Performance Analysis and Optimization of an Air–Water Heat Pump System for Residential Applications in a Subtropical Monsoon Climate

Building energy consumption, particularly in heating, ventilation, and air conditioning systems, represents a significant portion of global energy use, accounting for approximately 30–50% of total energy consumption in residential buildings [1]. With the increasing focus on energy efficiency and sustainability, heat pumps have emerged as a promising solution to reduce primary energy consumption [2]. These systems are particularly relevant in the context of nearly zero-energy buildings, where optimizing system performance is essential [3].

Ground-source heat pump (GSHP) systems have been widely adopted due to their high energy efficiency and environmental benefits [4]. The development of the GSHP technology in China has accelerated since the 1980s, supported by government initiatives and a growing demand for sustainable energy solutions [5]. However, the GSHP systems face challenges in certain climatic conditions, particularly in Northern China, where performance can be limited by the local environment [6]. Consequently, there has been a growing interest in air-source heat pumps (ASHPs) as an alternative, particularly in temperate regions where they are more suitable for distributed radiant floor heating [7].

ASHP systems offer several advantages, including lower installation costs and adaptability to various residential and commercial applications [8]. Recent studies have demonstrated that ASHP systems can achieve significant energy savings and reduce greenhouse gas emissions compared to traditional heating methods [9,10]. Moreover, the integration of ASHPs with renewable energy sources, such as solar power, has shown potential for further enhancing energy efficiency [11].

In addition to system performance, control strategies play a critical role in maximizing the efficiency of heat pump systems. Advanced control methods have been developed to improve the robustness and efficiency of both GSHP and ASHP systems. For instance, a multi-objective design optimization strategy has been proposed for vertical U-tube ground heat exchangers to minimize upfront costs and entropy generation simultaneously [12]. Experimental studies on various control strategies have demonstrated significant improvements in system performance under real-world conditions [13,14].

Heat pump technologies play a crucial role in modern energy systems, offering not only efficient heating and cooling solutions but also enhancing grid stability and flexibility [15]. For instance, advanced CO₂ gas cooler technology has shown significant efficiency improvements in harsh winter conditions [16], while solar-assisted heat pumps provide a sustainable option for remote and cold regions [17]. Each technology, however, comes with specific application scenarios and economic considerations. This study focuses on the air–water heat pump system for residential applications, aiming to support large-scale retrofitting and improve economic feasibility in specific regions.

The focus of this study is on the performance analysis and optimization of an air–water heat pump system designed for centralized floor heating in residential buildings located in a mid-latitude region of the Northern Hemisphere with a subtropical monsoon climate. The system's energy efficiency, heating capacity, and environmental adaptability are evaluated, with particular attention to the control strategies that govern its operation, including the defrosting process [18].

The schematic diagram of the proposed air–water heat pump system is shown in Fig. 1. The chiller system is composed of four independent subsystem units in parallel. It could be used for both central air conditioning and heat pump systems. The control of this large-scale unit is jointly determined by the expansion valve, fan speed, and the number of operating compressors. Since the heating demand for the four villas is continuous and stable, increasing the regulation proportion of the expansion valve and fan speed allows the use of fixed-frequency compressors, which are more cost-effective. The company has already developed a well-refined control strategy and detailed control standards to ensure efficient operation. The vapor subsystem unit mainly consists of a compressor, a four-way valve, a shell-and-tube heat exchanger (STHX), and an electronic swell valve. When the heating model is conducted in winter, hot water flowing through the STHX is produced for heating the floors. The refrigerant (R410A) in the chiller system is compressed from the compressor to the STHX (condenser) for heating the water from the beginning. Then, the throttling process is conducted in the electronic swell valve. Finally, the fluid comes to the compressor through the fin-tube evaporator that absorbs the heat from the air. In the water cycle, the water supplied from the valve absorbs the heat of the fluid in the shell tube heat exchanger. The auxiliary electric heater works if the hot water temperature is not high enough. When the water temperature falls below 45 °C, the auxiliary electric heater activates. After heating in the water tank, the hot water comes to the pipes in the floor panel heating system using the circulating pump. The floor panel heating systems heat the air in the villas separately in the ways of thermal radiation and thermal convection. Finally, the water goes back to the shell tube heat exchanger.

