Hindawi Journal of Robotics Volume 2022, Article ID 8656289, 11 pages https://doi.org/10.1155/2022/8656289 Research Article On the Performance and Efficiency of Surface Air Cooler Working under High Temperature and High Humidity Condition Juan Yang, Jing Yu , Shijing Wang, and Weidong Yan School of Intelligent Engineering Technology, Jiangsu Vocational Institute of Commerce, Nanjing, Jiangsu 211199, China Correspondence should be addressed to Jing Yu; firstname.lastname@example.org Received 12 August 2022; Revised 29 August 2022; Accepted 30 August 2022; Published 28 September 2022 Academic Editor: Shahid Hussain Copyright © 2022 Juan Yang et al. is is an open access article distributed under the Creative Commons Attribution License, which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited. Surface air cooler is widely used in refrigeration, air conditioning, chemical dehumidi‡cation, and other related ‡elds. Nowadays, the application of surface air cooler has been mature under standard working conditions, and more and more research has been carried out on its application. However, when conventional surface air coolers are directly used under high temperature and high humidity conditions (the temperature is higher than 30 C, both the humidity ratios are higher than 20 g/kg dry air), or under conditions with strict dehumidi‡cation requirements, the expected results are often not achieved. erefore, taking the con- ventional surface cooler as an example, the performance and e“ciency of dehumidi‡cation of the surface cooler under high temperature and high humidity are investigated speci‡cally in this paper. Based on the above model and experiment, the in•uence of operation parameters on heat and mass transfer of surface air cooler is analyzed. At the same time, the change law of heat and mass transfer capacity, dehumidi‡cation performance, and heat transfer e“ciency of a surface air cooler is specially investigated. Results proved that under high temperature and humidity conditions (34.8 C, 23.03 g/kg dry air), lower head-on wind velocity could realize deeper humidi‡cation e˜ects e˜ectively, while water velocity showed fewer impacts. As shown in the experiment, when water velocity increased to 1.2 m/s, and head-on wind velocity dropped to 1.06 m/s, the humidity ratio at the air outlet reduced to 8.88 g/kg while heat-transfer e“ciency reached 0.778; however, when head-on velocity increased to 2.5 m/s and water velocity rose to 1.2 m/s, humidity ratio at air outlet dropped to 13.17 g/kg only, and heat-transfer e“ciency reached merely 0.572. In conclusion, this paper has established an accurate and e˜ective mathematical model for analyzing the heat transfer and dehumidi‡cation characteristics of the surface cooler under high temperature and high humidity conditions, which provides a reference for the development and design of dehumidi‡cation surface cooler under special conditions. More importantly, the conclusion of this paper also points out that, di˜erent from the adjusting strategy under conventional working conditions, reducing the head-on wind speed is the best adjusting strategy when the air surface cooler is used in high temperature and high humidity working conditions. is adjusting strategy is not only conducive to promoting the air to achieve a better dehu- midi‡cation e˜ect but also can e˜ectively improve the e“ciency of the surface cooler. requirements, researchers increase pipe banks and enlarge 1. Introduction heat transfer area. Mei et al.  established a mathematical As a kind of e“cient and low-cost heat-transfer device, model of a cooling coil and carried out a simulation study on surface air coolers are highly popular in air-conditioning, the law of changes in the heat transfer capacity and dehu- dehumidifying, and chemical applications. midifying system of surface air cooler under di˜erent Current conventional surface air coolers are usually conditions such as changing only the chilled water •ow and designed for nominal conditions (the dry bulb temperature changing only the air