Application of a Phase-Change Material Heat Exchanger to Improve the Efficiency of Heat Pumps at Partial Loads

Abstract

Inverter-equipped heat pumps allow for increased energy efficiency. However, air conditioning (AC) systems often operate at low load ratios below where inverter control is effective, which reduces their energy efficiency. We developed an AC system that increases the apparent load ratio of the heat pump by using a phase-change material (PCM). Cooling and heating experiments were conducted with a PCM heat exchanger, which comprised aluminum plates and fins filled with paraffinic PCM. The result indicated a high heat transfer coefficient of >70 W/(m²·K). A simplified numerical model of the PCM heat exchanger as a lumped constant system was created based on the experiment. The calculations generally reproduced the experimental results, with root mean squared errors of 0.39 K for cooling and 0.84 K for heating, confirming their accuracy. Simulations were then conducted to evaluate the energy performance of the proposed system for the cooling season. While low load operation accounted for 39% of the total AC time for a non-PCM system, it was reduced to 2.7% for the proposed system. The proposed system demonstrated load ratios of 50–60% for most of the season, achieving an energy reduction of 11.4% owing to the improved efficiency at partial load ratios.

1. Introduction

While the promotion of renewable energy sources is important for achieving carbon neutrality, an equally necessary step is improving energy efficiency on the demand side through the use of heat pumps [1]. In Japan, air conditioning (AC) systems are commonly based on heat pumps equipped with inverters, which offer variable speed control so that the heat pump can operate efficiently not only at large load ratios close to the compressor capacity but also at partial load ratios [2]. However, inverter control becomes invalid at small load ratios below ~20% during which the compressor is simply turned on or off as necessary, which is inefficient [3]. During actual operation, heat pumps tend to be at low load ratios because AC systems are generally designed with excess capacity so that they can sufficiently cool and heat buildings during even the hottest and coldest days throughout the year. We performed field measurements showing that the energy efficiency of heat pumps in actual residential buildings decreases at low load ratios [4,5]. Therefore, it is necessary to improve the energy efficiency of heat pumps at low load ratios below the lower limit of inverter control to further reduce the energy consumption and CO₂ emissions of AC systems.

Thermal energy storage (TES) is a technology that can be used to adjust the temporal gap between energy supply and demand, which tend to peak at different times [6]. Unlike general sensible heat storage, in which water or concrete is utilized as the storage medium, phase-change materials (PCMs) can charge or discharge large amounts of heat at a constant temperature when they change phase [6]. Various studies have explored applying PCMs to buildings [7,8,9,10,11], such as mixing PCMs with flooring, wall, and roof materials, to extend the thermal capacity of buildings and stabilize the indoor thermal environment [12,13,14] or encapsulating PCMs in radiation panels to function as part of a thermally activated building structure [15,16]. Many studies have utilized PCMs to store natural thermal energy, such as solar heat [17,18] or cool air at night [19,20,21], to reduce the load and energy consumption of AC systems. Several studies have investigated precooling or prewarming PCMs at night, during which the energy supply tends to exceed demand, as a load leveling or peak shaving strategy [22,23]. Some researchers have proposed applications for PCMs in which heat pumps operate during energy-efficient hours, such as in the early morning in warm regions [24] and in the daytime in cold regions [25]. However, few studies have considered applying PCMs to improve the energy efficiency of inverter-equipped heat pumps at low load ratios.

The practical application of PCMs requires an increase in the heat transfer coefficient of the PCM itself or a heat exchanger. Several studies have reported that adding nanoparticles increases the apparent heat conductivity of a PCM [26]. Several researchers have investigated the heat exchange characteristics of functional PCMs enclosed in shell-and-tube type heat exchangers [27,28,29], triple tube heat exchangers [30,31], and porous foam materials forming wavy air channels [32]. While shell-and-tube type heat exchangers are commonly used for PCM applications where heat is exchanged with water [33,34,35], different heat exchangers have been investigated for buildings where heat is exchanged with air [10,36,37], such as the tube type [38,39], ball type [40], plate type [41], and finned plate type [42,43]. We also investigated using a packed bed of granulated PCM as a heat exchanger [44,45]. Although the granulated PCM had a large surface area that resulted in a large heat transfer coefficient with air, problems remained in terms of sealing and pressure drop for use in actual AC systems. Thus, research has been insufficient on PCM applications where heat is exchanged with air that achieve both a high thermal performance and a low pressure drop.

