I designed this experimental study to answer a practical question for electric vehicle thermal management: can R454C, a low-GWP zeotropic refrigerant, deliver robust cooling, battery cooling, combined cooling, and low-temperature heating in one integrated electric vehicle heat pump architecture? I was particularly interested in the interaction between refrigerant temperature glide, direct cabin evaporation, secondary coolant loops, refrigerant charge, superheat, frontal air velocity, compressor speed, and cabin outlet-air uniformity. Rather than treating the electric vehicle heat pump as a collection of isolated components, I treated it as a coupled system in which the refrigerant inventory, heat-transfer path, expansion control, and air-side flow distribution must be matched together.
The motivation for this work is straightforward. Electric vehicle air conditioning and heat pump systems are important auxiliary loads, and the transition from high-GWP refrigerants toward low-GWP alternatives is reshaping electric vehicle thermal management. R134a has a GWP of 1430, which is difficult to reconcile with long-term low-carbon electric vehicle requirements. R1234yf offers good compatibility but has limited low-temperature heating capability in some electric vehicle heat pump configurations. R290 has excellent thermodynamic properties but introduces A3 flammability constraints that are especially demanding for direct cabin circuits. R744 provides strong low-temperature heating but requires high-pressure components and more complex control. R454C offers a compromise: it is a zeotropic blend of R32 and R1234yf, has a GWP below 150, has zero ODP, and is classified as A2L. Its pressure level is closer to conventional vapor-compression systems than R744, which makes it attractive for electric vehicle heat pump integration, but its temperature glide and A2L safety classification require careful evaluation.
In my experimental work, I combined a direct-expansion cabin evaporator with a coolant-based secondary loop in the same integrated electric vehicle heat pump test bench. This allowed me to compare how temperature glide manifests itself in two different heat-transfer paths. In the direct-expansion path, refrigerant phase change occurs close to the cabin air, so flow maldistribution and local dryout can directly affect outlet-air temperature uniformity. In the secondary loop, coolant mixing and thermal inertia tend to smooth spatial temperature differences, but the loop introduces an additional heat-transfer temperature difference, pump power, and hydraulic complexity. For an electric vehicle, this trade-off matters because cabin comfort, battery thermal safety, packaging, refrigerant charge, and service safety all interact.
I structured the study around four operating families: single cabin cooling, single battery cooling through a chiller, combined cabin-and-battery cooling, and low-temperature heating. Within these families, I varied refrigerant charge, evaporator or chiller outlet superheat, frontal air velocity, and compressor speed. I also compared R454C with R134a in the same hardware and under the same control boundaries where possible. The comparison is not a claim of universal refrigerant ranking; rather, it is a system-level comparison under identical hardware and specified boundary conditions. This distinction is essential for electric vehicle heat pump design because compressor efficiency maps, heat exchanger circuitry, refrigerant charge, and control logic can change the apparent ranking of refrigerants.
Thermophysical Basis and Cycle-Level Expectations
R454C is a zeotropic mixture, so its phase-change temperature is not constant at a given pressure. I used the dew-point and bubble-point temperatures to define the temperature glide:
$$
\Delta T_{\mathrm{glide}} = T_{\mathrm{dew}} – T_{\mathrm{bubble}}
$$
The glide creates a temperature profile that can either improve or degrade heat exchanger performance depending on how the refrigerant temperature change matches the air-side or coolant-side temperature change. For a direct-expansion cabin evaporator, the glide is coupled to parallel-circuit flow distribution and local dryout. For a secondary-loop chiller, the coolant side can smooth the external temperature field, but the additional heat-transfer step reduces thermodynamic effectiveness.
The key properties I used to frame the electric vehicle heat pump comparison are summarized in Table 1. The table is not intended to rank refrigerants by a single property. Instead, it shows why R454C is a candidate for electric vehicle heat pump systems: moderate pressure, low GWP, zero ODP, and an A2L safety classification.
| Property | R134a | R454C |
|---|---|---|
| ODP | 0 | 0 |
| GWP | 1430 | 146 |
| Vapor density at reference state, kg/m3 | 11.2 | 16.2 |
| Liquid density at reference state, kg/m3 | 1317.6 | 1160.0 |
| Critical temperature, °C | 101.1 | 82.4 |
| Critical pressure, MPa | 4.06 | 4.63 |
| Normal boiling point, °C | -26.1 | -45.5 |
| ASHRAE safety class | A1 | A2L |
I performed a cycle-level comparison under fixed assumptions to clarify the theoretical trade-offs before examining experimental data. For cooling, I used an evaporation temperature of 0 °C, a suction superheat of 5 °C, a condensation temperature of 50 °C, and a condenser outlet subcooling of 5 °C. For heating, I used an evaporation temperature of -20 °C, a suction superheat of 5 °C, a condensation temperature of 45 °C, and a condenser outlet subcooling of 5 °C. I assumed a compressor isentropic efficiency of 60%, a volumetric efficiency of 90%, and negligible pressure drop. These assumptions are not a performance prediction for a specific electric vehicle heat pump; they are a controlled thermodynamic comparison.
