Copper Micro-channel Loop Thermosyphon. A thesis presented to. the faculty of. the Russ College of Engineering and Technology of Ohio University

Transcription

1 1 Copper Micro-channel Loop Thermosyphon A thesis presented to the faculty of the Russ College of Engineering and Technology of Ohio University In partial fulfillment of the requirements for the degree Master of Science Juan G. Flores-Lozada November Juan G. Flores-Lozada. All Rights Reserved.

2 2 This thesis titled Copper Micro-channel Loop Thermosyphon by JUAN G. FLORES-LOZADA has been approved for the Department of Mechanical Engineering and the Russ College of Engineering and Technology by Khairul Alam Moss Professor of Mechanical Engineering Dennis Irwin Dean, Russ College of Engineering and Technology

3 3 ABSTRACT Flores-Lozada, Juan G., M.S., November 2009, Mechanical Engineering Copper Micro-channel Loop Thermosyphon (121 pp.) Director of Thesis: Khairul Alam The purpose of this work was to use copper micro-channel tube to design, construct and analyze the performance of a loop thermosyphon prototype. The design stage included a pressure loss analysis to ensure working fluid circulation. Analytical, numerical and experimental analyses were performed to calculate the heat transfer coefficient and heat transfer and to compare the results. The analytical analyses consisted of estimations of the heat transfer coefficient and heat transfer with the fin theory, and flow over a flat surface. The numerical analysis was performed using the computational fluid dynamics software Fluent. The experimental work was conducted on a small-scale wind tunnel under constant air flow, 0.5 and 1.0 m/s, and different operating temperatures, ~50 C and ~70 C. Propane was chosen as the working fluid for the system due to its relatively low global warming potential, saturation temperature range, availability, ease of storage, and non toxic properties. This work confirmed the concept of a loop thermosyphon by providing an isothermal surface throughout the prototype. The isothermal fins increased the heat transfer by a maximum of 63% compared to the classical fin mode. Approved: Khairul Alam Moss Professor of Mechanical Engineering

4 4 ACKNOWLEDGMENTS I would like to thank my thesis advisor, Dr. K. Alam, for his guidance and the time he dedicated to this research project. In addition, thanks to the members of the thesis committee, Dr. F. Kraft, Dr. I. Urieli, for their valuable contributions to the project. Special thanks to Mr. J. Edwards, Mr. J. Pagán, Mr. R. Mulford and Mrs. S. Walker; without their help and contributions, this project would have been very difficult to complete. I would like to acknowledge financial support from the Department of Mechanical Engineering, the Research Challenge Program, the Center for Advanced Materials Processing and the International Copper Association. Finally, I would like to thank my family, especially my mom and my fiancée, for their encouragement and support throughout the completion of the project.

5 5 TABLE OF CONTENTS Page Abstract... 3 Acknowledgments... 4 List of Tables... 8 List of Figures List of Symbols Introduction Heat exchangers Parallel-flow heat exchanger Counter-flow heat exchanger Cross-flow heat exchanger Thermosyphon and heat pipe Micro-channel tubes Copper micro-channel production Mass flux in the micro-channels Working fluids Project objectives Prototype design and fabrication Stages Working fluid selection Pressure loss analysis Prototype fabrication... 44

6 Prototype construction Pressure resistance and leak prevention tests Wind tunnel fabrication Computer simulations and analytical analyses Computer simulations Gambit Fluent D general model Fluid and material properties Simulation results Air inlet at 0.5 m/s Air inlet at 1.0 m/s Analytical analyses Flow over a flat plate analysis Extended surface analysis Summary of results Experimental Stage Propane charging process Experimental set-up Experimental runs Experimental Results Experimental runs specific volume m 3 /kg... 76

7 Summary specific volume m 3 /kg Experimental runs specific volume m 3 /kg Summary specific volume m 3 /kg Experimental runs specific volume m 3 /kg Summary specific volume m 3 /kg Experimental runs prototype without propane Micro-channel tube bottom level temperature ~50 C - air flow at 0.5 m/s Micro-channel tube bottom level temperature ~50 C - air flow at 1.0 m/s Experimental heat transfer coefficient and heat transfer calculations Energy balance Thermosyphon effect comparison Chapter 6: Summary and Conclusions Future work References Appendix A: Propane Properties Used in Pressure Loss Analysis Appendix B: Prototype components specifications Appendix C: data for the Extended Surface Analysis Appendix D: Standard Operational Procedure Appendix E: Safety Evaluation Report

8 8 LIST OF TABLES Page Table 1: Various Working Fluids for Heat Pipes and Thermosyphons (Faghri 1995, The Chemical Engineers' Resource Page 2008) Table 2: Mass Flow and Mass Flux Estimations Table 3: Estimated Reynolds Numbers and Friction Factors for Propane at -30 C Table 4: Estimated Reynolds Numbers and Friction Factors for Propane at 0 C Table 5: Estimated Reynolds Numbers and Friction Factors for Propane at 50 C Table 6: Minor Pressure Loss Coefficients (K) Table 7: Pressure Losses Analysis Results for Propane at -30 C (Equation 6) Table 8: Pressure Losses Analysis Results for Propane at 0 C (Equation 6) Table 9: Pressure Losses Analysis Results for Propane at 50 C (Equation 6) Table 10: Boundary Conditions Used for Fluent Simulations Table 11: General Model Under-relaxation Factors Table 12: Air Properties used in Simulations Table 13: Wood and Copper Properties used in Simulations Table 14: Flow over a Flat Plate Analysis Results Table 15: Data from Experimental Run without Propane Table 16: Extended Surface Analysis Results Table 17: Heat Transfer Coefficient and Heat Transfer Results Table 18: Thermocouple Calibration Test (Air Flow and Micro-channel Tubes) Table 19: Thermocouple Calibration Test (Complete Set-up)

9 9 Table 20: Results for Experimental Runs with Propane at m 3 /kg Table 21: Results for Experimental Runs with Propane at m 3 /kg Table 22: Results for Experimental Runs with Propane at m 3 /kg Table 23: Experimental Heat Transfer Coefficient and Heat Transfer Results Table 24: Energy Balance Results Table 25: Heat Transfer Comparison for Thermosyphon and Classic Fin Operation Table 26: Propane Properties Used in Pressure Loss Analysis Table 27: Data Used for the Extended Surface Analysis with Air Flow at 0.5 m/s Table 28: Data Used for the Extended Surface Analysis with Air Flow at 1.0 m/s

10 10 List of Figures Page Figure 1: Parallel-flow Heat Exchanger Schematic Figure 2: Temperature Profile for Parallel-flow Heat Exchanger Figure 3: Counter-flow Heat Exchanger Schematic Figure 4: Temperature Profile for Counter-flow Heat Exchanger Figure 5: Cross-flow Heat Exchanger (Radiator) Schematic (Oliet, et al. 2007) Figure 6: Condensation and Evaporation in a Heat Exchanger Figure 7: Perkins Tube Schematic (Abel Chuang 2003) Figure 8: Thermosyphon and Heat Pipe Schematics Figure 9: VL800 Vectra Loop Thermosyphon (Beitelmal and Patel 2002) Figure 10: Copper Micro-channel Tube Example (Vaitkus 2008) Figure 11: Copper Micro-channel Loop Thermosyphon Prototype Figure 12: Pressure-enthalpy Diagram for Propane (Reynolds 1979) Figure 13: Change in Quality of Propane throughout a Complete Cycle. In the Saturation Diagrams the X-axis is the Enthalpy (kj/kg) and the Y-axis is the Pressure (MPa). Details are shown in Figure Figure 14: Actual Prototype Figure 15: Pressure Resistance and Leak Prevention Set Up Figure 16: Small-scale Wind Tunnel Figure 17: 3D Model Mesh in Gambit Figure 18: 3D Model with Colored Boundary Conditions in Fluent

11 11 Figure 19: Contours of Velocity in the Z Direction (Inlet = 0.5 m/s) Figure 20: Contour of Static Temperature for Inlet = 0.5 m/s (ºC) Figure 21: Contours of Velocity in the Z Direction (Inlet = 1.0 m/s) Figure 22: Contour of Static Temperature for Inlet = 1.0 m/s (ºC) Figure 23: Flow over a Flat Surface Schematic Figure 24: Extended Surface Schematic Figure 25: Temperature Comparison between Experimental Data and Equation 20 for Air at 0.5 m/s Figure 26: Temperature Comparison between Experimental Data and Equation 20 for Air at 1.0 m/s Figure 27: Reaching Charging Conditions Set-up Figure 28: Experimental Set-up Figure 29: Type-K Thermocouples attached to Loop Thermosyphon System (Air Inlet/Outlet not shown) Figure 30: Data for Experimental Run (0.007 m 3 /kg, evaporator at ~50 C, 0.5 m/s) Figure 31: Data for Experimental Run (0.007 m 3 /kg, evaporator at ~50 C, 1.0 m/s) Figure 32: Data for Experimental Run (0.007 m 3 /kg, evaporator at ~70 C, 0.5 m/s) Figure 33: Data for Experimental Run (0.007 m 3 /kg, evaporator at ~70 C, 1.0 m/s) Figure 34: Data for Experimental Run (0.009 m 3 /kg, evaporator at ~50 C, 0.5 m/s) Figure 35: Data for Experimental Run (0.009 m 3 /kg, evaporator at ~50 C, 1.0 m/s) Figure 36: Data for Experimental Run (0.009 m 3 /kg, evaporator at ~70 C, 0.5 m/s) Figure 37: Data for Experimental Run (0.009 m 3 /kg, evaporator at ~70 C, 1.0 m/s)

12 12 Figure 38: Data for Experimental Run (0.011 m 3 /kg, evaporator at ~50 C, 0.5 m/s) Figure 39: Data for Experimental Run (0.011 m 3 /kg, evaporator at ~50 C, 1.0 m/s) Figure 40: Data for Experimental Run (0.011 m 3 /kg, evaporator at ~70 C, 0.5 m/s) Figure 41: Data for Experimental Run (0.011 m 3 /kg, evaporator at ~70 C, 1.0 m/s) Figure 42: Data for Experimental Run No Propane (0.5 m/s) Figure 43: Data for Experimental Run No Propane (1.0 m/s) Figure 44: Copper Micro-channel Loop Thermosyphon Components Figure 45: Manifold (units-mm) Figure 46: T-Elbow Connector (units-mm) Figure 47: Return (units-mm) Figure 48: U-T Connector (units-mm) Figure 49: Copper Micro Channel Loop Thermosyphon Prototype Figure 50: Keithley Data Acquisition System (Tutorial) Figure 51: 1 x 8.25 Wind Tunnel Figure 52: 21.5-Quarts Metal Container Figure 53: Electric Variac Unit Figure 54: Temperature Control Unit Figure 55: 5% Lower Explosive Level (LEL) Portable Gas Leak Detector Figure 56: Propane Filling System (A) Figure 57: Propane Filling System (B)

13 13 LIST OF SYMBOLS Heat transfer to the air (W) Mass flow (kg/s) Latent heat of vaporization (kj/kg) Reynolds number for a pipe or tube (dimensionless) Fluid density (kg/m 3 ) Fluid velocity (m/s) Hydraulic diameter (m) Micro-channel area (m 2 ) Micro-channel perimeter (m) Dynamic viscosity (kg/m-s) Pressure loss by friction (Pa) Friction factor (dimensionless) Material roughness (µm) Pipe or tube length (m) Minor pressure loss (Pa) Minor loss coefficient (dimensionless) Pressure head (Pa) Pressure effect by buoyancy (Pa) Capillary pressure (Pa) Net pressure (Pa) Gravity (m/s 2 ) Micro-channel tube height (m) Gas specific weight (N/m 3 ) Micro-channel tube perimeter (m) Surface tension (N/m) Hydraulic radius (m) Micro-channel surface length (m)