The hot gas defrosting model is conducted using the four-way valve to switch the direction of the fluid flows. The refrigerant in the outdoor fin-tube heat exchanger condenses in the defrosting system for melting frost. Owing to the high heat storage capacity of the floor panel heating systems and the short time of defrosting, the thermal fluctuation in the heat pump system would not lead to an obvious change in the room temperature.

Each floor panel heating system operates independently, ensuring uniform heating for each villa. The area and structure of each villa are the same. Each villa is within easy reach of each other, and two of them are in one unit sharing one big garden. They are 280 m² and stand four storeys high respectively. The chiller is placed in the middle of those four villas.

For the vapor compression system, the external power is supplied by the compressor for the major components: evaporator, condenser, and expansion valve. The heat is absorbed from the environment in the evaporating process, while in the condenser the heat exchange occurs from the vapor compression system to the water source in the shell tube heat exchanger. The heat transfer between the chiller and the environment shown as Q_eva in Fig. 2 is conducted at a finite temperature difference, which is a major source of energy losses for the cycle.

To enhance EER, the following control strategies are proposed:

1. We need to optimize evaporating and condensing temperatures. To increase the evaporating temperature (at point 1) is to reduce the temperature difference between the evaporating temperature and the heat source temperature (ΔT_evap). To decrease the condensing temperature (at point 3) is to minimize the temperature difference between the condensing temperature and the cooling medium (air or water) temperature ΔT_cond. These adjustments will expand Q_cond (heat release) and simultaneously reduce W_comp.

2. We need to reduce compressor temperature difference loss. Minimizing the temperature difference between the suction and discharge stages (ΔT_suc-exh) is to improve the compressor's efficiency $(ηcomp)$⁠. The relationship can be expressed as Eq. (2)

   $EER∝ΔTevap-condΔTsuc-exh$

   (2)

3. ΔT_evap-cond is the temperature difference between the evaporating and condensing processes, and ΔT_suc-exh is the temperature difference between the suction and discharge processes in the compressor.

4. We need to optimize sub-cooling and superheating in the cycle. To increase sub-cooling (temperature reduction from point 3 to point 4) is to ensure more refrigerant enters the evaporator as a liquid thereby enhancing the evaporator's heat exchange efficiency. To control proper superheating (temperature change from point 5 to point 1) is to prevent excessive vapor entering the compressor, reducing unnecessary energy consumption.

According to Fig. 3, the control system of the air–water heat pump adjusts control parameters in real-time by calculating the sub-cooling temperature difference (ΔT_subc) and the EER_i using measurements of evaporating temperature (T_evap), ambient temperature (T_ambient), suction pressure (P_suc), and exhaust pressure (P_exh). If the CEER, ΔT_subc, and EER_i do not meet the required standards, the system adjusts compressor and fan speeds accordingly.

The t–dT defrosting method, based on defrosting time and the temperature difference between evaporating and ambient temperatures, is employed to enhance the system's adaptation to environmental changes. This method has been proven effective in maintaining stable indoor temperatures during defrosting. The defrosting process is triggered when the temperature difference (ΔT_evap-cond) exceeds 5 °C and persists for 30 min, or when the EER_daily is below the threshold (2.38) for defrosting. When EER_daily falls below a predefined threshold (derived from experimental observations), it suggests that frost accumulation has significantly reduced system efficiency. This threshold acts as a trigger for initiating the defrosting process.

The defrosting process consists of the following steps: the system first enters defrosting mode to initiate the operation. EER_daily is then monitored and calculated in real-time as an evaluation metric for system performance. During the defrosting process, the system adjusts the compressors and the four-way valve to switch the operating mode and complete defrosting. Next, the system evaluates whether the EER_daily falls below the target value: if it does, further adjustments to the compressors and other operating parameters are made; if the EER_daily is within the acceptable range, the current defrosting operation is maintained to ensure stable system performance. Each day has a calculated superheat degree centered around 8 °C with small fluctuations of ±1 °C.