The compressor pressure ratio is:
$$
\pi = \frac{p_{\mathrm{dis}}}{p_{\mathrm{suc}}}
$$
The coefficient of performance is:
$$
\mathrm{COP} = \frac{Q_{\mathrm{useful}}}{W_{\mathrm{comp}}}
$$
The volumetric refrigeration capacity is:
$$
q_v = \frac{q_m}{v_{\mathrm{suc}}}
$$
Table 2 shows the cycle-level results. I observed that R454C has higher suction pressure, higher suction density, lower pressure ratio, and higher volumetric refrigeration capacity than R134a under the selected assumptions. This suggests a potential to reduce compressor displacement or speed for a given capacity in an electric vehicle heat pump. However, the theoretical COP of R454C was lower than that of R134a in both the cooling and heating cases considered. I therefore did not use pressure level or volumetric capacity alone to judge system efficiency. The actual electric vehicle heat pump performance depends on compressor efficiency, heat exchanger matching, glide utilization, charge management, and control strategy.
| Parameter | R134a cooling | R454C cooling | R134a heating | R454C heating |
|---|---|---|---|---|
| Condensing pressure, MPa | 1.32 | 2.02 | 1.16 | 1.80 |
| High-side glide without pressure drop, °C | 0 | 6.31 | 0 | 6.58 |
| Evaporating pressure, MPa | 0.29 | 0.516 | 0.13 | 0.255 |
| Low-side glide without pressure drop, °C | 0 | 5.10 | 0 | 4.71 |
| Pressure ratio | 4.55 | 3.91 | 8.92 | 7.06 |
| Suction specific enthalpy, kJ/kg | 403.1 | 392.4 | 390.6 | 380.8 |
| Suction density, kg/m3 | 14.1 | 23.1 | 6.6 | 11.7 |
| Discharge specific enthalpy, kJ/kg | 449.1 | 443.6 | 468.1 | 455.4 |
| Discharge temperature, °C | 72.1 | 77.7 | 87.5 | 84.6 |
| Condenser outlet specific enthalpy, kJ/kg | 263.9 | 268.8 | 256.4 | 260.6 |
| Specific cooling or heating effect, kJ/kg | 139.2 | 123.6 | 212.5 | 194.8 |
| Volumetric cooling or heating capacity, kJ/m3 | 1958 | 2862 | 1403 | 2275 |
| Volumetric compression work, kJ/m3 | 646.8 | 1184 | 513.3 | 870.1 |
| COP | 3.03 | 2.42 | 2.73 | 2.61 |
For the electric vehicle heat pump, the practical conclusion from the cycle analysis is that R454C can reduce compressor speed or displacement for a given capacity because of its higher suction density and volumetric capacity, but its system COP must be evaluated experimentally with matched heat exchangers and controls. This is why I moved from thermodynamic screening to integrated electric vehicle heat pump testing.
Integrated Electric Vehicle Heat Pump Test Bench
I built the test bench to represent the key thermal paths of an electric vehicle heat pump without connecting a real battery pack. The bench included an environmental chamber, a refrigerant loop, a coolant loop, a direct cabin evaporator, a chiller for battery cooling, a water-cooled condenser, internal heat exchangers, electronic expansion valves, a compressor, and a multi-way valve network for mode switching. The environmental chamber controlled indoor and outdoor air temperature, humidity, air velocity, and coolant flow. The battery side was represented by a coolant temperature control system at the chiller outlet, so the results reflect steady component-level behavior rather than full vehicle dynamic load, battery thermal inertia, or road-induced disturbances.
The direct cabin cooling mode used a refrigerant-to-air evaporator. The single battery cooling mode, the battery branch of the combined cooling mode, and the heating mode used a refrigerant-to-coolant-to-air secondary path. This mixed architecture is important for an electric vehicle because it allows a direct-expansion cabin circuit and a safer secondary-loop battery circuit to coexist. It also allowed me to compare the air-side temperature uniformity of a direct evaporator with the smoothing effect of a coolant loop.
I used manual steady-state control of the compressor, electronic expansion valves, and coolant pump to isolate parameter effects. A test point was considered stable when the temperature fluctuation was within ±0.5 °C and the pressure fluctuation was within ±0.02 MPa over 10 minutes. The instrument accuracies were ±0.1 K for temperature, ±0.01% for pressure, ±0.01 kg/h for refrigerant and coolant flow, ±0.2 A for current, ±1 V for voltage, and ±0.5 g for refrigerant mass. The experimental matrix is summarized in Table 3. The matrix was designed to identify local trends and candidate control windows, not to fit a global optimum or extrapolate beyond the tested electric vehicle heat pump boundaries.
