As global climate governance accelerates, the refrigeration and air-conditioning industry is shifting from high global warming potential (GWP) working fluids toward low-GWP alternatives. The air-conditioning and heat-pump system of an electric car is a major auxiliary energy consumer, and refrigerant substitution has therefore become a key direction in electric car thermal management. China’s national plan for implementing the Montreal Protocol on Substances that Deplete the Ozone Layer (2025–2030) states that, starting from 1 July 2029, new M1-category vehicle air-conditioning systems using refrigerants with GWP greater than 150 will be prohibited. The conventional R134a has a GWP of 1,430 and can no longer satisfy future low-carbon requirements. In this work I investigate, from a first-person experimental perspective, how the zeotropic low-GWP refrigerant R454C behaves in a multimode heat-pump system intended for an electric car, and how temperature glide maps onto system performance and cabin outlet-air uniformity.
The available low-GWP routes include hydrofluoroolefins (HFOs), natural working fluids and refrigerant blends. R1234yf is reasonably compatible with existing hardware, but its low-temperature heating capability is limited. R290 has excellent thermophysical properties, yet its A3 flammability imposes strict charge limits, leak detection and safety protection on any direct cabin circuit, while a secondary loop adds heat-transfer temperature difference, pump power and complexity. R744 offers strong low-temperature heating, but its transcritical high pressure increases component pressure-rating and control difficulty. Blends can compromise among environmental impact, safety class, operating pressure, volumetric capacity and temperature glide. R454C is composed of R32 and R1234yf, has a GWP below 150, an ozone depletion potential (ODP) of zero and a safety class of A2L. Its pressure level is close to that of conventional vapor-compression systems, which gives it potential for electric car heat-pump applications, although A2L safety and high-pressure component adaptation still need to be assessed.
Existing studies have demonstrated the potential of R454C in residential heat pumps and in vehicle secondary-loop systems, but public experimental work on the direct-expansion cabin mode remains scarce. In a direct system, refrigerant phase-change temperature glide, the dryout location along each circuit and flow maldistribution map directly onto the air-side temperature field; in an indirect system, coolant mixing and thermal inertia smooth out spatial temperature differences, at the cost of additional secondary heat-transfer losses. At the same time, an electric car must switch among cabin cooling, battery cooling, combined cabin-and-battery cooling and low-temperature heating, and each mode has a different refrigerant inventory and control boundary. An “optimum” obtained under a single condition is therefore not automatically transferable. My differentiated contributions are: first, a single integrated test bench that simultaneously contains direct cabin evaporation and a coolant secondary loop, so that temperature glide can be compared across the two heat-transfer paths; second, a joint examination of refrigerant charge, superheat, frontal air velocity and compressor speed across multiple operating modes, together with an analysis of outlet-air uniformity in the direct-expansion system based on the evaporator outlet temperature field; and third, a direct limit-performance comparison between R454C and R134a in the same system. All recommended charge values, candidate superheat ranges and energy-saving percentages reported here are valid within the boundary of this particular test bench.
Theoretical Background and Thermophysical Properties
R454C is a binary blend of R32 and R1234yf with mass fractions of 21.5% and 78.5% respectively, a GWP of 146 and a safety class of A2L. Table 1 compares its main physical properties with those of R134a. The most relevant differences for an electric car heat pump are the higher vapor density, the lower liquid density, the lower critical temperature and the higher critical pressure of R454C, together with its much lower normal boiling point of −45.5 °C, which is favorable for low-temperature heating of an electric car.
| Property | R134a | R454C |
|---|---|---|
| ODP | 0 | 0 |
| GWP | 1,430 | 146 |
| Vapor density (kg/m³) | 11.2 | 16.2 |
| Liquid density (kg/m³) | 1,317.6 | 1,160.0 |
| 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 selected two representative cycles for a theoretical comparison, one for cooling and one for heating. The cooling condition was set to an evaporation temperature of 0 °C, an outlet superheat of 5 °C, a condensation temperature of 50 °C and an outlet subcooling of 5 °C. The heating condition was set to an evaporation temperature of −20 °C, an outlet superheat of 5 °C, a condensation temperature of 45 °C and an outlet subcooling of 5 °C. Both conditions assumed a compressor isentropic efficiency of 60% and a volumetric efficiency of 90%, with pressure drops neglected. The results are summarized in Table 2.
