The rapid integration of renewable energy sources such as wind and solar power into the electrical grid has introduced significant challenges related to grid stability and power quality. Energy storage technology, particularly electrochemical energy storage, has emerged as a critical solution to these challenges. Among the various storage technologies, lithium-ion battery systems are widely adopted due to their high energy density, reliability, and flexibility. However, the thermal management of large-scale battery packs remains a key technical barrier. Heat accumulation in battery cells accelerates aging, reduces cycle life, and, in extreme cases, leads to thermal runaway and catastrophic failures. Therefore, designing an efficient and low-energy cooling system for energy storage battery applications is of paramount importance.
Conventional cooling methods for energy storage battery systems include air cooling and liquid cooling. Air cooling offers simplicity and low cost, but its heat dissipation capability and temperature uniformity are inadequate for high-power battery systems. Liquid cooling, especially indirect liquid cooling using cold plates, provides superior thermal performance and is widely adopted in electric vehicles and small-scale storage stations. Nevertheless, the need for complex coolant channel designs, high coolant consumption, and the additional energy consumption of chiller units make these systems less effective from an energy-saving perspective. Direct immersion cooling, although excellent in temperature uniformity, requires expensive dielectric fluids and rigorous pressure control, limiting its scalability.
To overcome these limitations, this study proposes a novel cooling architecture that combines falling film evaporation with mechanical vapor recompression (MV R). The system utilizes water as the working fluid, leveraging its high latent heat and low cost. The evaporation process occurs in a vacuum falling-film chamber, where the coolant absorbs heat from the battery cells and phase-changes into vapor. The generated vapor is then compressed by a mechanical compressor to a higher pressure and temperature, enabling heat rejection to the ambient through a condenser. This approach eliminates the need for a separate refrigeration cycle, significantly reducing the overall power consumption.
The falling-film indirect liquid-cooling structure simplifies the design by removing the complex internal channels of traditional cold plates and substantially reduces the coolant inventory compared with immersion cooling. Importantly, the battery modules operate at atmospheric pressure while only the falling-film chamber is evacuated, ensuring safety and reliability. The MV R technology further enhances the temperature difference between the system and the cooling utility, thus achieving efficient heat dissipation. The objective of this paper is to present a detailed process design, simulate the system performance using Aspen Plus, compare it with a conventional chiller-based refrigeration system, and analyze the influence of key thermodynamic parameters.
System Architecture and Operating Principle
The complete MV R falling-film indirect liquid-cooling system consists of two main modules: the falling-film indirect liquid-cooling module and the MV R module. The falling-film module includes the battery module, a liquid distributor, and a vacuum falling-film chamber formed by heat-conducting plates. The MV R module comprises a steam compressor, a vapor-liquid separator, a condenser, a throttle valve, valves, and a circulation pump.
The operating sequence begins with evacuation of the entire system. After a vacuum condition is established, deionized water is introduced until the required inventory is reached. The compressor and circulation pump are then started. Water from the vapor-liquid separator and the condensed water from the condenser are pumped to the liquid distributor at the top of the falling-film chamber. The water is uniformly distributed over the falling-film plates, forming a thin liquid film. The film absorbs the heat generated by the battery cells through the solid wall and evaporates under the low pressure prevailing in the chamber. Since the pressure in the vapor-liquid separator is maintained below the pressure of the falling-film chamber, the generated vapor is drawn from the bottom of the chamber into the separator. In the separator, any entrained liquid droplets are removed; the liquid water returns to the distributor, while the vapor proceeds to the compressor.
In the MV R module, the vapor is compressed to a higher pressure and temperature. The superheated vapor then enters the condenser, where it transfers heat to an external cooling medium and condenses into liquid water. The condensed water passes through a throttle valve, reducing its pressure to the saturation pressure corresponding to the evaporation temperature in the falling-film chamber. The low-pressure cooling water is then returned to the falling-film module for another cycle. This closed-loop arrangement minimizes water consumption and ensures continuous operation.
