IAC-06-C2.7.6 MICRO HEAT SPREADERS BASED ON MICROCHANNELS FOR CONCENTRATED HEAT FLUXES ON SPACECRAFT SUBSYSTEMS.Mr. Rafael Rebolo Gómez SENER Ingeniería y Sistemas, S.A.. Tres Cantos. Spain. e-mail:
[email protected] Mrs. Paula Alvarez Legazpi SENER Ingeniería y Sistemas, S.A.. Tres Cantos. Spain. e-mail:
[email protected] Dr. Johan Steelant European Space Agency/ESTEC. Noordwijk. The Netherlands. e-mail:
[email protected] Dr. Ezequiel González Martínez Universidad Politécnica de Madrid. Madrid. Spain. e-mail:
[email protected] Dr. Benigno Lázaro Gómez Universidad Politécnica de Madrid. Madrid. Spain. e-mail:
[email protected] ABSTRACT A micro heat spreader (MHS) is a micro-fluidic device designed for thermal management of electronic components by means of controlled liquid convection in a closed loop. In contrast to classical fluid loops, the design and the optimization of a micro heat spreader is complex. It requires an adaptation and a good understanding of the fluid dynamic behaviour at this subscale-level along with the fabrication process of micro electro-mechanical systems. Based on analytical predictions, this new approach would increase the heat dissipation with an order of magnitude compared to passive cooling strategies. First, an overview will be given of what has been studied and developed so far on heat exchangers and micro-heat spreaders for dense heat fluxes. Their applicability towards a spacecraft environment will be assessed. Secondly, under design/optimisation aspects, where massive numerical tools are not applicable, engineering approximations based in simplified methods and engineering correlations need to be developed. Classical engineering correlations do not always seem to be applicable for micro-fluidic devices, and new ones from computational or experimental fluid dynamics (CFD and EFD respectively) have been obtained. Simple models featuring the physical behaviour of the MHS allow for sensitivity analysis of geometrical and operational parameters, driving to an optimum design. Finally, a MHS prototype based on microchannels, designed along the above mentioned considerations as a test demonstrator and using single phase fluid, will be discussed. Previously a less exigent design aimed to visualisation, rig calibration and priming process studies have been tested. 1 INTRODUCTION In electronic cooling, the primary critical metrics that it must be met is the device junction temperature, close related to the heat dissipation and method used for allowing it. As temperature in electronics is closely associated with operating efficiency and failure rates, management of thermal loads is necessary to ensure proper and reliable device performance. Elevated temperatures can adversely affect the operation, reliability, power handling capability, and achievable packaging density of the electronic devices. Since the electronics are smaller and faster, they produce more heat. This energy will be 1 and dissipated by them. As consequence of this intermediate heat sink with a relative high temperature. where the heat is transferred to it by forced convection. equipments and assemblies to the radiators. referred as the hot spot: CPU. increasing its enthalpy (with or without changing its phase. assessing their applicability within a spacecraft environment. Secondary Thermal Control System (STCS): Transfers the heat produced by a discrete group of hot spots to the radiators or intermediate heat sinks. in order to maintain the component temperature in the appropriate range. valves. the system will incorporate several subsystems not considered in this study. This component. one of the key parameters will be the thermal resistance. defined as the ratio of the temperature difference between junction and fluid intake bulk temperature to the dissipated heat. Therefore the heat shall be extracted directly from their surfaces by means of local spreaders. therefore the heat can be extracted globally from them (cold plates. The Thermal Control System (TCS) consists of the following sub-systems: Main Thermal Control System (MTCS): Transfers the heat produced by macro scale components. The second aim is to present those thermal-fluids aspects related to the problem (originated mainly by the use of micro scale flow paths). transferred to the radiators (heat pipes. The heat produced by these elements is very high and concentrated in comparison with their volume. . in the following the MHS. 2 2. in order to maintain an acceptable temperature for these components. The selected STCS. From literature survey. and dissipated by them. the heat dissipation systems must also improve accordingly. 