ARCTIC: A ROTARY COMPRESSOR THERMALLY INSULATED µcooler. Joshua D. Heppner David C. Walther Albert P. Pisano

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1 Proceedings of IMECE ASME International Mechanical Engineering Congress and Exposition November 5-11, 005, Orlando, Florida USA IMECE ARCTIC: A ROTARY COMPRESSOR THERMALLY INSULATED µcooler Joshua D. Heppner David C. Walther Albert P. Pisano Berkeley Sensor and Actuator Center University of California at Berkeley Berkeley, CA 9470 Phone (510) ABSTRACT Microscale cooling to date relies largely on passive on-chip cooling in order to move heat from hot spots to alternate sites. Such passive cooling devices include capillary pump loops (CPL), heat pipes, and thermosiphons. Recent developments for active cooling systems include thermal electric coolers (TECs) for heat removal. This paper focuses on the design of an active microscale closed loop cooling system that uses a Rankine vapor compression cycle cooling system. In this design, a rotary compressor will generate the high pressure required for efficient cooling and will circulate the working fluid to move heat away from chip level hot spots to the ambient. The rotary compressor will leverage technology gained from the Rotary Engine Power System (REPS) program at the UC Berkeley, most specifically the 367 mm 3 displacement platform. The advantage of a Wankel (Maillard) compressor is that it provides six compression strokes per revolution rather than a single compression stroke common to other popular compressors such as the rolling piston. The current Wankel compressor design will achieve a theoretical compression ratio of 8:1. The ARCTIC (A Rotary Compressor Thermally Insulated µcooler) system will be a microscale hybrid system consisting of some microfabricated (or MEMS) components including microchannels, in plane MEMS valves, and potentially MEMS temperature, pressure and flow sensors integrated with mesoscale, traditionally machined steel components, including the compressor itself. The system is designed to remove between 5-35 W of heat at 1000 rpm using R-134a but the system is easily scaleable through a speed increase or decrease of the compressor. Further, the current compressor design has a theoretical coefficient of performance (C.O.P.) of approximately, a significant improvement over comparable TECs with C.O.P.s of approximately Finally, a thermal circuit analysis determines that the time constant to achieve refrigeration temperature in 1 seconds is possible. INTRODUCTION The vapor compression cycle is the most common mechanical-type refrigeration cycle. The cycle and its variants are found anywhere from the common household refrigerator to a vehicle s air conditioning system. However, on the microscale the cycle has rarely been put to use as more often than not passive designs such as thermosiphons, heat pipes[1], and capillary pump loops (CPL) [], have been used. Active systems, such as thermoelectric coolers (TEC) [3], have been used as alternatives as well. These other options have found their way into the microscale in lieu of the vapor compression cycle for two reasons. The first reason applies only to the passive designs: the vapor compression cycle requires input power. Passive designs generally rely on a fluid s phase change ability to remove heat and to also provide motion. However, these designs are temperature limiters in reality and are limited to the fluid s thermal properties in regards to the final temperature of the cooled chip. The second reason vapor compression cycles have been limited on the microscale is the size of the system. Each of the previous designs, active or passive, have been minimized to some degree of success. While several pumps have been put to use on the microscale [4], they have been limited to the amount of pressure head which can be added to the fluid. Finally, compressors have been limited on the microscale mostly due to entropy considerations [5]. However, success at UC Berkeley in the Rotary Engine Power System (REPS) program has shown that engines can be operated at small scales [6]. This program has specifically demonstrated a 367 mm 3 displacement engine is capable of combustion. The cycle of this rotary engine is intake, compress, ignite, expand, and exhaust. A compressor would only require three of these: intake, compress, and exhaust. Therefore, transforming an engine into a compressor is not an ill-conceived notion. With the success of the Wankel engine to 1 Copyright 005 by ASME