volume, etc., and ‡nally put forward ° ° is 27 C, and the wet bulb temperature is 19.5 C), wind speed with a method for systematic energy conservation and ° ° 2.5 m/s, and inlet and outlet water temperature 7 C/12 C . optimization, that is, to adjust the chilled water •ow and the For the purpose of improving the heat transfer performance air volume simultaneously. Liu et al.  concluded, with the of surface air coolers to make them satisfy dehumidi‡cation control variable method, that the heat-transfer e“ciency of 2 Journal of Robotics surface air cooler would increase along with the increase of (1) *e surface air cooler is made of a regular ﬁnned- water velocity and decrease along with the increase of the pipe structure, with the air and the water ﬂow in a head-on wind velocity and the dehumidifying coeﬃcient. Li cross direction;  proposed, for the high temperature and high humidity (2) Condensate water generated in the heat-transfer conditions, to realize dehumidiﬁcation by adopting the two- process ﬂows away in time from the coil surface, thus level rotary dehumidifying plus direct evaporative cooling impacts from the cumulation of such condensate system. You et al.  analyzed, based on experiments, water in ﬁnned-pipe gaps are ignored; whether indirect evaporative cooling and energy-recovering (3) Contact thermal resistance and fouling resistance are system used in high temperature and humidity area can ignored; satisfy dehumidiﬁcation requirements. *e abovementioned literature researches show that there Based on such assumptions, a physical model of a ﬁn- are few studies on the dehumidiﬁcation performance and ef- ned-pipe surface air cooler is established. Figure 1(a) shows ﬁciency of conventional surface air coolers under high tem- the structure of the such model, while Figure 1(b) shows the perature and humidity conditions. However, high temperature internal structure of ﬁnned pipes. and humidity conditions are common in many applications, such as the air-conditioning system in the summer, especially 2.2. Mathematical Model. A mathematical model of the in the plum rain season, in regions south of the Yangtze River heat and mass transfer process of the surface air cooler , the solar-energy-based wastewater treatment system in the was established based on the physical model described chemical industry , the air-conditioning system in textile above. workshops , and other related industrial applications. In engineering applications, when a regular surface air cooler is 2.2.1. Heat-Transfer Model at Pipe Outer Side of Surface Air used under high temperature and humidity conditions (the Cooler temperature is higher than 30 C, both the humidity ratios are higher than 20 g/kg dry air), defects such as failure to reach (1) Energy-balance equation: dehumidiﬁcation requirements appear frequently, resulting in Q � q i − i − m i , the cumulation of humidity road. Moreover, surface air coolers 1 am a, a,out cw cw in (1) today generally focus on heat transfer but ignore dehumidi- Q � h × A × Δt . 1 s out out ﬁcation, which results in the common existence of poor heat- transfer eﬃciency when such surface air coolers are used for Mass-balance equation: dehumidifying . To meet the dehumidiﬁcation require- ments, namely, the requirements for high mass transfer, when m � q d − d , (2) cw am a,in a,out using a regular surface air cooler to treat air at high temperature and humidity, it is necessary to reform and optimize such in which Q is the heat transfer capacity kW between surface air cooler. At present, there are two optimization air at the outer pipe side and the outer pipe wall; methods: (1) to adjust the operating conditions of the surface q is the mass ﬂow of the air (kg/h); am air cooler, and (2) to change the structure of the conventional i , i are the enthalpy at the air inlet and outlet a,in a,out surface air cooler. Among them, the second method would (kJ/kg); increase the cost of equipment investment, resulting in low d d is the moisture content at air inlet and economic beneﬁts. *erefore, the ﬁrst method is more