Study on Surface Condensate Water Removal and Heat Transfer Performance of a Minichannel Heat Exchanger

: The condensate on the surface of the minichannel heat exchanger generated during air cooling substantially reduces the heat transfer performance as it works as an evaporator in the air-conditioning system. This has received much attention in scientiﬁc communities. In this paper, the e ﬀ ect of operating parameters on the heat transfer performance of a minichannel heat exchanger (MHE) is investigated under an evaporator working condition. An experimental MHE test system is developed for this purpose, and extensive experimental studies are conducted under a wide range of working conditions using the water-cooling method. The inlet air temperature shows a large e ﬀ ect on the overall heat transfer coe ﬃ cient, while the inlet air relative humidity shows a large e ﬀ ect on the condensate aggregation rate. The airside heat transfer coe ﬃ cient increases from 66 to 81 W / (m 2 · K) when the inlet air temperature increases from 30 to 35 ◦ C. While the condensate aggregation rate on the MHE surface increases by up to 1.8 times when the relative humidity increases from 50% to 70%. The optimal air velocity, 2.5 m / s, is identiﬁed in terms of the heat transfer rate and airside heat transfer coe ﬃ cient of the MHE. It is also found that the heat transfer rate and overall heat transfer coe ﬃ cient increase as the air velocity increases from 1.5 to 2.5 m / s and decreases above 2.5 m / s. Furthermore, a large amount of condensate accumulates on the MHE surface lowering the MHE heat transfer. The inclined installation angle of the MHE in the wind tunnel e ﬀ ectively enhances heat transfer performance on the MHE surface. The experimental results provide useful information for reducing condensate accumulation and enhancing microchannel heat transfer.


Introduction
The minichannel heat exchanger (MHE) is used as a novel condenser due to its advantages, such as high efficiency, compact structure, and low material cost [1,2]. It can be widely applied in automotive air conditioning [3,4] and thermoelectrics [5,6]. However, its shutter structure leads to problems such as the poor ability to discharge the condensate water and quick frosting, which limits its application, particularly in commercial refrigeration and air conditioning systems. Recently, MHEs have attracted more and more attention in the context of building air-conditioning and associated heat exchange elements.
The MHE concept was first proposed by Tuckerman and Pease [7] in the 1980s and, subsequently, widely used in large-scale integrated circuits. Swift et al. [8] developed an MHE that was used in common components (such as the photomultiplier tube). Wajs et al. [9] presented a heat exchanger with mini-jet and cylindrical construction. Li Hui et al. [10] utilized an MHE as a condenser in fully wet operating conditions. Therefore, it is necessary to explore the behavior and performance of systems operating under wet conditions and develop more efficient thermal systems. Saisai et al. [36] analyzed heat transfer characteristics and flow resistance characteristics. The results indicated that the influence of face velocity on heat transfer was significantly greater than that of water flow. Guojun et al. [37] studied the heat transfer and humidity characteristics on the surface of an MHE. The results showed that an increase in the installation angle enhanced removal of the surface condensate water and the installation angle positively influenced the heat transfer coefficient. Further, Lu et al. [38] pointed out that a non-uniform distribution of both the refrigerant and airflow leads to different degrees of performance attenuation of the heat exchanger. Zhang et al. [39] found that the application of an inclined fin layout (downwind/upwind) could improve the fin efficiency. Wei et al. [40] pointed out that under the sublimation condition, the pressure drop on the airside changed negligibly with time, and condensation first appeared on the leeward side. Mohammed et al. [41] found that the air RH, temperatures, and condensation water quantity significantly affected the heat transfer performance. Yin et al.'s [42] experimental results demonstrated that frost thickness increased in a parabolic manner with time and decreased with an increase in air velocity. Furthermore, the maximum heat transfer increased with an increase in air velocity.