| Test group | Purpose | Ambient or return condition, °C | Air velocity, m/s | Compressor speed, r/min | Superheat or charge setting |
|---|---|---|---|---|---|
| Charge mapping | Find a feasible charge window across modes | 43/41, 40/30, -7/-7 | 2.5 | 3000 to 4000 | Charge sweep |
| Single cabin cooling | Evaluate superheat and frontal air velocity | 43/41, 40/30, 55/30 | 2, 3, 4.5 | 3300 to 3700 | 5, 8, 12 °C superheat |
| Single chiller | Evaluate battery cooling branch | 43/41, 43/30 | 2, 4.5 | 3000 | 10, 15, 20 °C superheat |
| Combined cooling | Evaluate cabin and battery load sharing | 43/41, 43/34, 43/28 | 3 | 4300 to 5400 | 10 °C superheat |
| Heating | Evaluate low-temperature heating and superheat | -15/-15, -7/-7, 0/0, -25/-25, -20/-20 | 2 | 1900 to 5400 | 4, 8, 12 °C superheat |
For heating capacity, I used the air-side enthalpy method:
$$
Q_h = \dot{m}_{\mathrm{air}}\left(h_{\mathrm{out}} – h_{\mathrm{in}}\right)
$$
Compressor input power was calculated from synchronized voltage and current measurements:
$$
W_{\mathrm{comp}} = \frac{UI}{1000}
$$
The compressor-only heating COP was calculated as:
$$
\mathrm{COP}_h = \frac{Q_h}{W_{\mathrm{comp}}}
$$
For cooling, I used the analogous cooling capacity and compressor-only COP. I did not include fan power, coolant pump power, or auxiliary control power in the reported COP. Therefore, the COP values represent compressor-only system performance. For an electric vehicle, the net system COP should be defined as:
$$
\mathrm{COP}_{\mathrm{net}} = \frac{Q_{\mathrm{useful}}}{W_{\mathrm{comp}} + W_{\mathrm{fan}} + W_{\mathrm{pump}} + W_{\mathrm{aux}}}
$$
This distinction is important for electric vehicle heat pump design because frontal air fans and coolant pumps can change the optimum operating point, especially at high ambient temperature and high frontal air velocity.
I used the Moffat method for uncertainty propagation of indirectly measured quantities:
$$
R = R\left(X_1, X_2, X_3, \ldots, X_n\right)
$$
$$
\frac{\delta R}{R} =
\left[
\sum_{i=1}^{n}
\left(
\frac{\partial R}{\partial X_i}
\frac{\delta X_i}{R}
\right)^2
\right]^{1/2}
$$
The resulting relative uncertainties were 2.6% for cooling or heating capacity and 3.2% for COP. I used these thresholds when interpreting differences between test points. Small differences near the uncertainty level were treated as not distinguishable, while larger differences were used to identify robust trends. This conservative interpretation is important in electric vehicle heat pump testing because mode switching, valve hysteresis, and refrigerant inventory migration can produce local non-monotonic behavior.
Refrigerant Charge Matching Across Modes
Refrigerant charge is a system-level variable in a multi-mode electric vehicle heat pump. The same charge must support single cabin cooling, single battery cooling, combined cabin-and-battery cooling, and heating. If the charge is too low, the evaporator or chiller may be starved, superheat may rise, and capacity may fall. If the charge is too high, excess liquid may occupy condenser volume and receiver volume, condenser pressure may rise, and compressor power may increase. In a multi-mode electric vehicle heat pump, the problem is more complex because refrigerant inventory migrates between the cabin evaporator, chiller, condenser, receiver, and piping when the mode changes.
I defined a feasible charge window for each mode from the measured performance plateau. The approximate plateau ranges are summarized in Table 4. The single cabin cooling mode showed a plateau around 900 to 1100 g, the combined cooling mode around 850 to 1050 g, the single chiller mode around 800 to 1050 g, and the heating mode around 600 to 1100 g. The intersection of these ranges, together with pressure safety margin and repeatability, led me to select 1025 g as the recommended charge for this specific test bench.
| Operating mode | Approximate performance plateau, g | Main observed limitation outside plateau |
|---|---|---|
| Single cabin cooling | 900 to 1100 | Low charge causes evaporator starvation and high superheat |
| Single chiller | 800 to 1050 | Low charge reduces battery cooling capacity |
| Combined cabin and battery cooling | 850 to 1050 | Branch inventory migration and flow split sensitivity |
| Heating | 600 to 1100 | Wide plateau, but high charge raises high-side pressure |
I observed local non-monotonic changes in some indicators in the overcharge region. These changes came from competing effects: liquid fraction in the condenser, refrigerant inventory migration between modes, manual valve adjustment, and changes in effective heat exchanger area. A local fluctuation comparable to the measurement uncertainty was not treated as a stable optimum. Instead, I used the overlapping plateau and pressure safety margin to select the charge. This charge value should not be transferred directly to another electric vehicle heat pump with different piping volume, receiver size, or heat exchanger circuitry.
For charge sensitivity, I used the following local slope as a diagnostic:
$$
S_C = \frac{\partial \mathrm{COP}}{\partial m_{\mathrm{charge}}}
$$
The practical interpretation is that a robust charge should lie in a region where the COP slope is small in all major modes and where the high-side pressure remains below the safety limit with sufficient margin. In my test bench, 1025 g satisfied this compromise better than a charge optimized for a single mode. For electric vehicle heat pump control, this suggests that charge should be treated as a multi-mode design variable rather than a single-mode calibration constant.