suitable. a,in a,out outlet (kg/kg); In order to solve the above problems, in this paper, based on the conventional surface cooler structure, from the m is the mass ﬂow of the condensate water, cw perspective of changing the operating conditions, the heat namely, the dehumidiﬁcation capacity (kg/h); and mass transfer model of the surface cooler is established, i is the enthalpy of condensate water (kJ/kg.dry); cw and the corresponding experimental platform is also set up. h is the heat transfer coeﬃcient at outer pipe side *e dehumidiﬁcation performance and eﬃciency of the under wet conditions (W/m •K); surface cooler under the condition of high temperature and A is the heat transfer area out of the pipe (m ); high humidity are analyzed by means of both simulation and out experiment. *is paper aims to explore the heat and mass ∆t is the heat transfer temperature diﬀerence out out transfer law of surface coolers under high temperature and of the pipe, C. high humidity conditions through experiment and simu- It should be noted here that when air is cooled and lation and ﬁnd an eﬀective method to solve the poor de- dehumidiﬁed by the surface cooler, heat transfer and humidiﬁcation eﬀect under this special condition, to provide mass transfer are carried out together. *erefore, the reference for further research and design of surface cooler. part of energy taken away by condensate water also needs to be included in the energy balance equation. 2. Modeling (2) Heat transfer equation: 2.1. Physical Model. First of all, assumptions for the estab- *e heat transfer factor j proposed by Colburn  lishment of a physical model were given based on the actual is used to calculate the heat transfer process. *e situation of regular surface air cooler : mathematical model is expressed as follows: Journal of Robotics 3 S S f 2 (a) Outlet cooling water Outlet air Inlet air Inlet cooling water (b) Figure 1: Structure diagrams of surface air cooler (a, b). N is the number of pipe banks; h A 1 y 2/3 j Pr , C is the speci‡c heat of the wet air, kJ/kg•k. q C pa ma pa e model above is a heat transfer model under dry conditions. Because the ‡nned tube surface air cooler −1.2 1 − 1280 × N × Re L works under wet conditions, it is necessary to modify j ×0.0014 + 0.2618J , (3) 1 p −1.2 its heat transfer coe“cient. Under wet conditions, 1 − 5120 × Re the heat transfer coe“cient h is as follows: −0.15 h h × ζ × η , (4) s 1 ao −0.4 o J Re , P D where η is the e“ciency of the ‡nned surface, ao expressed as follows: in which Re is the Reynolds number based on η × A + A vertical pipe spacing; f f b η , (5) ao Re is the Reynolds number based on the outer A diameter of the pipe; Here, A is the side face area of the ‡n, m ; h is the heat transfer coe“cient at outer pipe side 2 2 A is the inter‡ns pipe surface area, m ; under dry conditions, W/m •K; A is the total side face area of ‡ns, m ; A is the head-on wind area, m ; 1 4 Journal of Robotics η is the ﬁn eﬃciency of the surface air cooler, f ⎧ ⎨ Pr ⎢ ⎥ ⎡ ⎢ ⎤ ⎥ ⎣ ⎦ V � × Re − 1000 × f w namely, the contact coeﬃcient . ⎩ 1/2 2/3 1 + 12.7(f/8) × Pr − 1 th(mh) 2/3 η � , ⎫ ⎬ mh × 1 + × c , ��� � (6) 2h m � , (9) λ δ l f V � h , f n s1 in which h— is the equivalent ﬁn height, m; f � , λ is the heat conductivity coeﬃcient of aluminum, 1.82 × log Re − 1.64 l w W/m•k; 0.01 Pr δ is the ﬁn thickness, m; c � , Pr In equation (4), ζ is the dehumidifying coeﬃcient of modiﬁed dry conditions : in which Pr is the Prandtl number of the water at r + t × C − t × C w a,m pw wa,out w mean temperature; ζ � 1 + pa Prw is the Prandtl number of the water at interior (7) wall temperature of the pipe; d − d a,m a,wa Re is the Reynolds number of the water at mean × , t − t temperature; a,m wa,out in which r is the latent heat of evaporation of the 2.2.3. Heat Transfer Model at Wall Face of the Surface Air water; Cooler t is the mean temperature of the air, C; a,m (1) Heat transfer equation : t is the temperature of the outer wall of the pipe, wa, out C; t − t wa,out wa,in Q � N × N × 2 × π × λ × L × , 3 1 c 1 d is the