The falling-film plate structure is illustrated conceptually: the battery cell is attached to a metallic plate (stainless steel, aluminum, or aluminum alloy) with thermal grease applied at the interface. The plate is shaped to form a closed vacuum chamber. A liquid distributor at the top ensures even wetting of the plate surfaces on both sides of the chamber. Because the coolant never directly contacts the battery cells, electrical safety is preserved, and the battery operates in a normal atmospheric environment.
Thermodynamic Modeling and Simulation
To evaluate the system performance, a steady-state simulation was carried out using Aspen Plus software. The working fluid was defined as water, and the IAPWS-95 property method was employed to calculate the thermodynamic properties of water and steam. The following simplifying assumptions were made: negligible heat losses in pipelines, neglecting the effect of non-condensable gases on heat transfer, and steady-state operation. The falling-film evaporation module and the vapor-liquid separator were lumped into a single evaporation block (EVAP) in the simulation. The compressor was modeled as a block (COMP), the condenser as COND, and the throttle valve as VALVE.
The compressor power consumption is calculated from the theoretical enthalpy rise under isentropic compression divided by the compressor efficiency:
$$W = \frac{m_v \Delta H_{id,v}}{\eta_c}$$
where \(W\) is the compressor power consumption in kW, \(m_v\) is the vapor mass flow rate in \(\text{kg·s}^{-1}\), \(\Delta H_{id,v}\) is the isentropic enthalpy rise of the vapor compression in \(\text{kJ·kg}^{-1}\), and \(\eta_c\) is the compressor efficiency.
Similarly, the heat absorbed in the evaporator (cooling load) is given by:
$$Q = m_w \Delta H_{w,e}$$
where \(Q\) is the evaporator heat transfer rate or cooling capacity in kW, \(m_w\) is the water mass flow rate in \(\text{kg·s}^{-1}\), and \(\Delta H_{w,e}\) is the enthalpy change of water during evaporation in \(\text{kJ·kg}^{-1}\).
In the present model, it was assumed that the entire water feed evaporates completely and that the vapor mass flow rate equals the liquid feed rate:
$$m_v = m_w$$
The initial design conditions are summarized in Table 1. The cooling load is set to 210 kW, the cooling temperature (evaporation temperature) is 303 K, the corresponding operating pressure is 4247 Pa, the falling-film water flow rate is 311 \(\text{kg·h}^{-1}\), and the compression temperature rise is 15 K. The compression temperature rise is defined as the difference between the discharge temperature and the suction temperature of the compressor.
| Parameter | Value |
|---|---|
| Cooling load (kW) | 210 |
| Cooling temperature (K) | 303 |
| Operating pressure (Pa) | 4247 |
| Falling-film water flow rate (kg·h−1) | 311 |
| Compression temperature rise (K) | 15 |
The simulation results under the design conditions are presented in Figure 1 of the original publication. From the simulated state points, the net cooling load is calculated to be 204.492 kW, considering a slight heat loss. The corresponding compressor power consumption is only 14.227 kW. This remarkably low power consumption is a direct consequence of the MV R principle, which reuses the latent heat of vaporization with a modest mechanical energy input.
Comparative Analysis with a Conventional Refrigeration System
To benchmark the proposed MV R falling-film indirect liquid-cooling system, a conventional room-level chilled water air conditioning product was selected as the reference case. The reference unit has a total cooling capacity of 210 kW, a sensible cooling capacity of 210 kW, an air flow rate of 53,000 \(\text{m}^3·\text{h}^{-1}\), and a specified coefficient of performance (COP) of 3.3 for the first-level energy efficiency refrigeration system. The comparison was performed at the same cooling load of 210 kW, and the results are listed in Table 2.
| Cooling temperature (K) | Latent heat (kJ·kg−1) | Cooling load (kW) | MV R compressor power (kW) | First-level refrigeration power (kW) | Energy savings (%) |
|---|---|---|---|---|---|
| 303 | 2429.67 | 204.492 | 14.227 | 61.967 | 77.04 |
The conventional system consumes 61.967 kW of electrical power when operating at first-level energy efficiency (COP = 3.3). In contrast, the MV R system requires only 14.227 kW for the compressor, resulting in energy savings of 77.04%. This substantial reduction in power consumption is attributed to the fact that the MV R system performs only a single vapor compression of water vapor, whereas the conventional refrigeration system operates with an extra refrigerant loop. The evaporative cooling water in the MV R system directly removes heat from the battery and rejects it through a condenser; no additional refrigerant cycle is required, thus eliminating the associated compression work.