From the above research, researchers have put a lot of effort into improving the performance of a minichannel heat exchanger as an evaporator. However, most researchers only focused on one factor and analyzed the influence of this factor on condensate removal and heat transfer performance of an MHE. The practical application is often simultaneously influenced by many factors. In this paper, the effect of influencing factors on the MHE performance was investigated. A certain range of parameters to determine a means of improving condensate removal, enhancing the heat transfer effect, and bettering the performance of an MHE was identified based on the experimental study. The specific inlet parameter values of air velocity and heat exchanger installation angle were quantitatively and experimentally studied to evaluate their influence on the MHE performance. Furthermore, the characteristics of the airside heat transfer, condensed water accumulation, and condensate removal from the MHE surface were predicted under the condensation condition using dynamic dip testing to assess condensate drainage behavior from the air-side surface of an MHE. This method provides highly repeatable data for real-time drainage. The results of the study provide useful information for engineers and researchers in the design and practical application of MHEs as evaporators. Figure 1 shows a schematic drawing of the MHE test rig, which consists of a wind tunnel, an MHE, condensate collection plate, chilled water supply system and a set of sensors and piping. The MHE is allocated in the wind tunnel, and the environmental state of the heat exchanger is simulated using an environmental test chamber that can provide all the required test conditions. Chilled water is used as the cold source for the MHE. The chilled water flows through the MHE. The chilled water inlet temperature is 5 • C. It cools the heat exchanger tube and removes heat via convective heat transfer from the windward-side air. An air-cooled chiller (model RO-06a) of a Chinese brand (Ri Ou) is used to provide cooling water. The rated cooling capacity is 16.9 kW. A mass flowmeter (LWGY-15) is installed in the straight water supply pipeline to measure the water flow rate. The armored thermocouple is installed on the water supply and return pipelines to measure the water supply and return temperatures. The water supply and return pipes are equipped with gate valves, which can adjust the water flow. The air-cooled chiller supplies the chilled water to the microchannel heat exchanger in the water-side for cooling air. A condensate collection plate is placed at the bottom of the heat exchanger to collect condensate water dripping into the bucket through the drainage pipe. The bucket is placed on a weighting meter (electronic scale), which records the weight of the condensate water. This weight is used to calculate the discharge rate of the condensate. The bucket is placed on a weighting meter (electronic scale), which records the weight of the condensate water. This weight is used to calculate the discharge rate of the condensate.  The size of the MHE used for our experimental study is 500 × 490 × 32 mm, and the minichannel diameter is 0.8 mm. Figure 2 shows the MHE structural diagram. The heat transfer area is 1.38 m 2 on the chilled water side and 8.33 m 2 on the airside. The inner diameter of the inlet and outlet copper pipes is 14 mm, the outer diameter of the header is 32 mm, and the length of the flat tube is 436 mm. The double-flow form of louver fins is adopted, wherein each flow section contained 24 rows of flat pipes with the header being positioned vertically. The chilled-water inlet and outlet are positioned on the same side of the header, and the connection mode of chilled-water is downward supply and upward return. The water mass flow rate is 1.38 kg/s. The two headers are welded with a movable bracket to allow for adjustment of the inclination of the heat exchanger. The measurement instrument used in the experiments mainly includes temperature and humidity sensors, T-type thermocouples, K-type thermocouples, micro-pressure differential meters, anemometers, liquid-turbine flow meter and electronic weighing meter. Table 1 lists the specifications of the used measurement instruments. In the experiments, the MHE test system is placed in a small room, and air from an outdoor air environment chamber is introduced into the wind tunnel using a variable-frequency fan. The air The size of the MHE used for our experimental study is 500 × 490 × 32 mm, and the minichannel diameter is 0.8 mm. Figure 2 shows the MHE structural diagram. The heat transfer area is 1.38 m 2 on the