moisture content of air at t , kg/kg; a,m a,m ln d /d × 1000 o i d is the moisture content of air at t , kg/kg; a,wa wa, out (10) C is the speciﬁc heat of water vapor, kJ/kg•k; pw where t is the temperature at the interior wall of w,in C is the speciﬁc heat of water, kJ/kg•k. the pipe, C; N is the number of pipe arrays; 2.2.2. Heat Transfer Model at Internal Side of Surface Air λ is the heat transfer coeﬃcient of copper, W/m•k; Cooler L is the total length of heat-transfer copper pipe, m. (1) Energy-balance equation: (2) Energy-balance equation: Q � C q t − t , 2 w mw w,out w,in Q � Q . (11) 3 1 (8) Q � h × A × Δt , 2 n in in *e interior wall temperature of the pipe, t is cal- wa,in culated with the equation above, as criteria for model in which Q is the heat transfer capacity at the in- iteration. ternal side of the pipe, kW; q is the mass ﬂow rate of water, kg/h; mw 2.2.4. Performance Parameter Model of the Surface Air t , tw are the inlet and outlet water tempera- ,out w,in ° Cooler. For the purpose of analyzing the heat and mass tures, C; transfer capacity, dehumidifying performance, and heat h is the heat transfer coeﬃcient inside the pipe, W/ transfer eﬃciency of the surface air cooler, the model m •k; outputs state parameters at the air outlet, heat transfer ca- A is the heat transfer area inside the pipe, m ; in pacity Q, dehumidiﬁcation capacity m , heat transfer co- cw ∆t is the heat transfer temperature diﬀerence inside eﬃcient K, contact coeﬃcient η , heat exchanger eﬃciency in f the pipe, C. η , and speciﬁc air consumption SAC . ao Heat transfer coeﬃcient K: (2) Heat transfer model: For the heat transfer coeﬃcient h inside the pipe, (12) K � . Gnielinski  formula is employed for solution as Δt · A follows: Journal of Robotics 5 Here, ∆t is the heat transfer temperature diﬀerence of (2) Indirect error: surface air cooler, and C; A is the heat transfer area of According to the method proposed by Moﬀat , surface air cooler, m . measuring uncertainties are deﬁned as diﬀerent Heat exchanger eﬃciency η : ao square roots between observed solid errors and random errors of the instrument. As for the ex- t − t a,in a,out η � . (13) ao periment described herein, directly measured un- t − t a,in w,in certainties are deemed as independent from each Speciﬁc air consumption refers to the amount of air other, uncertainties of indirect parameters, thus can consumed per unit of dehumidiﬁcation. It can reﬂect the be calculated as follows: ������������� � power consumption of the blower from the side. Generally, the higher the speciﬁc air consumption is, the higher the zf (15) u � u . power consumption of the blower will be. c xi zx i�1 ma SAC � . (14) cw Here, x is the directly measured parameter, f refers to the relationship between the indirectly measured parameter and directly measured parameter, and u is the direct mea- 2.3. Logic Block Diagram of the Model. *e mathematical surement error. According to the abovementioned method, model above consists of four major modules and multiple errors of all indirect parameters are ±5%, which proves the boundary conditions. Software EES was used for code editing, reliability of the test under the experiment. and a complete heat and mass transfer model of the ﬁnned-pipe surface air cooler was established for systematic stimulation 4. Results and Discussion research. See Figure 2 for the model calculation logic diagram. 4.1. Experimental Veriﬁcation of the Model. *is paper 3. Surface Air Cooler Experiment conducted a comparison study between experimental results and stimulated results so that to verify the accuracy of the 3.1. Establishment of the Testbed and Layout of Test Points. heat transfer model of the surface air cooler established In order to verify the accuracy of the model and examine the herein. actual inﬂuence of the changes in system operating pa- Figure 6 shows the comparison between the experi- rameters, an experimental bench as shown in Figure 3 is mental results and stimulated results of the outlet air dry built. *e setting and regulation of the working