2. will be in direct contact with the heat source (again. power electronic. the transmitter defines which MHS shall be operative or no operative in each moment. which monitors the hot spot temperature. transferred to the radiators or intermediate heat sinks from the MTCS. There are two aims on this technical note. Each MHS is placed in contact with the corresponding hot spot. while maintaining the component temperature below a given value. According to the thermal sensors signals. The first is to review the state-of-the-art of the existing MHS. which is focused on the thermohydraulic design of the hot component. sensor. drives the liquid from the spreader to the sink. there are not specific designs able to work within the requirements of high fluxes and relative long distances between hot and cold points. etc). Finally a short description of the MHS developed and their tested performance is presented.) Its mission is to ensure the dissipation of the concentrated high heat flux from the hot spots to an intermediate system from which heat is dissipated by the MTCS to the deep space. according to next paragraph is based on a working fluid pumped in a closed loop from the spreaders to the sink. The heat flux produced by these elements is low in comparison with their volume.1 Thermal Control System description The heat generated by the components in a satellite shall be effectively dissipated to the deep space.2 Secondary Thermal description Control System 2 SPACE APPLICATION Because the wide spectrum of possible applications it is difficult to define a particular optimum solution for the heat dissipation subsystem.efficiently removed in order to maintain the electronics temperature below a certain value. thus as the formulation and methods applicable in their resolution. because generality. and to the spreader again closing the circuit. in such way that heat is transferred from the hot spot to each MHS by conduction. Therefore a general description of the problem and the possible architecture will be presented only as a guide. The transmitter. Also. etc). understanding that many other solutions are possible. While the electronics improve their capabilities. The fluid inside the MHS takes the heat by forced convection. pipes). The operating or no-operating mode of each MHS can be commanded by a thermal sensor. depending on the design). etc. thanks to a hydraulic system (pump(s). phase change material. nevertheless any claims of an excellent merit figure have been taken into account. the jet impingement and the liquid cooling loops based on microchannels. etc. belonging to the MTCS. only some of the above systems are candidates because the high heat fluxes involved. Therefore heat will not be transferred directly to radiators. Maximum temperature difference on interface 40 C 2 Geometric Hot spot size 20 x 15 mm 2 Table Ia: MHS requirements Imposed by the spacecraft thermal control Thermal Sink temperature 60 C (equal to the cooling fluid inlet temperature) Geometric Minimum mass and volume (target of < 0. immersion cooling. etc. maintainability. liquid cooling loops. vapour chambers. mainly at laboratory level. Imposed by the cooled component (hot spot) Thermal Heat flux up to 150 W/cm .). loop HPs. These values no always are clearly defined in the papers. heat pipes. It is difficult to compare one to other because the different testing condition. There exits several reviews were more detailed description can be found [1]. This reduced set of requirements imposed to the current design is summarized in tables Ia to Ic. and some application zones have been traced (these regions can change as the number of represented devices increase. capillarity pumped HPs. The electronic box to be refrigerated will contain one or more components of high thermal flux (hot spots). Aiming to maintain generality in the MHS design. some of them proven and other at different states of development. and must be only taken as a guideline). geometries.06 kg and 40 x 3 40 x 10 mm excluding inlet/outlet) Table Ib: MHS requirements 3 . On the plots has been indicated a rectangular zone corresponding to the interest area for space application.Both philosophy and requirements of the MHS are dependent of the spacecraft/vehicle where it is going to be installed thus the mission to be fulfilled. even if it could be discussible. and inside it there is a hollow circle showing the target of this study. hot spot area pressure drop. a general scenario representative for a wide spectrum of applications is defined as: The STCS will use a heat sink. This set of scarce data has been plotted in the two following plots (figures 1 and 2). but limiting the possible scope of applications. Junction temperature equal or lower than 100 C. And investigation of the technical literature can provide some data about the performance of these devices. etc. In particular will be considered the heat pipes (HP) and its derivatives (miniature HPs. 2. but to an intermediate cold plate. etc [2 to 19]. Some of them are: doublets.3 Requirements The requirements imposed in this study to the system will be aimed to performance and operation. Therefore it has been extracted from the literature some merit figures such as heat fluxes. For the purpose of this paper. thermal resistance. being this cold plate part of the MTCS and with a higher temperature than the external radiators. sizes. neglecting such important aspects as life. in a distance range around half to one metre. considering the number as a record. Imposed by the spacecraft and the mission Operating temperature 40 to 80 C Non-operating temperature (-30 to 90 C) Minimum power consumption Operating at zero g No leakage Table Ic: MHS requirements 3 MHS REVIEW There exists an elevated number of devices and methods for cooling. mechanical/electrical environments. jet impingement. as cold point for rejecting heat. etc. ). During some years many of the discrepancies between macro and micro scale results were attributed to particular fluid behaviours. 