In this study, we utilized the TES concept through a high-performance PCM heat exchanger to operate a heat pump at a high apparent load ratio and avoid low-efficiency intermittent operation. To evaluate the effectiveness of our design, we fabricated a prototype heat exchanger containing about 1 kg of PCM and conducted experiments to observe the heat exchange characteristics and pressure loss at different air flow rates. Numerical simulations were carried out to evaluate the energy efficiency of the proposed AC system in different seasons and at partial load ratios compared to a conventional AC system without TES.

2. Design

Figure 1 shows conceptual diagrams of the proposed AC system in charge and discharge modes. A PCM heat exchanger is connected to the indoor unit, which is installed in a ceiling space with air ducts. For cooling operation during the summer, charge mode involves storing cold energy in the PCM, while discharge mode involves extracting the cold energy from the PCM. In charge mode, air taken from the room is first cooled in the indoor unit, which acts as an evaporator in the heat pump cycle. The air then passes through the PCM heat exchanger, which cools the PCM inside and is returned to the room as conditioned air. The air flow rate of the fan in the indoor unit is varied to remove the cooling load in the room and to cool the PCM within the rated value. Because the room and PCM are cooled simultaneously, the load ratio of the outdoor unit is increased, which increases the energy efficiency ratio (EER). In discharge mode, the fan in the indoor unit contributes to ventilation rather than the heat pump cycle. The air taken from the room is cooled by the PCM heat exchanger and circulated back to the room to remove the cooling load. The charge and discharge modes switch according to the outlet temperature (T_out) under the assumption that the PCM completes the phase change.

3. Experimental Setup

Figure 2 and Figure 3 show the dimensions and schematic diagrams, respectively, of the PCM heat exchanger prototype. The prototype was made of aluminum with a weight of 8.9 kg and external dimensions of 235 × 300 × 374 mm. Air passes through the heat exchange element filled with PCM (shaded in orange). The heat exchange element consisted of aluminum plates connected by aluminum fins. The 1 mm-thick aluminum plates had a hollow structure, with additional mesh-like fins arranged inside. The total heat exchange area was 4 m². Air exchanges heat with the heat exchange element by passing between the fins. PCM temperature (T_p) was measured at the junction of the PCM filling sections, as shown in Figure 3b, because of the limited space for inserting a thermocouple.

Table 1. Physical properties of the PCM.

Physical Properties	Value
Latent heat amount	188 kJ/kg
Melting point	17.8 °C
Freezing point	17.0 °C

presents the physical properties of the PCM used to fill the heat exchange element. The PCM mainly consisted of paraffin wax. Differential scanning calorimetry of a 1.7 mg sample determined the melting and solidification temperatures as 17.8 and 17 °C, respectively. The amount of latent heat was determined as 188 kJ/kg. The PCM was enclosed in the heat exchange element in liquid form with a filling weight of 0.8 kg.

Figure 4 shows a schematic diagram of the experimental system, including an exterior view. Air ducts were connected in front of and behind the PCM heat exchanger with plastic guides installed in the front duct to regulate the air. A chilled/warmed water coil and a fan were installed at the inlet and outlet of the air duct. The water in the coil was controlled to a specific temperature by a constant-temperature chiller. The PCM heat exchanger was covered with thermal insulation during the experiment. Figure 5 shows the measurement points, while

Table 2. Measurement devices and sensors.