conditions bulb temperature, the air outlet moisture content, and the are realized by using the enthalpy diﬀerence laboratory, as outlet water temperature. It can be clearly seen that the shown in Figure 4. prediction accuracy of the model for these three state pa- All experimental operating parameters, including air rameters is very high, and the error is within ±10%. inlet and outlet dry and wet bulb temperatures, pressure *erefore, the model of ﬁnned-pipe surface air cooler diﬀerence before and after the nozzle, inlet and outlet water established in this paper is reliable. Subsequently, the temperature, and water ﬂow rate, can be displayed and set dehumidiﬁcation performance and eﬃciency of surface air through the terminal control platform of the enthalpy dif- coolers under high temperature and humidity conditions ference laboratory, as shown in Figure 5. are analyzed in detail by combining simulation and experiment. 3.2. Test Conditions. To simulate engineering applications, high temperature and humidity fresh air condition in the 4.2. Analysis on Aﬀecting Factors. *is paper studied mainly summer of Nanjing were taken as inlet air parameters for the impacts of operating conditions on the dehumidiﬁcation simulation and experimental analysis, see Table 1 for ex- performance and eﬃciency of surface air coolers. Aﬀecting perimental test condition. parameters include mainly water ﬂow velocity and head-on wind velocity. As shown in Figure 7, when the water ﬂow velocity is 3.3. Experimental Data Processing and Error Analysis unchanged, both the temperature and moisture content of 3.3.1. Data Processing. In the experiment, primary mea- the air outlet dry bulb drop gradually along with the drop of sured parameters (t , t , t , and t ) can be ob- head-on wind velocity. In the contrast, when head-on wind a,in sa,in a,out sa,out tained directly by measurement, while indirect parameters velocity is unchanged, both the temperature and moisture (η , k, and SAC) of surface air cooler can be obtained from a content of the air outlet dry bulb increase gradually along ao calculation based on equations (1)∼(14). with the drop of water ﬂow velocity. By comparing the eﬀects of changes in water ﬂow ve- locity and changes in head-on wind velocity, we know the 3.3.2. Error Analysis. Experimental errors include instru- eﬀects of these two factors on the temperature and moisture ment errors and indirect errors. content of the air outlet dry bulb are opposite to each other. (1) Instrument errors are shown in Table 2. *us, preliminarily, reducing the head-on wind velocity or 6 Journal of Robotics Begin Input initial conditions: Physical property parameters: t and d etc a, in a, in. Structure parameters: s s s δ and N ect f, 1, 2, f 1. Boundary conditions: ω , ω , t s y w, in Initial assumption: t wa, in Water-side module : Air-side module: Assumed t t = t Assumed t a, out, wa, out wa, in w, out Calculate Q with the Calculate Q with energy energy-balance equation and mass equation ermodynamic calculation ermodynamic calculation of of surface air cooler: A surface air cooler: in, Δt , h , Q ' A , Δt , h Q ' in n 2 in in s 1 Calculate water outlet Reassign: Reassign: Calculate water outlet Reassign: t + t t + t temperature T wa, in wa, ina t + t a, out a, outa temperature T w,outa w, out w, outa a, outa t = t = wa, in = a, out w, out NO NO t t t t a,out a,outa w, out w, outa ≤ 0.001 ≤ 0.001 a,out w, out YES YES Calculate state parameters of internal and outer walls of pipe Calculate internal pipe wall temperature T wa, ina t t NO wa, in wa, ina ≤ 0.001 wa, in YES Output calculation results End Figure 2: Model calculation logic diagram. sa, out sa, in a, out sa, in Inlet air Outlet T T s s air P P 1 2 T T t t w,in w, out V Enthalpy difference chamber Water chilling unit Thermostatic water tank Figure 3: Test point layout of the experimental system. Mixing chamber Measuring chamber Pressure balancing chamber Nozzle Pressure balance chamber Blower Journal of Robotics 7 Figure 4: Physical photo of the experimental system. On Off On Off On Off 4 3 21 19.62°C 07.00 08.54 07.00 08.60 07.00 08.59 On Off Humidif Heating Blower i-cation 21.2% 21.8% 30.0 HZ On 032.0% 082. 