14 Surface temperature ( C) Mean temperature of the micro-channel tubes wall and the free air stream ( C) Thermal conductivity (W/m- C) Kinematic viscosity (m 2 /s) Prandtl number (dimensionless) Reynolds number for a flow over a surface (dimensionless) Nusselt number for a flow over a surface (dimensionless) Average heat transfer coefficient over a surface (W/m 2 - C) Flat plate surface area (m 2 ) Temperature difference between flat plate and air flow ( C) Temperature of the air flow ( C) Temperature at the base of the fin ( C) Temperature at the middle of the fin ( C) Temperature at the end of the fin ( C) Fin width (m) Fin thickness (m) Fin cross sectional area (m 2 ) Extended surface (fin) perimeter (m) Average heat transfer coefficient (W/m 2 - C) Temperature difference between location x at the fin and the air flow ( C) Temperature at location x ( C) Temperature difference between the fin base and the air flow ( C) Temperature difference between the fin end and the air flow ( C) Heat capacity (kj/kg- C) Outlet air temperature ( C) Inlet air temperature ( C) Outside surface area of the micro-channel tubes (m 2 ) Average temperature on the micro-channel tubes surface ( C) 14

15 15 Energy entering the system (W) Energy leaving the system (W) Energy accumulated within the system (W) Input voltage (V) Heater resistance (Ω) Watts Meters Pascal Degree Celsius Kilo grams Kilo Joules Seconds Liquefied petroleum gas with methylacetylene-propadiene

16 16 1. INTRODUCTION A heat exchanger is a thermal device that transports energy by heat transfer between moving fluids at different temperatures. Heat exchangers are widely used in many different applications and industries worldwide, varying from space, energy, heating, ventilation and air conditioning, and transportation. A radiator for cooling a car or truck engine, and a condenser for an air conditioner are typical heat exchangers in the automobile industry. This work is focused on developing a high-efficiency heat transfer component by using copper micro-channel tube in a thermosyphon system. The overall configuration could be used in condensers, plate-fin heat exchangers, or heat sinks in a variety of applications. 1.1 Heat exchangers Heat exchangers can be classified in many ways such as construction type, heat transfer processes, surface compactness, phase of the fluids, etc.. In this document, the classification used to describe the heat exchanger is by the flow arrangement of the fluids. When heat exchangers are classified by the flow arrangement, they can be divided into three main groups: parallel flow, counter flow, and cross flow (Kuppan 2000). The following sections briefly describe each one of the three main groups Parallel-flow heat exchanger A parallel-flow heat exchanger is where both the hot and the cold fluids travel parallel to each other by entering and leaving the heat exchanger at the same ends, so

17 17 they travel in the same direction. When comparing a heat exchanger with a parallel-flow arrangement with any other single pass heat exchanger with the same flow conditions (i.e., same entrance and exit flow temperatures, same mass flow for the hot and cold fluid), the parallel-flow heat exchanger has the lowest heat transfer effectiveness (for non condensing fluids). Although it has the lowest effectiveness, the parallel heat exchanger is preferred where the fluid temperature exceeds 1100ºC, when the warmer fluid must not reach the freezing point, for boiling applications, and for balanced heat exchanger (where the product of the heat capacity and mass flow is equal for both fluids) (Kuppan 2000). Figures 1 and 2 show the parallel-flow heat exchanger schematic and the temperature profile, respectively. Figure 1: Parallel-flow Heat Exchanger Schematic.

18 18 Figure 2: Temperature Profile for Parallel-flow Heat Exchanger Counter-flow heat exchanger A common definition for counter-flow heat exchanger is where both hot and cold fluids travel parallel to each other but in opposite directions. At one end of the heat exchanger, one fluid will enter the device while the other leaves the device, and a similar situation occurs at the other end. The counter-flow arrangement has the highest heat transfer effectiveness compared to other single pass heat exchangers with the same flow conditions (Kuppan 2000). Figures 3 and 4 represent the counter-flow heat exchanger schematic and temperature profile, respectively.

19 19 Figure 3: Counter-flow Heat Exchanger Schematic. Figure 4: Temperature Profile for Counter-flow Heat Exchanger Cross-flow heat exchanger A common definition for cross-flow heat exchanger is where both hot and cold fluid travel perpendicular to each other. With this configuration, many variations are possible, depending on construction constraints. Some of the variations can be each of the hot and cold fluids mixed, only one of the fluids mixed, and both fluids unmixed (Incropera and DeWitt 2002). Figure 5 represents one of the most common applications of cross-flow heat exchangers, an automobile engine radiator. Figure 6 represents the

20 20 temperature profiles of two special cases for heat exchangers: condensation and evaporation of a fluid. In this situation, when fluids with large latent heat values are used in the design; the heat exchange can be increased by keeping the heat exchange surface at a high or low constant temperature. The temperature profile in this situation, shown in Figure 6, is similar to what is expected to happen in the current heat exchanger component design. Kuppan (2000) mentions that, although the cross-flow heat exchanger has lower heat transfer effectiveness than the counter-flow heat exchanger, it is preferred when the heat exchanger unit has extended surfaces in its design. Simpler design and fabrication of the heat exchanger headers in a cross-flow heat exchanger provides an advantage over the other heat exchanger types (Kuppan 2000). Figure 5: Cross-flow Heat Exchanger (Radiator) Schematic (Oliet, et al. 2007).

21 21 Figure 6: Condensation and Evaporation in a Heat Exchanger. 1.2 Thermosyphon and heat pipe Thermosyphons and heat pipes are known for being part of the most efficient heat transfer devices today. The first design of a device working on the principle of a thermosyphon was developed by Jacob Perkins in 1836 in the United Kingdom, and it was called the Perkins Tube (Perkins 1836). A thermosyphon is similar to the Perkins Tube (Figure 7) since both designs need to have the evaporator section below the condenser section and use gravity as the driving force. The working principle of heat pipes and thermosyphons is that heat is transferred to the fluid in the evaporation section. This causes the fluid to boil and the vapor flows to the opposite (condensing) end where the temperature is lower. Heat is then extracted from the fluid in the condensing section. This leads to condensation in the cooler section, and the liquid flows back to the evaporation section. In heat pipes, a wick is used to provide the capillary action for the flow of the liquid.

22 22 Figure 7: Perkins Tube Schematic (Abel Chuang 2003). This evaporation and condensation process is repeated if the following condition is fulfilled. The temperature difference between the condensation and evaporation section reaches the sum of the temperature differences required to transfer heat through the pipe wall and the wick in both sections. In addition, the temperature field must create a vapor pressure which will make the vapor flow to the condensation section (Mills 1992). Thermosyphons, because they do not have a wick structure, can be used in applications where the condensation area is located above the evaporation area. In such cases, the condensed liquid can flow back to the evaporator with the help of gravity. Therefore, the capillary pumping action of the wick structure is not needed (see Figure 8).

23 23 Figure 8: Schematic of the Thermosyphon and Heat Pipe. Several complex designs have been achieved throughout the years using heat pipes and thermosyphons. Examples include the loop heat pipe and the loop thermosyphon. The main application of these loop thermosyphons have been in heat sinks for cooling of electronic components. For example, in 2001 Hewlett-Packard developed a loop thermosyphon prototype for one of their processors. The prototype was designed to handle a heat load of 80 W (see Figure 9). Different research groups have performed experiments on the following performance effects: inclination effects (Beitelmal and Patel 2002) and different working fluids (Pal, et al. 2002). These designs are intended to increase efficiency and decrease vibrations and maintenance costs (Hsu 2008). For this research project, the heat exchanger design prototype will emulate the loop thermosyphon operating principle with the added modification of a

24 24 fin containing a micro-channel array. The micro-channel array fins are made of copper for superior performance due to the high thermal conductivity. Typically these arrays are referred to as micro-channel tube in the industry. In this thesis, they will be also be referred to as micro-channel tube. Figure 9: VL800 Vectra Loop Thermosyphon (Beitelmal and Patel 2002). 1.3 Micro-channel tubes Although the idea of utilizing a wick structure to provide the capillary pumping in the heat pipe application goes back to the 1940s, it was not until 1984 that the combination of micro scale (diameter range mm) fabrication and twophase heat transport was proposed (Cotter 1984). Due to their geometry, microchannel tubes are devices with a high surface to volume ratio. This characteristic allows heat exchanger designers from the refrigeration industry to use less refrigerant mass in their applications, reduce costs, and increase performance in their units. For example, Delphi has designed condensers and evaporators that have aluminum micro-

25 25 channel tubes. The company claims their products are highly efficient and percent smaller and lighter than competitive products (Delphi Corporation 2008). An important characteristic of micro-channel tubes is that they increase both the heat exchange and efficiency of the heat exchanger designs. This characteristic allows the micro-channel heat exchanger to be smaller and yet have the same performance as a regular heat exchanger, or to get improved performance in the same volume as a conventional heat exchanger, as explained by the American Council for an Energy- Efficient Economy in its report of emerging technologies and practices (Sachs, et al. 2004) Copper micro-channel production Most micro-channel tubes are made of aluminum alloys since they are easier to fabricate than copper. Copper presents some advantages over aluminum. The main advantage for heat transfer applications is its thermal conductivity (401 W/m- C), which is almost double the thermal conductivity of aluminum (237 W/m- C). Another advantage that copper presents is that it is easier to join and repair than aluminum. Combining the properties of copper with the advantages of micro-channels seemed logical; but, until recently, no feasible process was developed to mass produce copper micro-channel tubes. Currently Ohio University is conducting research, led by Dr. F. Kraft, in order to develop a feasible method to produce copper micro-channel tubes. During this research, a novel copper micro-channel production procedure was developed. This procedure is based upon a multi-billet forward extrusion process which eliminates

26 26 the need for the billet to separate into different streams, thereby eliminating the work required for this flow separation and lowering stress on the die bridge (Vaitkus 2008). Extrusion dies have been designed and fabricated to produce copper micro-channel tubes, and research is being continued to improve and modify certain components of the extrusion process for mass production. Figure 10 shows a cross section of a copper micro-channel tube with its dimensions (as measured from a specific sample). This is one of the micro-channel tubes that has been produced at Ohio University, and is representative of the tubes used in this research project. Most micro-channel tubes consist of an array of tubes in a ribbon of solid material. For example, the copper micro-channel array shown here has nine tubes in the array. Figure 10: Copper Micro-channel Tube Example (Vaitkus 2008).

27 Mass flux in the micro-channels The mass flux through the micro-channels is influenced by the amount of heat rejected by the tubes. Since the working fluid condenses as it passes the channels, the amount of thermal energy ( ) rejected can be found from the following relationship: (1) where is the working fluid mass flow and is the working fluid latent heat of vaporization. In this research, the heat load applied to the system was estimated to be 50 Watts. To design the thermosyphon loop, heat losses and convection limitations require careful considerations. In this calculation, assumptions must be made concerning the heat rejected by the micro-channel tubes to the flowing air. The range was selected by assuming that the micro-channel tubes can reject the entire heat load input (ideal case, 50 Watts) as the upper limit, and rejecting around 20% of the heat load (assuming heat losses through the system, 10 Watts) as the lower limit. With these assumed limits, and applying Equation 1, the fluid mass flow range can be estimated, as shown in Equation 2. (2) The mass flux (kg/m 2 s) can then be estimated by dividing the mass flow by the total micro-channel cross sectional area. 1.4 Working fluids When searching for a working fluid for a thermosyphon or heat pipe, the temperature range of the application must be taken into account. This is due to the differences in the melting and boiling points of working fluids that can be used. Table

28 28 1 contains various working fluids with data of their critical temperature and pressure values, and the application range (Faghri 1995, The Chemical Engineers' Resource Page 2008). In addition to the phase change temperatures and the application range, choosing a working fluid is also influenced by the global warming potential of some of these fluids. The global warming potential can be defined as the capability of a gas to trap heat in the atmosphere over a 100-year period (Environmental Protection Agency 2006). For this reason a low global warming potential is preferred; other considerations include toxicity, availability, and ease of storing and handling. Table 1: Various Working Fluids for Heat Pipes and Thermosyphons (Faghri 1995, The Chemical Engineers' Resource Page 2008). Working Fluid Critical Temperature (ºC) Critical Pressure (MPa) Application Range (ºC) Ammonia to 100 Freon to 100 Acetone to 120 Ethanol to 130 Methanol to 130 Water to 200 Carbon Dioxide to 25 Propane to 90 (estimated) 1.5 Project objectives The main objectives of this research were to design and test a small efficient heat exchanger working on the loop thermosyphon principle. The heat exchanger was tested in a heat sink configuration, i.e., the micro-channel tubes were used as thermosyphons to

29 29 dissipate energy from a heat source. The goal was to use copper micro-channel tubes as ideal fins to provide an isothermal surface for heat transfer to air. A pressure loss analysis was performed to ensure the working fluid circulation within the system. The research was focused on proving the loop thermosyphon design concept and to validate the improvement in performance of an isothermal fin surface compared to a classical fin. The novel design developed in this study combines the advantages of thermosyphons and a micro-channel heat exchanger as follows: (i) The thermosyphon design eliminates the need to pump fluid through the microchannels. Since the pressure drop in micro-channels is quite high, the high pumping power for micro-channel flow is eliminated. (ii) The thermosyphon effect produces a uniform temperature in the micro-channel fins. This increases the fin efficiency to almost 100%. In comparison, an aluminum fin may have only 60% efficiency. Therefore, an increase in efficiency of about 60% over the classical fin can be expected. Unlike the conventional heat exchanger, thermosyphons transfer heat by phase change, so greater heat flux is achieved. The following steps were completed in order to satisfy the research objectives: Design stage: o A pressure loss analysis was performed on the system to ensure the working fluid circulation at a temperature range from -30ºC to 50ºC. o Available copper tubing/fittings were used and dimensions were selected with due consideration to pressure and temperature range.

