1、Atsushi Tsujimor is a professor at the Department of Mechanical Engineering, Kanto-gakuin University, Yokohama, Japan. Masakazu Ohnuki is an assistant professor at the Department of Mechanical Engineering, Kanto-gakuin University, Yokohama, Japan. g3Absorption Refrigeration Cycle Based On Capillary
2、Force Atsushi Tsujimori Dr.Eng Masakazu Ohnuki M.Sc. Member JSRAE ABSTRACT This paper presents an absorption refrigeration cycle which is driven by capillary pumping. Conventional absorption refrigeration cycles require solution pumps to transfer the solution from a low-pressure absorber to a high-p
3、ressure generator. In contrast, the cycle proposed in the present study involves a low-pressure generator and a low-pressure absorber. The low-pressure generator consists of a heat transfer tube with rectangular grooves, which provide a path for the refrigerant vapor, and a cylindrical capillary wic
4、k inserted into the heat transfer tube. The inside part of the wick and the cooling tubes constitute the absorber. The working solution in the absorber penetrates the wick, after which the refrigerant evaporates due to the heat load transmitted to the heat transfer tube. This evaporation creates a p
5、ressure difference between the evaporated refrigerant vapor and the solution at the liquid-vapor interface, and the high-pressure refrigerant vapor condenses into liquid phase as it cools inside the condenser. In this situation, the high-pressure liquid refrigerant expands via the expansion valve an
6、d evaporates into the evaporator, which results in a decrease in temperature. The low-pressure refrigerant vapor which evaporates into the evaporator is absorbed into the solution inside the absorber. In this cycle, only the refrigerant vapor generated in the generator and the liquid refrigerant in
7、the condenser are kept at high pressure. Therefore, there is no need to use a solution pump to transport the solution. For the purposes of this study, experimental equipment was manufactured in order to verify the effectiveness of the proposed cycle. The experimental equipment consisted of a condens
8、er, an evaporator, a low pressure generator and an absorber, where the generator and the absorber were separated by a porous wick. Furthermore, in order to simplify the experiment, the refrigeration test was performed by using a single-component refrigerant as the working fluid. The resulting decrea
9、se in pressure due to capillary pumping was between 3.3 and 4.7 kPa, and a decrease in temperature of 8.0 to 13.2 K was achieved. INTRODUCTION The increase of concentrated heat fluxes in electronic devices has determined the need for advanced cooling methods. However, the size of electronic componen
10、ts is constantly decreasing, which obstructs the mounting of cooling devices adjacent to the heat source. Heat transport loops that utilize latent heat 12345 have been proposed as a solution to this problem. In addition, we have developed a cooling method based on loop heat pipes 678. However, these
11、 cooling techniques are not as efficient when the ambient temperature is greater than the temperature of the target electronic components. In order to meet the cooling requirements, several types of refrigeration systems, such as vapor compression systems, absorption refrigerators 910, adsorption re
12、frigerators 11 and Stirling coolers 12 have been applied in electronic cooling. In contrast to heat pipes, such cooling devices require mechanical driving parts. Even absorption refrigerators, which are driven mainly by heat, require pumps in order to circulate the solution, which decreases their re
13、liability and applicability as general-purpose cooling devices. In this study, an absorption refrigeration cycle without solution pumps is proposed. In order to maintain the pressure difference between the condenser and the evaporator for the purpose of refrigeration, the generator and the absorber
14、are separated by a capillary wick, and only the generated refrigerant vapor is maintained at high pressure. LV-11-C044364 ASHRAE Transactions2011. American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). Published in ASHRAE Transactions, Volume 117, Part 1. F
15、or personal use only. Additional reproduction, distribution, or transmission in either print or digital form is not permitted without ASHRAES prior written permission.PROPOSED CYCLE The Duhring diagram of a conventional absorption cycle is shown in Fig. 1. The absorption cycle includes two pressuriz
16、ation stages, one of which corresponds to the evaporation and absorption range and the other corresponds to the generation and condensation range. A solution pump raises the solution pressure from the absorption range to the generation range and transfers the solution from an absorber into a generat