2% AHU2 Off Off Off 032. 0% Off Mixing Nozzle Blower 07. 01°C chamber 34. 82° C 13. 32°C 10. 30°C 1.670m3/h 15.0 HZ 28. 86° C 12. 81°C Off 24. 65Pa 152. 34Pa Home Setting Data Historical Real-time Curve 2 Curve 1 Report Report Indoors Side Figure 5: Control principle of the experimental system. improving the water ﬂow velocity can both reduce eﬀectively ﬁnned-pipe surface air cooler under high temperature and the temperature and moisture content of the air outlet dry humidity condition, namely, realizing deep dehumidiﬁca- bulb and realize lower moisture content at the outlet of the tion eﬀects. 8 Journal of Robotics Table 1: Experimental test condition. Parameters Value t / C 34.8 a,in t / C 28.9 sa,in ω /m/s 0.5∼3 ω /m/s 0.5∼1.2 t / C 7 w,in Table 2: Measuring parameters and instruments. Measuring parameters Instruments Model Precision Measuring range Temperature Pt100 platinum thermistor WZP-02 Level A −200∼420 C 3 3 Water •ow rate Electromagnetic •owmeter AXF040G ±0.35% m /h 0∼70 m /h Air •ow rate Di˜erential pressure gauge EJA110A ±0.25% FS 0∼1000 Pa 26 26 18 18 24 24 16 16 22 22 10% 14 14 20 10% 20 18 18 12 12 -10% -10% 16 16 10 10 14 14 8 8 12 12 8 10 12 14 16 18 12 14 16 18 20 22 24 26 Air outlet moisture content d (kg/kg dry air) Air outlet dry bulb temperature t (°C) 2-simulation results 2-simulation results (a) (b) 18 18 16 16 14 14 10% 12 12 -10% 10 10 8 8 8 10 12 14 16 18 Water outlet temperaturet (°C) s2-simulation results (c) Figure 6: Comparison between experimental results and simulation results. When the water •ow velocity was kept at 1.2 m/s and •ow velocity increased from 0.5 m/s to 1.2 m/s, the tem- the head-on wind velocity dropped from 3 m/s to 0.5 m/s, perature of the air outlet dry bulb dropped from 15.49 C to the temperature of the air outlet dry bulb dropped from 13.11 C, while the air humidity ratio dropped from 10.57 g/ ° ° 19.04 C to 10.4 C, while the air humidity ratio dropped kg to 8.88 g/kg, dropped by 16%. Moreover, increasing the from 13.15 g/kg to 7.45 g/kg, dropped by 43.3%. When the water •ow velocity will consume more power from the head-on wind velocity was kept at 1.06 m/s and the water water pump. In consequence, reducing the head-on wind Air outlet dry bulb temperature t (°C) 2-experimental results Water outlet temperaturet (°C) s2-experimental results Air outlet moisture content d 2-experimental (kg/kg dry air) results Journal of Robotics 9 32 20 140 1.0 0.8 0.6 0.4 4 0.2 0 0.0 0.5 1.0 1.5 2.0 2.5 3.0 0.5 1.0 1.5 2.0 2.5 3.0 Head-on wind velocity (m/s) Head-on wind velocity (m/s) Simulation η (1.2m/s) Simulation t (1.2m/s) Experimental t (1.2m/s) Simulation K (1.2m/s) 2 2 f Simulation t (0.9m/s) Experimental t (0.9m/s) Simulation η (0.9m/s) Simulation K (0.9m/s) 2 2 Simulation t (0.7m/s) Experimental t (0.7m/s) Simulation η (0.7m/s) Experimental K (1.2m/s) 2 2 Simulation t (0.5m/s) Experimental t (0.5m/s) Simulation η (0.5m/s) 2 2 Experimental K (0.9m/s) Simulation d (1.2m/s) Experimental d (1.2m/s) Simulation K (0.7m/s) 2 2 Experimental K (0.7m/s) Simulation d (0.9m/s) Experimental d (0.9m/s) Simulation K (0.5m/s) 2 2 Experimental K (0.5m/s) Simulation d (0.7m/s) Experimental d (0.7m/s) 2 2 Figure 8: E˜ects of the head-on wind velocity and water •ow Simulation d (0.5m/s) Experimental d (0.5m/s) 2 2 velocity to heat transfer performance. Figure 7: E˜ects of the head-on wind velocity and water •ow velocity to air outlet parameters. 