30 30 o The geometry of the prototype was designed so it could be tested in a wind tunnel. Fabrication stage: o The prototype was built with the dimensions obtained from the design stage. o Pressure resistance and leak tests were performed on the prototype joints. o A small scale wind tunnel was built to experimentally test and record temperature data to validate the thermosyphon behavior of the prototype. Analytical estimations: o A three dimensional model was prepared with Fluent software and the flow across the micro-channel tubes was simulated to obtain heat transfer coefficient and heat transfer estimates. o An extended surface (fin) analysis and flow over a flat surface analysis were performed to obtain heat transfer coefficient and heat transfer estimates. Experimental estimations: o Heat transfer coefficient and heat transfer were estimated from the experiments and compared with the results of the analytical estimations.

31 31 o The micro-channel tubes classical fin behavior was compared with the thermosyphon behavior to validate the benefits of the design.

32 32 2. PROTOTYPE DESIGN AND FABRICATION STAGES This chapter covers all the aspects related to the copper micro-channel loop thermosyphon design. The initial concept was to use a bank of heat pipes or thermosyphons (arranged and staggered configurations) where the evaporation section would be exposed to the hot fluid, and the condenser to the cool fluid in the system. A primary application could be an automobile radiator. In such an application, the thermosyphon principle could be readily used because of the following: Micro-channel tubes would increase the contact area with the working fluid. The thermosyphon would have a stable vertical orientation. It would be possible to use gravity as the driving force for the circulation of the working fluid. A patent disclosure on this idea had been submitted by committee members (Dr. Alam and Dr. Urieli) in 2008 and this project was then undertaken. Changes were made to the initial design until a final decision was taken to design a loop thermosyphon. The design of the loop thermosyphon was assisted by discussions with Mr. J. Edwards of Global Cooling (Edwards 2009). The prototype (Figure 11) consists of three copper micro-channel tubes, four U- bend copper tubes, three copper tees, two copper elbows, and copper connecting tubes As discussed earlier, copper micro-channel tube is a new product being developed at Ohio University. Therefore, the supply of these was limited at the design stage. Except for the copper micro-channel tubes, all the other components were obtained commercially. These connecting tube components are made of 12.7 mm copper pipes (drawn-type L), and the

33 fittings are copper fittings for the specified pipes. Details related to the components can be found in the prototype fabrication section (2.4). 33 Figure 11: Copper Micro-channel Loop Thermosyphon Prototype. 2.1 Working fluid selection A variety of fluids were considered. The main requirements were the operating temperature range, availability, global warming potential, and possible safety and health hazards. Since this was to be tested in the heat sink configuration, the desired operating range for this project was established to be from -30ºC to 50ºC. Refrigerant 744 (CO 2 ) was an option since its global warming potential (GWP) is unity. R744 had to be discarded because its critical temperature (T cr ) is below 50ºC (T cr = 31.1ºC). Thus the thermosyphon effect would be lost in working conditions above the critical point. Ammonia seemed to be a good candidate, but because of its corrosiveness and toxicity, it was also discarded. Although propane is not a common working fluid for thermosyphons, it was chosen for this project. The selection was based on the

34 34 following criteria. It satisfies the operating temperature range, it has a relatively low global warming potential (3) (ToolBox 2005), it is readily available and, it is not toxic. For the research, 99.5% pure propane was obtained. Figure 12 shows a pressureenthalpy (P-h) diagram for propane, where the typical charging test conditions and the operating range for the design are specified. It is also necessary to select the test conditions for the charging of the system to be out of the saturation region; since the pressure and temperature in a single phase provides a measurement of the specific volume of the working fluid in the system.

35 Figure 12: Pressure-enthalpy Diagram for Propane (Reynolds 1979). 35

36 Pressure loss analysis The approach taken to ensure the working fluid circulation within the loop thermosyphon design is explained in this section. Once a heat load is applied to the evaporator section, propane will start evaporating and since the gas is lighter than the liquid it will tend to rise. To ensure the working fluid circulation the following conditions must be satisfied: have a liquid column (to provide a pressure head) which will overcome the friction losses on the return path, which includes the losses in the micro-channel tubes and other minor losses. In addition to pressure losses by friction, the pressure head, buoyancy, and capillary forces need to be taken into account. Figure 13 shows how the propane travels along the prototype with lines of different colors representing the different sections taken into account for the pressure loss analysis. In addition, the figure shows each stage, with their propane quality (vapor mass percent in the saturated fluid) values estimated for the analysis. In this figure, the propane cycle within the saturation zone is assumed to be at a temperature of approximately 50ºC. The orange lines represent the evaporator section where the propane is going from a quality of 0 to 1. The red lines represent the return path for the propane where the fluid is assumed to be saturated vapor (quality = 1). The blue lines represent the condenser section (micro-channels) where the propane is going from a quality of 1 back to 0 (saturated liquid).

37 37 Figure 13: Change in Quality of Propane through a Complete Cycle. In the Saturation Diagrams the X-axis is the Enthalpy (kj/kg) and the Y-axis is the Pressure (MPa). Details are shown in Figure 12. Theoretically, the application of different heat loads at the evaporator section will cause different mass fluxes. Since the system capacity can be tested only during the experimental stage of the project, the mass flux was not known in the design phase. The mass flux is an important part of the pressure losses analysis. Using the mass flux and the fluid density, the velocity of the fluid can be estimated. The Reynolds number, the friction factor, and the pressure losses by friction and minor losses can then be calculated. Due to the unpredictability of the mass flux, a range of mass fluxes were estimated from Equation 2 and the total micro-channel cross sectional area. Since propane was chosen as

38 38 the working fluid for the system, its latent heat (h fg ) value at 50ºC was used for the calculations (h fg = kj/kg). The cross sectional area per channel is 1 mm 2, each tube contains nine of these channels, and by having three tubes, the total cross sectional area is 27 mm 2. The mass flux range for the prototype was chosen based on Table 2. Table 2: Mass Flow and Mass Flux Estimations. Q (Watts) (rejected at tubes) Mass flow (kg/s) (total) Mass flux (kg/m 2 s) (per channel) x x Each mass flux value (1 and 6 kg/m 2 s) was analyzed at different operating temperatures (-30 C, 0 C and 50 C). The analysis was performed using a Microsoft Excel spreadsheet. The input cells in the spreadsheet are the mass flux, micro channel hydraulic diameter (d h ), number of channels, saturation properties of propane (dynamic viscosity, specific volume, and estimated quality at each section), copper tube diameter, and estimated minor loss coefficients (for the properties used for the analysis see Appendix A). The hydraulic diameter is given by, (3) where is the micro-channel cross sectional area and is the microchannel perimeter. The equations used to estimate the pressure losses are presented by White (1999). Equation 4 is used to calculate the Reynolds number, a dimensionless quantity, which is then used to define the flow as laminar or turbulent.

39 39 (4) where is the fluid density, is the fluid velocity, is the hydraulic diameter and is the dynamic viscosity. For a circular pipe Re d <2300 is considered laminar and Re d >4000 is considered turbulent. Reynolds numbers larger than 2300 and smaller than 4000 are considered to be in transition. The pressure loss (P friction ) by friction was estimated with Equation 5: (5) The friction factor (f ) is estimated using the Moody s diagram, the Reynolds number, and the relative roughness (ε/d) for turbulent flows. For laminar flows, only the Reynolds number is needed in order to find the friction factor (f = 64/Re d ). When a flow is in transition, no reliable friction factor can be calculated. To estimate the friction pressure losses when a transition flow is obtained, both methods (laminar and turbulent flow) were applied and the larger estimated value was chosen. This approach ensured that the pressure losses were not underestimated. For this project, the roughness value used for the copper tubes was taken to be 1.5 µm (ToolBox 2005). Tables 3-5 give the calculated Reynolds numbers and the estimated friction factors for the specific operating temperatures, mass fluxes per micro channel, and Reynolds numbers at the different prototype sections. It can be concluded that the flow in the micro-channels is always laminar because of its small hydraulic diameter.

40 40 Table 3: Estimated Reynolds Numbers and Friction Factors for Propane at -30 C. Mass flux kg/m 2 s (per channel) Micro channel Bottom Manifold Return Path Top Manifold Re d f Re d f Re d f Re d f Table 4: Estimated Reynolds Numbers and Friction Factors for Propane at 0 C. Mass flux kg/m 2 s (per channel) Micro channel Bottom Manifold Return Path Top Manifold Re d f Re d f Re d f Re d f Table 5: Estimated Reynolds Numbers and Friction Factors for Propane at 50 C. Mass flux kg/m 2 s (per channel) Micro channel Bottom Manifold Return Path Top Manifold Re d f Re d f Re d f Re d f In addition to pressure losses by friction, minor pressure losses (P minor ) were taken into account. These losses are due to sudden expansions, contractions, and changes in flow directions (elbow, U-bend and tee fittings). Equation 6 (White 1999) was used to estimate the minor pressure losses (P minor ). (6)

41 41 K i represents the minor losses coefficient and it is a dimensionless value used to quantify the minor pressure losses in a system. For the project prototype, six different minor loss coefficients were estimated. These were for the sudden expansion (from the micro channel to the copper tube), for the U-bends, for the tees with a branched flow, for the tee with a through flow, for the 90-degree elbows, and for the sudden contraction (from the copper tube to the micro channel). Table 6 contains the estimated values for the minor pressure loss coefficients. Table 6: Minor Pressure Loss Coefficients (K). Condition Sudden expansion (from micro channel to copper tube) Minor pressure loss coefficient 1.0 U-bend fitting 1.5 Tee (branched flow) degree elbow 1.5 Tee (through flow) 0.9 Sudden contraction (from copper to micro channel tube) 0.4 Estimation source Equation (White 1999) The Engineering ToolBox (ToolBox 2005) The Engineering ToolBox (ToolBox 2005) The Engineering ToolBox (ToolBox 2005) The Engineering ToolBox (ToolBox 2005) Equation (White 1999) To finalize the pressure losses analysis, buoyancy and capillary forces were calculated in order to evaluate their effects within the system. As explained in the previous section, the pressure head provided by the liquid column needs to overcome the friction and minor losses in addition to buoyancy and capillary effects (see Equation 7).