| Parameter | Cooling, R134a | Cooling, R454C | Heating, R134a | Heating, R454C |
|---|---|---|---|---|
| Condensing pressure (MPa) | 1.32 | 2.02 | 1.16 | 1.80 |
| High-side glide, no pressure drop (°C) | 0 | 6.31 | 0 | 6.58 |
| Evaporating pressure (MPa) | 0.29 | 0.516 | 0.13 | 0.255 |
| Low-side glide, no pressure drop (°C) | 0 | 5.10 | 0 | 4.71 |
| Pressure ratio | 4.55 | 3.91 | 8.92 | 7.06 |
| Suction enthalpy (kJ/kg) | 403.1 | 392.4 | 390.6 | 380.8 |
| Suction density (kg/m³) | 14.1 | 23.1 | 6.6 | 11.7 |
| Discharge enthalpy (kJ/kg) | 449.1 | 443.6 | 468.1 | 455.4 |
| Discharge temperature (°C) | 72.1 | 77.7 | 87.5 | 84.6 |
| Condenser outlet enthalpy (kJ/kg) | 263.9 | 268.8 | 256.4 | 260.6 |
| Specific refrigerating effect (kJ/kg) | 139.2 | 123.6 | 212.5 | 194.8 |
| Volumetric refrigerating effect (kJ/m³) | 1,958 | 2,862 | 1,403 | 2,275 |
| Volumetric compression work (kJ/m³) | 646.8 | 1,184 | 513.3 | 870.1 |
| COP | 3.03 | 2.42 | 2.73 | 2.61 |
Under these assumptions, R454C exhibits a higher suction pressure and suction density, a lower pressure ratio and a higher volumetric refrigerating capacity than R134a, which indicates a potential for reducing compressor displacement or speed while delivering the same capacity in an electric car. However, its theoretical COP is 2.42 and 2.61 for the cooling and heating conditions respectively, lower than the 3.03 and 2.73 of R134a. System energy efficiency therefore cannot be judged from operating pressure or volumetric capacity alone; the actual result also depends on compressor efficiency, heat-exchanger matching, exploitation of temperature glide and control strategy.
R454C exhibits an equilibrium temperature glide of approximately 4.7–6.6 K across the evaporation and condensation two-phase regions. Temperature glide can reduce heat-transfer irreversibility when the refrigerant temperature variation matches that of the air or coolant, but it can also couple with maldistribution among parallel circuits and with dryout at different locations, producing wall-temperature and outlet-air-temperature differences. In a direct-expansion evaporator these differences propagate directly to the cabin air side, whereas in a secondary loop the coolant mixing has a smoothing effect. This is the physical basis for the separate discussion of system performance and outlet-air uniformity that follows.
Experimental Setup and Test Matrix
To study the cooling, heating and outlet-air uniformity behavior of R454C in an integrated thermal management system for an electric car, I built the test bench shown schematically in the architecture. The environmental simulation chamber can regulate indoor and outdoor air temperature, humidity, velocity and coolant flow rate. The system consists of a refrigerant circuit and a coolant circuit, and multi-way valves switch among cabin-only cooling, battery-only cooling, combined cabin-and-battery cooling and cabin heating. Cabin-only cooling uses a refrigerant-to-air direct-expansion circuit; battery-only cooling, the battery branch of the combined mode and the heating mode use a refrigerant-to-coolant-to-air secondary loop. This hybrid architecture allows the two heat-transfer paths to be compared on the same hardware platform.