Measurement Item	Device	Sensors	Range	Accuracy
Air temperature	HIOKI data logger *1	Thermocouples [Type T, Class1]	~100 °C	±0.5 °C
PCM temperature	HIOKI data logger *1	Thermocouples sheath type	−200 to +300 °C	±0.5 °C
Air velocity	Testo480 multifunction measuring instrument *2	Testo Hot-wire anemometer	0 to 20 m/s	±0.03 m/s
Pressure loss	OMRON differential pressure station *3		−500 to 500 Pa	±0.2 Pa

Manufacturer information: *1 HIOKI E.E. CORPORATION (Nagano, Japan), *2 Testo SE & Co. KGaA (Titisee-Neustadt, Germany), *3 OMRON Corporation (Kyoto, Japan).

presents the specifications of the measurement devices and sensors. The temperature, differential pressure, and wind speed were measured every 30 s. The inlet air temperature (T_in) and outlet air temperature (T_out) were measured by thermocouples placed at nine points near the front and back surfaces, respectively, of the fins. The differential pressure was measured before and after the heat exchange element. The air velocity was measured by a one-dimensional hot-wire anemometer placed at the center of the duct cross-section.

Experiments were performed to evaluate first the cooling process and then the heating process. For the cooling process, the PCM was solidified, and the T_in was controlled at 10 °C by using chilled water, which was set at 5 °C in the coil. The cooling process was completed once the PCM temperature (T_p) dropped below 12 °C. For the heating process, T_in was increased to 30 °C by the chiller under the assumption that cold energy was extracted from the PCM in ventilation mode without the aid of the heat pump. The heating process continued until T_p > 20 °C, at which point the PCM was completely melted.

Table 3. Experimental cases.

Properties	Case 1	Case 2	Case 3
Air velocity [m/s]	0.5	1.0	1.5
Air flow rate [m3/h]	32	64	95

presents the experimental cases 1–3. The air velocity was set to 0.5, 1.0, and 1.5 m/s, respectively, by adjusting the fan frequency. The air flow rate was calculated as the product of the measured air velocity and cross-sectional area and was 32, 64, and 95 m³/h, respectively. Each case was tested at least twice between May and July 2024.

4. Experimental Results

The amount of heat exchanged by the PCM heat exchanger (Q) was calculated as follows:

$$Q=c_a{\rho }_{a}V(T_{\mathit{in}}-T_{\mathit{out}}),$$

(1)

where c_a is the specific heat of air [kJ/(kg·K)], ρ_a is the density of air [kg/m³], and V is the air flow rate [m³/s]. V, T_in, and T_out were measured in the experiments. The overall heat transfer coefficient (U) was calculated as follows:

$$U=Q/(A_{ex} \times ∆T),$$

(2)

where A_ex is the heat exchange area [m²], and ∆T is the difference between T_p and T_out [K].

Figure 6 shows the spatial distributions of T_in and T_out measured at nine points on each side in case 3. During the cooling process, T_in had a range of 11.9–13.9 °C and was up to 2 °C lower at the center than at the edges. This can probably be attributed to the air velocity distribution in the duct cross-section. Although T_out was relatively low in column 3, it was generally maintained around 17 °C, which corresponds to the phase change temperature, and had a smaller range than T_in. During the heating process, T_in varied by only 0.8 °C, which was less than the variation during the cooling process. This suggests that supplied air with a low density helped equalize the velocity entering the heat exchange element. T_out again stayed near the phase change temperature during the heating process. These results indicate that the air flow entering the PCM heat exchanger was regulated by the fins, which facilitated a uniform heat exchange and led to adequate heat extraction at the outlet side. Below, the mean of the values obtained at the nine measurement points is used to represent T_in and T_out.

Figure 7 shows the relationship between the face air velocity and pressure drop in the heat exchange element. The face air velocity was calculated by dividing V by 0.09 m², which is the surface area of the heat exchange element. At a face air velocity of 0.1–0.3 m/s, the pressure drop was only 2–9 Pa and was generally below 10 Pa. This pressure drop is much smaller than that of typical fin-tube heat exchangers, which require a face air velocity of approximately 3.0 m/s to achieve U values of 55–60 W/(m²·K) [46].