42 42 The buoyancy and capillary effects are taken into account since they can prevent the liquid from flowing from the copper micro-channels into the evaporator. The pressure loss equation can be written as: (7) For the design to properly operate the net pressure (P net ) must be positive. The pressure head provided by the liquid column was estimated with Equation 8, where ρ represents the fluid density, g represents the gravitation constant (9.81 m/s 2 ), and h represents the fluid column height. (8) Some propane, in the vapor phase, may prevent the flow of liquid propane out of the micro-channel by buoyancy forces against the liquid propane. For this reason the pressure exerted, by the gas to the liquid, at the exit of the micro-channel is estimated using Equation 9 (Edwards 2009). (9) The specific weight (γ) of the propane, in the gas form, is multiplied by the channel perimeter (m), resulting in an estimate of the pressure at that location. The capillary effects were estimated using the surface tension (σ) values at the different temperature conditions of the fluid, and divided by the hydraulic radius of the micro-channel. Equation 10 (Kraus and Bar-Cohen 1983) provides an estimate of the pressure difference between the liquid and vapor phases of a fluid. (10)

43 43 The results for the pressure loss analysis are shown in Tables 7-9. The information contains the results of P net at the different temperatures and mass fluxes per channel. It was expected that the pressure losses would increase, due to both friction and minor pressure losses, as the mass flux was increasing. The increment is mainly due to the increase in velocity of the flow (see Equation 5 and Equation 6). Since the head provided by the liquid column remains constant, the upper limit of operation must satisfy Equation 7. As the results show, the design fulfills the conditions to provide flow circulation within the loop thermosyphon system. Table 7: Pressure Losses Analysis Results for Propane at -30 C (Equation 6). Mass flux kg/m 2 s (per channel) P head (Pa) P friction (Pa) P minor (Pa) P buoyancy (Pa) P capillary (Pa) P net (Pa) Table 8: Pressure Losses Analysis Results for Propane at 0 C (Equation 6). Mass flux kg/m 2 s (per channel) P head (Pa) P friction (Pa) P minor (Pa) P buoyancy (Pa) P capillary (Pa) P net (Pa)

44 44 Table 9: Pressure Losses Analysis Results for Propane at 50 C (Equation 6). Mass flux kg/m 2 s (per channel) P head (Pa) P friction (Pa) P minor (Pa) P buoyancy (Pa) P capillary (Pa) P net (Pa) Prototype fabrication The prototype was fabricated with special emphasis in joining techniques, so leaks were eliminated. After fabrication, the prototype was thoroughly tested for leaks and pressure resistance Prototype construction The prototype was built by Mr. R. Mulford, the Mechanical Engineering Machine Shop and Lab Coordinator. All the copper couplings were procured and the copper pipes were cut to their specific length. For all dimensions of the prototype components, refer to Appendix B. The process used to join all the components of the prototype was silver brazing. Silver brazing was chosen instead of soft soldering because of its higher mechanical strength. Due to the thin walls of the top and bottom manifolds (0.1 mm), the copper micro-channel tubes were pushed in slightly into the top and the bottom manifolds to provide greater contact area for bonding strength. To ensure that the micro-channel joints will sustain both temperature and pressure variations without failure or leaks, silver brazing was the preferred choice. During the brazing process, rings of Silvaloy 35 (Ag 35%, Cu 26%, Zn 21% and Cd 18%) were used to fit all the joints. The rings were made

45 45 by bending the material around smaller tubing than the prototype tubing to be joined. This allowed the rings to be tight during the heating stage. The joints were fluxed to remove oxides from the material. Finally, the joints were heated with MAPP gas (propyne 30%, propadiene 14%, propylene 43%, propane 7%, and butane 6%) in a torch to have the rings melted and fill the joints by capillary action (Mulford 2009). The prototype is shown in Figure 14. As shown in the figure, an extra copper tube was added to the 3D computer design for charging. Since the prototype needs to be filled multiple times with the working fluid for different experiments, a charging system has to be included to the system. The charging system contains a refrigeration Schrader valve and a tee section to attach a pressure gauge. The Schrader valve is used to evacuate and fill the system, while the pressure gauge is used to monitor the pressure during the charging process. Figure 14: Actual Prototype.

46 Pressure resistance and leak prevention tests Pressure tests were performed to validate that the material did not lose its strength properties and would withstand the pressure loads during the experimental stage. The prototype was connected to a CO 2 tank and filled with the gas to an initial pressure of 5.5 MPa at ambient temperature (~18ºC). The first test consisted of leaving the prototype charged for 24 hours and search for any decrease in pressure or leaks in all the joints. The first test was completed successfully since neither a pressure decrease nor any leaks were found. The second test consisted of heating the prototype to test the heating effects on the joints. The temperature was raised to 200ºC, using two propane torches and heating the prototype evenly. This increased the pressure to 6.9 MPa. During the test, the pressure and temperature were kept constant for 20 minutes to asses any pressure drops or leaks in the joints. The second test was also successful. Figure 15 shows the prototype connected to the CO 2 system and the thermocouple.

47 47 Figure 15: Set Up for Pressure Resistance and Leak Prevention. 2.5 Wind tunnel fabrication In order to test the performance of the prototype, a special, small-scale wind tunnel was fabricated. The wind tunnel was fabricated with wood. Since wood has a low thermal conductivity (0.173 W/m-ºC), it acts as a thermal insulator in addition to providing the desired cross sectional area. The cross section for the wind tunnel is 25.4 mm x 222 mm. This cross sectional area provides approximately 12.7 mm on each side of the micro-channel tubes, providing a smaller area for the air flow, so that the temperature difference between inlet and outlet air would be at least a few degrees. Figure 16 shows the wind tunnel. A flow meter (rotameter) was attached to the wind tunnel, as well as a base to support the prototype (at the point where the pressure gauge and charging valve

48 48 are connected). The air flow to the wind tunnel is provided from a pressurized air line. The red hose is connected to the air line at one end, and to the inlet of the flow meter at the other end. The yellow hose is then connected from the outlet of the flow meter to the entrance of the wind tunnel. A mesh was attached to the entrance of the wind tunnel in order to distribute the air flow evenly through the wind tunnel. Since the air flow is laminar within the wind tunnel, an air mixer was placed after the thermosyphon. The air mixer consists of a mesh made of aluminum such that it mixes the outlet air so that an average temperature can be measured. Figure 16: Small-scale Wind Tunnel.

49 49 3. COMPUTER SIMULATIONS AND ANALYTICAL ANALYSES Computer simulations were performed to determine the heat transfer characteristics of the copper micro-channel tubes. The results were compared to analytical calculations. From the analyses and the simulations, two values (heat transfer coefficient and heat transfer) were estimated in order to characterize the system. The heat transfer coefficient and heat transfer were calculated for air velocities of 0.5 m/s and 1.0 m/s. The results were then compared to obtain a reliable estimate of the values. 3.1 Computer simulations As part of the research, a computer model was developed to simulate the system performance. The computational fluid dynamics (CFD) software package Fluent was used for this task. Fluent software package contains software called Gambit, which was used to prepare 3-dimensional (3D) models. A more detailed explanation of both programs is given in the following sections Gambit Gambit is software that allows users to create two-dimensional (2D), threedimensional (3D) models, and meshes. These meshes are then used in computational fluid dynamics software and other applications to solve problems numerically. Gambit contains a graphical user interface (GUI) which makes the software easier to use and to build models under study. The usual steps to follow when building a model in Gambit are the following: (1) create the geometry, (2) specify the zones, and (3) mesh the geometry.

50 50 In the first step, Gambit allows the user to construct any 2D or 3D geometry to closely represent the real model geometry. Once the geometry is completed, the user needs to specify the zones in the geometry. In this step the volumes are defined as solid or fluid. Boundary conditions must be assigned to the area surfaces of the walls and the inlets as well as the outlets. Examples of the boundary conditions are walls, pressure inlet/outlet, velocity inlet, symmetry, etc. The boundary conditions used for the simulations are explained in more detail in Section 3.1.3, and are defined in Fluent. The third step is used to define how the geometry should be meshed. At this stage, the user can define the types of elements, distance between nodes, and other features in Gambit. Gambit also allows the user to define different types of meshes for different surfaces and/or volumes Fluent Fluent is computational fluid dynamics software (Fluent Inc. 2001) that allows users to solve fluid and heat transfer problems in complex geometries. The software has the ability to model and solve many types of fluid flow problems. Once a mesh is imported to Fluent, the user has to define all the properties of the system. These properties include existing, or user-defined, properties of solid materials and fluids for the volumes in the system. For the model surfaces, the user needs to define boundary conditions such as wall temperature, velocity inlet, and pressure, within other conditions depending on the model. The properties and conditions used in Fluent for the model are explained in more detail in the following section.

51 D general model For simulations of air flowing at 0.5 and 1.0 m/s, similar geometry, mesh, material, and air properties were used. The only difference between the two simulations is the velocity inlet boundary condition assigned to the air flow. The experiments were run in the small-scale wind tunnel, so the air chamber was modeled using the crosssectional area of the wind tunnel. The 3D model volume is defined by a height of 190 mm, a width of 25.4 mm, and a length of 100 mm. The model volume was defined as a fluid, which for the simulation is the air flowing across the copper micro-channel tubes. The height was given by the length of the copper micro-channel tubes in the prototype. Figure 17 shows the 3D model mesh created for the simulation in Gambit. As shown in the top view of the model, a special mesh was created for the simulation. A Gambit feature, called sizing function, was used in order to create a mesh which started with smaller elements at the channels walls, and increased their size linearly. This type of mesh was chosen because it allows Fluent to have a higher resolution in the area of interest, which in this case are the micro-channel tubes.

52 52 Figure 17: 3D Model Mesh in Gambit. The 3D model contains approximately 1.3 million elements throughout the mesh. Once the model was built, and the boundary conditions assigned and meshed, it was exported to Fluent (see Figure 18). The boundary conditions used in the model were Velocity Inlet (for the air flow entrance, shown in blue), Pressure Outlet (for the air flow outlet, shown in red), and Wall (for the wind tunnel walls and the micro-channels, shown in white). Table 10 contains the values used for the boundary conditions during the simulations. In the Fluent user manual, information can be found regarding the choice of the values for each boundary condition depending on the model. For example, a turbulence intensity of 1% or less is generally considered low and turbulence intensities greater than 10% are considered high (Fluent Inc. 2001, Fluent Ansys 2006). For the simulations, turbulence intensities of 8% and 5% were chosen for the inlet and outlet,

53 53 respectively. These values within the recommendations of the user manual assuming the fluid would be less turbulent at the outlet. The inlet velocity, pressures and walls were defined by the experimental conditions, listed in Table 10. Figure 18: 3D Model with Colored Boundary Conditions in Fluent. Table 10: Boundary Conditions Used for Fluent Simulations. Boundary Conditions Simulation 1 (0.5 m/s) Simulation 2 (1.0 m/s) Velocity Inlet Pressure Outlet Walls wood Walls micro-channels 0.5 m/s (22 C) Turbulence intensity 8% Atmospheric pressure Turbulence intensity 5% Constant temperature (22 C) Constant temperature (55 C) 1.0 m/s (22 C) Turbulence intensity 8% Atmospheric pressure Turbulence intensity 5% Constant temperature (22 C) Constant temperature (55 C)

54 54 In addition to boundary condition definitions, under-relaxation parameters were assigned to the model. Under-relaxation parameters are used to define how much the actual iteration will influence the next one. Although the Fluent user manual recommends using the default values since they are near optimal for the largest possible number of cases (Fluent Ansys 2006), the momentum, turbulence kinetic energy, and turbulence dissipation rate factors were slightly decreased to prevent oscillations in the solution. Table 11 contains the values used for the model. Table 11: General Model Under-relaxation Factors Under-relaxation Parameter Factor Pressure 0.3 Density 1 Body Forces 1 Momentum 0.5 Turbulence Kinetic Energy 0.4 Turbulence Dissipation Rate 0.4 Turbulent Viscosity 1 Energy Fluid and material properties All the material properties used in the simulations are the default properties assigned by Fluent, except for the copper thermal conductivity. The micro-channel tubes are made of 99.99% copper, which has a thermal conductivity of approximately 401

55 55 W/m-ºC at 55ºC (ToolBox 2005). The default value in Fluent is 387 W/m-ºC; so it was changed to 401 W/m-ºC. Table 12 shows the air properties used during the simulations. The wood and copper properties are shown in Table 13. Table 12: Air Properties used in Simulations. Thermal Density Specific Heat Viscosity Fluid [kg/m 3 Conductivity ] (C p ) [J/kg-ºC] [kg/m-s] [W/m-ºC] Air x10-5 Table 13: Wood and Copper Properties used in Simulations. Material Density [kg/m 3 ] Specific Heat (C p ) [J/kg-ºC] Thermal Conductivity [W/m-ºC] Wood Copper Simulation results Air inlet at 0.5 m/s This section presents the results obtained for the general model simulation with the air inlet velocity equal to 0.5 m/s. The simulation converged after 305 iterations. Figure 19 shows the velocity contours for the air in the Z direction. Note that most values are negative because the air is flowing in the negative direction of the Z axis. The highest value for the air velocity is m/s, and this occurred between the wood wall and the

56 56 last micro-channel facing the flow. Figure 20 shows the contours of the static temperature in Celsius scale. The average air temperature at the outlet of the model was 25.0ºC. The calculated total heat transfer rate from the copper micro-channels to the air is 10.0 Watts. The calculated average heat transfer coefficient of the air flowing across the microchannel tubes is 18.7 W/m 2 -ºC. These two values were compared with the results from the analytical analyses and the experimental data. Figure 19: Contours of Velocity in the Z Direction (Inlet = 0.5 m/s).