chilled water side and 8.33 m 2 on the airside. The inner diameter of the inlet and outlet copper pipes is 14 mm, the outer diameter of the header is 32 mm, and the length of the flat tube is 436 mm. The double-flow form of louver fins is adopted, wherein each flow section contained 24 rows of flat pipes with the header being positioned vertically. The chilled-water inlet and outlet are positioned on the same side of the header, and the connection mode of chilled-water is downward supply and upward return. The water mass flow rate is 1.38 kg/s. The two headers are welded with a movable bracket to allow for adjustment of the inclination of the heat exchanger. The measurement instrument used in the experiments mainly includes temperature and humidity sensors, T-type thermocouples, K-type thermocouples, micro-pressure differential meters, anemometers, liquid-turbine flow meter and electronic weighing meter. Table 1 lists the specifications of the used measurement instruments. velocity is measured using an anemometer (SYSTEM-6242). At the exit of the wind tunnel, the anemometer measures nine points, which are equally allocated on the cross-section of the wind tunnel. Then, the average velocity of the nine measurement points is considered as the studied air velocity. The temperature and humidity sensors are used to monitor the air temperatures and humidity before and after the MHE. The heat transfer capacity, heat transfer coefficient, and condensate removal rate are calculated to analyze the heat transfer characteristics and surface condensate drainage characteristics. The inlet air temperature in the experiment varies from 30 to 35 °C, the air velocity ranges from 1.5 to 3.0 m/s (i.e., air mass flow rate varies from 12.5 to 24.9 kg/s), and the installation angle varies from 0 to 20°.

Calculations
The main performance parameters of the MHE are heat transfer, overall heat transfer coefficient, and condensation water removal rate. The heat transfer is calculated based on Equation (1). The air enthalpy is calculated by adapting the NIST(National Institute of Standards and Technology) Refprop [43] simulation tool and verified by a psychrometric chart, as shown in Figure 3. For a given air  In the experiments, the MHE test system is placed in a small room, and air from an outdoor air environment chamber is introduced into the wind tunnel using a variable-frequency fan. The air velocity is measured using an anemometer (SYSTEM-6242). At the exit of the wind tunnel, the anemometer measures nine points, which are equally allocated on the cross-section of the wind tunnel. Then, the average velocity of the nine measurement points is considered as the studied air velocity. The temperature and humidity sensors are used to monitor the air temperatures and humidity before and after the MHE. The heat transfer capacity, heat transfer coefficient, and condensate removal rate are calculated to analyze the heat transfer characteristics and surface condensate drainage characteristics. The inlet air temperature in the experiment varies from 30 to 35 • C, the air velocity ranges from 1.5 to 3.0 m/s (i.e., air mass flow rate varies from 12.5 to 24.9 kg/s), and the installation angle varies from 0 to 20 • .

Calculations
The main performance parameters of the MHE are heat transfer, overall heat transfer coefficient, and condensation water removal rate. The heat transfer is calculated based on Equation (1). The air enthalpy is calculated by adapting the NIST(National Institute of Standards and Technology) Refprop [43] simulation tool and verified by a psychrometric chart, as shown in Figure 3. For a given air temperature and RH, the enthalpy and absolute moisture content can be obtained from the NIST Refprop and verified by the psychrometric chart, and can then be used to calculate the heat transfer and condensation water production rate, respectively. The heat transfer on the air-side of the MHE is calculated by the following equation: where m  is dry air mass flow rate., kg/s, h1, and h2 are air specific enthalpy at inlet and outlet, The heat transfer on the air-side of the MHE is calculated by the following equation: where . m a is dry air mass flow rate., kg/s, h 1, and h 2 are air specific enthalpy at inlet and outlet, respectively, kJ/kg. h c is a specific enthalpy of the condensate.
. m w is the condensate mass flow rate and is calculated by: where, w 1 and w 2 are moisture content of the air at the inlet and outlet of the wind tunnel, kg water/kg of dry air.
where m is wet air mass flow rate, kg/s, . m a is dry air mass flow rate, kg/s, p is air density, kg/m 3 . U is air average velocity, m/s, A c is the wind cross area, m 2 .