Single Cabin Cooling Performance
In the single cabin cooling mode, I varied evaporator outlet superheat and frontal air velocity. The superheat definition was:
$$
SH = T_{\mathrm{evap,out}} – T_{\mathrm{sat}}\left(p_{\mathrm{evap}}\right)
$$
At 43 °C ambient, increasing superheat from 5 °C to 12 °C changed cooling capacity only from about 7.65 kW to 7.77 kW, a maximum difference of about 1.3%. The COP changed from about 1.98 to 2.02, a maximum difference of about 2.0%. Both differences were close to or below the resolved threshold, so I did not identify a precise optimum near 10 °C. At 40 °C ambient, the maximum capacity difference was about 1.4%. The COP at 12 °C superheat was 1.87, about 5.6% higher than 1.77 at 8 °C superheat. This difference was slightly above the combined COP uncertainty, but I had to weigh it against the outlet-air uniformity deterioration discussed later.
| Condition | Superheat, °C | Cooling capacity, kW | COP | Interpretation |
|---|---|---|---|---|
| 43 °C ambient | 5 to 12 | 7.65 to 7.77 | 1.98 to 2.02 | Capacity and COP changes are small |
| 40 °C ambient | 8 | Not resolved separately | 1.77 | Reference point |
| 40 °C ambient | 12 | Not resolved separately | 1.87 | Higher COP but uniformity must be checked |
Frontal air velocity had a much stronger effect than superheat in the high-ambient cooling tests. At 43 °C ambient, increasing frontal air velocity from 2 m/s to 4.5 m/s increased cooling capacity from 5.96 kW to 7.65 kW, an increase of 28.3%. The COP increased from 1.33 to 2.16, an increase of 62.4%. At 55 °C ambient, the capacity increase was smaller, from 1.31 kW to 1.41 kW, an increase of 8.0%, while the COP increased from 0.53 to 0.69, an increase of 30.2%. The smaller gain at 55 °C reflected the exhaust temperature limit and the reduced ability of the front-end heat exchanger to reject heat to the ambient air.
| Ambient, °C | Frontal air velocity, m/s | Cooling capacity, kW | COP | Relative capacity change | Relative COP change |
|---|---|---|---|---|---|
| 43 | 2.0 | 5.96 | 1.33 | Reference | Reference |
| 43 | 4.5 | 7.65 | 2.16 | +28.3% | +62.4% |
| 55 | 2.0 | 1.31 | 0.53 | Reference | Reference |
| 55 | 4.5 | 1.41 | 0.69 | +8.0% | +30.2% |
I also observed diminishing returns as frontal air velocity increased. The gain from 2 m/s to 3 m/s was higher than the gain from 3 m/s to 4.5 m/s. A local marginal benefit can be expressed as:
$$
MB(v) = \frac{\partial \mathrm{COP}}{\partial v_{\mathrm{air}}}
\approx
\frac{\Delta \mathrm{COP}}{\Delta v_{\mathrm{air}}}
$$
For an electric vehicle heat pump, this means that increasing frontal air flow is very effective up to a point, but the fan power penalty must be included in the net energy balance. The compressor-only COP improvement does not automatically translate into vehicle-level energy savings if the fan power rises faster than the compressor power falls. My results therefore support a control strategy in which high ambient cooling prioritizes front-end heat rejection, but the fan speed should be selected using a net COP objective rather than compressor COP alone.
For this test bench, I recommend an initial superheat control window of about 5 to 8 °C for direct cabin cooling when COP and outlet-air uniformity are considered together. This is not a universal optimum. It should be adjusted with load, evaporator circuitry, refrigerant charge, and anti-liquid-slug protection requirements. The important electric vehicle heat pump lesson is that superheat is not only a compressor protection variable; in a direct-expansion cabin evaporator it also affects two-phase heat-transfer area and air-side temperature distribution.
Single Chiller and Battery Cooling Performance
In the single chiller mode, I varied chiller outlet superheat and frontal air velocity. When chiller outlet superheat increased from 10 °C to 20 °C, cooling capacity decreased from 4.71 kW to 3.90 kW, a reduction of 17.4%, which exceeded the capacity uncertainty. The COP was 1.33 at 10 °C superheat and 1.35 at 15 °C superheat, a difference of only 1.5%, which was not distinguishable. At 20 °C superheat, the COP fell to 1.27. I therefore identified 10 to 15 °C as a candidate superheat range for this condition, but I did not claim 15 °C as a precise optimum.
| Chiller superheat, °C | Cooling capacity, kW | COP | Observation |
|---|---|---|---|
| 10 | 4.71 | 1.33 | High capacity, lower superheat |
| 15 | Between 4.71 and 3.90 | 1.35 | COP within uncertainty of 10 °C case |
| 20 | 3.90 | 1.27 | Capacity and COP both degrade |
Frontal air velocity again had a strong effect. When frontal air velocity increased from 2 m/s to 4.5 m/s, chiller cooling capacity increased from 3.90 kW to 4.40 kW, an increase of 12.8%, and COP increased from 1.27 to 1.77, an increase of 39.4%. This change was clearly above the test uncertainty. The mechanism was the same as in cabin cooling: improved front-end heat rejection reduced high-side pressure and compressor power. For an electric vehicle battery cooling mode, the practical recommendation is to increase front-end air flow when the net benefit is positive after fan power is included, and to keep superheat within a range that avoids both liquid slugging risk and excessive capacity loss.
| Frontal air velocity, m/s | Chiller cooling capacity, kW | COP | Relative capacity change | Relative COP change |
|---|---|---|---|---|
| 2.0 | 3.90 | 1.27 | Reference | Reference |
| 4.5 | 4.40 | 1.77 | +12.8% | +39.4% |
The chiller results also highlight the difference between direct and indirect heat transfer in an electric vehicle heat pump. The chiller is separated from the cabin air by a coolant loop, so the air-side temperature field is smoothed by coolant mixing. However, the additional heat-transfer step introduces a temperature difference between refrigerant and coolant, and the coolant pump consumes power. The design question is therefore not whether the secondary loop is inherently better or worse, but whether the safety, charge reduction, and packaging benefits justify the thermodynamic penalty for a given electric vehicle architecture.