57 57 Figure 20: Contour of Static Temperature for Inlet = 0.5 m/s (ºC) Air inlet at 1.0 m/s This section presents the results and figures obtained for the general model simulation with the air inlet velocity equal to 1.0 m/s. The simulation converged at iteration number 310. Figure 21 shows the velocity contours for the air in the Z direction. The highest value for the air velocity is 1.3 m/s and, again, it occurs between the wood wall and the last micro-channel tube facing the flow, and right after the flow comes in contact with the first micro-channel. Figure 22 shows the contours of the static temperature in Celsius scale. The average air temperature at the outlet of the model was 24.3ºC. The calculated total heat transfer rate from the copper micro-channel tubes to the air is 14.7 Watts. The calculated average heat transfer coefficient of the air flowing across the micro-channels is 27.7 W/m 2 -ºC.

58 58 Figure 21: Contours of Velocity in the Z Direction (Inlet = 1.0 m/s). Figure 22: Contour of Static Temperature for Inlet = 1.0 m/s (ºC).

59 Analytical analyses This section presents the methodology and results of the analytical analyses. Two different approaches were implemented for the analyses. The two approaches are: flow over a flat surface and the extended surface (fin) analysis Flow over a flat plate analysis For the following analysis, all the equations and properties were obtained from Fundamentals of Heat and Mass Transfer (Incropera and DeWitt 2002). Figure 23 shows the schematic used for the analysis where the micro-channel tube length (l) is 13 mm and the micro-channel tube height (h) is 190 mm (the dimensions of one micro-channel tube). The air temperature (T ) is taken as 22ºC and the plate surface (T s ) has a constant temperature of 55ºC. The heat transfer and heat transfer coefficient are calculated for air velocities ( ) of 0.5 m/s and 1.0 m/s. The analysis neglects radiation effects and assumes steady state. The properties for air were obtained at the mean temperature of the microchannel tubes wall and the free air stream (T film = (22+55)/2 = 38.5 C). The properties that are needed for the analysis were the thermal conductivity (k = 26.9x10-3 W/m ºC), the kinematic viscosity (ν = 16.69x10-6 m 2 /s), and the Prandtl number (Pr = 0.7).

60 60 Figure 23: Flow over a Flat Surface. To determine the heat transfer coefficient and the heat transfer with this approach, an appropriate Reynolds number must be defined. Equation 11 is the formula widely used to estimate Reynolds ( ) numbers for flows over a plate. Equation 12 gives the definition of the Nusselt ( ) number and the formula to estimate it for this specific situation. (11) From Equation 12 the following relationship is obtained.. (12) To estimate the heat transfer rate for one side of the micro-channel, Equation 14 is used. (13) (14) where is the flat plate surface area and is the temperature difference between the flat plate and the air flow. Both sides of the channel are exposed to the flow and the prototype consists of three micro-channels; therefore, the result was multiplied by six in

61 61 order to obtain an estimate of the overall heat transfer of the system. Since the configuration of the copper micro-channels in the actual prototype does not allow the middle and back micro-channel tubes to be fully exposed to the flow, this approach was expected to overestimate the values for the heat transfer coefficient and the heat transfer. This was taken into consideration when comparing the results with other approaches. Table 14 contains the results for the analysis. Table 14: Flow over a Flat Plate Analysis Results. Estimated Quantity Air velocity 0.5 m/s Air velocity 1.0 m/s [W/m 2 - C] [W] Extended surface analysis This analysis is also known as the fin analysis. Fins are widely used for situations where it is desired to increase the available surface area for heat transfer by convection. In this case the experimental data presented in Section 4.7 was used to estimate the heat transfer coefficient and heat transfer from the micro-channel tubes to the air flow at 0.5 and 1.0 m/s. Since two of the micro-channel tubes are not fully exposed to the air flow, the temperature readings might lead to calculation errors in this analysis. For this reason the experimental data used for the analysis and presented in Table 15, are average

62 62 temperature values at steady state, recorded from the Front micro-channel tube as described in Chapter 4. In addition, values for the heat transfer coefficient and heat transfer are presented using the average temperature reading for the bottom level (T b ), middle level (T m ), and top level (T L ), including all the micro-channel tubes. T refers to the average air flow temperature. Table 15: Data from Experimental Run without Propane. Experimental Data [ C] Air velocity 0.5 m/s Air velocity 1.0 m/s All formulas and properties for the analysis were obtained from the Fundamentals of Heat and Mass Transfer (Incropera and DeWitt 2002). Figure 24 is a schematic for the extended surface analysis. The assumptions made for the analysis are as follows. A micro-channel tube is analyzed as a rectangular extended surface 13 mm wide (w), 190 mm long (L) and 2 mm thick (t). The radiation effects are neglected, and the system is at steady state.

63 63 Figure 24: Extended Surface Schematic. For extended surfaces with a uniform cross sectional area, Equation 15 is used to estimate the temperature profile of the surface, in this case the copper micro-channel tube. (15) where T is the temperature at the surface, is the average heat transfer coefficient, P fin is the fin perimeter, k is the fin conductivity, and A c is the fin cross sectional area. For convenience, Equation 16 is introduced and substituted into Equation 15 to simplify the problem. (16) where T(x) is the temperature at a specific location x. Once the substitution is made, Equation 17 is obtained. where is, (17) (18)

64 64 From the micro-channel tube data, the perimeter (P fin ) and the cross sectional area (A c ) were estimated with the following formulas: and, To solve Equation16, two boundary conditions were needed. In this problem these are: Boundary Condition 1: Boundary Condition 2: For the chosen boundary conditions the solution for the problem becomes Equation 19. (19) For this specific situation, the main unknown is the heat transfer coefficient. The measured temperature data for each case was plotted with the solution for 38 points (every 5 mm) using Equation 20. (20) The value of m was then adjusted to fit the measured temperature data (refer to Appendix C). When the estimated temperature values fitted the measured temperature values, Equation 18 was used to find the average heat transfer coefficient ( ). Figure 25 shows a comparison for the results of Equation 20 and the experimental data for air flowing at 0.5 m/s. The results of Equation 20 closely matched those measured during the experimental run. The value of m was set to 10.4 to achieve the temperature profile. For the calculations with air flow at 1.0 m/s, the same approach was used. A value of 11.7 for m

65 was found to be the best fit for the data measured during the experiment for the same air flow conditions. Figure 26 shows the results of Equation 20 compared to the data Temperature (C) Equation 20 Measured Data Micro channel Lentgh (m) Figure 25: Temperature Comparison between Experimental Data and Equation 20 for Air at 0.5 m/s Temperature (C) Equation 20 Measured Data Micro channel Length (m) Figure 26: Temperature Comparison between Experimental Data and Equation 20 for Air at 1.0 m/s.

66 66 Table 16 summarizes the results for both air flow analyses. In addition, the table contains the values calculated by using the average temperature of each level with all three microchannel tubes, and the overall heat transfer. The heat transferred ( ) to the air was calculated with Equation 21: (21) Table 16: Extended Surface Analysis Results. Estimated Quantity Front tube temperature data Air velocity 0.5 m/s Air velocity 1.0 m/s All tubes average temperature data Air velocity 0.5 m/s Air velocity 1.0 m/s As shown in Table 16, the heat transfer coefficient and heat transfer values differ from each other although they have the same air flow conditions. As explained previously, the arrangement of the micro-channel tubes in the prototype could have influenced the results.

67 Summary of results Results of the computer simulations and the analyses are presented in this section. As discussed earlier, the heat transfer coefficient and heat transfer values were calculated with three different methods, (1) a computer simulation of the system, (2) assuming flow over a flat surface, and (3) with an extended surface analysis. The results are shown in Table 17. Table 17: Heat Transfer Coefficient and Heat Transfer Results. Analysis Heat transfer coefficient (W/m 2 - C) Air velocity 0.5 m/s Heat transfer (W) Heat transfer coefficient (W/m 2 - C) Air velocity 1.0 m/s Heat transfer (W) Fluent Flat Plate Extended surface It was confirmed that the heat transfer coefficients for flow over a flat plate and extended surface (with the front tube temperature data) analyses were higher than the value obtained with Fluent. This was mainly because these two analyses assumed the system had only one micro-channel tube, while the computer simulation considered all three micro-channel tubes. The overall heat transfer values show how the thermosyphon effect helps to increase the system performance. For the extended surface analysis, the heat transfer was lower than for flow over a flat plate analysis and the computer simulation. From the three calculations, the extended surface analysis is the least reliable due to the

68 68 uncertainty of the data used for the calculations, while the flow over a flat plate calculation provided a general idea of the characteristics of the system. On the other hand, the Fluent simulation is considered the most accurate estimation due to the availability of the computer model to simulate the conditions of the system.

69 69 4. EXPERIMENTAL STAGE This chapter contains all the information related to the experimentation stage of this research project. A Standard Operating Procedure (SOP) and a Safety Evaluation Report (SER) were first prepared. The SOP explains the steps that should be followed to reproduce the experiment. The document also explains personnel preparation, safety measures, and experimental hardware and software needed to successfully complete the experiments. The SER is a document where the main hazards of the experiment are identified. Once the hazards were defined, actions to minimize, or avoid possible damage, were taken. Both documents are provided in Appendix D and E. 4.1 Propane charging process Prior to propane charging, the system was evacuated with a vacuum pump to remove unwanted fluids within the system. Then, the prototype was filled up with 99.5% pure propane to the test conditions given in Section 4.3. The desired charging test conditions were achieved by overfilling the prototype with propane and then heating the system to reach the specific temperature and pressure in the super heated region. Reaching the charging test conditions in the super heated region ensures that the propane is at the desired specific volume (refer to Figure 12). For this research the selected range for the specific volume is m 3 /kg. Figure 27 shows a stage of the charging process. For a detailed explanation of the charging process refer to Appendix D.

70 70 Figure 27: Set Up for Reaching the Charging Conditions. 4.2 Experimental set-up This section contains the information related to the experimental hardware used to complete the experiments. Only three copper micro-channel tubes were available at the design stage. To obtain an appropriate airflow across them, a small-scale wind tunnel was built (refer to Section 2.5). A smaller cross-sectional area in the wind tunnel increases the temperature difference between the air inlet and outlet. This improves temperature measurement accuracy. Figure 28 shows the experimental set-up for the project. The wind tunnel has a flow meter in series to measure the air flow entering the experiment. In addition to the wind tunnel, the apparatus includes a Keithley 2700 MMDA system, a temperature control unit, a voltage control unit (variac), a multi meter, an electric heater (not shown in the photo), type-k thermocouples, and fume hood. Due to the fact that propane is heavier than air, an aluminum duct was attached to the fume hood in order to help evacuate any propane from a leakage in the system.

71 71 Figure 28: Experimental Set-up. Fourteen type-k thermocouples were used to collect data during the experimental runs. Thermocouples were placed at the air inlet and outlet. One thermocouple was placed at the evaporator area, one at the midpoint and one at the top of the return path. The remainder nine thermocouples were placed at the top, middle and bottom of each micro-channel tube (see Figure 29). An extra thermocouple was attached to the evaporator area and connected to the temperature control unit in order to prevent the system from overheating.