The heat transfer on the water-side of the MHE is calculated by the following equation: where C ρω is the specific heat capacity of water, m w is the mass flow rate of water, ∆T m is the temperature difference between the water inlet and outlet and is calculated by: According to the heat balance method, we can know that: The overall heat transfer coefficient of the MHE is calculated using the following equation: where, K is the overall heat transfer coefficient; A is the air side surface area of the heat transfer, m 2 ; ∆T m is the logarithmic mean temperature difference of heat transfer in the heat exchanger, • C.
where ∆T 1 -the temperature difference between the hot fluid inlet and the cold fluid outlet. ∆T 2 -the temperature difference between the hot fluid outlet and the cold fluid inlet. ln-the natural logarithm. The Reynolds number on the air side is calculated using the following equation: where p is the density of the air, v is air velocity across minimum wind cross area, m/s, µ is the dynamic viscosity of air, L p is the louver pitch, m. The condensation removal rate on the MHE surface is calculated by: . m w1 = ∆m ∆t (11) where ∆t is the period selected when calculating the average condensate removal rate at a certain moment, s; ∆m is the condensate weight change, kg. The surface condensate accumulation rate . m w2 on the MHE was calculated using the following equation: The experimental measurement error can be calculated based on the sensor accuracy using error propagation. According to the literature [44], the experimental result R is assumed to be calculated from a set of independent variables, which can be represented by: The uncertainty of the measurement of a single variable for the experimental results can be given by: The experimental uncertainty of the final result R can be determined by combining uncertainties of individual terms by a root-sun-square method: Based on the accuracies of the measurement devices shown in Table 1, the uncertainty of the heat transfer coefficient and water removal rate can be calculated via Equations (13)- (15). Based on the experimental data, the uncertainties of the air-side heat transfer rate, water-side heat transfer rate, water removal rate, overall heat transfer coefficient, mass flow rate of air, mass flow rate of water, enthalpy of air and moisture content are 4.5%, 3.5%, 4.5%, 5.5%, 2.2%, 0.7%, 3.1% and 2.9%, respectively.

Results and discussion
In order to minimize the measurement error in the air-side measurement, the heat balance between the water and air sides of the heat exchanger is calculated and checked. Figure 4 shows the comparison of the heat transfer rate between the water and air sides. The results showed that the discrepancy between the water and air-side heat transfer rate is less than 9.3% over the studied conditions. Energies 2020, 13, 1065 8 of 19 Based on the accuracies of the measurement devices shown in Table 1, the uncertainty of the heat transfer coefficient and water removal rate can be calculated via Equations (13)- (15). Based on the experimental data, the uncertainties of the air-side heat transfer rate, water-side heat transfer rate, water removal rate, overall heat transfer coefficient, mass flow rate of air, mass flow rate of water, enthalpy of air and moisture content are 4.5%, 3.5%, 4.5%, 5.5%, 2.2%, 0.7%, 3.1% and 2.9%, respectively.