1.0 300 0.8 250 velocity shows more advantages in heat transfer and dehumidi‡cation. e above analysis focuses on the in•uence of state 0.6 200 parameters that explain the overall heat transfer and de- humidi‡cation e˜ects of the surface air cooler. In addition, it 0.4 150 is equally important to analyze the performance parameters. As shown by curves in Figure 8, when the water •ow velocity 0.2 100 was the same, the heat coe“cient dropped gradually while the contact coe“cient increased gradually, along with the reduction of head-on wind velocity; when the head-on wind 0.0 50 velocity was the same, heat transfer coe“cient decreased 0.5 1.0 1.5 2.0 2.5 3.0 gradually while contact coe“cient increased gradually along Head-on wind velocity (m/s) with the reduction of water •ow velocity. In terms of overall Simulation SAC (1.2m/s) Experimental SAC ( 1.2m/s) e˜ects, reducing the head-on wind velocity or water •ow Simulation SAC (0.9m/s) Experimental SAC ( 0.9m/s) velocity goes against the increase of the heat transfer co- Simulation SAC (0.7m/s) Experimental SAC ( 0.7m/s) e“cient but helps increase the contact coe“cient. erefore, Simulation SAC (0.5m/s) Experimental SAC ( 0.5m/s) changing condition parameters could not prove whether the Simulation η ( 1.2m/s) Experimental η (1.2m/s) ao ao surface air cooler is good or not in heat transferring or Simulation η ( 0.9m/s) Experimental η (0.9m/s) ao ao dehumidifying. Moreover, this paper studied mainly the Simulation η ( 0.7m/s) Experimental η (0.7m/s) ao ao dehumidi‡cation performance of surface air coolers under Simulation η ( 0.5m/s) Experimental η (0.5m/s) ao ao high temperature and humidity conditions, which makes the Figure 9: E˜ects of the head-on wind velocity and water •ow mere considering of heat transfer coe“cient one-sided velocity to heat transfer e“ciency. relatively. As shown in Figure 9, when the water •ow velocity kept e“ciency of surface air cooler and make the overall de- the same, the reduction of head-on wind velocity made the humidi‡cation and heat transfer process closer to the heat transfer e“ciency improve signi‡cantly and SAC theoretical limit (for example, when the head-on wind dropped gradually; when the head-on wind velocity kept the velocity decreased from 2.5 m/s under standard working same, the increase of water •ow velocity made the heat condition of regular surface air cooler to 1.06 m/s under transfer e“ciency and SAC improved slightly. conditions described in the experiment herein, η in- According to the above analysis, reducing the head-on ao creased by 36% and SAC decreased by 30.5%). Moreover, wind velocity could not only realize the dehumidi‡cation reducing the head-on wind velocity could reduce SAC, e˜ects but also improve e˜ectively the heat transfer Air outlet dry bulb temperature t (°C) Air outlet humidity ratio d (g/kg dry air) Heat transfer eﬃciency η ao 2 Heat transfer coefficient K (w/m .k) Contact coefficient η Speciﬁc air consumption SAC 10 Journal of Robotics Table 3: Comparison of surface coolers in diﬀerent applications. *e head-on Air humidity ratio after *e heat Types of surface cooler application wind dehumidiﬁcation by surface exchanger eﬃciency speed (m/s) cooler (g/kg dry air) of the cooler η ao Surface cooler for chiller units used in conventional 1.8 9.5 0.694 air-conditioning  Surface cooler for evaporative chiller units used in hot 1.8 12.7 0.617 and humid areas  Surface cooler for dehumidifying high temperature and 2.0 16.44 0.525 high humidity air in wastewater recovery treatment  1.8 12.68 0.621 Surface air cooler in this paper 0.5 7.45 0.867 and the reduction of SAC means lower airﬂow of surface 5. Conclusion air cooler under the same dehumidiﬁcation requirements, In this paper, the performance and eﬃciency of the surface which is beneﬁcial to reduce energy at the air-delivery side, that is, to reduce the power consumption of the cooler are investigated in detail under high temperature and high humidity conditions. By establishing the heat and mass blower. Although increasing the water ﬂow velocity could slightly promote the dehumidiﬁcation and heat transfer transfer model of the surface cooler and combining it with the performance test of the surface cooler, the simulation eﬀects and performance, such an increase resulted in higher power consumption of the water pump. *erefore, and experimental test of the surface cooler are carried out. *e inﬂuence of working condition parameters on the de- from the perspective of working condition regulation, signiﬁcantly reducing the head-on wind speed is more