72 72 Figure 29: Type-K Thermocouples attached to Loop Thermosyphon System. For temperature readings above 0ºC, the limit of error of the thermocouples is 2.0ºC or 0.75%, the greater applies (Omega Engineering 2009). These are absolute errors and they can produce significant errors in the experiment. However, the results of interest are the temperature differences. Therefore, a comparative calibration check was performed over the temperature range of 25 to 75 C. The entire set of thermocouples was compared while measuring an isothermal, heated, metal block. Special emphasis was given to the 9 thermocouples that measure the temperature of the micro-channel tubes,

73 73 and the two thermocouples that measure the inlet and outlet temperatures. The results are shown in Tables 18 and 19. It is evident that the thermocouples at the air inlet and oulet differ by no more than 0.06 C, and in the expected range of operation (~25 C), they differ by approximately0.01 C. The thermocouples for the micro-channel tubes have a standard deviation of 0.2 C or less. All thermocouples show very small deviations from each other. Therefore, the comparison of microchannel temperatures is quite accurate, and the measurement of temperaure rise of the air flow is very accurate. Table 18: Thermocouple Calibration Test (Air Flow and Micro-channel Tubes). Temperature ( C) Air flow ( C) Micro-channel tubes ( C) Inlet Outlet Difference Low High Average Standard deviation Table 19: Thermocouple Calibration Test (Complete Set-up). Temperature ( C) Complete set-up ( C) Low High Average Standard deviation

74 Experimental runs A total of fourteen experiments were performed. Experiments were run with the prototype charged at three different specific volumes (0.007, and m 3 /kg), two different air flow speeds (0.5 and 1.0 m/s), and a temperature of approximately 50ºC at the evaporator, for a total of six experiments. Although the operating range for the prototype was set lower than 50ºC, the same experiments were repeated after increasing the evaporator temperature to approximately 70ºC. The primary objective of these experiments was to test the design capacity, and to check for dry-out conditions in the prototype (when the evaporator region does not have liquid available). Two additional experiments were performed with the prototype without propane. These experiments were conducted in order to evaluate the heat transfer, and to compare the temperature profiles of the micro-channel tubes when they operate as a fin. The experiments without propane were performed with air flow at 0.5 and 1.0 m/s until the temperature at the bottom level of the micro-channel tubes reached approximately 50ºC. For a detailed explanation of the steps to perform the experimental runs refer to Appendix D. The data acquisition system was set to record the temperature readings every 1 second. For a specific thermocouple the readings are separated by 14 seconds. Since the objective of the research is to study the performance of the prototype at steady state, the difference in time for the readings does not affect the results of the study. In all of the figures and tables of the Chapter 5, the nomenclature used to identify a thermocouple reading is Return Middle and Return Top (for the return path at the corresponding locations), Heater Area (for the evaporator section, and Air Inlet and

75 75 Air Outlet (for the air temperature measurements). For the micro-channel tubes section, these are identified with two words. The first word refers to the position of the microchannel tube in relation to air flow, and the second word refers to the height location. For example, Front Bottom refers to the front micro-channel tube at the bottom level; Mid Middle refers to the mid micro-channel tube at the middle level, etc.

76 76 5. EXPERIMENTAL RESULTS The following sections present the experimental data collected during the experiments. Each figure in the following sections presents the temperature reading in degrees Celsius in the Y-axis, while the X-axis contains the corresponding reading number. 5.1 Experimental runs specific volume m 3 /kg The first set of experiments to be performed was with the prototype charged with propane at a specific volume of m 3 /kg. Figures 30 to 33 show the data for the experimental runs. Temperature (C) Prototype Air outlet Air inlet Reading Number Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Figure 30: Data for Experimental Run (0.007 m 3 /kg, evaporator at ~50 C, 0.5 m/s).

77 77 Temperature (C) Air outlet Prototype Air inlet Reading Number Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Figure 31: Data for Experimental Run (0.007 m 3 /kg, evaporator at ~50 C, 1.0 m/s). Temperature (C) Prototype Air outlet Air inlet Reading Number Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Figure 32: Data for Experimental Run (0.007 m 3 /kg, evaporator at ~70 C, 0.5 m/s).

78 78 Temperature (C) Reading Number Prototype Air outlet Air inlet Figure 33: Data for Experimental Run (0.007 m 3 /kg, evaporator at ~70 C, 1.0 m/s). Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet The thermosyphon behavior is clearly observed in the plots. In Figure 30 all the thermocouples attached to the micro-channel tubes measured approximately the same temperature during the experiments. The average temperature for the readings was 53.7ºC with a standard deviation of 0.2ºC. For air flow inlet and outlet temperature, the difference is bigger for the experimental runs with an air flow velocity of 0.5 m/s (Figures 30 and 32), than for the experiments with an air flow velocity of 1.0 m/s (Figures 31 and 33). This result was expected since twice the amount of air is passing across the micro-channel tubes. As mentioned previously, although the operating range for the prototype was for temperature values lower than 50 C, experiments were run with a higher temperature to check for dry out conditions in the evaporator section. Figure 32 shows the data for the first experimental run with the evaporator temperature of ~70 C. It is important to notice the Back Middle temperature reading. Right after the experimental run started, all the thermocouples (attached to the prototype) measured approximately the same temperature

79 79 except for the Back Middle reading. Although the reason is unknown, the cause might have been a bubble blocking the propane flow in one or more micro-channels in the back tube. Eventually, the temperature measurement got closer to the other readings of the prototype. Figure 33 presents the data for the thermocouple measurements during the experimental run with the evaporator at ~70 C, with the air flow velocity at 1.0 m/s. During this experiment all thermocouples placed at the prototype measured a similar temperature except for Mid Middle. The thermocouple was measuring a lower temperature up to the reading number ~65, afterwards the measurements were fairly similar to the other thermocouples on the prototype Summary specific volume m 3 /kg Table 20 presents a summary of the results for the experiments with the prototype charged with a specific volume of m 3 /kg. Table 20: Results for Experimental Runs with Propane at m 3 /kg. Experimental Data (0.007 m 3 /kg) Evaporator at ~50 C (0.5 m/s) Evaporator at ~50 C (1.0 m/s) Evaporator at ~70 C (0.5 m/s) Evaporator at ~70 C (1.0 m/s) Average prototype temperature ( C) Standard deviation ( C) Average air inlet temperature ( C) Standard deviation ( C) Average air outlet temperature ( C) Standard deviation ( C)

80 80 For the experiments with the evaporator at ~50ºC the standard deviation for the recorded temperature values is 0.2 and 0.5ºC, for air flow at 0.5 and 1.0 m/s, respectively. These results are important since they demonstrate the thermosyphon behavior of the prototype. For the experiments with the evaporator at ~70ºC the thermosyphon behavior is also demonstrated by the low standard deviation values (1.2 and 1.3ºC). More important is the fact that no dry-out occurred in the evaporator at this operating temperature. 5.2 Experimental runs specific volume m 3 /kg Figures 34 to 37 present the data for the experimental runs with the prototype charged at a specific volume of m 3 /kg. As shown in the figures, the readings of the thermocouple placed in the Heater Area (evaporator section), are higher than those of the rest of the prototype sections. Temperature (C) Air outlet Prototype Air inlet Reading Number Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Figure 34: Data for Experimental Run (0.009 m 3 /kg, evaporator at ~50 C, 0.5 m/s).

81 81 Temperature (C) Air outlet Prototype Air inlet Reading Number Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Figure 35: Data for Experimental Run (0.009 m 3 /kg, evaporator at ~50 C, 1.0 m/s). Temperature (C) Reading Number Prototype Air outlet Air inlet Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Figure 36: Data for Experimental Run (0.009 m 3 /kg, evaporator at ~70 C, 0.5 m/s).

82 82 Temperature (C) Reading Number Prototype Air outlet Air inlet Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Figure 37: Data for Experimental Run (0.009 m 3 /kg, evaporator at ~70 C, 1.0 m/s). Figure 36 presents the data for the experimental run with the evaporator running at an average temperature of ~70 C, and air flow at 0.5 m/s. During the experiment, two thermocouples attached to the Front micro-channel tube, recorded temperatures lower than the average of the prototype. The reason could be bubbles obstructing the liquid to flow out of the micro-channels to the evaporator section. Due to the temperature differences, the standard deviation was 2.3 C for all the thermocouples attached to the prototype, the highest of all the experimental runs with the prototype charged with propane at m 3 /kg. Figure 37 presents the data for the experimental run with the evaporator running at an average temperature of ~70 C, and an air flow at 1.0 m/s. During this experiment, only the thermocouple located at the heater recorded a notable different temperature compared to the rest of the prototype temperature readings. As shown in the figure, this time all the micro-channel tubes worked properly.

83 Summary specific volume m 3 /kg Table 21 presents a summary of the results for the experiments with the prototype charged with a specific volume of m 3 /kg. Table 21: Results for Experimental Runs with Propane at m 3 /kg. Experimental Data (0.007 m 3 /kg) Evaporator at ~50 C (0.5 m/s) Evaporator at ~50 C (1.0 m/s) Evaporator at ~70 C (0.5 m/s) Evaporator at ~70 C (1.0 m/s) Average prototype temperature ( C) Standard deviation ( C) Average air inlet temperature ( C) Standard deviation ( C) Average air outlet temperature ( C) Standard deviation ( C) The reason for the evaporator temperature readings to be higher than the rest of the prototype readings might be directly related to the heater location. While reaching the charging test conditions, the heater was separated from the prototype. When the prototype was charged, the heater and the insulation were put on the prototype in preparation for the experimental runs. No direct contact between the thermocouple and the heater was desired but this might have happened while attaching the insulation tape. This affects the standard deviation values for the experimental runs. For example, if the evaporator temperature measurements are not taken into account, the standard deviation values for

84 the remainder temperature measurements (11) of the prototype would be 0.6, 0.5, 1.9 and 1.3ºC, respectively Experimental runs specific volume m 3 /kg Figures 38 to 41 show the data for the experimental runs with the prototype charged with propane at a specific volume of m 3 /kg. The temperature measurements were relatively close during all the tests with maximum standard deviation values 0.8 and 1.1ºC, for air flow at 0.5 and 1.0 m/s, respectively. Temperature (C) Prototype Air outlet Air inlet Reading Number Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Figure 38: Data for Experimental Run (0.011 m 3 /kg, evaporator at ~50 C, 0.5 m/s).

85 85 Temperature (C) Air outlet Prototype Air inlet Reading Number Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Figure 39: Data for Experimental Run (0.011 m 3 /kg, evaporator at ~50 C, 1.0 m/s). Temperature (C) Prototype Air outlet Air inlet Reading Number Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Figure 40: Data for Experimental Run (0.011 m 3 /kg, evaporator at ~70 C, 0.5 m/s).

86 86 Temperature (C) Reading Number Prototype Air outlet Air inlet Figure 41: Data for Experimental Run (0.011 m 3 /kg, evaporator at ~70 C, 1.0 m/s). Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Figure 39 shows the data for the experimental run with the evaporator running at an average temperature of ~50 C, and air flow at 1.0 m/s. The temperature measurements in the evaporator area are slightly higher than the rest of the measurements of the prototype. Figure 40 presents the data for the experimental run with the evaporator running at an average temperature of ~70 C, and air flow at 0.5 m/s. The figure shows that at steady state operation the temperature measurements of the thermocouples attached to the prototype were consistent with thermosyphon operation. Before reaching steady state, the Front Bottom and Front Middle thermocouples had some unexpected behaviors. This can be an effect of bubbles blocking the liquid propane inside the micro-channels. Figure 41 presents the data for the experimental run with the evaporator running at an average temperature of ~70 C, and air flow at 1.0 m/s.

87 Summary specific volume m 3 /kg Table 22 presents a summary of the results for the experiments with the prototype charged with a specific volume of m 3 /kg. Table 22: Results for Experimental Runs with Propane at m 3 /kg. Experimental Data (0.007 m 3 /kg) Evaporator at ~50 C (0.5 m/s) Evaporator at ~50 C (1.0 m/s) Evaporator at ~70 C (0.5 m/s) Evaporator at ~70 C (1.0 m/s) Average prototype temperature ( C) Standard deviation ( C) Average air inlet temperature ( C) Standard deviation ( C) Average air outlet temperature ( C) Standard deviation ( C) The temperature measurements at the evaporator section were very similar to the rest of the prototype during the previous four experiments. The results support the possibility of contact between the heater and the thermocouple at the evaporator section during the experiments with the prototype charged with propane at a specific volume of m 3 /kg. The difference is evident in a comparison of the standard deviation values. For the m 3 /kg experiments, the maximum standard deviation value was 2.3ºC, while for the m 3 /kg experiments, the maximum standard deviation value was 1.1ºC.

88 Experimental runs prototype without propane Experiments were performed with the prototype running without propane. Two experiments were performed with an average temperature of ~50ºC, at the bottom level of the micro-channel tubes. The system was tested with air flow velocities of 0.5 and 1.0 m/s. These two experiments were performed to compare the micro-channel tubes behaving as a classical fin versus the isothermal fin provided by the thermosyphon operation Micro-channel tube bottom level temperature ~50 C - air flow at 0.5 m/s Figure 42 shows the data for the experimental run with no propane inside of the prototype, and at an air flow velocity of 0.5 m/s. The data show that the heater area temperature is approximately 30ºC above the average of the bottom level temperature. The average temperature, during steady state operation, for the bottom level was 53.4ºC; for the middle level 36.0ºC; and for the top level 28.7ºC. The air flow inlet and outlet average temperature, at steady state conditions, were 21.9ºC and 23.8ºC, respectively.