Results and discussion
In order to minimize the measurement error in the air-side measurement, the heat balance between the water and air sides of the heat exchanger is calculated and checked. Figure 4 shows the comparison of the heat transfer rate between the water and air sides. The results showed that the discrepancy between the water and air-side heat transfer rate is less than 9.3% over the studied conditions.   Figure 5 shows the variation of the air outlet temperature under different working conditions. The inlet air temperature varies from 30 to 35 °C and the RH changes from 50% to 70%. It was found that the air outlet temperature decreased rapidly in the first 10 min as the thermal energy was removed by the chilled water. Then the air outlet temperature dropped slowly and reached a stable temperature after 25 minutes. Although the relative humidity and air inlet temperature have a bit of an effect on the time to reach the stable condition, as shown in the figure, the heat transfer process could reach a stable condition after 25 minutes in all studied cases.  Figure 5 shows the variation of the air outlet temperature under different working conditions. The inlet air temperature varies from 30 to 35 • C and the RH changes from 50% to 70%. It was found that the air outlet temperature decreased rapidly in the first 10 min as the thermal energy was removed by the chilled water. Then the air outlet temperature dropped slowly and reached a stable temperature after 25 min. Although the relative humidity and air inlet temperature have a bit of an effect on the time to reach the stable condition, as shown in the figure, the heat transfer process could reach a stable condition after 25 min in all studied cases.  Energies 2020, 13, 1065 9 of 17 Figure 6a shows the variation of the overall heat transfer coefficient of the MHE under different working conditions. It was found that the overall heat transfer coefficient increased with air inlet temperature. This is caused by the increase of temperature difference between the air inlet and cold water, which substantially enhances the air-side heat exchange in the forced convective heat transfer. Therefore, the overall heat transfer coefficient changes with the air inlet temperature. As the inlet air temperature increased from 30 to 35 • C, the overall heat transfer coefficient increased from 72.5 to 82.5 W/m 2 ·K at an RH of 60%. However, the maximum overall heat transfer coefficient was found at the RH of 60% for each temperature. These changes in the overall heat transfer coefficient could be explained from the air-side heat transfer. As shown in Figure 6b, the air-side heat transfer rate of the MHE increased with the air inlet temperature. However, it was observed that the air-side heat transfer reached the maximum value at RH of 60% for all studied temperatures. This might be due to the fact that the increasing of RH causes the increasing of latent heat exchange and the total heat transfer, but at the same time, the condensate accumulation on the MHE surface, as shown in Figure 7, which results in the decrease of heat transfer. The combined effect results in the optimal heat coefficient occurring at RH of 60%.  Figure 6a shows the variation of the overall heat transfer coefficient of the MHE under different working conditions. It was found that the overall heat transfer coefficient increased with air inlet temperature. This is caused by the increase of temperature difference between the air inlet and cold water, which substantially enhances the air-side heat exchange in the forced convective heat transfer. Therefore, the overall heat transfer coefficient changes with the air inlet temperature. As the inlet air temperature increased from 30 to 35 °C, the overall heat transfer coefficient increased from 72.5 to 82.5 W/m 2 ·K at an RH of 60%. However, the maximum overall heat transfer coefficient was found at the RH of 60% for each temperature. These changes in the overall heat transfer coefficient could be explained from the air-side heat transfer. As shown in Figure 6b, the air-side heat transfer rate of the MHE increased with the air inlet temperature. However, it was observed that the air-side heat transfer reached the maximum value at RH of 60% for all studied temperatures. This might be due to the fact that the increasing of RH causes the increasing of latent heat exchange and the total heat transfer, but at the same time, the condensate accumulation on the MHE surface, as shown in Figure  7, which results in the decrease of heat transfer. The combined effect results in the optimal heat coefficient occurring at RH of 60%.    