Figure 8 shows the temperature, Q, and U profiles of case 3 during the cooling and heating processes. During the cooling process, T_p dropped to 18 °C after 10 min and remained constant for 17 min during which the PCM solidified. T_out dropped earlier than T_p but stayed constant with T_p during the phase change period. This stable behavior for T_out resulted in Q remaining constant at 140 W. U increased with smaller ∆T and reached a maximum of 61 W/(m²·K) during the phase change period. During the heating process, T_p increased to 17 °C after 3 min and remained stable for 7 min, during which the PCM melted. Although an increase in T_out was suppressed during the phase change period, the difference with T_p was greater than during the cooling process. Q was 220–250 W during the phase change period, which was larger than in the cooling process and can be attributed to the larger difference between T_in and T_out in the heating process. On average, U was approximately 28 W/(m²·K) during the phase change period. After the phase change, both T_p and T_out rose to approach T_in. However, the rise in T_p was suppressed to around 25 °C for 12 min after 20–25 min. This behavior was probably due to the delayed melting of the PCM deep inside the heat exchanger. This suggests that T_p was affected by the heat capacity of the surrounding PCM. In addition, the fin structure in the PCM heat exchanger may make natural convection difficult to occur. This meant that the low heat conductivity of the liquid phase had a relatively large impact. Therefore, PCM in the junction may have partially delayed or not contributed to the heat exchange. Thus, we decided to determine the effective weight that contributed to the heat exchange, which we then used to develop the model for our later numerical simulations.

To predict T_out after the heat exchange, we formulated empirical formulas for calculating U based on the experimental results. Figure 9 shows the relationships between T_p and U obtained in the cooling and heating processes. The values observed every 30 s are plotted on the figure. Large values of U were observed at temperatures of 17 and 18 °C, which correspond to the phase change temperature of the PCM. During the cooling process, U was 10–20 W/(m²·K) above the phase change temperature and increased sharply when T_p was below 18 °C. U became small again below the phase change temperature with an average value of 4.1 W/(m²·K). During the heating process, U was around 20 W/(m²·K) below the phase change temperature and increased drastically when T_p exceeded 17 °C. While U decreased to 10 W/(m²·K) after the phase change, it rose again in several cases, which may be due to the melting of the PCM deep inside the heat exchange element. The formulation procedure was as follows.

(1)

The maximum U and corresponding T_p for each case were averaged to obtain U = 69 W/(m²·K) at T_p = 17.60 °C during the cooling process and U = 77 W/(m²·K) at T_p = 17.55 °C during the heating process.

(2)

During the phase change, the relationship between U and T_p can be approximated as a power law that asymptotically approaches 17.57 °C during the cooling process and 17.60 °C during the heating process, where a bi-square weighting function is applied to prevent the influence of outliers.

$$U=16.17{(T_p-17.57)}^{-0.31}.$$

(3)

$$U=24.80{\left({17.60-T}_p\right)}^{-0.24}.$$

(4)

When compared with the experimental results, both Equations (3) and (4) showed a high explanatory power with R² > 0.7.

(3)

After the phase change, U decreased linearly from the maximum U value with a temperature difference of 0.2 K to constant values of 4.1 W/(m²·K) during the cooling process and 9.5 W/(m²·K) during the heating process, on average.