The compressor, electronic expansion valve and coolant pump were controlled manually in steady state in order to isolate the effect of each parameter. The bench was not connected to a real battery pack; instead, a coolant temperature-control system simulated the battery outlet coolant temperature. The results therefore reflect steady-state component-level behavior and do not include vehicle-level dynamic load, battery thermal inertia or road-condition disturbances.
All instruments were calibrated before the experiments. The accuracies of temperature, pressure, refrigerant flow rate, coolant flow rate, current, voltage and electronic balance were ±0.1 K, ±0.01%, ±0.01 kg/h, ±0.01 kg/h, ±0.2 A, ±1 V and ±0.5 g respectively. The experimental matrix is given in Table 3. Condition 1 determined a reasonable charge interval, while conditions 2–5 examined frontal air velocity, compressor speed and evaporator or chiller outlet superheat. After stabilization, data were collected continuously, with the criterion being a temperature fluctuation of no more than 0.5 K and a pressure fluctuation of no more than 0.02 MPa over 10 min. The three-level sweeps only identify local trends and candidate ranges; they are not used to fit a global optimum or to extrapolate.
| No. | Test | Ambient / return air or inlet coolant (°C) | Air velocity (m/s) | Compressor speed (r/min) | Air flow (m³/h) | Evaporator or chiller outlet superheat (°C) |
|---|---|---|---|---|---|---|
| 1 | Charge test (cabin AC, single chiller, AC + chiller, heating) | 43/41, −7/−7 | 2.5 | 3,000–3,500 | 468 | — |
| 2 | Single AC performance | 40/30, 43/41, 55/30 | 2, 3, 4.5 | 3,300–3,700 | 468 | 5, 8, 12 |
| 3 | Single chiller performance | 43/30, 43/41 | 2, 3, 4.5 | 3,500–4,300 | — | 10, 15, 20 |
| 4 | AC + chiller | 43/34, 43/28 | 3 | 4,700–5,400 | 468 | 10 |
| 5 | Heating performance | −15/−15, −7/−7, 0/0, −25/−25, −20/−20 | 2 | 1,900–5,400 | 255–440 | 4, 8, 12 |
Heating performance was calculated using the air-side enthalpy-difference method. The internal condenser heating capacity is obtained from the air mass flow rate and the inlet and outlet moist-air specific enthalpy difference:
$$Q_h = \dot{m}_{air}\left(h_{out} – h_{in}\right)$$
where \(Q_h\) is the heating capacity in kW, \(\dot{m}_{air}\) is the air mass flow rate measured by the nozzle flow meter in kg/s, and \(h_{in}\) and \(h_{out}\) are the inlet and outlet moist-air specific enthalpies of the internal condenser in kJ/kg.
Compressor input power was calculated from simultaneously sampled voltage and current:
$$W = \frac{UI}{1000}$$
where \(W\) is the compressor input power in kW, \(U\) is the compressor input voltage in V and \(I\) is the compressor input current in A. The COP in this work only uses compressor input power as the energy input and does not include fan, coolant pump or auxiliary control power. The system heating COP is therefore:
$$COP = \frac{Q_h}{W}$$
Cooling capacity and cooling COP were computed in an analogous way from the refrigerant-side enthalpy difference:
$$Q_c = \dot{m}_{ref}\left(h_{out} – h_{in}\right), \qquad COP_c = \frac{Q_c}{W}$$
The Moffat method was used to propagate the uncertainty of indirectly measured quantities:
$$R = R\left(X_1, X_2, X_3, \ldots, X_n\right)$$
$$\frac{\delta R}{R} = \left[\sum_{i=1}^{n}\left(\frac{\partial R}{\partial X_i}\cdot\frac{\delta X_i}{R}\right)^{2}\right]^{1/2}$$
where \(R\) is the target quantity calculated from the measured parameters, \(X_i\) is the \(i\)-th independent measured parameter and the subscript \(i\) is the measurement-variable index. From the basic instrument accuracy, the relative uncertainties of cooling or heating capacity and of COP were calculated to be 2.6% and 3.2% respectively.