Combined Cabin and Battery Cooling
In the combined cabin-and-battery cooling mode, I adjusted the flow split between the cabin evaporator and the chiller. Three representative conditions produced total cooling capacities of 7.92 kW, 6.78 kW, and 7.29 kW, with COPs of 1.68, 1.44, and 1.32, respectively. The results show that shifting cooling load toward the chiller secondary loop tends to reduce COP because of the additional heat-transfer temperature difference and coolant-side losses. Because the three points also changed branch loads simultaneously, the results indicate mode-level differences in this test bench rather than a clean single-variable sensitivity.
| Combined test point | Total cooling capacity, kW | COP | Interpretation |
|---|---|---|---|
| Point A | 7.92 | 1.68 | Higher cabin share and favorable load split |
| Point B | 6.78 | 1.44 | Intermediate load split |
| Point C | 7.29 | 1.32 | Higher secondary-loop load share |
For an electric vehicle heat pump controller, the combined cooling mode is a multi-objective problem. The controller must satisfy cabin comfort, battery thermal safety, compressor envelope, high-side pressure limit, and system energy consumption. I express the combined cooling allocation problem as a constrained objective:
$$
\min J =
\frac{W_{\mathrm{comp}} + W_{\mathrm{fan}} + W_{\mathrm{pump}}}
{Q_{\mathrm{cabin}} + Q_{\mathrm{battery}}}
+
\lambda_1 \left(T_{\mathrm{cabin}} – T_{\mathrm{cabin,target}}\right)^2
+
\lambda_2 \left(T_{\mathrm{battery}} – T_{\mathrm{battery,target}}\right)^2
$$
Here, the first term represents the inverse of net system COP, and the penalty terms enforce cabin and battery temperature targets. In practice, the controller should not maximize instantaneous COP at the expense of battery safety or cabin comfort. The charge and superheat settings must also remain feasible across mode transitions. This is a central challenge for electric vehicle heat pump systems using zeotropic refrigerants: the best charge and control window in one mode may not be the best in another.
Low-Temperature Heating Performance
Heating performance is a critical differentiator for electric vehicle heat pumps. I tested superheat effects at -15 °C, -7 °C, and 0 °C. At -15 °C, -7 °C, and 0 °C, the maximum heating capacity differences caused by superheat changes were about 0.9%, 1.3%, and 0.7%, respectively, all below the combined capacity uncertainty. At -15 °C, the COP values at 4 °C and 8 °C superheat were 2.64 and 2.70, a difference of 2.3%, which was not distinguishable. At -7 °C, the maximum COP difference among three points was about 4.1%, still comparable to the 4.5% threshold. At 0 °C, the COP at 4 °C superheat was 3.33, which was 10.6% higher than at 8 °C superheat and 13.3% higher than at 12 °C superheat. This difference was distinguishable and supports lower superheat control at 0 °C.
| Ambient, °C | Superheat range, °C | Maximum heating capacity difference | COP observation | Interpretation |
|---|---|---|---|---|
| -15 | 4 to 12 | About 0.9% | 2.64 at 4 °C, 2.70 at 8 °C | Not distinguishable |
| -7 | 4 to 12 | About 1.3% | Maximum difference about 4.1% | Near uncertainty threshold |
| 0 | 4 to 12 | About 0.7% | 3.33 at 4 °C | Lower superheat is favored |
I also examined the low-temperature operating boundary. At -25 °C, the outlet air temperature was 22.4 °C, which would be difficult to accept for cabin heating. At -20 °C, the outlet air temperature exceeded 40 °C and the COP was close to 2.5. At -15 °C, the outlet air temperature was 53.2 °C and the COP was about 2.6. The test bench therefore maintained effective low-temperature operation down to -20 °C without vapor injection. At -25 °C, the heating capacity was insufficient. This boundary depends on compressor displacement, minimum suction pressure protection, heat exchanger frost state, and air flow. It should not be interpreted as the thermodynamic limit of R454C.
| Ambient, °C | Outlet air temperature, °C | COP | Practical interpretation |
|---|---|---|---|
| -25 | 22.4 | Not sufficient for the target | Heating capacity insufficient |
| -20 | Above 40 | About 2.5 | Effective low-temperature operation |
| -15 | 53.2 | About 2.6 | Strong heating margin |
For electric vehicle heating, the low-temperature boundary is not only a refrigerant property issue. It is a system integration issue involving compressor speed limits, suction pressure protection, evaporator frost, refrigerant distribution, and cabin air flow. My results show that R454C can extend useful electric vehicle heat pump operation to -20 °C in this test bench without vapor injection, but the design must still verify the compressor envelope and high-side pressure margin.
