure. However, as the discharge pressure increases, the benefit of the ejector diminishes due to its limited effect on further increasing the compressor inlet pressure, resulting in reduced relative gains in COP and power savings.

In terms of exergy efficiency, TCPE displays more complex behavior. Although the ejector increases the system’s heat utilization potential, it also introduces significant exergy destruction, especially within the ejector itself, which was found to have the highest exergy loss (15.8 kJ∙kg⁻¹) among all components. While TCPE generally shows higher exergy efficiency at higher discharge pressures, this improvement is not uniform. In the optimal COP condition, TCPE exhibits lower exergy destruction across most components compared to TCP, especially in the compressor. However, at the maximum outlet water temperature condition, TCPE’s overall exergy destruction is actually higher due to the substantial loss introduced by the ejector. This highlights a trade-off between thermal enhancement and exergy degradation caused by introducing the ejector.

From an economic standpoint, the cost–benefit analysis shows that the system payback period remains nearly unchanged (~1.5 years) for both circulatory modes and across different operating states. Although the ejector increases the production value by raising the outlet temperature, this advantage is largely offset by its additional manufacturing cost. Thus, while the ejector contributes to performance enhancement, its effect on economic returns is limited under the test conditions in this study.

In summary, the ejector significantly improves system thermal performance and enables greater COP at optimized conditions, demonstrating technical effectiveness in heat pump enhancement. However, its benefits are diminished at higher discharge pressures, and it introduces notable exergy losses and only marginal economic gains. Future work should focus on optimizing the ejector design to minimize internal losses and enhance entrainment performance or explore adaptive control strategies to selectively engage the ejector under favorable conditions, thereby maximizing its net benefit.

4. Conclusions

This study experimentally compares the performance, exergy efficiency, and economic viability of a transcritical CO₂ heat pump system operating in two circulatory modes: with and without an ejector. Based on the analysis, the following conclusions are drawn:

• The integration of an ejector significantly enhances system thermal performance. Compared with the conventional system (TCP), the ejector-assisted system (TCPE) achieves a maximum increase of 7.7 °C in outlet water temperature, reduces the average compressor power consumption by 4.8%, and improves the system’s COP by 6.5%.
• A positive correlation is observed between the system’s exergy production and the outlet water temperature; that is, higher outlet temperatures lead to greater exergy output.
• Exergy destruction and exergy efficiency exhibit significant sensitivity to discharge pressure, with their trends primarily governed by variations in compressor power consumption.
• Although the ejector improves overall system performance, it also leads to increased manufacturing costs. Consequently, the impact of the circulatory mode on economic performance is limited, with the average system payback period remaining approximately 1.5 years.

5. Future Work and Perspectives

Future research on transcritical CO₂ heat pump systems should focus on optimizing the ejector structure to improve the entrainment ratio and internal flow characteristics, reducing internal energy losses through component refinement and the advanced control of key operating parameters, and exploring new circulation configurations to achieve lower energy consumption and higher system efficiency.