2: Different cooling concepts for small hot spots refrigeration and dense heat fluxes. Most of them do not fulfil any requirements. Length is a measurement of the distance between hot source and cold sink. Some of the more evident advantages are [2]: Increased effectiveness by integration of cooling system with payload Increased freedom in locating electronic or science payload Removal of large heat fluxes over large distances Effect of temperature in the fluid must be considered (increment of 40 C between inlet and outlet is normal. should be considered in a practical application). In this study has been selected as MHS based on microchannel pumped liquid cooling system with a single-phase working fluid. and the traditional non-slip and non-temperature jump boundary conditions. The origin of discrepancies between usual predictions and testing (after neglecting ill defined tests) is that in microscale must be considered the following effects: Fluid and important thermal entry length are From the plots. in the interest area both microchannels cooling device and loop heat pipes (some experts considers these loop HP values as questionable ones) could cover the thermal performance. The data have been extracted from technical literature and can be considered as the state of the art. for the aforementioned range of duct sizes (dh > 50 microns). etc. 4 1000 Host spot heat flux (W/cm 2) LHP mLHP mHP HP MHS DESIGN 100 10 1 0.) After the analysis of the literature and some CFD testing the conclusion. piping. etc.01 0. accumulators. volume. Also. Neither heat flux nor temperature is constant along the pipe wall and around the perimeter. The lines try to identify application areas. 10 1 0 5 10 Hot spot area (cm 2) 15 20 Fig. the design of heat exchangers based on microchannels is not as straightforward as in meso/macro scale.1 Length (m) 1 10 Fig. phenomena neglected in macro flow were proposed as explanation for the discrepancies (non newtonian behaviour. being the pipe length of the order of a centimetre). Main difficulties arise because the classical engineering approach based on correlations does not apply. is that no strange behaviours are found in the flow (except perhaps the early transition to turbulent). 1: Different cooling concepts for small hot spots refrigeration and dense heat fluxes. For active systems. Today.1000 Hot Spot Heat flux (W/cm 2) mchannel mHP mLHP 100 mjet LHP Target Precision in the temperature control of electronics/payload by micropump flow control Ability to function in adverse and zero gravity Of course there also are some drawbacks as to be an active system with the possible increase on weight because additional equipment (pumps. Comparison of MHS based on microchannels with HPs. being experimental devices. A final trade-off will depend of particular application and architecture.5 mm. which put in doubt the applicability of the classical Navier-Stokes equations. etc. electric double layer. In this paper a microchannel is considered as a duct (whatever cross section) with hydraulic diameters between 10 microns and 0. 4 . functionally there is not limitation for this length (nevertheless constrain because mass. In a similar way. which could be applicable to simple engineering models. based on average channel bottom wall temperature.1 Heat transfer and pressure drops α n−d The approach followed in this work has been to obtain correlations by using CFD and published results. μ and cp are density. d 0.2537⎜ ⎟ ) ⎝H⎠ ⎝H⎠ ⎝H⎠ ⎝H⎠ ⎝H⎠ 2 3 4 5 Nu = (Nu ∞ + Nu op )⋅η fin h ⋅ dh k fluid Where the Nusselt number is defined as Nu = The pressure drop at inlet/outlet manifolds must be computed by CFD and the data can be correlated in a simple way in order to be applied in the models. operating point and geometry. The Nusselt number for thermally developed and laminar flow is a function of the channel aspect ratio.9564⎜ ⎟ . ρ. for the pressure drop in the channels has been taken the following correlations: ΔPμc = 4.776 + 0.2 ⋅ ⎛ Re⋅ Pr⋅ h ⎞ ⎜ L⎟ ⎝ ⎠ Nu op = 0.426 ⎟ L ⎞ ⎛ ⎜ ⎟ ⎜ 1 + 2.8 with K1 and K2 from numerical analysis (or handbook if the geometrical configuration is not too much complex). L and W are the footprint size. fluid inlet temperature and evacuated heat.586*H – 0. also maintains the classical expression with slightly modifications from CFD results (Hf refers to the ratio between fin and channel width). equal to the ratio between channel depth and width.5 ⎟ ⎟ H ⎜⎝ kw f ⎠ tgh( N ) ⎠ ≈ 1.1⋅ ⎛ Re⋅ Pr⋅ h ⎞ ⎜ ⎟ L⎠ ⎝ 0. dynamic viscosity and specific heat for the fluid. 5 ⎟ ⎜⎜ kw f ⎠ ⎝⎝ ⎠ η fin for pumping power. average channel lateral walls heat flux and mean fluid bulk temperature and properties: −0. This resistance can be split in three terms: where 5 . Nu.1.4⎜ Re n ⎟ ⎟ ⎜ ⎟ Dh ⎟ ⎜ ⎜ ⎟ ⎝ ⎠ ⎝ ⎠ ( ) ( ) and α = 96.9467⎜ ⎟ .0.3553⎜ ⎛1⎞ ⎛1⎞ ⎛1⎞ ⎛1⎞ ⎛1⎞ ⎟ + 1.467 d 1 + 0. the sub index ∞ to ducts with dh/L → 0 and the op one to the operation point (takes into account entry length). These pressure drops were written as: ΔPin / out = ⎛ K 1 2 ρvin / out ⋅ ⎜ K1in / out + 2in / out ⎜ 2 Re in / out ⎝ ⎞ ⎟ ⎟ ⎠ being h the convection coefficient. for the Nusselt number.1364 ⋅ = 1.5 ⎞ ⎛⎛ k f ⎜ ⎜ Nu ⎞ 1 + H ⎟ ⎟ tgh ⎜ 0. a the channel width and Q the evacuated heat.0347*H2 + 0. H.0233⎜ ⎟ ⎟ Dh ⎟ ⎟ ⎜ ⎜ ⎝ ⎠ = λ ⋅ α = ⎜1 + ⋅α 0.4.1364 ⋅ ⎝ 0.0007*H3 While the operation is a function of Reynolds and Prandtl number and channel relative length.8 ⎛ L ⎞ ⎞ ⎛ ⎜ ⎜ Re n ⎟ ⎟ ⎜ 0. In previous expressions.5 N ⎛⎛ k f ⎞ ⎜ ⎜ Nu ⎞ 1 + H ⎟ ⎟ ⎟ H 0. 