According to Fig. 3, ΔT_sub and EER_i are monitored in real-time to adjust the number of operating compressors, the expansion valve opening, and fan speeds. Dynamic parameter adjustments are made based on the real-time temperature and pressure feedback:

• Ambient temperature (>5 °C): operates with less compressors and low fan speed to save energy.

• Ambient temperature (0–5 °C): increases the expansion valve opening to enhance heating efficiency.

• Ambient temperature (−7 °C to 0 °C): activates partial auxiliary electric heaters to boost heating capacity.

• Ambient temperature (≤ −7 °C): activates auxiliary electric heaters and strengthens defrost control.

In summary, the system dynamically adjusts the compressor operation, expansion valve opening, and fan speeds based on variations in temperature, pressure, and EER. Long-term operational data are recorded and analyzed to dynamically optimize control parameters, improving the system's long-term efficiency and stability.

The t–dT temperature control strategy [19,20] is based on two key parameters—time and temperature difference. When the temperature difference between the evaporator surface and the ambient environment becomes increasingly larger, it indicates a decline in the evaporator's heat exchange efficiency and the accumulation of thicker frost layers on its surface. To address various control scenarios, we have incorporated CEER as a comprehensive evaluation criterion. The CEER is a critical reference parameter for the control strategy of this system. It is defined as the ratio of the total heat absorbed by the evaporator and released by the floor under defrost shutdown conditions to the energy consumed during operation. As frost accumulation increases, system efficiency decreases, resulting in a gradual reduction in the combined heat release from the floor and heat absorption by the evaporator. Consequently, Q_i decreases progressively.

The system was subjected to a 93-day field test at a villa in Hefei, Anhui Province, China, covering the period from Nov. 11, 2022, to Feb. 20, 2023. During this period, the system operated under real-world winter conditions, providing heating for four villas. Table 1 presents the rated working conditions for the chiller under both cooling and heating modes. The rated working condition is shown in Table 2. The system performance was monitored using a multi-sensor data acquisition system, which included thermocouples for measuring temperatures listed in Table 1, pressure transducers for capturing suction and exhaust pressures (P_suc and P_exh), and power meters for tracking energy consumption. A data logger with a sampling frequency of 1 min was employed to ensure high-resolution performance analysis.

Based on the total building area of four villas (280 m² × 4) and the local climate conditions in Hefei, the required heating demand under different ambient temperatures was estimated using historical heating data, and full-load performance tests were conducted, as shown in Table 3. The results demonstrate that the system operating at full capacity can meet the heating demands of the four villas under various ambient temperatures.

Combining the data in Table 1, the chiller provides a rated heating capacity of 140 kW, which exceeds the maximum required heating demand of 141.36 kW for the four villas. This ensures sufficient capacity to maintain indoor temperatures at 20 °C under varying outdoor conditions, including extreme cold. Additionally, the EER of 3.5 in heating mode reflects high system efficiency, reducing energy consumption and making it particularly advantageous for long-term residential heating applications with lower operational costs.

The heating mode specifies an evaporating temperature of 0 °C and a condensing temperature of 60 °C, aligning with the requirements for efficient operation in colder climates (down to −15 °C ambient temperature). Furthermore, the rated water flowrate of 0.172 m³/(h/kW) ensures consistent heat transfer performance, making the system well-suited for residential floor heating applications requiring stable water temperatures.

Throughout the field test, key performance metrics, including suction and exhaust pressures, condensing and evaporating temperatures, heating capacity, and indoor temperatures, were monitored and recorded. The control strategy's effectiveness, particularly in defrosting cycles, was evaluated based on these data.