2 build upon, it is the goal of this work to set out to build compressors which will aim to reach microscale sizes. The 367 mm 3 engine platform which the initial compressor design will be built upon has a footprint of.5 cm x 3.0 cm or about the size of 6 IC (integrated circuit) chips placed side by side. Further, a wankel compressor can be designed to achieve large pressure ratios. To enable a vapor compression cycle with coefficients of performance (C.O.P.) around a pressure ratio of 8:1 is necessary, where C.O.P. is defined as: C. O. P. = HeatLift PowerIn This is the inherent disadvantage TEC s have. They suffer from parasitic conduction of heat back to the cool side of the device because they are conducting current to achieve the desired cooling effect, limiting their C.O.P. Vapor compression cycles can minimize this by using insulating layers to limit parasitic conduction losses while maintaining their high C.O.P. Finally, to combat the increased entropy limiting compressors on the microscale, the wankel rotary compressor can achieve six compression strokes per revolution. The wankel compressor has a pair of intake and exhaust ports which allows each rotor face to go through two compression strokes per revolution, leading to six strokes total. This allows the wankel compressor to output six times the mass flow rate of other more common compressors such as the rolling piston, which has only one compression stroke per revolution. This paper will discuss the design of a complete vapor compression refrigeration cycle for the microscale. This will include the design of individual parts as well as a thermal circuit analysis using lump capacitance modeling. The thermal circuit analysis will be developed not only to show the performance of the system but also analyze parameters which will lead to optimal performance of the system. NOMENCLATURE A Cross-sectional area heat travels through, m c Specific heat for constant pressure, J / kg K p C. O. P. Coefficient of Performance C Thermal capacitance, J / K h Heat transfer coefficient, W / m K i Electrical current, A k Thermal conductivity, W / m K K Constant L Conducting length, length heat will travel, m m Mass, kg P Pressure, N / m, atm Q Heat, W r p r v Pressure Ratio Compression Ratio R Thermal resistance, W / K T Temperature, K t Time, s 3 V Volume, m Voltage, V Greek symbols ε γ σ Emissivity Ratio of specific heats Stefan-Boltzman Constant, W / m K Miscellaneous subscript symbols cond conv in out rad s sur Thermal conduction Thermal convection Into the compressor Out of the compressor Thermal radiation Surface of device Surrounding air ARCTIC DESIGN The vapor compression cycle is composed of four essential steps. The active part of the system, the compressor, receives uncompressed refrigerant and does work on this vapor by compressing the gases to the desired pressure. The compressor releases the vapor into the passive part of the system, beginning with the condenser. In the condenser, the refrigerant goes through a phase change, transforming from a vapor to a liquid. In the process, the refrigerant releases heat to the environment. Once the refrigerant leaves the condenser it is in a high pressure, high temperature liquid state prepared to enter the expansion valve. The expansion valve expands the liquid into a low pressure, low temperature state. The refrigerant moves through the evaporator absorbing heat by again changing phases, this time from liquid to vapor. Then the cycle repeats [7]. A review of this cycle can be seen in Figure 1. The goal for the overall system design, which is shown in Figure, is a flat device. By utilizing MEMS technology the system will be able to minimize it height so that it resembles a TEC. This resemblance will also include the fact that there will be a hot side, the condenser, and a cold side, the evaporator. To 4 Copyright 005 by ASME

3 minimize heat leakage between the components insulation will also be used. Port Transformation High Pressure, High Temperature Vapor Compressor Low Pressure, Low Temperature Vapor Condenser Evaporator Figure 3: Transformation from two ports to two sets of ports. Note that compressor ports on the right are not optimized but shown in this manner for clarity High Pressure, High Temperature Liquid Low Pressure, Low Temperature Liquid Expansion Valve Figure 1: Vapor Compression Cycle The ports shown in Figure 3 are actually not optimal as they allow cross-talking between ports. Cross-talking occurs when the rotor allows a direct path for the fluid from the inlet port to the exhaust port. In order to design optimal ports one wants to cover the exhaust ports when the rotor opens the inlet ports. An optimal port design is shown in Figure 4. This design minimizes cross-talking and allows the compressor to achieve a compression ratio, r v defined as, Cold Side Evaporator Compressor Hot Side Figure : ARCTIC system schematic Condenser Compressor design The design of the Wankel rotary compressor