0.1. These expressions show the influence of the different parameters involved in design: fluid properties. * (1 . would be Neglecting entry effect the expressions can be simplified and written in non dimensional way as: ΔPμc a 3 ⎡ μ ⎤⎡ Q ⎤ L ⎢ ⎥⎢ ⎥ ⎢ ρc p ⎥ ⎣ ΔT ⎦ W ⎣ ⎦ = 2 α ⎡ (1 + H f )(1 + H ) ⎤ ⎢ 8⎢ ⎣ H3 ⎥ ⎥ ⎦ for pressure and ∏ μc a 3 ⎡ μ ⎤⎡ Q ⎤ 2 L ⎢ ⎥ 2 ⎢ ⎥ ⎢ (ρc p ) ⎥ ⎣ ΔT ⎦ W ⎣ ⎦ = 2 α ⎡ (1 + H f )(1 + H ) ⎤ ⎢ 8⎢ ⎣ H3 ⎥ ⎥ ⎦ η fin is the fin efficiency. was extracted from CFD the following correlations. Nu∞ = 1. dh the hydraulic diameter and k the thermal conductivity.7012⎜ ⎟ + 0. Per instance. Total pressure ΔPMHS = ΔPin + ΔPμch drop + ΔPout .2 MHS thermal resistance α α μ 1 2 L with ρv c f c f = n−d = n−d 2 Dh Re Dh ρvDh The thermal resistance definition is the classical one based on junction temperature. Operation point 6 . when restrictions to the MHS (such as maximum junction where the entry length effect have been neglected in order to simplify the Nussetl expression. θ ≈ a conv 1 H (1 + H f ) −1 a 1 H (1 + H f ) −1 η ≈ η 2 LW Nu ⋅ k f (1 + H ) LW Nu ∞ k f (1 + H )2 Table II: Qualitative effect of the different geometrical. but that implies a variation in pressure drop and pumping power. As a guide. working fluid and operation parameters in the MHS performance. This value should be analyzed in detail later. ↑ means a parameter increase. Low separation between channels improves the design (at least up to a minimum in terms of θconv). ↓ a decrease and a double arrow an increase/decrease with an exponent higher than 1. L and W are dependent of the hot spot’s size. When these parameters are fixed. B. Assuming an allowable ΔP for the pressure drop (inlet/outlet pressure drop neglected) this can be rewritten as: θ heat a 3 ΔP ⎛ μ ⎜ ⎜ ρc ⎝ p ⎞L ⎟ ⎟W ⎠ = 2 α (1 + H ) (1 + H f ) 8 H3 or θ heat (a 3 ∏ ) 1/ 2 1/ 2 ⎡⎛ μ ⎞ L ⎤ ⎟ ⎥ ⎢⎜ ⎜ (ρc )2 ⎟ W ⎥ ⎢⎝ p ⎠ ⎦ ⎣ ⎡α (1 + H )2 (1 + H f =⎢ H3 ⎢8 ⎣ )⎤ ⎥ ⎥ ⎦ 1/ 2 if the constraint is the pumping power. being only function of the channel aspect ratio. Structural or manufacturing considerations dictate this value (about 50 to 75 μm is considered a practicable value). is dictated by structural considerations. that is a function of the k w LW minimum thickness between MHS bottom and channel lower wall and the material. e. there is a value that minimized the total thermal resistance.θ MHS = θ MHS T j − T inlet with some mathematical manipulation and a simple model where only transversal conduction is considered can be rewritten as: A. and the target (requirements) imposes them. Distance between MHS’s interface and bottom side of the channel. θ cond ≈ Geometry Q Q = θ cond + θ conv + θ heat = T j − Tw + Tw − T exit T exit − T inlet + Q Q Parameter Π ↓ ↓ ↑↑ ⎯ ↓ ↓ ↓ ↓ ⎯ ⎯ ↓↓ ↓↓ ↑ ↑ ΔPμc ↓ ↓ ↑↑ ⎯ ↓ ↓ ↓ ↓ ⎯ ⎯ ↓ ↓ ↑ ↑ θcond ⎯ ⎯ ⎯ ↓ ↑ ↓ ⎯ ⎯ ⎯ ↓ ⎯ ⎯ ⎯ ⎯ θconv ↓ ↓ ↓ ⎯ ↑ ↓ ⎯ ⎯ ↓ ↓ ⎯ ⎯ ⎯ ⎯ θheat ↓ ↓ ↑↑ ⎯ ↓ ↓ ↓ ↓ ⎯ ⎯ ↑ ↓ ↓ ↓ Working Fluid H↑ Hf ↓ a↓ e↓ L↓ W↑ μ ↓ (ρc p ) ↑ kf ↑ kw ↑ Q↓ ΔT ↑ ΔPμc ↑ Π↑ e 1 . Minimum value is considered around 1 mm. though the benefices decrease at higher H. With the above analysis and the supposition that both the heat to be dissipated and the maximum increase in cooling fluid temperature are given. As a guide the aspect ratio would be greater than 3 or 4. Minimum thickness (compatible with structural integrity and deformations) and maximum thermal conductivity are desirable. Pressure can be minimized if the microchannel is split in branches (branch length: L/n). C. can be stated that: It is advantageous to select channel aspect ratios as high as possible. the lower Hf needs to be selected. Microchannel width has a great impact on pressure loss and pumping power. or written in nondimensional way: θ conv L ⋅ W ⋅ k f a ≈ 1 H (1 + H f ) −1 η Nu ∞ (1 + H )2 With the above equations is very easy to plot the effect of the different parameters or write some qualitative table as Table II. θ heat ≈ ΔT fluid 1 ≈ Q ρc pϕ In the previous expressions the thermal resistance can be changed with a variation in the mass flow. 98 10 -3 4. ( ) 54 % .06 0. Only the convective and heat resistances have been presented because conduction one is a constant value proportional to MHS bottom thickness.39 0. such as low viscosity and high thermal capacity and conductivity. For the fluid.08 0. High atmospheric boiling point or low vapour pressure at the operating temperature. In particular requirements 1.3 Working fluid Water Ammonia -77 100 (4) 545 5230 0.temperature and fluid entry temperature) are imposed. that the thermal capacity.65 W/m-K Viscosity 4. Though the working range is above 60 C (temperature from the returning fluid from sink). In general.12 0. Electrical conductivity (it is important if the fluid comes in direct contact with the electronics.35 Acetone -93. No or minimal regulatory constraints (environmentally friendly. Low freezing point and burst point (in this case < -30 C) 3. and possibly biodegradable) 8. Cond.8 10-3 ( ) 22 % .94 10 Tfreezing (Tpouring). 4 y 6 are important.7 10-4 Kg/m-s From the previous analysis can be deducted some requirements for the selection of the working fluid. 4. 