Figure 7a shows the condensate aggregation rate under different working conditions. As the RH increased from 50% to 70%, the condensate aggregation rate increased by up to 1.8 times. This was because more moisture was contained in the air with high RH at a given temperature. When the air was cooled, more moisture was condensed from the air with high HR. It was also found that the condensate aggregation rate increased with the temperature. This could also be explained by the moisture contained in the air. At the high temperature, more moisture was contained in the air,  Figure 7a shows the condensate aggregation rate under different working conditions. As the RH increased from 50% to 70%, the condensate aggregation rate increased by up to 1.8 times. This was because more moisture was contained in the air with high RH at a given temperature. When the air was cooled, more moisture was condensed from the air with high HR. It was also found that the condensate aggregation rate increased with the temperature. This could also be explained by the moisture contained in the air. At the high temperature, more moisture was contained in the air, although the RH is the same. As shown in Figure 7a, the condensate aggregation rate was significant. Under a condition of inlet air temperature of 35 • C and 70% RH, the condensate water accumulation rate reached 30.2 g/s. This could lead to the rapid blockage of the MHE fins, which might significantly affect the heat transfer rate of the MHE. Figure 7b shows pictures of condensate accumulation on the MHE surface taken by a high definition camera under the working condition of inlet air temperature, 30 • C and RH 50%. It was observed that the condensate starts to accumulate after 20 min when the system reached a stable condition. The condensation amount was significant after 30 min. The condensate on the MHE surface could gradually block the airflow, which in turn significantly affected the heat transfer between the air and MHE surface. This explained why the air-side heat transfer reached an optimal value at the RH of 60% and decreased as the RH increased beyond 60%. Figure 8 shows the variation in the air outlet temperature, MHE air-side heat transfer and overall heat transfer coefficient, and condensate rate under different air velocities. The inlet air temperature and relative humidity were set at 33 • C and 60%, respectively. The air velocity varied from 1.5 to 3.0 m/s in steps of 0.5 m/s. As shown in Figure 8a, both the air outlet temperature and air-side pressure drop increase with increased air velocity. As the air velocity increased from 1.5 m/s to 3.0 m/s, the air outlet temperature increased by 3.7 • C. This rise in the air outlet temperature was mainly because increasing the air velocity increased the air mass flow rate across the MHE and reduced the heat transfer time as the air flowed through the MHE. The air-side pressure drop was found to be increased by 61.1%. This was mainly caused by the high speed of the air.  Figure 9 shows the schematic drawing of the MHE installation in the wind tunnel. The air inlet temperature and relative humidity were maintained at 33 °C and 60%, respectively.  Figure 8b shows that both the air-side heat transfer and the overall heat transfer coefficient increased initially and dropped after the air velocity was above 2.5 m/s. As the air velocity increased, both air Reynolds number and the amount of air contacting the MHE surface increased, which enhanced the air-side heat transfer. The dependence of the overall heat transfer coefficients on the Reynolds number on the air-side of the MHE is similar to the air velocity (as is shown in Figure 8c) However, an increase in air velocity shortens the heat exchange time, which led to the reduction of the air-side heat transfer. So the air-side heat transfer and overall heat transfer coefficient were affected by these two factors. As the air velocity increased from 1.5 to 2.5 m/s, the first factor played the major role, and hence, both the air-side heat transfer and the overall heat transfer coefficient increased. As the air velocity is above 2.5 m/s, the latter factor played the major role, and hence, the air-side heat transfer and the overall heat transfer coefficient dropped. The optimal air velocity in the studied cases was around 2.5 m/s. Figure 8d shows the variation of the condensate formation and removal rate. It was found that the condensate formation rate was much higher than the removal rate. This explained the condensate aggregation on the MHE surface, as shown in Figure 7b. This condensate aggregation also increased Energies 2020, 13, 1065 13 of 17 the pressure drop of the air. It was also found that the highest condensate rate and removal rate appeared at the air velocity, 2.5 m/s. This could be explained from the air-side heat transfer. The results presented in Figure 8 clearly indicated that the air velocity significantly influenced the air-side heat transfer characteristics. Figure 9 shows the schematic drawing of the MHE installation in the wind tunnel. The air inlet temperature and relative humidity were maintained at 33 • C and 60%, respectively.  Figure 10a shows the air-side heat transfer and overall heat transfer coefficient under different inclined installation angles in the vertical direction. It was found that the total heat transfer did not change significantly, while the overall heat transfer coefficient increased slightly as the inclined installation angle increased from 0 to 20 °C. This increase in the overall heat transfer coefficient might be due to the condensate aggregation rate. As the inclined angle increased, the condensate aggregation rate decreased as the condensate dropped to the collection plate due to the gravity force. This can be evidenced by the pressure drop, as shown in Figure 10b. In Figure 10b, the air outlet temperature decreased as the inclined installation angle increased. This indicated that the air-side heat transfer increased as the inclined installation angle increased from 0 to 20 °C. As shown in Figure  10c, the condensate water accumulation rate did not change significantly if the measurement error was considered as the inclined installation angle changed from 0° to 10°. However, the condensate accumulation rate reduced substantially from 1.01 to 0.55 g/s as the inclined installation angle changed from 10° to 20°. This was because the inclined installation enhanced the condensate removal rate caused by gravity.  