5. Numerical Model

5.1. Setup

A numerical model was developed to predict T_out from the PCM heat exchanger. The temperature distribution in the PCM heat exchanger was not considered under the assumption that the air flow is uniform across the cross-section. The heat balance in the PCM heat exchanger can then be represented as a lumped constant system where the internal PCM and exterior aluminum are a single unit:

$$(W_{\mathit{ale}}+W_{\mathit{pe}}){\displaystyle \frac{{\mathit{dH}}_{p+\mathit{al}}}{\mathit{dt}}}=U{A}_{\mathit{ex}}(T_{\mathit{out}}-T_p),$$

(5)

where W_pe and W_ale are the effective weights of the PCM and aluminum [kg], respectively, contributing to the heat exchange. The heat capacity of the PCM heat exchanger can be expressed as the change in enthalpy per the total effective weight of PCM and aluminum (H_p+al) [kJ/kg]. H_p+al can be converted to T_p by the enthalpy method. The amount of heat exchanged with air can then be calculated by using U as determined from the experimental results. The heat balance on the air side can be represented by the heat from advection and heat exchanged by the PCM heat exchanger:

$$c_a{\rho }_{a}V(T_{\mathit{out}}-T_{\mathit{in}})=U{A}_{\mathit{ex}}(T_p-T_{\mathit{out}})+Q_{\mathit{loss}},$$

(6)

where the heat storage of air can be neglected as its heat capacity was sufficiently smaller than that of the PCM heat exchanger. Equation (6) can then be used to determine T_out. Q_loss is the heat loss and is calculated as follows:

$$Q_{\mathit{loss}}= K_{loss}{A}_{sur}\left(\left(T_{\mathit{in}}+T_{out}\right)/2-T_{amb}\right),$$

(7)

where K_loss is the heat loss coefficient identified as 2.0 W/(m²·K) from experimental validation, A_sur is the surface area of the PCM heat exchanger (=4.0) [m²], and T_amb is the ambient air temperature [°C].

Figure 10 shows the relationship between temperature and enthalpy for the combined mass of PCM and aluminum as given by the enthalpy method. The latent heat due to the phase change of the PCM can be obtained from the linear relationship at 17–18 °C. The enthalpy change outside the phase change temperature can be calculated by using the mean specific heat of the PCM heat exchanger:

$$c_{p+\mathit{al}}=(W_{\mathit{pe}} \times c_p+W_{\mathit{ale}} \times c_{\mathit{al}})/(W_{\mathit{pe}}+W_{\mathit{ale}})$$

(8)

where c_p+al, c_p, and c_al are the specific heats of the PCM heat exchanger, PCM, and aluminum [kJ/(kg·K)].

As shown in Figure 8, the PCM in the junction may have been partially delayed or not contributed to the heat exchange. To represent the phenomenon in the model, W_pe and W_ale were determined according to the following procedure. To identify W_ale, experiments were performed using the same PCM heat exchanger prototype as before, but without PCM. Then, W_ale was calculated by setting W_pe = 0 in Equations (5) and (8) and varying the weight of aluminum in 0.1 kg increments to reproduce the experimentally obtained T_out with the smallest error. Figure 11 shows the root mean squared error (RMSE) between the experimental and calculated T_out at different aluminum weights. RMSE was minimized at 5.83 kg, which corresponded to 72% of the total aluminum weight of the prototype (8.1 kg). Thus, W_ale was set to 5.83 kg for the following simulations. Similarly, W_pe was determined as the weight of the PCM at which the experimental T_out was reproduced with the smallest error. As shown in Figure 12, the smallest RMSE between the experimental and calculated T_out was obtained at a PCM weight of 0.53 kg, which was equivalent to 66% of the total weight of the prototype. Thus, W_pe was set to 0.53 kg.

5.2. Verification

The accuracy of the numerical model was verified under the same boundary conditions as in the experiment, using W_ale and W_pe, as determined previously. Figure 13 shows the relationship between the experimental and calculated T_out. The calculated T_out was sometimes higher than the experimental T_out during the cooling and lower during the heating process. The calculated T_out generally remained constant while the experimental T_out varied under the influence of T_in. Deviations were observed in cases 1 and 2, which had relatively small air volumes. The experimental results suggested that the phase change occurred in a distributed manner in the prototype, but the numerical model assumed that the heat exchanger was a lumped constant system. This had the clear effect of regulating the temperature during the phase change in the calculated results compared with the experimental observations. In cases with a smaller air volume, the phase change progressed more slowly, which may have facilitated the dispersion of heat extracted from the PCM to the air during the experiment. Overall, the average RMSE for all cases was 0.39 K for the cooling process and 0.84 K for the heating process, indicating relatively good reproducibility. Thus, the numerical model showed reasonable accuracy for evaluating seasonal or annual system behavior despite its simplicity. Further validation may be needed with modified heat exchangers that reflect the ideal heat exchange behavior of a lumped constant system.