Outlet-Air Temperature Uniformity in Direct Expansion
The most important zeotropic-specific result in my study was the effect of superheat on cabin outlet-air temperature uniformity. R454C has an equilibrium temperature glide of roughly 4.7 to 6.6 °C in the evaporating and condensing two-phase regions. The glide itself is set by pressure and mixture composition, so changing outlet superheat does not directly change the inherent glide at the same pressure. However, in a direct-expansion evaporator, the two-phase region shortens as superheat increases, and the dryout or superheated region expands. This reduces local two-phase heat-transfer area and makes branch-to-branch flow maldistribution more visible in the air-side temperature field.
I used the maximum-minus-minimum outlet-air temperature difference as a simple uniformity metric:
$$
\Delta T_{\mathrm{air}} = \max\left(T_{\mathrm{out},i}\right) – \min\left(T_{\mathrm{out},i}\right)
$$
I also considered the coefficient of variation and a dimensionless uniformity index:
$$
CV_T = \frac{\sigma_T}{\bar{T}_{\mathrm{out}}} \times 100\%
$$
$$
U_T = 1 – \frac{\sigma_T}{\Delta T_{\mathrm{ref}}}
$$
At 43 °C ambient, increasing superheat from 5 °C to 12 °C increased the outlet-air temperature difference from about 3.1 to 3.4 °C to about 3.8 to 4.1 °C across the three measurement groups. At 40 °C ambient, the difference increased from about 3.4 to 3.7 °C to about 4.6 to 5.0 °C. I interpret this as a coupled result of temperature glide, local dryout or superheated region expansion, and non-uniform refrigerant and air flow distribution. It is not correct to attribute the effect solely to one component evaporating more than another.
| Ambient, °C | Superheat, °C | Outlet-air temperature difference, °C | Interpretation |
|---|---|---|---|
| 43 | 5 | About 3.1 to 3.4 | Better uniformity |
| 43 | 12 | About 3.8 to 4.1 | Degraded uniformity |
| 40 | 5 | About 3.4 to 3.7 | Better uniformity |
| 40 | 12 | About 4.6 to 5.0 | Stronger degradation |
The practical implication is that direct-expansion electric vehicle heat pump control should not optimize superheat only for COP or capacity. A low superheat may improve two-phase coverage and outlet-air uniformity, but it can increase liquid carryover risk. A high superheat may protect the compressor but can degrade cabin comfort and effective evaporator utilization. For this test bench, 5 to 8 °C superheat provided a better compromise between COP and outlet-air uniformity. The exact window should be calibrated for each electric vehicle evaporator and refrigerant charge.
The contrast between direct expansion and the secondary loop is also important. In the chiller, coolant mixing and thermal inertia smooth spatial temperature differences, so the air-side temperature field is less sensitive to local refrigerant dryout. However, the secondary loop introduces an additional heat-transfer temperature difference and pump power. For an electric vehicle heat pump, the choice between direct cabin evaporation and a secondary cabin loop is therefore a system-level trade-off among comfort uniformity, refrigerant charge, safety, and efficiency.
R454C and R134a Comparison Under the Same Hardware
I compared R454C and R134a in the same test bench under given control boundaries. This comparison should be read carefully: the two refrigerants were not separately optimized with different compressor displacements, heat exchanger circuits, or controllers. Some conditions were not strictly equal-capacity or equal-speed. The results therefore describe system behavior under the same hardware and specified control boundaries, not a universal ranking of intrinsic refrigerant performance.
The COP improvement was calculated as:
$$
\mathrm{Improvement}_{\mathrm{COP}} =
\left(
\frac{\mathrm{COP}_{\mathrm{R454C}}}
{\mathrm{COP}_{\mathrm{R134a}}}
– 1
\right)
\times 100\%
$$
The corresponding compressor power reduction at equal cooling load, ignoring auxiliary power, was estimated as:
$$
\mathrm{Reduction}_{W} =
\left(
1 –
\frac{\mathrm{COP}_{\mathrm{R134a}}}
{\mathrm{COP}_{\mathrm{R454C}}}
\right)
\times 100\%
$$
At 43 °C and 40 °C single cabin cooling conditions, R454C produced COPs of 1.98 and 1.73, compared with 1.65 and 1.52 for R134a. This corresponds to COP improvements of 20.0% and 13.8%, and approximate compressor power reductions of 16.7% and 12.1% at equal cooling load. R454C reached similar outlet-air temperatures at lower compressor speed, which is consistent with its higher suction density and volumetric capacity. This advantage can change with compressor efficiency maps and heat exchanger matching.