Superheat at the evaporator or chiller outlet was defined as the difference between the outlet temperature and the saturation temperature at the local pressure:
$$\Delta T_{sh} = T_{out} – T_{sat}\left(p_{out}\right)$$
For a zeotropic blend, the relevant saturation temperature depends on the local composition and on whether the state is bubble-point or dew-point. The equilibrium temperature glide is:
$$\Delta T_{glide} = T_{dew}\left(p\right) – T_{bubble}\left(p\right)$$
This glide is the key quantity that couples refrigerant-side phase change with air-side temperature uniformity in the direct-expansion evaporator of an electric car.
Refrigerant Charge Determination for Multimode Operation
The charge results for each mode in condition 1 are summarized in Table 4. The performance plateau intervals were approximately 900–1,100 g for cabin-only cooling, 850–1,050 g for combined cooling, 800–1,050 g for single-chiller operation and 600–1,100 g for heating. When the charge is insufficient, the evaporator or chiller is under-fed, the superheat rises and the cooling capacity decreases. As the charge increases, liquid supply and subcooling improve. Once the overcharge region is entered, the additional liquid occupies volume in the condenser and receiver, the condensing pressure and compressor power rise, and manual electronic expansion valve adjustment, refrigerant inventory migration among modes and changes in effective heat-exchanger area compete with one another. Some indicators therefore show a non-monotonic increase followed by a decrease. The effect of charge deviation from the design value on cooling and heating capacity and efficiency is equipment- and condition-dependent; local fluctuations of the same order as the measurement uncertainty were not treated as stable turning points.
| Mode | Performance plateau interval (g) | Dominant limiting factor |
|---|---|---|
| Cabin-only cooling, direct expansion | 900–1,100 | Evaporator liquid supply and superheat stability |
| Combined cabin and battery cooling | 850–1,050 | Inventory split between the two branches |
| Single chiller, secondary loop | 800–1,050 | Chiller outlet superheat and coolant temperature |
| Heating, secondary loop | 600–1,100 | Low-side pressure and condenser liquid hold-up |
| Recommended charge for this bench | 1,025 | Overlap of plateaus plus pressure safety margin |
Taking into account the plateau intervals of all modes, the pressure safety margin and repeatability, I selected 1,025 g as the recommended charge for this test bench. This value should not be transferred directly to a system with different pipe volume or heat-exchanger geometry, and it should be re-established for each electric car platform.
Single Cabin Cooling Performance
The effect of superheat in cabin-only cooling is shown in Table 5. At 43 °C, when the superheat was increased from 5 °C to 12 °C, the cooling capacity varied between 7.65 and 7.77 kW, a maximum difference of about 1.3%, and the COP varied between 1.98 and 2.02, a maximum difference of about 2.0%. Neither exceeded the corresponding resolvable threshold, so an exact optimum near 10 °C cannot be identified from these data. At 40 °C, the maximum difference in cooling capacity was about 1.4%; the COP at 12 °C superheat was 1.87, about 5.6% higher than the 1.77 at 8 °C, slightly above the combined COP uncertainty, but this must be weighed against the degradation of outlet-air uniformity discussed below.
| Condition | Variable | Cooling capacity (kW) | COP | Relative change |
|---|---|---|---|---|
| 43 °C, superheat 5→12 °C | Superheat | 7.77→7.65 | 2.02→1.98 | ≈1.3% / 2.0% (not resolvable) |
| 40 °C, superheat 8→12 °C | Superheat | — | 1.77→1.87 | +5.6% in COP |
| 43 °C, air velocity 2→4.5 m/s | Frontal air velocity | 5.96→7.65 | 1.33→2.16 | +28.3% / +62.4% |
| 55 °C, air velocity 2→4.5 m/s | Frontal air velocity | 1.31→1.41 | 0.53→0.69 | +8.0% / +30.2% |
The influence of frontal air velocity is much more significant. At 43 °C, when the velocity was increased from 2 m/s to 4.5 m/s, the cooling capacity rose from 5.96 kW to 7.65 kW, an increase of 28.3%, and the COP rose from 1.33 to 2.16, an increase of 62.4%. At 55 °C, limited by discharge temperature, the cooling capacity rose only from 1.31 kW to 1.41 kW, an increase of 8.0%, and the COP from 0.53 to 0.69, an increase of 30.2%. Increasing the air velocity enhances frontal heat rejection and lowers the high-side pressure and pressure ratio, but the gain in the 2–3 m/s stage is larger than in the 3–4.5 m/s stage, showing diminishing marginal returns. These results do not include frontal fan power, so the net vehicle-level energy saving of an electric car still requires further verification.