0 C Tboiling 100 C Density 983 3 kg/m Cp 4185 J/kg-K Ther. units are in IS.2 100 ( ) 780 2130 0. H2O+ NH3 (1) -46 100 (5) 900 4300 0.59 -4 0. high thermal capacity and high thermal conductivity are advantageous for the MHS. H2O+EG (2) -40 107. the requirements may vary depending on the type of application. Good thermo-physical properties (high thermal conductivity and specific heat and low viscosity as already stated).02 0. From the above requirements some are more critical than others for the considered application in this study.1 1018 3600 0. 4. including water as reference. 3.14 0. 2.90 10 2. 7 . non-toxic. Table III shows typical properties for the liquids that have been selected for comparison. low viscosity. ( ) Pressure ∼ 4 to 5 Bar 1 2 3 4 5 Table III: Working fluid typical properties at 60 C. improvement in the first will be more important in the final result (this reason justified the greater weight when selecting a fluid of the thermal conductivity.52 10 -3 1. ρcp). while the others can be relaxed. The following plot shows the typical behaviour of thermal resistance for fixed pressure loss and geometry (fixed H and Hf) as a function of the channel width. Economical 9. following there is a list of some general requirements for single phase flow [20]: 1.1 XCELTM 500 < -40 766 2330 0.20 0.16 0.04 0. high thermal conductivity is advantageous (in consideration with working fluid compatibility). 3: Example of the existence of an optimum for thermal resistance (conduction thermal resistance not plotted because dependence of channel geometry is almost negligible). than θ conv is greater θ heat in the interest area (to the right of the minimum). Properties are given at the operating temperature (around 60 C) and they correspond to 1 Bar (else indicated). Non-corrosive to materials of construction 7. High flash temperature point and auto-ignition 6. ( ) Pressure ∼ 100 Bar .2 0.16 5 GALDEN HT130 < -40 130 1650 1212.00 0 50 100 150 200 a (microns) 250 300 350 400 conv heat conv+heat Fig. kfluid.18 0. 2.8 1057 3340 0. This precludes the use of pure water (otherwise and excellent election) and some other fluids must be considered.( ) 54 % . the fluid needs to withstand without freezing up to –30 C. For the MHS material.15 10 -4 8. or if it leaks out of a cooling loop or is spilled during maintenance).5 10 -4 1. Therefore.38 H2O+PG (3) -40 106.13 -4 5. a. Good mechanical and thermal stability (specially under severe environments) 5.10 0. 0. 83 6. for equal channel geometry. Influence of the allowable pressure drop (equivalent to mass flow). that precludes the use of copper. because can be higher than the corresponding saturation one. While in macroscale heat exchanger. While the fluid heat capacity has a direct effect on θ heat .00 2. Applied to a microchannel. Effect of splitting the channels (1 or 2 branches).08 0. in order avoiding boiling (this design is for single phase flow MHS) to be pressurized (around 100 Bar).24 1. Inspection of the table shows that water and ammonia are the best fluids under the point of view of performance.73 31. 0.00 θheat/(θheat)water (Π constant) θconv/(θconv)water Table IV: Qualitative effect of the working fluid in some merit figures.18 5.32 4. see figure 3). (compatible with aluminium) and fluids based glycol (per instance TYFOCOR L50 compatible with both aluminium and copper).00 4.00 1.00 1.77 2. perhaps the better option would be the mixture ammonia/water.Using a simple model as described above.77 4.11 0. leading to boiling even if fluid bulk temperature is far from saturation. but manufacturing method can limit the possible geometries. the product ρ ⋅ c p usually is one of the main criteria for fluid selection.00 1.e. This ratio does not take into account the necessary increase in MHS bottom wall thickness in order to compensate the higher pressurization.50 13.30 1.65 1. On the other hand.42 1.00 1.73 5.18 2.32 1.75 0.00 13.07 0.39 1. The second one requires. Nevertheless the first one cannot be used because the operating temperature range imposed by the requirements. 8 . In the fluid selection is very important the aspect of compatibility between fluid and MHS material.4 Optimization From the table. 4. for a given fluid and hot spot size. with better conductivity than aluminium (this last compatible wit ammonia). the temperature at the bottom channel wall must be checked.00 1.01 1.29 0.15 1. For this work two fluids are considered as appropriate for the application.00 3. Water ΔPμc/( ΔPμc)water Π/( Π) water θcond/(θcond)water θheat/(θheat)water (ΔP constant) Ammonia H2O+NH3 H2O+EG 0.75 1. Therefore it is needed to perform special analysis for corrosion inhibitors and to study the long-term fluid stability in a hostile environment (mainly radiation during ten or fifteen years of expected life). This pressurization translates into potential leakage and more robust structural design (and therefore a thicker MHS increasing conduction thermal resistance) and at the end additional weight.00 1.40 1. From analysis can be selected an optimum geometry.12 Thermal resistance (K/W) 0.10 1. On one hand the θ conv is inversely proportional to kfluid.06 0.67 1.09 0. convection coefficient). this number means that half a fin can be corrode in one year.00 3.63 5. and typically the sensibility is greater than the one of θ heat (mainly because θ heat is smaller than θ conv .73 5.65 1. pumping power and thermal resistance (Table IV).89 8.29 0.06 GALDEN XCELTM HT130 500 3. in this particular design is more important the fluid thermal conductivity. 