began by using the same rotor and housing from the REPS 367 mm 3 Wankel rotary engine. This design used parameters set forth in previous papers [6], [8]. In order to transform the engine to a compressor this required a redesign of the ports. A rotary engine requires two ports: one inlet port and one exhaust port. A rotary compressor does not require an ignition stroke or an expansion stroke like a rotary engine does. Thus, a rotary compressor can compress the vapor on both sides of the housing as the housing is symmetric. To allow compression on both sides of the compressor a second set of ports must be added, so that there are two sets of inlet ports and two sets of exhaust ports, as shown in Figure 3. V r v = V where V is the volume. This particular design achieves a compression ratio of 3.3. Assuming a constant entropy compression cycle this will enable a high pressure of 3.6 atm as given from the refrigeration tables for R-134a. If the working fluid is air the high pressure would be 5.3 atm for a reversible adiabatic process. This is found using the following wellknown relationship, γ PV = K where P is the pressure, K is a constant, and γ is the ratio of specific heats. Knowledge of the high pressure will set the pressure ratio, r p, defined as: P in out out r p =. Pin Pressures of 3.6 atm are most often not enough to achieve the desired heat lift. This is because the pressure and the temperature are interrelated in a given state; by compressing the refrigerant the temperature raises. To allow the heat to be absorbed into the environment the temperature must be greater than that of the ambient. Due to these constraints all positive displacement compressors and pumps require check valves, one-way fluid valves, to prevent back flows into the compressor 3 Copyright 005 by ASME

4 Exhaust Opening pressure loss and thus enthalpy loss across the expansion valve [1]. The expansion valve will be designed to utilize such characteristics. Inlet Closing a) b) Figure 4: Optimal port design to eliminate cross-talking. a) Shows inlet just closing with exhaust completely closed. b) Shows exhaust just opening with inlet completely closed. and the Wankel compressor is no different. Thus, the check valves are designed so that they open at the desired pressure and close as fast as possible. Check valves will allow the system to reach pressures up to 9.1 atm. Design of the passive parts The passive parts of the system include the condenser, expansion valve, and evaporator. Each of these parts can be made using microfabrication techniques. However, none of the parts have yet been optimally designed; thus, there is no discussion here about the correct number of turns or the size of the channels to reduce pressure drop. The condenser, shown in Figure 5, is a serpentine structure with fins. This tubular structure takes advantage of the microscale s large surface area to volume ratio in order to increase heat transfer to the environment. This device can be made in two manners. Through a process called polymolding [9] or by using an optimized silicon DRIE process [10] and capping the etched channels by bonding a second substrate to the device layer. Figure 6: MEMS expansion chamber The evaporator, shown in Figure 7, is very similar in design to the condenser. The only exception is that the material between the serpentine is not removed. This additional material increases the heat capacity of the evaporator and allows heat to be absorbed into the fluid from the sides as well as the bottom. Again the advantage of having microscale channels here allows the fluid to quickly absorb heat, minimizing the necessary size of the evaporator. It is interesting to note that the insulating layer shown in Figure 7 is imperative to the system design as the results will show later. A MEMS sensor suite of temperature and pressure sensors can be added to the package as well. Such sensors will enhance system performance, minimizing unnecessary power to the system once the system has reached the desired heat lift. Figure 5: Conceptual design of ARCTIC condenser The expansion valve can made using a simple MEMS device, the diffuser. Figure 6 shows the work done by the author and others [11] on a fuel vapor delivery system for the MEMS REPS project. The expansion valve in a vapor compression system requires efficiency. An idealized vapor compression system assumes the enthalpy entering the expansion valve is the same going out. To achieve this, the entrance loss must be minimized. Work has been done to show that a half angle of 7 is the most effective in minimizing Figure 7: Conceptual design of ARCTIC evaporator and accompanying insulation layer ARCTIC ANALYSIS In order to determine the performance of ARCTIC, an analysis of the refrigeration cycle has been completed. Once parameters are determined for the refrigeration cycle then an investigation of the heat transfer from element to element can be completed. This analysis of the heat transfer was completed using a thermal circuit model. Such modeling splits the thermal system elements into their electrical analogs which allow one to 4 Copyright 005 by ASME