For this bench, high-ambient direct cooling should first guarantee frontal heat-rejection capability. The evaporator outlet superheat should be controlled initially in the 5–8 °C range as a compromise between COP and outlet-air uniformity, and then corrected according to load and liquid-slugging prevention requirements, rather than being treated as a fixed optimum across conditions.
Single Chiller Performance
The single-chiller results are summarized in Table 6. When the chiller outlet superheat was increased from 10 °C to 20 °C, the cooling capacity dropped from 4.71 kW to 3.90 kW, a decrease of 17.4%, exceeding the capacity uncertainty. The COP was 1.33 and 1.35 at 10 °C and 15 °C respectively, a difference of only 1.5% that is not resolvable, and dropped to 1.27 at 20 °C. Only the 10–15 °C range can therefore be identified as a candidate interval for this condition; 15 °C cannot be declared the exact optimum.
| Condition | Variable | Cooling capacity (kW) | COP | Relative change |
|---|---|---|---|---|
| Superheat 10→20 °C | Superheat | 4.71→3.90 | 1.33→1.27 | −17.4% in capacity |
| Superheat 10 °C vs 15 °C | Superheat | — | 1.33 vs 1.35 | +1.5% (not resolvable) |
| Air velocity 2→4.5 m/s | Frontal air velocity | 3.90→4.40 | 1.27→1.77 | +12.8% / +39.4% |
When the frontal air velocity was increased from 2 m/s to 4.5 m/s, the cooling capacity rose from 3.90 kW to 4.40 kW, an increase of 12.8%, and the COP rose from 1.27 to 1.77, an increase of 39.4%. This change is clearly larger than the test uncertainty, because enhanced frontal heat rejection lowers the high-side pressure and compressor power. For the single-chiller mode of an electric car, the frontal air flow should be increased while accounting for the additional fan power, and the superheat should be controlled in a reasonable interval between liquid-slugging prevention and capacity degradation.
Combined Cabin and Battery Cooling
In condition 4, the flow distribution between the evaporator and chiller branches was adjusted to split the cooling capacity between the cabin and the battery. The three groups had total cooling capacities of 7.92, 6.78 and 7.29 kW and COPs of 1.68, 1.44 and 1.32 respectively, as listed in Table 7. When cooling capacity is shifted toward the chiller secondary loop, the COP generally decreases, reflecting the secondary heat-transfer temperature difference and the coolant-side loss. Because the three points changed the branch loads simultaneously, the results only indicate the mode differences of this bench and do not isolate a single causal parameter.
| Case | Total cooling capacity (kW) | COP | Observation |
|---|---|---|---|
| 1 | 7.92 | 1.68 | Higher cabin share |
| 2 | 6.78 | 1.44 | Intermediate split |
| 3 | 7.29 | 1.32 | Higher chiller share, larger secondary loss |
The trend indicates that, in an electric car with simultaneous cabin and battery thermal loads, the secondary loop is convenient for packaging and safety but carries an efficiency penalty. The control strategy should therefore avoid unnecessary capacity transfer to the coolant loop when the battery load can be met by other means.