The ambient temperature was measured in real-time using detection equipment installed around the villas in the suburban area of Hefei. Additionally, the daily average values of humidity and wind speed obtained from the Open-Meteo website were plotted together in Fig. 4. By comparing these data with the system's EER, we can draw the following conclusions:

The real-time temperature, humidity, and wind speed data significantly impact the performance of the heat pump system. Lower temperatures reduce the efficiency of the refrigerant cycle by narrowing the temperature gradient (ΔT_evap-cond) and increase the likelihood of frost formation. High humidity accelerates frost accumulation on the outdoor coil, leading to frequent defrost cycles that lower overall efficiency. Wind speed influences heat transfer rates; moderate wind enhances performance, while high wind may cause uneven frost distribution. These environmental factors dynamically affect the system's energy efficiency and operational stability, highlighting the need for adaptive control strategies that integrate real-time weather data to optimize defrosting and improve system performance.

The system's performance during the field test is summarized in Fig. 5, which combines the results into a single figure with six subparts as follows:

• Figure 5(a) shows the variations in daily suction and exhaust pressures, which were found to be directly influenced by ambient temperatures. The compressor operated within a safe compression ratio (π) range of 2.67–5.79, ensuring stable performance.

• Figure 5(b) displays the daily condensing and evaporating temperatures. These temperatures were adjusted in response to changes in ambient conditions, helping to maintain the desired heating capacity.

• Figure 5(c) illustrates the heating capacity (Q_H) of the system, which varied between 10.10 kW and 33.50 kW. Since the heat pump compressors in this study do not operate with variable frequency but are controlled by adjusting the number of active compressors, the maximum instantaneous heating output under full load for a single compressor is 45.43 kW, and for four compressors operating at full load, it is 127.52 kW. The Q_H was negatively correlated with ambient temperature, indicating the system's adaptability to external temperature fluctuations.

• Figure 5(d) presents the EER_daily, which ranged from 2.32 to 4.71, with an average of 3.46 over the test period. This demonstrates the system's high efficiency in varying winter conditions.

• Figure 5(e) shows the inlet and outlet water temperatures, which fluctuated based on the heating demand. The water temperature measurement point of T_w,in,daily is located at the inlet of the water tank, and the measurement point of T_w,out,daily is located at the outlet of the water tank. The system's ability to adapt to changing heating requirements is reflected in these temperature variations.

• Figure 5(f) depicts the indoor temperatures maintained by the system. T_0.1 represents the temperature measured at a height of 0.1 m above the ground, T₁ represents the temperature measured at a height of 1 m above the ground, T_1.5 represents the temperature measured at a height of 1.5 m above the ground, and T₂ represents the temperature measured at a height of 2 m above the ground. Despite outdoor temperatures ranging from −10 °C to 14 °C, the indoor temperatures remained stable at around 19 °C, demonstrating the system's effectiveness in maintaining thermal comfort.

According to Fig. 6, it can be observed that the heat exchange rates on both the hottest and coldest days show a negative correlation with real-time temperature. The correlation is more evident on the hottest day since there was no defrosting interference. In contrast, the correlation on the coldest day is less obvious due to defrosting operations that occurred at the 5th, 14th, and 20th hours, which disrupted the correlation pattern.

The field test results confirm that the air–water heat pump system is highly effective in maintaining stable indoor temperatures and achieving a high energy efficiency ratio under real-world winter conditions. The optimized control strategy, particularly the defrosting process, minimized energy losses and ensured consistent heating capacity throughout the test period.

In Tables 4 and 5, the key experimental data under extreme weather conditions and the performance parameters compared with other air-source heat pumps both demonstrate that, while ensuring the required heating output, the effective control of the air-water heat pump system has significantly improved its EER. Compared to systems by Heo et al. and Yu et al., the proposed system excels in adaptability, stability, and broad applications. With a heating capacity of 70.21–121.36 kW, it's ideal for large-scale residential communities. A 93-day field test in −10 °C to 14 °C showed steady indoor temperatures at 19 °C and EER_daily of 2.32–4.71. The t–dT defrosting approach enhanced frost management, ensuring long-term efficiency and reliability in challenging climates.

In this paper, based on the experimental data of the air–water heat pump system and the analysis of the tests in the winter season, the following main conclusions can be drawn:

1. The control strategy of the heat pump system consists of regulating the fan speed of the finned tube heat exchanger and on/off control of the compressor. This operation is effective in maintaining stable P_suc and P_exh. It ensures that the system operates within safe parameters, prevents overloading of the compressor, and maintains Q_H in the range of 0.4–0.8 MPa for P_suc and 2.1–2.6 MPa for P_exh. Robust control strategy of the operating process to adapt to different ambient temperatures.