| Mode and condition | Refrigerant | Compressor speed, r/min | Suction pressure, MPa | Discharge pressure, MPa | Outlet air temperature, °C | COP |
|---|---|---|---|---|---|---|
| Single cabin cooling, 43/41 °C | R454C | 3700 | 0.574 | 2.95 | 11.39 | 1.98 |
| Single cabin cooling, 43/41 °C | R134a | 8000 | 0.304 | 1.817 | 13.43 | 1.65 |
| Single cabin cooling, 40/27 °C | R454C | 3100 | 0.563 | 2.968 | 9.99 | 1.73 |
| Single cabin cooling, 40/27 °C | R134a | 8000 | 0.283 | 1.821 | 10.43 | 1.52 |
| Heating, -15 °C | R454C | 4800 | 0.218 | 2.85 | 58.15 | 2.43 |
| Heating, -15 °C | R134a | 3000 | 0.12 | 0.457 | 9.21 | 3.45 |
| Heating, -7 °C | R454C | 3500 | 0.297 | 2.762 | 56.79 | 2.60 |
| Heating, -7 °C | R134a | 8000 | 0.116 | 1.257 | 48.10 | 1.71 |
The low-temperature heating comparison shows the importance of operating boundaries. At -15 °C, R134a reached a suction pressure of 0.12 MPa at 3000 r/min, and the outlet air temperature was only 9.21 °C under the protection boundary. R454C reached an outlet air temperature of 58.15 °C at 4800 r/min. At -7 °C, R454C produced a COP of 2.60, higher than 1.71 for R134a. These comparisons do not mean that R454C is universally superior in heating COP; the R134a point at -15 °C had a much lower outlet temperature, so the useful heating capacity was limited. The system-level conclusion is that R454C extended the no-vapor-injection operating boundary to -20 °C in this electric vehicle heat pump test bench, while R134a was constrained by low suction pressure and insufficient heating delivery at the tested boundary.
R454C is not a drop-in replacement without engineering changes. It is A2L and has a higher high-side pressure than R134a in several conditions. An electric vehicle heat pump using R454C should recheck pressure rating, compressor displacement, heat exchanger circuitry, refrigerant charge, leak detection, ventilation, and electrical protection. The design should also consider service and repair procedures for A2L refrigerants in electric vehicle thermal management systems.
System-Level Interpretation for Electric Vehicle Heat Pumps
My experimental results suggest that the optimization of a zeotropic low-GWP refrigerant in an electric vehicle heat pump should shift from single-parameter optimization to coordinated matching of charge, superheat, heat-transfer capacity, and system heat-transfer path. A charge that is best for cabin cooling may not be best for battery cooling. A superheat that maximizes COP may degrade cabin outlet-air uniformity. A frontal air velocity that improves compressor COP may increase fan power and reduce net vehicle energy efficiency. These interactions are especially important in electric vehicle heat pump systems because the same hardware must operate across cooling, heating, dehumidification, battery cooling, and combined modes.
I summarize the main control and design windows in Table 9. These values are specific to my test bench and should be treated as starting points for electric vehicle heat pump development rather than universal calibration constants.
| Design or control item | Observed recommendation | Reason |
|---|---|---|
| Multi-mode refrigerant charge | 1025 g for this test bench | Overlapping performance plateau plus pressure safety margin |
| Direct cabin cooling superheat | About 5 to 8 °C | Compromise between COP and outlet-air uniformity |
| Chiller superheat | About 10 to 15 °C for the tested condition | Capacity retention and COP within uncertainty |
| Frontal air velocity | Increase until net benefit diminishes | Strong high-side pressure reduction but marginal returns and fan power penalty |
| Low-temperature heating boundary | Effective to -20 °C without vapor injection | Outlet air above 40 °C and COP near 2.5 |
| Refrigerant safety | A2L review and pressure rating verification | Higher pressure than R134a in several conditions |
I also recommend a structured electric vehicle heat pump design checklist. The checklist should include refrigerant safety, pressure envelope, charge migration, direct-expansion flow distribution, superheat control, front-end heat rejection, secondary-loop pump power, compressor speed envelope, and mode transition logic. Table 10 presents this checklist in a compact form.
| Area | Design question | Why it matters for an electric vehicle heat pump |
|---|---|---|
| Refrigerant safety | Are A2L leak detection, ventilation, and electrical protection sufficient? | R454C is A2L and requires specific service and safety measures |
| Pressure envelope | Are components rated for the highest high-side pressure across all modes? | R454C high-side pressure can exceed R134a in some conditions |
| Charge management | Does the charge remain feasible in cabin, battery, combined, and heating modes? | Refrigerant inventory migrates between branches during mode switching |
| Direct evaporator circuitry | Is refrigerant distribution uniform enough at low and high superheat? | Glide and maldistribution affect outlet-air temperature uniformity |
| Superheat control | Does the controller balance COP, liquid carryover, and air-side uniformity? | Superheat changes two-phase area and dryout location |
| Front-end heat rejection | Is fan power included in the net COP objective? | Higher air velocity improves compressor COP but may increase auxiliary load |
| Secondary loop | Does coolant mixing justify the added temperature difference and pump power? | Secondary loops smooth temperature but reduce thermodynamic effectiveness |
| Compressor envelope | Does the compressor map cover low-temperature heating and high-ambient cooling? | Electric vehicle heat pump operation spans wide pressure ratios and speeds |
From a modeling perspective, I found it useful to separate the electric vehicle heat pump performance into three layers. The first layer is the refrigerant cycle: pressure ratio, suction density, compressor efficiency, and COP. The second layer is the heat exchanger and air-side behavior: temperature glide, flow distribution, dryout, and outlet-air uniformity. The third layer is the vehicle system: fan power, pump power, cabin load, battery load, mode transitions, and safety constraints. A design that is optimal in the first layer may be suboptimal in the third layer. This is the main reason I emphasize coordinated matching rather than a single “best” charge, superheat, or air velocity.