Figure 10a shows the air-side heat transfer and overall heat transfer coefficient under different inclined installation angles in the vertical direction. It was found that the total heat transfer did not change significantly, while the overall heat transfer coefficient increased slightly as the inclined installation angle increased from 0 to 20 • C. This increase in the overall heat transfer coefficient might be due to the condensate aggregation rate. As the inclined angle increased, the condensate aggregation rate decreased as the condensate dropped to the collection plate due to the gravity force. This can be evidenced by the pressure drop, as shown in Figure 10b. In Figure 10b, the air outlet temperature decreased as the inclined installation angle increased. This indicated that the air-side heat transfer increased as the inclined installation angle increased from 0 to 20 • C. As shown in Figure 10c, the condensate water accumulation rate did not change significantly if the measurement error was considered as the inclined installation angle changed from 0 • to 10 • . However, the condensate accumulation rate reduced substantially from 1.01 to 0.55 g/s as the inclined installation angle changed from 10 • to 20 • . This was because the inclined installation enhanced the condensate removal rate caused by gravity.  Figure 10a shows the air-side heat transfer and overall heat transfer coefficient under different inclined installation angles in the vertical direction. It was found that the total heat transfer did not change significantly, while the overall heat transfer coefficient increased slightly as the inclined installation angle increased from 0 to 20 °C. This increase in the overall heat transfer coefficient might be due to the condensate aggregation rate. As the inclined angle increased, the condensate aggregation rate decreased as the condensate dropped to the collection plate due to the gravity force. This can be evidenced by the pressure drop, as shown in Figure 10b. In Figure 10b, the air outlet temperature decreased as the inclined installation angle increased. This indicated that the air-side heat transfer increased as the inclined installation angle increased from 0 to 20 °C. As shown in Figure  10c, the condensate water accumulation rate did not change significantly if the measurement error was considered as the inclined installation angle changed from 0° to 10°. However, the condensate accumulation rate reduced substantially from 1.01 to 0.55 g/s as the inclined installation angle changed from 10° to 20°. This was because the inclined installation enhanced the condensate removal rate caused by gravity.

Conclusions
In this study, an MHE test system was developed to investigate the influence of inlet air temperature, air velocity, and inclined installation angle of MHE on the surface condensate aggregation/removal and heat transfer performance of the MHE used as evaporators. Some specific conclusions were drawn as below: • Both air inlet temperature and relative humidity showed a large effect on the overall heat transfer coefficient and condensate aggregation rate. As the inlet air temperature increased from 30 to 35 °C, the overall heat transfer coefficient increased from 72.5 to 82.5 W/(m2·K) at 60% RH.
An optimal heat transfer coefficient was found at 60% RH for each temperature. The condensate aggregation rate on the MHE surface increased with both air inlet temperature and relative humidity.
• The air velocity also showed a significant effect on the heat transfer characteristics of the MHE.

Conclusions
In this study, an MHE test system was developed to investigate the influence of inlet air temperature, air velocity, and inclined installation angle of MHE on the surface condensate aggregation/removal and heat transfer performance of the MHE used as evaporators. Some specific conclusions were drawn as below: • Both air inlet temperature and relative humidity showed a large effect on the overall heat transfer coefficient and condensate aggregation rate. As the inlet air temperature increased from 30 to 35 • C, the overall heat transfer coefficient increased from 72.5 to 82.5 W/(m 2 ·K) at 60% RH. An optimal heat transfer coefficient was found at 60% RH for each temperature. The condensate aggregation rate on the MHE surface increased with both air inlet temperature and relative humidity.

•
The air velocity also showed a significant effect on the heat transfer characteristics of the MHE. The outlet air temperature and pressure drop across the MHE increased as the air velocity increased from 1.5 to 3 m/s. However, analysis of air-side heat transfer, overall heat transfer coefficient