5 use an electrical circuit model code such as Spice by Cadence. From a thermal perspective this is the same as treating each element as a point with mass and utilizing lump capacitance modeling. Thermal circuit modeling provides the added convenience of using industrial code for solving the system. Refrigeration Cycle The refrigeration cycle to be modeled is one which uses R- 134a as the refrigerant. R-134a has a critical temperature of K which determines the evaporator temperature. The compressor is expected to achieve a pressure ratio of 8:1. This pressure ratio will set the condenser temperature to vary between and 315K depending on the position of the fluid. Allowing the compressor to rotate at 1000 rpm ARCTIC can lift approximately 7 W of heat across a temperature difference of 51K from ambient. If the compressor is 80% efficient and the DC Motor powering the compressor is 80% efficient as well, the system will require approximately 13W. This gives a C.O.P. for the system of.08. It is necessary to note that applications requiring more heat lift can be handled by ARCTIC by increasing the rotation speed of the compressor. For example an application requiring 54W of heat lift will require the ARCTIC compressor to rotate at 000 rpm. Assumptions and Equations for a Thermal Circuit Model A lumped capacitance analysis requires that each of the elements in a system have a node which represents the average temperature across the device, thus no single element has a temperature gradient. Thermal circuit modeling requires turning each aspect of the thermal regime into its electrical circuit analog. Every mode of heat transfer, that is conduction, convection and radiation, become resistors. Each element also has a thermal capacitance due to its mass. Temperature and heat are analogous to voltage and current respectively. These analogs are summarized in Table 1 [13] [14]. The analysis assumes that the fluid immediately reaches its steady state temperature. Thus, the fluid in the evaporator immediately reacts as if it has been compressed, expanded, and can now transform into a vapor. Also, the expansion valve is not modeled as it is assumed that the expansion valve can efficiently expand the gas adiabatically. While this is not precise, the literature generally treats the expansion valve this way in refrigeration systems [7]. Finally, interface resistances are not considered here. Analysis The thermal circuit model analysis must begin with the determination of the circuit to be solved, shown in Figure 8. This begins with the determination of what ground represents. In this thermal model, ground will represent the ambient temperature, 98K. This temperature is chosen because the compressor is initially at ambient temperature. Thus, ground determines the initial condition of the system. The voltage sources dictate what temperature the fluid is within the element. In the evaporator the temperature is set to 47 K, in the compressor it is set to the highest condenser temperature, 315 K, and finally in the compressor it is set to 95 K. Table 1: Summary of Electrical-Thermal Circuit Analogs Electrical Analog Voltage, Current, i Resistor, R Capacitor, C Thermal Analog Temperature, T Heat transfer rate, Q Conduction, R cond = R cond ka Convection, Representative equation L 1 R conv = R conv ha Radiation, R R rad rad hrad A Thermal Capacitance, C 1 =, where h = εσ ( T + T )( T + T rad C = mc p s sur Figure 8: Thermal Circuit Model The capacitors linked to the voltage sources represent the capacitance of the fluid inside the element. The resistors linked to the voltage source represent the heat transfer from fluid to the element. The capacitors in the lower half of Figure 8 linked to ground, represent the capacitance of the individual element while the resistor in parallel represents the heat being leaked to the ambient from the element. Finally, the power being introduced into the system through the DC motor is represented by the current source. The values for each of the circuit s sur ) 5 Copyright 005 by ASME