89 89 Temperature (C) Reading Number Figure 42: Data for Experimental Run No Propane (0.5 m/s). Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Micro-channel tube bottom level temperature ~50 C - air flow at 1.0 m/s Figure 43 presents the data for the experimental run with no propane inside of the prototype, and at an air flow velocity of 1.0 m/s. Similar to the previous experiment, the heater area temperature is approximately 30ºC hotter than the average of the bottom level temperature. It should be noted that the temperature history is different from the previous experiment. It does not have a temperature increment at the beginning of the run. The main reason is that between the experimental runs the only difference was the air flow velocity, so once the data was recorded for the experiment with air flow velocity of 0.5 m/s, the air flow velocity was just increased, and the new data were recorded. Since the research focus is at the operation of the system at steady state conditions, this fact does not affect the results of the experimental run. The average temperature, during steady state operation, for the bottom level was 49.1ºC; for the middle level 32.9ºC; and for the top level 27.1ºC. The air flow inlet and outlet average temperature, at steady state conditions, were 21.9ºC and 23.1ºC, respectively.

90 90 Temperature (C) Reading Number Heater Area Back Bottom Back Middle Back Top Mid Bottom Mid Middle Mid Top Front Bottom Front Middle Front Top Return Middle Return Top Ain Inlet Air Outlet Figure 43: Data for Experimental Run No Propane (1.0 m/s). The temperature profiles obtained from the two experiments performed without propane inside of the prototype were similar. These temperature profiles show the classical fin behavior of the system. The main difference between the experiments was the decrease in temperature measurements for the micro-channel tubes at an air flow of 1.0 m/s compared to 0.5 m/s. In addition the inlet/outlet temperature difference decreased from 1.9ºC to 1.2ºC, for air flow at 0.5 and 1.0 m/s, respectively. 5.5 Experimental heat transfer coefficient and heat transfer calculations To compare the heat transfer coefficient calculated with the analyses and the computer simulation, heat transfer coefficients were calculated from the experimental data with air flow velocities of 0.5 and 1.0 m/s. Equation 22 is used with the experimental properties in order to find the average heat transfer coefficient ( ). (22)

91 91 In the above equation, the mass flow rate ( ) for the experimental runs were 3.4 and 6.8 g/s, for air velocities of 0.5 and 1.0 m/s, respectively. The air specific heat capacity ( ) for the experimental conditions is kj/kg- C. The outside surface area ( ) of the micro-channel tubes exposed to the air flow was estimated to m 2. The air inlet temperature ( ), air outlet temperature ( ) and the micro-channel tubes temperature ( ) were obtained from the steady state conditions of the experimental runs. The experimental heat transfer coefficient was calculated for each experimental run. The average heat transfer coefficient, for the experiments with an air flow velocity of 0.5 m/s, was 22.2 W/m 2 -ºC. On the other hand, for the experiments with an air flow velocity of 1.0 m/s, the average for the heat transfer coefficient was 34.6 W/m 2 -ºC. The average heat transfer values for the experiments were 9.4 and 12.9 W, for the air flow velocities of 0.5 and 1.0 m/s, respectively. Table 23 presents the calculated values for each experimental run in addition to the average values.

92 92 Table 23: Experimental Heat Transfer Coefficient and Heat Transfer Results. Charging test conditions (m 3 /kg) Heat transfer coefficient (W/m 2 - C) Air flow velocity 0.5 m/s Heat transfer (W) Heat transfer coefficient (W/m 2 - C) Air flow velocity 1.0 m/s Heat transfer (W) Average From Table 23 it can be observed that the highest values for the heat transfer coefficient and heat transfer were obtained during the experimental runs with the prototype charged with propane at m 3 /kg. The situation is similar for air flow velocities of 0.5 and 1.0 m/s. The results suggest that charging the prototype with a higher specific volume would increase the performance of the system. On the other hand, more experiments should be run, with different charging test conditions and at different operating temperatures to find an optimal charging conditions point. Since finding the optimal charging conditions is out of the scope of this research, these experiments were not performed. In comparison to the computer simulation, these heat transfer coefficients values are higher. One reason for this could be the exposure of the top and bottom manifolds to the air flow during the experimental runs. The manifold s exposure to air flow might

93 affect the experimental result calculations by adding surface area available to exchange heat with the air Energy balance An energy balance was performed to the system in order to validate the analysis. The system losses were expected to be relatively high mainly due to the prototype geometry. The area of the prototype outside the wind tunnel accounted for approximately two thirds of the entire device. Although this area was insulated, especially the heater area, it lost heat to the surroundings. Higher temperatures than the surroundings could be registered on the insulation surface and on the pressure gauge. This temperature difference between these surface areas and the surrounding air would cause natural convection to occur. Some heat was also lost to the wind tunnel parts in contact with the device. Equation 23 is the general energy balance equation. (23) For the system at steady state, the energy accumulated ( ) is zero. The energy entering the system ( ) is the heat load applied with the electric heater at the evaporator section. The energy leaving the system ( ) can be divided between the heat transferred from the micro-channel tubes to the air, and the heat loss to the surroundings. Equation 23 becomes Equation 24 for the system under study. (24) where is the heat transferred to the air flow and is the heat loss to the surroundings. The term can be calculated with the following equation:

94 94 (25) where represents the voltage applied to the electric heater, and represents the heater resistance. For the experimental runs with the heater temperature running at approximately ~50ºC, the voltage applied to the heater ranged from 55.2 to 55.9 volts. The heater resistance ranged from 69.9 to 70.1 Ω. Table 24 presents the results of the energy balance for the experimental runs. Table 24: Energy Balance Results. Charging test conditions (m 3 /kg) (W) Air flow velocity 0.5 m/s (W) (W) (W) Air flow velocity 1.0 m/s (W) (W) Average The heat losses are relatively high in the system, as expected. For the experimental runs with an air flow velocity of 0.5 m/s, an average of 21% of the heat load was transferred to the air; for an air velocity of 1.0 m/s, an average of 29% of the heat was transferred to the air. Actions to decrease the heat loss term can be taken for new loop thermosyphon designs. Minimizing the return path area and the charging extension area will help

95 decreasing the heat loss. For new designs and the existing prototype, better insulation can be used for the area exposed to the surroundings Thermosyphon effect comparison Table 25 shows the values of the heat transferred to the air during the experimental runs in addition to a comparison between the thermosyphon and the classic fin operation of the micro-channel tubes. Table 25: Heat Transfer Comparison for Thermosyphon and Classic Fin Operation. Charging test conditions (m 3 /kg) (W) Air flow velocity 0.5 m/s Improvement over classical fin Air flow velocity 1.0 m/s (W) Improvement over classical fin % % % % % % No Propane (classical fin case) 6.5 n/a 8.3 n/a From Table 25 it is evident that the thermosyphon operation improves the micro-channel tubes performance. The heat transferred to the air increased by an average of 45% (0.5 m/s) and 55% (1.0 m/s), when the prototype was charged with propane in comparison to the classical fin operation. The charging test conditions at a specific volume of m 3 /kg provided the highest improvement for the experiments with 62% (0.5 m/s) and 63% (1.0m/s). The results for the experiments with charging test conditions at a specific

96 96 volume of m 3 /kg were the lowest of the experiments. These results might have been affected by flow blockage and the heater temperature measurement.

97 97 CHAPTER 6: SUMMARY AND CONCLUSIONS A loop thermosyphon was designed, analyzed, fabricated, and experimentally tested. The prototype contained three novel copper micro-channel tubes in its design. Propane was selected as the working fluid and it worked successfully in the application. A pressure loss analysis was performed to ensure the working fluid circulation within the system. The micro-channel tube heat transfer coefficient and overall heat transfer for the thermosyphon operation were calculated and compared with the fin mode of operation. The major results from this study are: A loop thermosyphon using micro-channel tubes was designed using a pressure loss analysis. The analysis proved to be a useful design tool for this type of thermosyphon. Novel copper micro-channel tubes were used for the thermosyphon, and they also served as fins in a heat sink configuration. The fins were virtually isothermal, and therefore were nearly 100% efficient. The isothermal fins increased the heat transfer by a maximum of 63% compared to the classical fin mode. Heat transfer coefficient values were calculated analytically as well as experimentally. The results yielded average heat transfer coefficients of 22.3 and 31.8 W/m 2 -ºC, for air flow at 0.5 and 1.0 m/s, respectively. The experiments performed on the heat sink system confirmed the loop thermosyphon behavior of the micro-channel tubes. The largest standard deviation for the temperature measurements on the surface of the micro-channel tubes was

98 98 1.1ºC when the loop thermosyphon was operating at approximately 50ºC. In contrast, the temperature decreased by more than 20ºC over the length of the micro-channel tubes when they were operated in the classical fin mode. 6.1 Future work The copper micro-channel loop thermosyphon prototype worked quite well as a thermosyphon. However, there were some discrepancies in the temperature readings, e.g., lower readings were observed in one of the micro-channel tubes during some of the experiments. This resulted in lower heat transfer to the air. The possible cause could be flow obstruction due to bubble formations in the channels. Since the reason is not clear, future research needs to focus on how to identify and prevent this event. Another consideration for future work should be the addition of more microchannel tubes in the heat sink and the addition of classical fins to the micro-channel tubes to extend the surface area; the goal would be to increase the heat transfer in the heat exchanger. Propane can be a working fluid for heat sinks that are used in electronic cooling. However, it is not suitable for heat exchangers such as radiators. Therefore, experiments should be performed with different working fluids to determine and compare their performance in a micro-channel thermosyphon.

99 99 REFERENCES Abel Chuang, Po-Ya. An Improved Steady-State Model of Loop Heat Pipes Based on Experimental and Theoretical Analyses. PhD Thesis, University Park : Pennsylvania State University, Beitelmal, Monem H., and Chandrakant D. Patel. Two-Phase Loop: Compact Thermosyphon. HP Laboratories Report, Palo Alto: Hewlett-Packard Company, Cotter, T.P. "Principles and Prospects of Micro Heat Pipes." 5th International Heat Pipe Conference. Tsukuba, Delphi Corporation. Delphi. August 21, (accessed June 8, 2009). Edwards, Jesse, interview by Juan Flores. Heat Exchangers Design Approach (February 17, 2009). Environmental Protection Agency. High Global Warming Potential (GWP) Gases. November 28, (accessed October 2, 2009). Faghri, Amir. Heat Pipe Science and Technology. Washington, D.C.: Taylor & Francis, Fluent Ansys. Fluent Documentation , Fluent Inc. Fluent Documentation , Hsu, Li-Chieh. Device of Micro Loop Thermosyphon for Ferrofluid Power Generation. United States of America Patent January 26, Incropera, Frank P, and David P DeWitt. Fundamentals of Heat and Mass Transfer 5th Ed. New York: John Wiley & Sons, Kuppan, T. Heat Exchanger Design Handbook. New York: Marcell Dekker, Inc, Mills, ANthony F. Heat Transfer. Concord: R.R. Donnelley & Sons, 1992.