4: Example of possible optimization with a simple model. the MHS geometry can be optimized. and this fact will depend mainly of fluid thermal conductivity (i.73 2.86 Acetone 2. can be estimated the effect of the fluid in the pressure loss.05 0 50 100 150 a (microns) 200 250 300 100 kPa (1 branch) 50 kPa (1 branch) 25 kPa (1 branch) 100 kPa (2 branch) 50 kPa (2 branch) 25 kPa (2 branch) Fig.24 1.89 2.31 1.71 H2O+PG 4. The variation in the parameter θcond/(θcond)water for the ammonia and its solution is only due to material compatibility. Normal criterion for compatibility defines it as excellent when corrosion ratios are less than 2 mils per year.10 3. with a single phase design. Figures 4 to 7 show some examples of the optimization plots that can be obtained.10 0. the thermal conductivity influences two aspects. The optimization criteria can be based on minimum thermal resistance for given allowable pressure drop (or pumping power) or minimum pressure drop for a given thermal resistance. without a deep compatibility analysis: the water/ammonia mixture With the model described above. filament of ∅0. 90 and 100 microns (microphotographs obtained by TEKNIKER). Follow a short description for each one.Per instance in our first design was selected electroerosion as the method for microchannel manufacturing. 5.08 K/W.1 MHS-1 Visualized priming process and to check flow uniformity and bubble absence inside the MHS. 9 The objective of this model was: . Δ P (kPa) 100 H:3 (1 branch) H:3 (2 branch) H:4 (1 branch) H:4 (2 branch) 10 0 50 100 a (microns) 150 200 250 Fig. 5: Example of possible optimization with simple models. For the second model with improved performance was selected for manufacturing a micro milling process. Also the channel width was driven by available filament diameters. General plot showing thermal resistance isocontours as a function of the channel width and aspect ratio. 8: Electroerosion method. In this case the channel depth was increased and the fins were thinner. Hyperbolae correspond to manufacturing constraints. The shadowed area shows the interest zone. for constant thermal resistance of 0. Effect of split the channels with varying channel aspect ratio. 2 and 3. To test manufacturing procedures. Fin deformations as a function of its thickness. such as limited fin aspect ratio (4 can be obtained by electroerosion in aluminium and 87 with micromilling). However tests showed that because the high thermal properties of the selected alloy (and lower mechanical ones) the minimum thickness was around 110 microns (figure 8). however the channel width increased (because the milling tool thickness). 7: Example of possible optimization with simple models. Fig. Final channel width will be a compromise between performance and manufacturing. 1000 Fig. for a given fin thickness and pressure drop. named MHS-1. Based in previous experience with normal aluminium alloys it was expected to obtain fin thickness around 70 microns and depths about 400 microns. It is at the right side of the minimum in order to avoid the sharp thermal resistance increase on the left.1 mm. from left to right 70. Fig. 5 MHS MODELS Along the project were developed three different models. 6: Example of possible optimization with simple models. Typical values for the spreaders are summarized in the Table V. with small differences in the inlet and outlet sections. Main differences with the previous ones was the higher channel aspect ratio.8 MHS-3 41x29x7.1 0. Visualization model. equivalent to a power of 450 W. The microscale zone was similar to MHS-1. obtained by the change of the manufacturing procedure from electroerosion to micromilling. In this rig a key component was the heater. 5. was made a new MHS aiming to improve pressure drop and heat transfer characteristics. and a redesign of the inlet/outlet manifolds (figure 11).2 split 130 3. this element represents the hot component and must be able to put a uniform heat flux on the MHS greater than 150 W/cm2. 6 TEST BENCH Fig. For testing the MHS models was necessary to develop a test bench. Fig.2 MHS-2 Fig.6 0.6 Table V: Geometrical definition for MHS 1 to 3. Other aspect that required special attention was the characterization of the thermal interface material (TIM). Prototype of microchannel heat spreader 2.2 split 130 3.To test hydraulic models To test and tune the test bench The model had a glass cover.3 MHS-3 Based on the previous experience. 11: MHS-2 and MHS-3. and because it the inlet/outlet was located in its bottom part.1 0. increasing its thickness and worsening its thermal resistance. These topics are extended in the following paragraphs. MHS-1 Size (w/o connectors) Type a (microns) H (b/a) Hf (wf/a) 50x50x23. This was the first prototype aimed to thermal performance. The size was higher than the requirements (Table V).8 MHS-2 37x29x7. 10: MHS-2. 5. The microchannel manufacturing was also made by electroerosion and the union of the two components (upper and lower cover) was made with friction stir welding (figure 10). The union of the different parts was done by gluing and by pressure fasteners (figure 9). 9: MHS-1. 10 .7 split 145 6. the heater has eight 0.2 Heater description A hydraulic loop that is composed by a deionised water1 reservoir. The heater consists in a copper block. the MHS. General view. An electrical autotransformer.5 mm below the MHS interface. depending of heater thermal isolation). 