6 components and the values necessary to calculate them are shown in Table. The heat transfer coefficients for the two-phase flows in the evaporator and condenser transferring heat to the plate itself are determined using an analysis similar to that describe by Lou [15]. While heat transfer coefficients change as the quality of the vapor varies, the lowest heat transfer coefficient is included for the thermal circuit analysis. This ensures a worst case performance scenario for the analysis. The heat transfer coefficient for the heat transferred from the condenser to the environment is calculated using a free convection analysis for the upper surface of a heated plate [13]. The model is built using Spice exactly as shown in Figure 8. To enable a transient analysis, the voltage and current sources must introduce a step input to the system. The step input causes the capacitors to react as the heat transferred Table : Values for circuit components and the necessary values for the calculation Conduction Resistors Surface area k L Evaporator Insulation Compressor to Environment Condenser to Environment Evaporator/Compressor Interface Compressor/Condenser Interface DC Motor Wire E Convection Resistors h A Evaporator Fluid Compressor Fluid Condenser Fluid Condenser To Environment Capacitance density volume Mass cp Evaporator E E Compressor E E Condenser E E Evaporator fluid E E Compressor fluid E E Condenser fluid E E through the capacitor is defined as: dt Q = C dt Thus, the capacitor will only react with a change in temperature over time. Using a transient Euler analysis, Spice can output the systems response to the fluid instantly being changed to the steady state temperature. RESULTS In order to determine system performance, individual parameters of the system were varied while the rest were maintained constant. The evaporator is the one aspect of a refrigeration system which can affect the performance of the system the most. If too much heat is leaked to the environment then the system will not have enough capacity remaining to cool an attached device. By varying the resistance of the evaporator insulation according to the thermal conductivity values of different insulation material, the system performance can be determined. The results of this analysis are shown in Table 3. From this analysis it is found the evaporator must have insulation. For the case with no insulation the evaporator loses its capability of lifting heat as 6.3 W of the available 7 W heat lift is lost to the environment. Higher end insulation materials such as the aluminum cryogenic insulation material can reduce the heat leaked to.% of the total heat lift. However, less expensive insulation materials such as aerogel do nearly as well, reducing the heat leaked to.3% of the total. As Figure 9 shows there is only slightly less heat loss when using reflective aluminum insulation. Materials such as foam and aerogel lie on the graph at the point of diminishing returns. Resistance (W/K) Table 3: Insulating material and amount of heat lost to environment Insulation Material Heat leaked (W) no insulation 6.3 glass fiber 1. foam aerogel layer of reflective aluminum evacuated Aerogel Evaporator Insulation vs Heat Loss Layered Aluminum insulation Foam Heat Loss (W) No insulation Figure 9: Evaporator Insulation Resistance vs. Heat Loss To design the compressor it must be determined if insulation would be necessary for system effectiveness. From the model it was determined that insulation would not be useful. 6 Copyright 005 by ASME

7 Table 4 shows that insulating the compressor would lead to high operating temperatures mainly due to the input power of the dc motor which would not be able to escape. Thus, heat would be absorbed by the fluid traveling thru the compressor, reducing the efficiency of the system. Considering this effect, the compressor must be made out of conducting material. Two such effective compressor materials are Aluminum and Ceramic which raise the temperature of the compressor to no higher that 305.6K. The heat absorbed by the refrigerant at this temperature is negligible. Table 4: Compressor Materials and Insulation vs. Operating Temperature Compressor/ Insulating Material Temperature (K) Aluminum 98.5 Ceramic Teflon Insulation 41 Foam Insulation 463 ARCTIC s overall system performance based on temperature is shown in Figure 10. This analysis uses all the optimal values found through the entire analysis and previously shown in Table. The plot shows how quickly the system responds with time. The time constant to cool the evaporator to operating temperature is 1 seconds. This can be dropped further though by operating the compressor at a higher speed which will drive more refrigerant through the evaporator in a given time. Finally, it is valuable to note that adding insulation between the compressor and the evaporator will be imperative. The temperature gradient between layers is large, 10,000 K/m, and thus a layer providing insulation is necessary to ensure that heat leaked from the evaporator