6. Simulations

6.1. Conditions

Energy simulations were performed by applying the developed numerical model to an office space with windows on the south side, as shown in Figure 14. An AC system was constructed by using the Life Cycle Energy Management (LCEM) tool ver. 3.10 distributed by MLIT [47], as shown in Figure 15. The system was assumed to be a variable refrigerant flow (VRF) system comprising eight indoor units and one outdoor unit, where a PCM heat exchanger was connected to each indoor unit. The energy consumption was calculated at 5 min intervals by using cooling loads from May 1 to October 31 obtained from the Building Energy Simulation Tool (BEST) Program [48] as the input. The weather data in Tokyo, standard values of heat gains through structure and windows, and internal loads were used to calculate the cooling load.

Table 4. Specifications of the AC system.

(a) Outdoor Unit		(b) Indoor Units
Number of outdoor units	1	Number of indoor units	8
Rated cooling capacity [kW]	65	Rated cooling capacity [kW]	7.3
Rated power consumption [kW]	17.8	Rated power consumption [kW]	0.3
Rated COP	3.64	Rated air volume [m3/h]	1170

presents the specifications of the AC system, which were determined so that the outdoor unit operated at an average load ratio of 30% in July. The eight indoor units were assumed to exhibit the same behavior, and the set of one indoor unit and PCM heat exchanger shared one-eighth of the cooling load. The inlet air temperature (T_in) after the indoor unit and air flow rate (V) as determined by the cooling load were used as inputs to calculate the outlet air temperature (T_out) after the PCM heat exchanger. The charge and discharge modes were switched based on T_out, as shown in Figure 1. In charge mode, the outdoor unit was operated to cool the room space and the PCM at the same time. When T_out fell below 15 °C, the system switched to discharge mode, in which the room space was cooled by the charged PCM, and the outdoor unit did not operate. When T_out rose above 21 °C, the system switched back to charge mode.

6.2. Optimization of the PCM Weight

The amount of PCM was optimized by varying the effective weight of the heat exchanger for an indoor unit (W_ale + W_pe) obtained in the previous experiments (case A) by four times (case B), eight times (case C), and 12 times (case D). We assumed that the PCM heat exchanger could be extended in the direction of air flow without changing the thickness of the heat exchange element and the pressure loss. Figure 16 shows the monthly integrated energy consumption for each case. The energy consumption tended to decrease in cases with larger weights. Although cases C and D had similar energy consumptions, case C had a smaller amount of heat insufficient for the cooling load than case D. Thus, the optimal PCM weight to satisfy the cooling load of the target room was eight times the experimental weight (i.e., case C). This value was applied in the following simulations.

6.3. Thermal Behavior on a Representative Day

Figure 17 shows the changes in temperature and amounts of heat of the AC system on May 1, where Q_room is the heat to cool the room, Q_c-PCM is the heat charged to the PCM, and Q_d-PCM is the heat discharged from the PCM. In charge mode, T_p was maintained around 17 °C, owing to solidification of the PCM, and then decreased to 15 °C, owing to changes in the sensible heat. T_out accordingly dropped until it fell below 15 °C, which was the lower limit for the AC system to switch to discharge mode. In discharge mode, T_p was maintained around 17 °C as the PCM melted and then rose to 20 °C. The charge and discharge times for each cycle were 25 and 47 min, respectively, on average. The charge time, during which the outdoor unit was required to operate, was approximately one-third of the total AC time. Q_c-PCM was larger than Q_room at 8:00 because the PCM was cooled from its initial temperature of T_p = 26 °C. After 8:00, Q_c-PCM was added to Q_room at the same rate in charge mode, while Q_d-PCM was always equivalent to Q_room in discharge mode because the outdoor unit was not operating. The maximum load was approximately 15 kW even when the heat for cooling the PCM was added to the indoor cooling load, which was small relative to the total cooling capacity of the eight indoor units (58 kW).