Heating Performance and Low-Temperature Operating Boundary
The heating results are summarized in Table 8. At −15 °C, −7 °C and 0 °C, the maximum differences in heating capacity caused by superheat variation were about 0.9%, 1.3% and 0.7% respectively, all below the combined capacity uncertainty. At −15 °C, the COP at 4 °C and 8 °C superheat was 2.64 and 2.70 respectively, a difference of 2.3% that is not resolvable. At −7 °C, the maximum COP difference among the three points was about 4.1%, still of the same order as the 4.5% threshold. At 0 °C, the COP at 4 °C superheat was 3.33, 10.6% and 13.3% higher than at 8 °C and 12 °C respectively, and this difference is resolvable. The −15 °C to −7 °C experiments therefore only give a candidate interval of 4–8 °C, while the 0 °C condition supports lower superheat control.
| Ambient temperature (°C) | Superheat range (°C) | Heating capacity variation | COP | Remark |
|---|---|---|---|---|
| −15 | 4–8 | ≈0.9% | 2.64–2.70 | Candidate interval only |
| −7 | 4–12 | ≈1.3% | Max. difference ≈4.1% | Not fully resolvable |
| 0 | 4 vs 8 vs 12 | ≈0.7% | 3.33 at 4 °C superheat | +10.6% and +13.3% vs higher superheat |
| −25 | — | — | — | Outlet air 22.4 °C, insufficient |
| −20 | — | — | ≈2.5 | Outlet air above 40 °C |
| −15 | — | — | ≈2.6 | Outlet air 53.2 °C |
The temperature dependence was used to probe the low-temperature operating boundary. At −25 °C the outlet air temperature was 22.4 °C, which can hardly meet the cabin heating demand of an electric car. 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. This bench can therefore operate down to −20 °C without vapor injection, but the heating capacity at −25 °C is insufficient. This boundary is affected by compressor displacement, the minimum suction pressure protection, heat-exchanger frosting state and air flow, and should not be regarded as the thermodynamic limit of R454C.
Outlet-Air Temperature Uniformity
The equilibrium temperature glide of R454C is set by pressure and circulating composition; changing the outlet superheat does not directly change the intrinsic glide at the same pressure. In a direct-expansion evaporator, the refrigerant in the two-phase region warms along the flow path, and local flow-distribution differences cause individual circuits to dry out at different locations. When the superheat increases, the two-phase heat-transfer region shortens, the dryout and superheated region expands and the local heat-transfer coefficient decreases, so wall-temperature differences among circuits are more easily mapped into outlet-air temperature stratification. Studies of multi-circuit evaporators have shown that refrigerant maldistribution can be characterized by outlet superheat and by the air outlet temperature field. In an indirect chiller, coolant mixing and thermal inertia weaken spatial temperature differences, and temperature glide may reduce heat-transfer irreversibility when the flow directions are matched, but this requires a secondary heat-transfer temperature difference and pump power.
The effect of superheat on outlet-air temperature uniformity was evaluated using the difference between the maximum and minimum outlet-air temperatures among the evaporator measurement points, as summarized in Table 9. At 43 °C, when the superheat was increased from 5 °C to 12 °C, the temperature difference of the three measurement groups increased from about 3.1–3.4 °C to 3.8–4.1 °C. At 40 °C, the difference increased from about 3.4–3.7 °C to 4.6–5.0 °C. The increase in temperature difference should therefore be interpreted as the coupled result of temperature glide, expansion of the local dryout and superheated region and maldistribution of refrigerant and air flow, rather than as the effect of a single component evaporating more. In summary, 5–8 °C superheat gave the best outlet-air uniformity on this bench.
| Ambient condition | Superheat 5 °C | Superheat 8 °C | Superheat 12 °C |
|---|---|---|---|
| 43 °C | ≈3.1–3.4 °C | ≈3.5–3.8 °C | ≈3.8–4.1 °C |
| 40 °C | ≈3.4–3.7 °C | ≈4.1–4.5 °C | ≈4.6–5.0 °C |
The uniformity metric used here can be written as:
$$\Delta T_{air} = T_{air,max} – T_{air,min}$$
A smaller \(\Delta T_{air}\) corresponds to a more uniform cabin supply temperature. Because the direct-expansion evaporator feeds the cabin of an electric car directly, this metric is closely related to passenger thermal comfort and to the risk of local cold or warm spots.