Influence of Key Thermodynamic Parameters
The performance of the MV R falling-film indirect liquid-cooling system is strongly affected by two operational variables: the cooling temperature (i.e., the evaporation temperature of the falling film) and the compression temperature rise. Figure 5 in the original paper illustrates the relationship between these parameters and the compressor power consumption. Our simulation results follow the same trends as those reported in previous studies.
For a fixed compression temperature rise of 12 K, when the cooling temperature increases from 283 K to 313 K, the compressor power decreases by 1.311 kW, corresponding to a 10.7% reduction. This is because a higher cooling temperature means that the vapor entering the compressor has a higher suction temperature. For a given temperature rise, the pressure ratio decreases with increasing suction temperature, leading to lower compression work. Additionally, as shown in Figure 6 of the original publication, the vapor density increases with cooling temperature, which reduces the volumetric flow rate entering the compressor. Consequently, the compressor work decreases. Therefore, within the constraints of the battery temperature requirements, selecting a higher cooling temperature helps minimize the compressor load.
On the other hand, the compressor power increases monotonically with the compression temperature rise. At a cooling temperature of 303 K, raising the compression temperature rise from 10 K to 15 K increases the compressor power by 4.883 kW, a 52.3% increase. This is because a larger temperature rise implies a larger pressure ratio, which directly increases the specific compression work. The compression temperature rise should be selected carefully to balance the required heat rejection temperature and energy consumption.
Figure 7 in the original paper shows the combined effect of these parameters on the overall energy saving relative to the conventional system. Higher cooling temperatures improve the energy saving percentage. For example, at a compression temperature rise of 12 K, the energy saving increases from 80.4% to 82.5% when the cooling temperature rises from 283 K to 313 K. Conversely, increasing the compression temperature rise reduces the energy saving. At 303 K cooling temperature, an increase from 10 K to 15 K causes the energy saving to decline from 83.9% to 77.0%. These trends emphasize that the system should be operated at the highest feasible cooling temperature and the lowest required compression temperature rise to maximize efficiency.
Economic Analysis
Beyond the technical performance, the economic viability of the MV R falling-film indirect liquid-cooling system was assessed using two metrics: the annualized investment cost (AC) and the payback period (PBP). The annualized cost approach distributes the initial capital investment over the system lifetime and includes operational and maintenance costs. The payback period measures the time required to recover the initial investment through annual savings.
The annualized investment cost is expressed as:
$$AC = C_{acc} + C_{mc} – S_a + C_{rec}$$
where \(C_{mc}\) is the annual maintenance cost, taken as 10% of the annualized capital cost \(C_{acc}\). The annualized capital cost is calculated from the initial total capital cost \(C_{ccm}\) using the capital recovery factor:
$$C_{acc} = C_{ccm} \times \frac{i(1+i)^n}{(1+i)^n – 1}$$
where \(i\) is the interest rate (assumed 10%) and \(n\) is the system lifetime (assumed 15 years). The annual residual value \(S_a\) is computed as:
$$S_a = S \times \frac{i}{(1+i)^n – 1}$$
where \(S\) is the residual value, taken as 20% of \(C_{ccm}\). The annual electricity cost \(C_{rec}\) is:
$$C_{rec} = C_{ee} \times W \times t$$
where \(C_{ee}\) is the average electricity price (0.8 yuan/kWh), \(W\) is the system power consumption in kW, and \(t\) is the annual operating time (3000 h).