100 Mulford, Randy, interview by Juan Flores. Brazing Process for Copper Micro-channel Loop Thermosyphon Prototype (10 5, 2009). Oliet, C., A. Oliva, J. Castro, and C.D. Pérez-Segarra. "Parametric studies on automotive radiators." Applied Thermal Engineering 27, 2007: Omega Engineering. Omega Engineering Technical Reference (accessed October 1, 2009). Pal, Aniruddha, Yogendra K. Joshi, Monem H. Beitelmal, Chandrakant D. Patel, and Todd M. Wenger. "Design and Performance Evaluation of a Compact Thermosyphon." IEEE TRANSACTIONS ON COMPONENTS AND PACKAGING TECHNOLOGIES, December 2002: Perkins, Jacob. United Kingdom Patent April Reynolds, William C. Thermodynamic Properties in SI: Graphs, Tables and Computational Equations for forty substances. Stanford: Department of Mechanical Engineering Stanford University, Sachs, H., et al. Emerging Energy-Savings Technologies and Practices for the Buildings Sector as of Report, Washington, D.C.: American Council for an Energy- Efficient Economy, The Chemical Engineers' Resource Page (accessed October 2, 2009). ToolBox, The Engineering. The Engineering ToolBox (accessed May 26, 2009). Vaitkus, Victor L. A Process for the Direct Hot Extrussion of Hollow Copper Profiles. Master's Thesis, Athens: Ohio University, White, Frank M. Fluid Mechanics 4th Edition. Boston: McGraw-Hill,

101 APPENDIX A: PROPANE PROPERTIES USED IN PRESSURE LOSS ANALYSIS 101 Table 26: Propane Properties Used in Pressure Loss Analysis. Saturation property Temperature -30 C 0 C 50 C Liquid specific volume (m 3 /kg) 1.8 x x x10-3 Vapor specific volume (m 3 /kg) 2.6 x x x10-2 Liquid density (kg/m 3 ) Vapor density (kg/m 3 ) Liquid dynamic viscosity (kg/m-s) 1.7 x x x10-5 Vapor dynamic viscosity (kg/m-s) 6.6 x x x10-6 Surface tension (N/m) 1.4 x x x10-3 Vapor specific weight (N/m 3 )

102 102 APPENDIX B: PROTOTYPE COMPONENTS SPECIFICATIONS All the schematics for the copper micro-channel loop thermosyphon components with its dimensions. As shown in Figure 44 the prototype s assembly contains eight different components. The manifolds, returns and connectors are 12.7 mm copper pipes (drawntype L), with an inner diameter of 13.8 mm and an outer diameter of 15.9 mm. Each component was cut to the specific lengths as shown from Figure 45 to Figure 48. The u- return bends, elbows and tees are 12.7 mm copper fittings. Figure 44: Copper Micro-channel Loop Thermosyphon Components.

103 103 Figure 45: Manifold (units-mm). Figure 46: T-Elbow Connector (units-mm). Figure 47: Return (units-mm).

104 Figure 48: U-T Connector (units-mm). 104

105 105 APPENDIX C: DATA FOR THE EXTENDED SURFACE ANALYSIS Table 27: Data Used for the Extended Surface Analysis with Air Flow at 0.5 m/s. Length Equation 20 Experiment T bot= 51.9 C x T(x) T Measured T mid= 33.7 C T top= 28.2 C T air= 21.9 C θ b = 30.0 C θ L = 6.3 C L= 0.19 m m= P= 0.03 m A c = m k= 401 W/m C h= 24.6 W/m 2 C q= 6.2 W

106 106 Table 28: Data Used for the Extended Surface Analysis with Air Flow at 1.0 m/s. Length Equation 20 Experiment T bot= 47.6 C x T(x) Measured T mid= 30.9 C T top= 26.6 C T air= 21.9 C θ b = 25.7 C θ L = 4.7 C L= 0.19 m m= P= 0.03 m A c = m k= 401 W/m C h= 31.1 W/m 2 C q= 6.0 W

107 107 APPENDIX D: STANDARD OPERATIONAL PROCEDURE Department of Mechanical Engineering Standard Operating Procedure for "Copper Micro-channel Loop Thermosyphon" Last Updated: 9/21/09 Contributor(s): Juan G. Flores Khairul Alam Israel Urieli Frank Kraft Randy Mulford Jesús Pagán Adriane Mohlenkamp

108 108 I. Test System Description The main objective of the proposed research is to design, model and test a more efficient and smaller heat exchanger radiator compared to actual radiator designs. The radiator design will consist of copper micro channel tubes functioning with the similar principle of a loop thermosyphon. Propane has been chosen as the working fluid. Once the system is charged a heat load is applied and air is passed through the copper micro channels to extract the heat from the system. Figure 49: Copper Micro Channel Loop Thermosyphon Prototype. II. Personnel Preparations A. Required Personnel Training The person performing the test of the system must have knowledge using Keithley 2700 Multi Meter/Data Acquisition system; in addition, the person should know how to read measurements from an air flow meter (rotameter). Regarding the use of propane the person must be careful and able to follow the safety handling instruction for a flammable gas tank. B. Required Safety Equipment The room where the experiment will be run must be well ventilated and no fire flames or equipment which can cause a spark can be in the room. A gas leak detector must be working at high sensitivity all time in order to detect leaks in the system. A fire extinguisher must be in the room due to the flammability of

109 109 propane. The person must wear eye and face protection at all time during the experiment. C. Required Tools/Analytical Equipment Data Acquisition system (see Figure 50) Figure 50: Keithley Data Acquisition System (Tutorial). Type K thermocouples Wind tunnel with a 1 x 8.25 cross sectional area (see Figure 51) Air hoses Safety gloves Air flow meter (rotameter) Pressure gauge Figure 51: 1 x 8.25 Wind Tunnel.

110 110 Electric heater Propane 99.5% pure Quick Coupler CGA-510 to ¼ NPT ¼ NPT to ¼ SAE connector Refrigeration charging hose Refrigeration charging hose with valve Teflon tape Electric hot plate 21.5-quarts metal container (see Figure 52) Figure 52: 21.5-Quarts Metal Container. Vacuum pump Electric variac unit (see Figure 53) Figure 53: Electric Variac Unit.

111 111 Multi Meter Temperature control unit (see Figure 54) Figure 54: Temperature Control Unit. Portable gas leak detector (see Figure 55) Figure 55: 5% Lower Explosive Level (LEL) Portable Gas Leak Detector F Ceramic Tape Insulation D. Required Procedures Safety Handling of Propane tank Operating instructions of the Data Acquisition system. III. Project/Site Preparations A. Required Services 110 V power outlet Potable water Compressed air line Fume hood ventilation system

112 112 B. Required Equipment/Supplies N/A C. Required Operators and/or Technical Assistance One operator to conduct the experiment and a safety watch person Date of Final Safety Review Board Evaluation and Certification of Project/Site Safety Certification Completed: Signature and Date

113 113 IV. Operation Attention! If at any moment throughout the experiment the propane gas leak detector alarm goes off, shut down all electrical equipment except for the ventilation system and evacuate the room for 30 minutes and conduct a gas leak test before performing any new experiment. 1. Wear safety glasses and face shield at all time during the experiment. 2. Gather all the equipment and tools that are needed and have them ready for use. 3. Measure and record the resistance of the electric heater. 4. Assemble the propane filling system A (prototype side) (see Figure 56). Figure 56: Propane Filling System (A). 5. Connect the prototype to the propane filling system A and then to the vacuum pump to evacuate the system. 6. Caution! Make sure no fire flames or any other devices which can cause a spark are in the room. 7. Important! Start the fume hood ventilation system in the room. 8. Assemble the propane filling system B (see Figure 57).

114 114 Caution! Make sure another person helps turning over the tank and keeping it upside down during the filling process 9. Turn the propane tank up-side down with the help of an assistant. This person must stay during the whole filling process. 10. Wear the safety gloves. Figure 57: Propane Filling System (B). 11. Connect the propane filling systems A and B to start the filling process. 12. Open the propane tank valve and let the charging system to be filled with liquid propane. Close the propane tank valve. Important! Use the portable gas leak detector to check for any leakage on the propane tank and filling hose system fittings. In case of any leakage the room must be evacuated for 30 minutes.

115 13. Insert the brown hose into the fume hood suction section and open the bleeding valve until liquid propane comes out of the hose. Close the bleeding valve immediately after liquid propane flows out of the hose. 14. Open the filling valve to let the liquid propane enter the prototype. Make sure that the prototype is at a lower position than the tank so all the liquid propane can enter the prototype. 15. Close the filling valve. 16. Important! Repeat steps 10, 11 and 12 only once, this will provide enough propane to the prototype in order to reach the charging conditions. 115 Caution! Some gas propane might still be in the charging system lines, make sure the gas propane will be extracted by the ventilation system before disconnecting them. 17. Relocate the propane tank at a safe place far from the experimenting area. Important! Use the portable gas leak detector to check for any leakage on the prototype fittings. In case of any leakage the room must be evacuated for 30 minutes. 18. Fill the 21.5-quarts metal container with enough water to cover most of the prototype and secure the metal container so it does not slip on top of the hot plate. Caution! Steps 16 and 17 require full attention from the operator. Make sure the water temperature is not increased in steps larger than 10 C and the pressure is kept below 4.5MPa. If the pressure exceeds 5 MPa the system must be shut down. Keep the portable gas leak detector on and close to the 21.5-quarts metal container. 19. Caution! Turn on the electric hot plate and heat the metal container with the prototype inside up to the boiling point of the water. Do this step incrementing the water temperature by 10 C. 20. Attention! Monitor the pressure gauge and release some gas at every stage of the previous step. Remember the charging conditions are 97.5 C and 4.2MPa for a specific volume of m 3 /kg. 21. Once the charging conditions are reached, turn off the electric hot plate and let the prototype cool down to ambient temperature. Important! Do not remove the prototype from the metal container until ambient temperature is reached. Caution! When the time has come to move the prototype from one place to another be careful the prototype does not fall or is put under any type of mechanical load.

116 Important! Use the portable gas leak detector to check for any leakage on the prototype fittings. In case of any leakage the room must be evacuated for 30 minutes Attach the electric heater to the evaporator section of the prototype; insulate the evaporator with the rest of the prototype which will not be exposed to the air flow. 23. Connect all the thermocouples from the prototype and the wind tunnel to the data acquisition system. 24. Open the air flow valve and make sure to record the air flux measurements for future calculations. 25. Once the air is flowing steadily, start the data acquisition system and make sure all thermocouples are recording data accordingly (double check with an external thermocouple or thermometer). 26. Connect the heater to the temperature control unit and the cables to multi meter in order to measure the voltage input. 27. Turn on the electric heater to apply a steady heat load to the system at a maximum temperature of 50 C. 28. Collect all the temperature measurement data until the system reaches a steady state operation. Caution! The whole prototype will become hotter once the heat load is applied; avoid contact with it until the experiment is over. 29. Turn off the electric heater and let the system cool down. 30. For different air flow or heat load experiments repeat steps Close the air flow valve. V. Project/Site Cleanup 1. Once all experiments are performed, disconnect and organize all the thermocouples. Turn off the data acquisition system. 2. Important! Turn off any other equipment except for the gas leak detector. 3. Important! Discharge the prototype (release all the propane through the ventilation hood, do not leave the system with propane inside) 4. Clean and organize the room. 5. Return all equipment and tools to their respective locations.

117 117 APPENDIX E: SAFETY EVALUATION REPORT Department of Mechanical Engineering Safety Evaluation Report for "Copper Micro-channel Loop Thermosyphon" Last Updated: 9/21/2009 Contributor(s): Juan G. Flores Khairul Alam Israel Urieli Frank Kraft Randy Mulford Jesús Pagán

118 118 Initial Safety Certification - Failure Mode and Effects Analysis Worksheet (Adapted from Cincinnati Machine PFMEA) I. Name / Description of Test System, including SOP 1 : (include photos of system and location) 1 A written step-by-step experimental procedure or standard operating procedure (SOP) must exist in order to judge its safety, so that has to be completed before the FMEA can be completed. Consider including scheduled safety inspections or checks in the experimental procedure. The SOP must be included in the DFSR or appropriately referenced. The main objective of the proposed research is to design, model and test a more efficient and smaller heat exchanger radiator compared to actual radiator designs. The radiator design will consist of copper micro channel tubes functioning with the similar principle of a loop thermosyphon. Propane has been chosen as the working fluid. Propane fulfilled the requirements of the application s temperature range and in addition contains a relatively low global warming potential (GWP = 3) (ToolBox 2005). A prototype has been built and will be filled with propane and tested in a wind tunnel for heat transfer performance studies. Figure 1: Loop Thermosyphon Prototype. Key Contact / Phone Juan G. Flores Date of Initial FMEA July 10, 2009 Core Team: Date of Initial System Demonstration August 5, 2009 Location: Stocker Center 015 Safety Review Board Approval / Date SOP name: SOP_Copper Micro-channel_Loop Thermosyphon_ Other Approval (if required)

119 119 II. Hazard Identification Discussion During the design process a main hazard was identified: working fluid flammability. Since a prototype was built by the process of brazing, the joints safety was uncertain for two reasons: the pressure at which the system will be exposed and for any gas leakage through the joints. The main hazard was identified since propane will be used as the system working fluid. Propane is considered a flammable gas and safety precautions must be taken when experimenting with it. Figure 2: Propane purity and flammability labels.