R454C versus R134a in the Same Hardware
Table 10 compares R454C and R134a. Both refrigerants were tested on the same bench, but they were not separately optimized in terms of compressor displacement, heat-exchanger circuiting and controller, and some conditions were not strictly capacity-matched or speed-matched. The comparison therefore reflects system behavior under identical hardware and given control boundaries, and is not equivalent to a general ranking of intrinsic working-fluid performance.
| Mode | Refrigerant | Ambient / return air (°C) | Compressor speed (r/min) | Suction pressure (MPa) | Discharge pressure (MPa) | Outlet air (°C) | COP |
|---|---|---|---|---|---|---|---|
| Single cabin cooling | R454C | 43/41 | 3,700 | 0.574 | 2.95 | 11.39 | 1.98 |
| R134a | 43/41 | 8,000 | 0.304 | 1.817 | 13.43 | 1.65 | |
| Single cabin cooling | R454C | 40/27 | 3,100 | 0.563 | 2.968 | 9.99 | 1.73 |
| R134a | 40/27 | 8,000 | 0.283 | 1.821 | 10.43 | 1.52 | |
| Heating | R454C | −15/−15 | 4,800 | 0.218 | 2.85 | 58.15 | 2.43 |
| R134a | −15/−15 | 3,000 | 0.12 | 0.457 | 9.21 | 3.45 | |
| Heating | R454C | −7/−7 | 3,500 | 0.297 | 2.762 | 56.79 | 2.60 |
| R134a | −7/−7 | 8,000 | 0.116 | 1.257 | 48.10 | 1.71 |
At 43 °C and 40 °C cabin-only cooling, the COP of R454C was 1.98 and 1.73, respectively 20.0% and 13.8% higher than the 1.65 and 1.52 of R134a. If the comparison is converted at the same cooling load and auxiliary power is neglected, the compressor power could be reduced by about 16.7% and 12.1%. R454C required a lower compressor speed to reach a similar outlet-air temperature, consistent with its higher suction density and volumetric capacity, but this advantage will change with the compressor efficiency map and heat-exchanger matching.
The low-temperature heating comparison reflects the operating boundary of the bench rather than an absolute refrigerant limit. At −15 °C, the suction pressure of R134a at 3,000 r/min had already dropped to 0.12 MPa; limited by the protection boundary, the outlet-air temperature was only 9.21 °C, whereas R454C at 4,800 r/min delivered 58.15 °C. At −7 °C the COP of R454C was 2.60, higher than the 1.71 of R134a. Combined with the temperature sweep, R454C extended the no-vapor-injection operating boundary of this bench to −20 °C. However, R454C is an A2L fluid with a higher high-side pressure than R134a, and it is not a drop-in replacement requiring no modification; engineering application must re-check pressure rating, compressor displacement, heat-exchanger circuiting, charge, leak detection, ventilation and electrical protection.
Design and Control Implications for Electric Car Heat Pumps
The experimental evidence supports a shift in optimization philosophy for electric car heat-pump systems using zeotropic low-GWP blends. Instead of single-parameter optimization, the charge, superheat, heat-transfer capacity and system heat-transfer path should be matched in a coordinated way. Table 11 summarizes the practical implications that I derive from the experiments.