The payback period is calculated using the formula:
$$PBP = \frac{ \ln \left[ 1 – \frac{C_{ccm}}{B_j}(i-r) \right] }{ \ln \left( \frac{1+r}{1+i} \right) }$$
where \(B_j\) is the savings in the first year and \(r\) is the discount rate, set to 6% according to bank interest rates.
Table 3 lists the initial investment costs for the major components of the MV R system. The total initial capital cost is estimated at 17,200 yuan, which includes the steam compressor, condenser, vapor-liquid separator, throttle valve, circulation pump, and associated valves. This cost is lower than that of the conventional system, which was estimated at 24,200 yuan, because the MV R system eliminates one heat exchanger associated with the separate refrigerant loop. A detailed economic comparison between the two systems is presented in Table 4.
| Component | Quantity | Price (yuan) |
|---|---|---|
| Steam compressor | 1 | 12,000 |
| Condenser | 1 | 3,000 |
| Vapor-liquid separator | 1 | 1,000 |
| Throttle valve | 1 | 400 |
| Circulation pump | 1 | 500 |
| Valve | 1 | 300 |
| Total | 17,200 |
| Item | MV R system | Reference system |
|---|---|---|
| Initial capital cost \(C_{ccm}\) (yuan) | 17,200 | 24,200 |
| Annualized capital cost \(C_{acc}\) (yuan) | 2,262 | 3,182 |
| Annual maintenance cost \(C_{mc}\) (yuan) | 227 | 319 |
| Annual electricity cost \(C_{rec}\) (yuan) | 34,145 | 148,721 |
| Annualized investment cost \(AC\) (yuan) | 36,525 | 152,069 |
| Payback period \(PBP\) (year) | 0.163 | – |
The results show that the MV R system saves 114,576 yuan annually in electricity costs compared with the conventional system. The annualized capital cost and maintenance cost are also slightly lower, resulting in an annualized cost savings of 115,544 yuan. The computed payback period is only 0.163 years, meaning that the extra investment (which is actually negative, since the MV R system has a lower initial cost) is recovered within approximately two months. This outstanding economic performance is a direct consequence of the dramatically reduced compressor power requirement.
Conclusion and Future Perspectives
In this study, a novel MV R falling-film evaporative indirect liquid-cooling system for energy storage battery thermal management was designed and analyzed. The system employs water as the working fluid, and the falling-film evaporation mechanism combined with mechanical vapor recompression enables efficient heat removal with remarkably low energy consumption. The following conclusions can be drawn from the simulation and economic analysis.
First, at a cooling temperature of 303 K and a cooling load of 210 kW, the MV R compressor power is only 14.227 kW. Compared with a conventional first-level energy efficiency refrigeration system with a COP of 3.3, the proposed system achieves an energy saving of 77.04%. The elimination of a separate refrigerant cycle is the main reason for this substantial improvement.
Second, the system performance is highly sensitive to the cooling temperature and compression temperature rise. Raising the cooling temperature reduces the compressor pressure ratio and volumetric flow, thereby reducing power consumption. Lowering the compression temperature rise also reduces the pressure ratio and hence the compressor work. To maximize efficiency, the system should be operated at the highest allowable cooling temperature and the lowest permissible compression temperature rise that satisfies the heat rejection requirements.
Third, from the economic perspective, the MV R falling-film indirect liquid-cooling system has a lower initial capital cost than the conventional system and significantly reduces annual electricity expenses. The annualized cost is 36,525 yuan versus 152,069 yuan for the reference system, and the payback period is as short as 0.163 years (about two months). These results demonstrate the strong economic attractiveness of the proposed technology.
Future work should focus on the detailed thermal and flow characteristics of the falling-film evaporation module, including temperature uniformity across the battery pack and the underlying heat and mass transfer mechanisms. Experimental validation and optimization of the falling-film distributor will also be essential to translate the theoretical advantages into practical, reliable cooling systems for large-scale energy storage battery installations.