The thermal conductivity and thermal contact conductance of several thermoplastic and thermosetting polymers have been studied over a range of interface pressures and temperatures. The temperature range for the thermal conductivity study varied from 10 to 100°C (50 to 212°F). The study showed that ultra high molecular weight (UHMW) polyethylene had the highest thermal conductivity through the range of temperatures and also had the highest thermal conductance values at an interface temperature of 20°C (68°F). The thermal contact conductance study was conducted over a pressure range of 510-2760 kPa (75-400 psi) and a temperature range of 20-40°C (68-104°F). The conductance values for UHMW polyethylene ranged from 1095.3 to 1659.4 W/m 2 K (192.9 to 292.2 Btu/h ft 2 °F), whereas the thermal conductivity remained constant at 0.45 W/m K (0.26 Btu/h ft °F) throughout the range of temperatures. Polycarbonate and Teflon® had the next highest thermal conductance values at the same interface temperature. The thermal contact conductance values for polyethylene, Teflon, and phenolic polymers were measured at an elevated temperature of 40°C (104°F). The thermal contact conductance values for both Teflon and phenolic increased with increasing temperature, whereas the values for UHMW polyethylene decreased due to their unique chain structure at the higher temperature. The polymers were chosen because of their widespread engineering interest applications. Nomenclature A a = apparent area A r = real area E = modulus of elasticity E' = (E^/t^O -v\) + E,(\ -F = flatness H c = hardness H e = elastic hardness (E'/\/2) X m h c = thermal contact conductance h e -dimension elastic conductance h p = dimension plastic conductance k = thermal conductivity m = asperity slope P d = profile slope PIH C = dimensionless plastic contact pressure PIH e = dimensionless elastic contact pressure P r = profile roughness R = roughness W = waviness Subscripts a "=• average p = polymer q -root mean square s -roughness subst = substrate Presented as Paper 95-0421 at the AIAA 33rd Aerospace Sciences

The heat ow across a metal/polymer interface is a very important problem in many modern engineering applications. A thermal joint conductance model that employs the surface mechanics of a contact interface in conjunction with an existing elastic thermal contact conductance model was developed. In developing the model, an elastic contact hardness term was derived to predict the actual contact area of a metal/polymer interface under loading. The model predicts a microscopic resistance region where the interface resistance is dominant and a bulk resistance region where the thermal conductivity of the polymer is dominant. An experimental apparatus was fabricated, and a successful experimental program was conducted. New experimental data were gathered on different polymeric specimens over a pressure range of 138-2758 kPa (20-400 psi). The experimental data were compared to the proposed thermal joint conductance model. It was found that the proposed model predicted the data quite well. The data followed the predicted trends for both the microscopic and bulk resistance regions. Nomenclature A a= apparent area of contact, m 2 A r = real area of contact, m 2 a c = contact radius, m a 0 = Hertzian contact radius, m c = half-width of plane contact area, melastic modulus of polymer, Pa E s = elastic modulus of substrate, Pa H c = contact microhardness, Pa H e = Mikic elastic hardness (E 0 = p 2)£ m, Pa H ep = polymer elastic hardness, Pa h bulk = thermal bulk conductance, W/m 2 K h c = thermal contact conductance, W/m 2 K h e = dimensionless elastic conductance h j = thermal joint conductance, W/m 2 K h j exp = experimentally calculated joint conductance, W/m 2 K h micro = thermal microscopic conductance, W/m 2 K J .t / = creep compliance k ux = thermal conductivity of ux meter, W/m K k p = thermal conductivity of polymer, W/m K k s = harmonic mean thermal conductivity, W/m K L = load, N m ab = mean absolute asperity slope, rad n = number of contact spots per unit area of apparent contact, m ¡2 P = apparent pressure, Pa P=H c = dimensionless plastic contact pressure P=H e = dimensionless elastic contact pressure P m = mean pressure at interface, Pa Q = heat rate, W q = heat ux through ux meter, W/m 2 R b = bulk thermal resistance, K/W R g;1 = gap resistance at interface 1, K/W R g;2 = gap resistance at interface 2, K/W R j = joint resistance, K/W R micro = microscopic thermal resistance, K/W R t;c = contact resistance, K/W R 1 = thermal resistance for upper interface, K/W R 2 = thermal resistance for lower interface, K/W T c = temperature of specimen, K T sl = temperature of lower interface of specimen, K T su = temperature of upper interface of specimen, K T 1 = temperature at surface 1, K T 2 = temperature at surface 2, K t = elastic layer (polymer) thickness, m t f = polymer thickness after loading, m t 0 = polymer thickness before loading, m t ¤ = critical polymer thickness, m Y = separation distance between contacting surfaces, m Y .t / = relaxation modulus " p = strain on polymer in the vertical directioņ = dimensionle...

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The use of dimple technology for improvement in friction factors and enhancement of heat transfer has been attracting the attention of many scientists and engineers. Numerical and experimental studies have shown there is a positive improvement (two-fold on average) in Nusselt number when dimpled surfaces are compared to flat plates, and this improvement is achieved with pressure drop penalties that are small when compared to other more intrusive types of turbulence promoters. When arrays of specific dimple geometry are used, pressure drop penalties are roughly equivalent to the heat transfer improvement. This, at least theoretically, will enable the design of smaller heat transfer devices such as heat sinks, which are especially appealing in those applications where size is an important design factor. A literature review of numerical modeling and experiments on flow over dimpled surfaces was performed, and key parameters and flow structure were identified and summarized. With these premises, a numerical model was developed. The model was validated with published experimental data from selected papers and fine tuned for channel flow within the laminar flow regime. Subsequently, the model was employed for a specific application to heat sinks for microelectronic cooling. This paper, then, provides a comparative evaluation of dimple technology for improving heat transfer in microelectronic systems.