17、or. However, only a pair composed of a high pressure condenser and a low pressure evaporator is needed in order to achieve refrigeration. Therefore, the high-pressure solution inside the generator is needed only for the purpose of generating high-pressure refrigerant vapor. The cycle proposed in thi
18、s study is shown in Fig. 2. In contrast to a conventional absorption cycle, the absorber and the generator operate in the same pressure range. Figure 1 Duhring diagram of conventional absorption cycle Figure 2 Absorption refrigeration cycle proposed in this study Temperature PressurePLPHTe Tc T gPre
19、ssurePLPHTemperatur TeTcTg 2011 ASHRAE 365In other words, a solution pump which transfers the low-pressure solution to the high pressure region becomes unnecessary. In order to realize such a cycle, the maximum pressure difference Pmaxof the liquid phase and the gas phase shown in Eq. 1 is maintaine
20、d by using a porous material (wick) in the generator which separates the vapor channels and the solution reservoir inside the absorber. The working solution in the absorber penetrates the wick, after which the refrigerant evaporates due to the heat load transmitted to the heat transfer tube. This ev
21、aporation creates a pressure difference between the evaporated refrigerant vapor and the solution at the liquid-vapor interface. The high-pressure refrigerant vapor condenses into liquid phase as it cools inside the condenser. Next, the high-pressure liquid refrigerant expands via the expansion valv
22、e and evaporates into the evaporator, which produces a decrease in temperature. Finally, the low-pressure refrigerant vapor, which evaporates into the evaporator, is absorbed into the solution in the absorber after it is cooled. In this cycle, only the refrigerant vapor generated in the generator an
23、d the liquid refrigerant in the condenser are maintained at high pressure. Therefore, there is no need to use a solution pump to transport the solution from the absorber to the generator. Here, the pressure difference Pmaxconsists of the pressure difference P between the condenser and the evaporator
24、 in the expansion stage and the decrease in pressure caused by the flow of the refrigerant and the solution during the cycle. Because the decrease in pressure caused by the flow of the refrigerant and the solution depends on the system configuration, only the pressure difference Pwin the wick is con
25、sidered in Eq. 2. edP4max= g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3 (1)g3g3wLHPPPPP =maxg3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3 (2)g3g3where g3g3 GKAtPwwsw= g3g3g3g3g3g3g3g3
26、g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3(3)g3g3The energy equations for the evaporator, the condenser, the generator and the absorber are shown below. g3g3 ()eieorehhGQ = g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g
27、3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3(4)g3g3g3 ( )cogorchhGQ = g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3 (5)g3g3g3wswsssssgorghGhGhGQ += g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3 (6)g3g3g3wswsss
28、sseorahGhGhGQ += g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3g3 (7)g3g3acgeGQQQ +=+ g3g3g3g3g3g3g3g3g3g3g3g3g3g3 (8)g3EXPERIMENTAL SETUP Experimental equipment In this study, a test regarding refrigeration based on capillary pumping was performed as the first step towards the realization o
29、f the refrigeration system proposed in the previous section. The experimental equipment shown in Fig. 3 consists of a combined generator-absorber, a condenser, an evaporator and an expansion valve, where the cross section of the generator-absorber is shown in Fig. 4. In this setup, the low pressure
30、generator consists of a heat transfer tube with rectangular grooves which provide paths for the refrigerant vapor, and the absorber consists of a capillary wick and cooling 366 ASHRAE Transactionstubes. The capillary wick is cylindrical in shape and is inserted into the heat transfer tube, while the
31、 inside part of the wick and the cooling tubes constitute the absorber. The high pressure in the condenser and the low pressure in the evaporator are measured by pressure transducers, and the temperatures are monitored by thermocouples. Figure 3 Test configuration Figure 4 Cross section of generator
32、-absorber PPGenerator-absorberEvaporator Condenser Cooling water Cooling water Resistance heater Resistance heater Vapor channelHeat transfer tubeWickCooling tubeReservoir 2011 ASHRAE 367Testing method Generally, in the absorption refrigeration cycle the working fluid is a refrigerant-solvent pair.