is minimized. Temperature (K) Temperature Response Time Time (seconds) Evaporator Condenser Compressor Figure 10: Plot of temperature response time for ARCTIC components CONCLUSIONS This paper has suggested a design for a vapor compression system, ARCTIC. The vapor compression cycle is minimized through the use of MEMS structures, most specifically the check valves, condenser, evaporator, expansion valve, and a sensor suite. The prior accomplishment of the MEMS REPS 367 mm 3 has further enabled a flat design resembling a TEC device with a cold side and a hot side. The performance of this refrigeration unit is 7 W of heat lift at a temperature change of 51 K and a compressor operating at 1000 rpm. The heat lift is scaleable thru compressor speed allowing ARCTIC to be used for a variety of applications. The most intriguing aspect of the design is the efficiency of the system which theoretically can achieve C.O.P. of.08 with the compressor and dc motor set being merely 64% efficient. Furthermore, such a system has been found to be feasible through a thermal circuit model analysis. This analysis has determined the necessity of insulation around the evaporator and between the evaporator and compressor. The addition of simple insulation materials such as foam and aerogel can improve system performance by three orders of magnitude in terms of heat leaked to the environment. The system performs well dynamically with a time constant for cool down of 1 seconds which can be enhanced with startup sequences which initially flood the evaporator with more refrigerant than necessary. The application space for ARCTIC varies. Many electronics such as optics and processors operate more effectively and efficiently at lower temperatures. There are applications in the biological field which require a refrigerator to cool small areas. Finally, the device could be used as a next generation cooling system for computers which are emitting more heat with design advancement making air cooled designs less feasible or intolerably large. ACKNOWLEDGMENTS The authors would like to thank the members of the REPS team and BMAD at the University of California at Berkeley for their useful contributions. Most specifically Chris Hogue, Robert Azevedo, Fabian Martinez, and Mitchell Swanger for there suggestions and analytic help. REFERENCES [1] B. Suman, N. Hoda, "Effect of Variations in Thermophysical Properties and Design Parameters on the Performance of a V-shaped Micro Grooved Heat Pipe," International Journal of Heat and Mass Transfer, vol. 48, pp. 090, 005. [] Pettigrew, K.I., 00, MEMS Based Capillary Pumped Loops for Integral Thermal Management, PhD Thesis, University of California at Berkeley. [3] Yang, R., G. Chen, J. Snyder, J-P. Fleurial, Multistage Thermoelectric Micro Coolers 8th Intersociety Conference on 7 Copyright 005 by ASME

8 Thermal and Thermommechanical phenomena in Electronic Systems, San Diego, May 30-Jun 1, 00. [4] Garimella, Suresh V., V. Singhal, Single-Phase Flow and Heat Transport in Microchannel Heat Sinks First International Conference on Microchannels and Minichannels, Rochester, April 4-5, 003. [5] Jong, S., How Difficult is it to Make a Micro Refrigerator, International Journal of Refrigeration, 7, pp [6] Fu, K., A, Knobloch, F. Martinez, D. C. Walther, C. Fernandez-Pello, A.P. Pisano, D. Liepmann, K. Miyaska, and K. Maruta Design and Experimental Results of Small-Scale Rotary Engines IMECE/MEMS-394, Proc. ASME 001 International Mechanical Engineering Congress and Exposition (IMECE), New York, November 11-16, 001. [7] Dossat, Roy J., 1997, Principles of Refrigeration, Prentice Hall, Upper Saddle River. [8] Yamamoto, K, 1981, Rotary Engine, Toyo Kogyo Co, Hiroshima. [9] Talbot, N.H., 1999, Polysilicon Micromolding of Closed-flow Passages for the Fabrication of Multifunctional Microneedles, PhD Thesis, University of California at Berkeley. [10] Martinez, F.C., N. Chen, M. Wasilik, A. P. Pisano, Optimized Ultra-DRIE for the MEMS Rotary Engine Power System EMN04, Paris, October 0-1, 004. [11] Sun, C.-L. Dynamic Modeling of a Thermally-Driven Micro Diffuser Pump, Proc. ASME 003 International Mechanical Engineering Congress and Exposition (IMECE), Washington D.C, November 16-1, 003. [1] Olsson, A., G. Stemme, E. Stemme, Numerical and Experimental Studies of Flat-Walled Diffuser Elements for Valve-Less Micropumps Sensors and Actuators, vol. 84, issue 1, pp. 165, 000. [13] Incropera, F. P., D.P. DeWitt, 1996, Introduction to Heat Transfrer, John Wiley & Sons, New York. [14] Senturia, S.D., 001, Microsystem Design, Kluwer Academic Publishers, Boston. [15] Lou, Z.D., A Dynamic Model of Automotive Air Conditioning Systems Proc. 005 SAE World Congress, Detroit, April 11-14, Copyright 005 by ASME