For a direct-expansion electric vehicle heat pump, I would prioritize evaporator circuit design and refrigerant distribution at the early design stage. The outlet-air temperature field is a sensitive indicator of refrigerant maldistribution, and it can reveal problems that are not visible in average capacity or COP. For a secondary-loop battery cooling branch, I would prioritize chiller approach temperature, coolant flow distribution, and pump power. For a combined cooling mode, I would prioritize a control strategy that maintains battery safety while avoiding unnecessary secondary-loop heat-transfer losses.
Limitations and Future Work
I interpret the results within the boundaries of the test bench. The battery was represented by a coolant temperature control system, so the study does not include real battery thermal inertia, cell-level temperature gradients, or dynamic road loads. The compressor, expansion valves, and pump were manually controlled to isolate parameter effects, so the results do not include production controller dynamics. The charge recommendation is specific to the piping volume and heat exchanger geometry of this electric vehicle heat pump test bench. The refrigerant comparison used the same hardware and did not include separate compressor or heat exchanger optimization for each refrigerant.
Future work should extend the present steady-state findings to transient mode switching, frost and defrost behavior, real battery thermal load, and vehicle-level net energy consumption. A useful next step would be to combine the experimental trends with a dynamic electric vehicle heat pump model that includes refrigerant inventory migration, cabin thermal mass, battery thermal mass, fan power, and pump power. Another important direction is to optimize the direct-expansion evaporator circuitry specifically for zeotropic glide, so that the air-side temperature profile benefits from the glide rather than being degraded by maldistribution. Finally, the A2L safety and high-pressure requirements of R454C should be evaluated together with packaging and serviceability in production electric vehicle thermal management systems.
Conclusions
I draw the following conclusions from this integrated electric vehicle heat pump study.
- In the multi-mode electric vehicle heat pump, the feasible refrigerant charge differs by mode. The single cabin cooling, single chiller, combined cooling, and heating modes showed different performance plateaus. For this test bench, 1025 g provided the best compromise when the overlapping plateau and pressure safety margin were considered together. The local non-monotonic behavior in the overcharge region came from competing effects of condenser liquid fraction, mode-to-mode inventory migration, valve adjustment, and effective heat exchanger area.
- High-ambient cooling performance was more sensitive to frontal heat rejection than to superheat within the tested range. At 43 °C ambient, increasing frontal air velocity from 2 m/s to 4.5 m/s increased cooling capacity by 28.3% and COP by 62.4%. At 55 °C ambient, the improvements were smaller because of exhaust temperature limits. The gain showed diminishing returns, so the electric vehicle heat pump controller should use a net COP objective that includes fan power.
- In direct cabin cooling, superheat affects outlet-air temperature uniformity. Increasing superheat from 5 °C to 12 °C increased the outlet-air temperature difference from about 3.1 to 3.7 °C to about 3.8 to 5.0 °C. I interpret this as a coupled result of temperature glide, local dryout or superheated region expansion, and non-uniform refrigerant and air flow distribution. A 5 to 8 °C superheat window provided a better compromise between COP and uniformity in this test bench.
- In the single chiller mode, increasing superheat from 10 °C to 20 °C reduced cooling capacity by 17.4%. The COP at 10 °C and 15 °C superheat was not distinguishable within uncertainty, while 20 °C superheat reduced COP. Increasing frontal air velocity from 2 m/s to 4.5 m/s increased chiller capacity by 12.8% and COP by 39.4%.
- The combined cabin-and-battery cooling mode showed that shifting load toward the secondary loop tends to reduce COP because of the additional heat-transfer temperature difference and coolant-side losses. The controller should treat combined cooling as a constrained multi-objective problem involving cabin comfort, battery safety, compressor envelope, and net energy consumption.
- R454C maintained effective low-temperature heating to -20 °C without vapor injection in this electric vehicle heat pump test bench. At -20 °C, the outlet air temperature exceeded 40 °C and the COP was close to 2.5. At -25 °C, the outlet air temperature fell to 22.4 °C, indicating insufficient heating capacity. This boundary depends on compressor displacement, suction pressure protection, frost state, and air flow.
- Under the same hardware and given control boundaries, R454C achieved 13.8% to 20.0% higher COP than R134a in the 40 to 43 °C cooling conditions, with approximate compressor power reductions of 12.1% to 16.7% at equal cooling load. R454C also reached useful heating delivery at lower temperatures than R134a in the tested boundary. These results do not make R454C a drop-in replacement; A2L safety, high-side pressure, compressor displacement, heat exchanger circuitry, charge, and leak detection must be rechecked for each electric vehicle heat pump platform.
- For zeotropic low-GWP refrigerants in electric vehicle heat pumps, the optimization approach should move from single-parameter tuning to coordinated matching of charge, superheat, heat-transfer capacity, and system heat-transfer path. The direct-expansion path requires attention to flow distribution and outlet-air uniformity, while the secondary-loop path requires attention to approach temperature and pump power. This coordinated approach is the most reliable route to efficient, comfortable, and safe electric vehicle heat pump systems using R454C and other low-GWP zeotropic blends.