33、However, in this study we focus on the verification of refrigeration by capillary pumping. Therefore, we used only a refrigerant for the working fluid in order to simplify the experiment. The experimental conditions are shown in Table 1 (in this experiment, TFE was chosen as the refrigerant). The te
34、mperature of the cooling water inside the condenser was set to 25C(298K),g3 and the temperature of the cooling water for the absorber (low temperature condenser in this case) was varied between 0C(273K) and 10C(283K) in order to allow the low-pressure refrigerant vapor to condense. In these experime
35、nts, we measured the differences in pressure and temperature between the condenser and the evaporator. Table 1. Experimental Conditions Refrigerant TFECooling water (condenser) 25C(298K) Cooling water (absorber: low temperature condenser) 0 to 10C (273K to 283K) Diameter of wick pores 20m RESULTS AN
36、D DISCUSSION Calculation Results Figures 5 and 6 show the calculation results for the condition of maximum pressure difference produced by capillary pumping for different pore diameters. In these calculations, the evaporation temperature is constant at 10C, and the weak solution temperature is also
37、constant at 60C. Figure 5 shows the changes in the condensation temperature and the pressure difference P between the condenser and the evaporator. The pressure difference P is given as the difference Pmax- Pw,and the condensation temperature is determined by calculating PH(= P + PL) from Eq. 2. As
38、clear from Fig. 5, as the pore diameter becomes smaller, the pressure difference increases, and the condensation temperature increases following the increase of the enthalpy of the generated refrigerant vapor. The condensation temperature is 20.4C when the pore diameter is 30m, which increases to 53
39、.8C when the pore diameter is 2m. When the pore diameter is smaller than 10m, the cooling COP in Fig. 6 decreases as the pore diameter becomes smaller. However, the cooling COP is almost constant when pore diameter is larger than 10m. The absorption refrigeration cycle proposed in this study is driv
40、en by capillary pumping without the need for mechanical solution pumps. Therefore, the concentration of the solution is automatically determined, and the concentration of a strong solution decreases (from 0.77 to 0.49 in this case) as the pore diameter decreases from 30m to 2m. As a result, the circ
41、ulation ratio increases and the temperature difference between the weak solution and the strong solution becomes smaller, which limits the loss of sensible heat. In the range of smaller pore diameters (less than 10m), the effect of increasing the enthalpy of the generated refrigerant vapor is larger
42、 than that of decreasing the loss of sensible heat of the solution for the purpose of decreasing the cooling COP. In the subsequent calculations, the temperature of the weak solution is chosen as the target parameter. Also, the evaporation temperate is 10C and the pore diameter is constant at 20m. T
43、he temperature of the weak solution represents the temperature necessary for the generation of refrigerant vapor from the solution, and the calculation results are shown in Figs. 7 and 8. As seen from Fig. 7, both the pressure difference, which is caused by capillary pumping, and the condensation te
44、mperature decrease as the temperature of the weak solution increases. This is due to the fact that the surface tension of the solution decreases as the solution temperature increases. In addition, as seen from Fig. 8, the temperature of the strong solution increases, accompanied by the decrease of t
45、he condensation temperature. The generator and the absorber are separated only by the wick, and the weak solution in the generator must be cooled to the absorption temperature in the absorber in order to ensure that the sensible heat in the weak solution is not recovered in this cycle. This affects
46、the cycle performance and decreases the cooling COP. 368 ASHRAE Transactions01020304050600 5 10 15 20 25 30 35Pore diameter de (m)Pressuredifferenceg17363P (kPa)0102030405060Condensationtemperature Tc(g16989)Fig. 5 Effect of pore diameter on pressure difference and condensation temperature (Evaporat
47、ion temperature = 10C) 0.50.520.540.560.580.600.20.40.60.810 5 10 15 20 25 30 35CoolingCOPStrongsolutionconcentrationsPore diameter de (m)Fig. 6 Effect of pore diameter on concentration of strong solution and cooling COP (Evaporation temperature = 10C) Condensation temperature Pressure difference Co
48、ncentration of strong solution Cooling COP 2011 ASHRAE 369024681055 60 65 70 75 80 85 90Weak solution temperature Tws (g16989)Pressure difference g17363P (kPa)051015202530Condensation temperatureTc(g16989)Fig. 7 Effect of temperature of weak solution on pressure difference and condensation temperatu
49、re (Evaporation temperature = 10C) 00.20.40.60.8155 60 65 70 75 80 85 90Weak solution temperature Tws (g16989)Solutionconcentration0.30.40.50.60.7CoolingCOPFig. 8 Effect of temperature of weak solution on solution concentration and cooling COP (Evaporation temperature = 10C) Condensation temperature Pressure difference Concentration of strong solution Concentration of weak solution Cooling COP 370 ASHRAE TransactionsExperimental Results Several experiments were performed in order to verify the proposed ref