| Design or control item | Observed behavior | Recommended approach |
|---|---|---|
| Refrigerant charge | Different plateau intervals among modes | Use the overlap of mode plateaus plus a pressure safety margin; 1,025 g on this bench |
| Direct-expansion superheat | COP weakly sensitive, uniformity strongly sensitive | Use 5–8 °C as the initial window and correct for load and liquid-slugging risk |
| Frontal air velocity | Large gain but diminishing returns | Prioritize frontal heat rejection at high ambient; account for fan power |
| Secondary-loop capacity share | COP decreases as capacity shifts to the coolant loop | Minimize unnecessary transfer when the battery load allows |
| Low-temperature heating | Effective down to −20 °C without vapor injection | Retain a vapor-injection option for ambient below −20 °C |
| Safety and pressure rating | A2L fluid, higher high-side pressure than R134a | Re-check pressure rating, leak detection, ventilation and electrical protection |
The compressor speed and capacity relationship can be expressed through the volumetric capacity:
$$Q_c = \eta_v V_d n \rho_{suc} \Delta h_{evap}$$
where \(\eta_v\) is the volumetric efficiency, \(V_d\) is the compressor displacement, \(n\) is the rotational speed, \(\rho_{suc}\) is the suction density and \(\Delta h_{evap}\) is the specific refrigerating effect. The higher suction density of R454C explains why a lower speed can deliver comparable capacity in an electric car, which is also consistent with the measured speed differences in Table 10.
The energy-saving ratio relative to R134a can be expressed as:
$$\varepsilon_{COP} = \frac{COP_{R454C} – COP_{R134a}}{COP_{R134a}}$$
and the corresponding compressor-power reduction at equal cooling load is:
$$\varepsilon_{W} = 1 – \frac{COP_{R134a}}{COP_{R454C}}$$
Applying these relations to the measured values at 43 °C and 40 °C yields the 20.0% and 13.8% COP improvements and the 16.7% and 12.1% power reductions reported above.
Conclusions
I carried out steady-state multimode performance and outlet-air uniformity experiments on an R454C electric car heat-pump bench that combines direct cabin evaporation with a coolant secondary loop, and compared R454C with R134a on the same bench. The following conclusions can be drawn.
First, the reasonable charge interval differs among cabin-only cooling, single-chiller, combined cooling and heating modes. Combining the performance plateaus of all modes gave a recommended charge of 1,025 g for this bench. The local non-monotonic variation in the overcharge region results from competition among condenser liquid fraction, inventory migration between modes, valve adjustment and changes in effective heat-exchanger area.
Second, high-ambient cooling performance is sensitive to frontal heat-rejection conditions. At 43 °C, increasing the frontal air velocity significantly raised both cooling capacity and COP, with a maximum cooling capacity of 7.77 kW and a maximum COP of 2.16, but the gain diminished at high velocity, indicating clear diminishing marginal returns of frontal heat-transfer enhancement. In contrast, over a superheat range of 5–12 °C the COP changed only slightly. At −20 °C the heating outlet-air temperature exceeded 40 °C with a COP close to 2.5, while at −25 °C the outlet-air temperature fell to 22.4 °C.
Third, the temperature glide of the zeotropic refrigerant strongly affects air-side temperature uniformity in the direct-expansion evaporator. When the superheat was increased from 5 °C to 12 °C, the outlet-air temperature difference increased from about 3.1–3.7 °C to about 3.8–5.0 °C. The underlying mechanism is the combined effect of temperature glide, expansion of the local dryout and superheated region and maldistribution of refrigerant and air flow. Superheat control in a direct-expansion system must therefore consider COP, effective two-phase heat-transfer area and outlet-air temperature uniformity simultaneously.
Fourth, under identical hardware and given control boundaries, the COP of R454C at 40–43 °C cooling conditions was 13.8%–20.0% higher than that of R134a, and the system maintained effective low-temperature operation down to −20 °C without vapor injection. Combined with the theoretical analysis, the higher suction density and volumetric capacity of R454C favor a reduction in compressor displacement or speed for a given capacity in an electric car. The multimode charge matching, direct-expansion temperature-uniformity control and system-level component coordination reported here can serve as a reference for the design and control of other low-GWP zeotropic refrigerant heat-pump systems for electric car applications.
For future work, the steady-state component-level findings should be extended to vehicle-level dynamic load, battery thermal inertia and road-condition disturbances, and the fan and pump power should be included in a net energy analysis so that the real energy-saving potential of R454C in an electric car can be quantified under realistic operating profiles.