humidiﬁcation performance of surface coolers under high temperature and high humidity conditions is mainly ex- conducive to improving the dehumidiﬁcation eﬀect of plored, and the following conclusions are reached: surface air coolers under high temperature and humidity conditions and reducing energy consumption. (1) *e heat and mass transfer model established in this paper has high accuracy (the error of predicting air temperature and humidity can be maintained within 4.3. Comparison and Analysis with Current Surface Coolers. ±5%). *erefore, this model can provide a reference for As shown in Table 3, the ﬁnal dehumidiﬁcation eﬀect of the the subsequent optimization design of surface coolers surface cooler and the comparison of heat exchanger eﬃ- used in extreme working conditions such as high ciency under diﬀerent application conditions are shown. temperature and high humidity air treatment. When applied to conventional working conditions, it is (2) Factor analysis shows that the impact of the head-on obvious that the surface air cooler can have both a better wind speed on the heat transfer and dehumidiﬁcation dehumidiﬁcation eﬀect (keeping the air humidity ratio performance of the surface cooler is greater than that of below 10 g/kg dry air) and good heat exchanger eﬃciency the water velocity. *erefore, in engineering applica- (about 0.7) at a higher head-on wind speed. However, when tions, it is more beneﬁcial to improve the energy ef- the surface cooler is used in high temperature and high ﬁciency of the system by regulating the head-on wind humidity conditions, continuing to choose the same head- speed to meet the actual dehumidiﬁcation demand. on wind speed as in the conventional conditions can no longer meet better dehumidiﬁcation and high eﬃciency, just (3) Performance analysis shows that, compared with the as shown in the literature  and literature . *e re- conventional working conditions, when the surface search results in this paper also show that at a higher head- cooler with the same design parameters is used in on wind speed (1.8 m/s), the humidity ratio of the outlet air high temperature and high humidity working con- of the surface cooler is also maintained at 12.68 g/kg dry air, ditions, the dehumidiﬁcation eﬀect of the surface and the heat exchanger eﬃciency is also lower (only 0.621). cooler and the heat exchanger eﬃciency will be However, another analysis result of this paper also shows signiﬁcantly reduced. Under this condition, the that in conditions of high temperature and humidity, the air design changes of the heat exchanger to reduce the humidity ratio after dehumidiﬁcation by surface cooler can head-on wind speed can eﬀectively improve the be eﬀectively reduced by reducing the head-on wind speed, dehumidiﬁcation eﬀect, improve the heat exchanger thus achieving a better dehumidiﬁcation eﬀect. At the same eﬃciency, and reduce the gas consumption ratio to time, high heat exchanger eﬃciency can be maintained reduce the power consumption. (0.867). *e conclusion of this paper enriches the research of *erefore, by comparing with the application of existing surface coolers under extreme conditions of high temper- surface coolers, it can also be concluded that reducing the ature and humidity and has reference signiﬁcance for the head-on wind speed has a good improvement eﬀect on the development and design of surface coolers under such problems of poor treatment eﬀect and low eﬃciency of the conditions. However, this paper only investigated the per- heat exchanger when the surface cooler is treating high formance and eﬃciency of the surface cooler. In the future, temperature and high humidity air. Journal of Robotics 11  Y. Z. Wu, Principles and Equipment of Refrigeration, Xi’an the dehumidiﬁcation performance of the surface cooler Jiaotong University Press, Xi’an, China, 1997. structure optimization will be further studied by using the  V. 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Journal of Robotics
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Published: Sep 28, 2022