6.4. Effects of the PCM Heat Exchanger

Figure 18 compares the load ratios of the outdoor unit on the same day. Without PCM, low load ratios of <20% were observed throughout the day, suggesting a low energy efficiency. With PCM, the load ratio increased to 50–60%, except for in the late afternoon at around 18:00. Figure 19 shows the load ratio frequencies at 10% intervals and the EER for each load ratio during the whole cooling season. Without PCM, low load ratios of <20% (EER < 3.4) corresponded to an AC time of 680 h, which was 38.6% of the total AC time. With PCM, low load ratios of <20% were reduced to 2.7% of the total AC time. Instead, load ratios of 50–60% (EER = 5.4) took up most of the AC time. Thus, the PCM heat exchanger clearly helped increase the load ratio of the AC system and achieve a high EER. Seasonal EER (SEER) with PCM was 4.97, which was 19% higher than that without PCM. Figure 20 shows the monthly energy consumption of the whole AC system with and without PCM, including the outdoor and indoor units. With PCM, a reduction in energy consumption was observed in all months. The reduction was relatively small in July and August when the cooling load was large and there was little available capacity to cool the PCM. The total reduction in energy consumption for the cooling season was calculated as 11.4%, which indicates that the PCM heat exchanger effectively improved the partial load efficiency of the AC system.

7. Conclusions

We developed a new AC system that uses a PCM heat exchanger to increase the apparent load ratio of the heat pump. Experiments with a prototype heat exchanger demonstrated that the outlet air temperature close to the phase change temperature was obtained during the melting or solidification of the PCM, which led to a U value of greater than 70 W/(m²·K). The heat exchange element also achieved a small pressure drop of less than 10 Pa, suggesting that the PCM heat exchanger could achieve a high thermal performance while maintaining a small pressure drop. The relationship of U to T_p during phase change was formulated as power-law approximations with R² > 0.7 for both the cooling and heating processes. Currently, U is calculated from T_out, which is the output obtained from the experiment. It will be necessary for future works to provide more physical explanations of the relationship with the input conditions.

A numerical model was developed for predicting the outlet air temperature of the PCM heat exchanger. The effective weights of the PCM and aluminum contributing to the heat exchange were determined as 0.53 kg and 5.83 kg, respectively. The numerical model reproduced the experimental outlet air temperatures relatively well, with the average RMSEs of 0.39 K for cooling and 0.84 K for heating, although some deviations were observed during the phase change, which may be attributed to simplifying assumptions such as a uniform phase change by the PCM. The numerical model was then used in simulations to compare the energy consumption of the proposed AC system with and without the PCM heat exchanger. The results showed that using the PCM heat exchanger increased the partial load ratio, avoiding low-efficiency intermittent operation, to approximately 50% in charge mode while allowing the outdoor unit to stay off in discharge mode. The higher partial load ratio in charge mode increased the EER, which is a key indicator of energy efficiency. The total energy consumption was reduced by 11.4% demonstrating the effectiveness of the proposed system.

In this study, the PCM heat exchanger was successfully applied to repeat the charge and discharge modes, while ensuring the room is air-conditioned. In the future, the prototype developed in this study should be extended for use in actual AC systems in buildings. The shape of the prototype is required to be improved to further use the heat capacity of PCMs and prevent the delayed heat transfer deep inside the junction of the prototype, which will cause the PCM to act as a lumped constant system function in the numerical model. In addition, the heat exchange element will be expanded to meet the air flow rate required in the room space, while maintaining high thermal performance and small pressure drops.