11 . and the heater. This value is corrected with a heat flux dependent correlation. the first one was used with MHS-1 and MHS-2 and it was fitted with 500 W (1x300 W and 2x100 W cartridges). a filter. 1 In the pedestal (the prismatic upper part). the outlet section measurement and the return to the water reservoir closing the loop. the second one was used in MHS-3 and it was able to deliver up to 700 W (1x300 W + 2x 200 W) in order to obtain the target of 450 watts evacuated in the MHS (heat losses resulted in 30 % and 20 % of the electrical power. the flow rotameter. The measurement module consists in: Pressure: differential between inlet/outlet and absolute at inlet. The test bench (figure 12) consists in: Fig. the electrical power measurement zone. where thermocouples will be installed. then the inlet measurement section. Inside the block there are three thermal cartridges. 8 (10 for TIM characterisation) thermocouples in the heater. in the MHS contact area (figure 14). 13: Test bench. which presses the MHS against the hot surface by means of a plate and screws on the columns. a tap. 2 thermocouples at outlet.5 mm and 6. These holes with a depth varying from 6 to 10 mm allow the measurement of temperature and its extrapolation to the interface in order to estimate the junction temperature. Two heaters were built along the project. Temperature: 2 thermocouples in the fluid at inlet. Can be seen the three holes for the thermal cartridges and the upper pedestal. 14: Numerical model of the heater. Fig. 12: Test bench. The heater is fitted on a ceramic base with a four columns structure. The inlet and outlet sections where enthalpy increase and pressure drop are measured are shown in conjunction with the heater and MHS (not visible because isolation layers and the fixing structure) Water was used during test because simplicity and security. The entire heater is isolated aiming to reduce convective heat losses. shaped in such way that a uniform heat flow is obtained at the top.1 General rig description 6. a pump discharging in a pressurized damping tank. the MHS will be fitted.5 mm holes at two levels: 1. This correction was obtained from numerical computation. accounting for the distortions in the temperature field caused by the thermocouples.6. On the top. Volumetric flow Fig. 6.14 TIM Thermal Resistance (K/W) 0. on the right the upper one. The rig was mounted repeatedly in order to characterize the variations induced in the paste application by the manual procedure.12 0. 17: TIM characterization. 12 . As can be seen the dispersion in the temperature range of 350 to 380 K goes from 0. the value measure for the thermal resistance is: θ set −up = θ MHS + θ TIM . 470 460 450 440 430 420 410 400 390 380 -10 -5 0 5 10 15 20 Fig.3 Thermal Interface Material (TIM) In a real application. it corresponds to the set-up value minus 0.10. ΔT in the TIM corresponds to the temperature jump in ‘X’ equal to zero. 0. the temperature gradient and therefore estimation for the heat flux at the interface surface can be extracted from the thermocouple readings. as TIM material was chosen a commercial thermal paste. the lower one is the nominal value provided by the TIM manufacturer and it corresponds to a 0.08 0. The upper line corresponds to the expected resistance during the first working hours. On the left the lower block. Then it is necessary to know the θ TIM in order to make a realistic estimation for the MHS thermal resistance.02 0. therefore when MHS thermal resistance is presented.00 300 350 400 TIM Temperature (K) 450 Fig. θ TIM = 0. (a) (b) Fig. After the dilemma of which value must be selected for TIM.Also. An aluminium adaptor with two pieces was fitted to the heater.03.03 K/W is selected. after several cycles and more than 200 working hours.06 0. the estimation of temperature jump across the interface and the global results for the TIM thermal resistance. that subtracted from θ set −up will provide a conservative approach giving a greater value for θ MHS . The dots correspond to experimental data obtained for the different rig assemblies. With the experimental set-up.003 inch thickness. 16: Example of temperature variation across the aluminium blocks.04 0. In the following. allowing plotting the ratio ΔT/Q ( θ TIM ). Can be seen the holes for the thermocouples used for obtaining the ΔT across the TIM. In figure 17 have been plotted two lines. For simplicity.10 0. it was decided to take the minimum one. and the thermal gradients in the upper and lower part were obtained and extrapolated to the interface.04 to 0. The thermal paste was applied manually on both interface surfaces. This value is based in the long term operation. 15a & b: Aluminium blocks for TIM calibration. Figures 15 to 17 show the aluminium blocks. the MHS will be mounted on the hot component with a TIM in order to reduce the contact thermal resistance. Test 1 0.1 0. Fig. (K/W) 0.14 Thermal resistance (no TIM). MHS-3.5 0. The test were done up to a heat flux of 120 W/cm2 (figures 19 to 21). 19: Pressure drop in the prototypes. as expected from the improvement in the design.06 Test 1.000 0. Values of 0.000 Test-2 Theoretical Target 0.02 0.5 0. MHS-3 th.015 0. 20 and 22).7 Total Pressure drop (+rig) (Bar) 0.3 0.14 0.025 Fig.015 Mas flow (kg/s) 0. 