2. The system is able to maintain stable T_cond and T_evap despite fluctuations in ambient temperature. Comparison of indoor and ambient temperatures shows its adaptability and stability to the environment. The Q_H,daily of the system ranged from 10.10 kW to 33.50 kW, and the EER_daily ranged from 2.32 to 4.71, showing a high level of efficiency. These results confirm that an excellent control strategy ensures that indoor temperatures are consistently maintained at around 19 °C, even when outdoor temperatures vary widely.

3. The condensing temperature maintains a relatively stable temperature difference from the ambient temperature, ensuring stable heat absorption during evaporation. When the ambient temperature is above zero, the condensing temperature remains relatively stable, resulting in consistent heat exchange and stable water supply temperatures. However, when the ambient temperature drops below zero, defrosting issues arise, and the indoor heating demand increases. As a result, the condensing temperature rises significantly, leading to an increase in heat exchange and inlet/outlet water temperatures.

4. The system has good environmental adaptability. Q_H and EER fluctuate more on the hottest days of winter compared to the coldest days of winter shown in Table 4, but are still within acceptable limits. This adaptability is attributed to effective control of water flow, fan speed, compressor ON/OFF, and expansion valve which accommodates differences in heat leakage caused by changes in ambient temperature. The integrated judgment of defrosting timing in the control strategy saves energy and increases the Q_H,daily output.

5. In Fig. 6, the system demonstrates excellent heating capacity stability, maintaining output between 31 kW and 33.5 kW, even under fluctuating and extremely cold temperatures. However, the EER is highly sensitive to ambient temperature, with a noticeable efficiency drop during the coldest periods. This behavior highlights the importance of designing the system with optimized defrosting strategies and high-efficiency compressors to mitigate efficiency losses in extreme conditions. The overall performance indicates that the system is suitable for cold climates, ensuring stable heating while maintaining acceptable efficiency during peak loads.

6. In regions with hot summers and cold winters, heating methods include gas boilers, coal-fired systems, and heat pump air conditioning. In the area studied in this paper, the prevalent heating methods are gas boiler heating and centralized coal-fired heating. For heating terminals, both radiator systems and floor-embedded piping systems are commonly used. This study compares air-source heat pump (ASHP) floor heating with gas boiler radiator heating and centralized coal-fired boiler floor heating. The economic comparison of the three heating methods is shown in Table 6. This study shows ASHP floor heating, while having a higher initial cost, achieves 38.25% lower annual costs than gas boilers and 14.44% lower than coal-fired systems. ASHP is eco-friendly and government-supported, making it the best economic and sustainable choice.

Overall, the control strategy of the air–water heat pump system proves to be highly effective in managing key operational parameters, keeping stable and efficient heating performance. These findings highlight the system's potential for providing reliable and efficient heating solutions in residential applications.

The work of this paper is financially supported by Yangzi Air Conditioner Co. Ltd., in Anhui Chuzhou and the Independent Innovation Specialized Fund Project of Anhui Province (Grant No. 2012XKJH0003).

There are no conflicts of interest.

The datasets generated and supporting the findings of this article are obtainable from the corresponding author upon reasonable request.

h =

specific enthalpy, J/kg

N =

cumulative number

P =

pressure, Pa

Q =

heat exchange, W

T =

temperature, °C

W =

power consumption, W

CEER =

comprehensive energy efficiency ratio

EER =

energy efficiency ratio

ΔT =

temperature difference, °C

  ΔP =

pressure difference, Pa

η =

efficiency, %

π =

compression ratio

p =

pressurization

r =

refrigerant

w =

water

i =

ith component

j =

jth component

H =

heating

R =

heat pump

ambient =

ambient air

cond =

condensing/condenser

daily =

daily average value

env =

environment

evap =

evaporating/evaporator

exh =

exhaustion

in =

inlet

out =

outlet

suc =

suction