190 W 0. such as HiThermTM005 (∼0. They present the junction temperature for the nominal heat flux and fluid temperature inlet.020 0.000 MHS-1 MHS-2 MHS-3 MHS-1 th. The values include the additional pressure drop due to inlet and outlet measurement sections. Table VI resumes the main results. MHS-3 th. 0.010 0.020 0.42 Bar.025 Fig.010 0.010 0. 0.06 MHS-2 Test-1 0.015 Mass Flow (kg/s) 0.020 0.12 0.0 0. 20: Pressure drop in the prototypes.2 0. MHS-2 provided the expected thermal performance. reduced the thermal resistance and pressure drop to very low levels.020 0. The picture shows the instant when channels are just filled. 0.03 K/W for the TIM). 21: MHS-2 Thermal resistance (after subtracting 0.005 0.025 Mas flow (kg/s) Fig.09 Bar and 0. Comparing with figure 21.03 K/W for the TIM). 18: Visualization result from MHS.10 0.000 MHS-1 MHS-2 MHS-3 MHS-1 th.04 K/W with a pressure drop of 0.00 0. As in 19 but with the associated rig pressure drop subtracted. 0.025 0. (K/W) 0.4 0.018 kg/s and a pressure drop of 0.02 kg/s were obtained with heat fluxes up to 180 W/cm2 (figures 19.2 0.3 0.1 0. with a water flow of 0. Fig.06 K/W.08 0. MHS-2 th.028 K/W).08 0.010 0. 350 W Theoretical Target 0.005 0. obtaining thermal resistances of 0.0 0.00 0. and the maximum heat flux that it is possible 13 .10 0. Thermal resistance (w/o TIM).015 Mass Flow (kg/s) 0. 7 EXPERIMENTAL RESUTS The principal results from MHS-1 were the checking of the hydraulic design and the manufacturing techniques (figure 18).005 0.4 0.6 0.04 0.6 0. the modifications incorporated to the design leaded to a considerable improvement.04 Test 1. MHS-2 th.005 0.02 0.Total Pressure drop (Bar) The selected value for the TIM agrees with standard material used in space industry.12 MHS-3. 22: MHS-3 Thermal resistance (after subtracting 0. responsible for the welding process. We are in debt with the people from the Laboratorio de Mecánica de Fluidos (LAMF) from the E. 1. 8. Mudawar.019 0. 11.019 0.020 0. “Micro/nano spacecraft thermal control using a MEMS-based pumped liquid cooling system”. responsible of the manufacturing of the MHSs and the people from LOKTER. “Electronic component cooling using surfactant solutions flow through micro-channels”. Applied Thermal Engineering 25 (2005) 6.M. MHS performance without TIM Mass flow (kg/s) MHS-2 MHS-3 MHS-3 0. they are time consuming.A: Shedd. 5.03 K/W) 9 REFERENCES 8 CONCLUSIONS In relation with the possibility to evacuate the high heat fluxes in space. Also some manufacturing techniques have been tested. “Two phase flow in high-heat-flux micro-channel heat sink for refrigeration cooling applications” Part I & II. K. Universidad Politécnica de Madrid. though CFD codes can be used for design.T. Simons. de Ingenieros Aeronáuticos. This work has shown the possibility to use these conventional engineering tools in the design of heat exchangers based on microchannels. Nov 2005 2. Cambridge.G. But there are yet some open points in relation with the working fluid and material compatibility. (2004) 7. These manufacturing methods can be applied in space. N 4. Part I: Heat transfer data using FC-72”. and without expensive computations. Alberto José Herrero and Mr. This work has been partly funded by the Spanish Ministry of Industry through PROFIT. for final performance verification The optimization process is better performed using simple engineering models. Kim. Lee and I.07 0.020 0. Birur et al. “Spray impingement cooling with single. “Advances in High Performance Cooling for Electronics”.06 0. “Single nozzle spray cooling heat transfer mechanisms”. but new strategies must be considered for massive production. J. Referring to the design methodology. K. MHS performance with TIM (0. Malamut. Clemens J.020 DP (kPa) 50 50 10 10 θ MHS 0. Robert E. Pautsch and T. International Journal of Heat and Mass Transfer.06 0. “Single phase liquid cooled microchannel heat sink for electronic packages”. 4.04 Tinlet (C) 60 60 60 60 Q (W) 450 650 450 1000 q (W/cm2) 150 225 150 325 Tj (C) 87 100 78 100 Table VIa.04 0. able to represent the effect of the multiple parameters involved in the design. Zhang et al. the study has shown that it is possible to do that. Horacek. Sabino Azcarate from the Fundación TEKNIKER. Leemaster and S. thus as the long term stability of the fluid in the space environment. J. 2003. International Journal of Heat and Mass Transfer 48 (2005). “Microscale thermal engineering of electronic systems”.C. Electronic Cooling.020 DP (kPa) θ MHS + Tinlet (C) Q (W) q (W/cm2) Tj (C) θ TIM 0.and multiple-nozzle arrays.Y.019 0. Goodson. International 14 . Lasance. where the MHS cost is insignificant. G. Massachusetts Institute of Technology.T. Proceedings of Rohsenow Symposium on Future Trends of Heat Transfer. Massachusetts. H.S. B. 3. It is better to use them in: Getting correlations as a virtual test bench For the detailed analysis of particular configurations And of course. A. Vol. The authors want to express their gratitude to Mr.09 0.07 50 10 10 60 60 60 450 450 575 150 150 175 101 92 100 Table VIb. Kiger and J.to evacuate without overpass the maximum junction temperature. Mass flow (kg/s) MHS-2 MHS-2 MHS-3 MHS-3 0. Miniature loop heat pipe for electronic cooling 20. S. Yu. “Experimental study on silicon micro-heat pipe arrays”. R. 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S. “Thermofluiddynamic study of a flat plate closed loop pulsating heat pipes”. Applied Thermal Engineering 25 (2005) 21. J. Lin. Ponnappan. Sartre and M. Launay et al.