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ENHANCED HEAT TRANSFER FOR ELECTRONIC DEVICESOPERATING INSIDE ENCLOSURES

bySorin Witzman, Tristano Nicoletta, Bonnie Mack

BNR Ottawa, Canadatel (613) 763 2688

ABSTRACT

The paper contains an extended experimental study of temperature fields for populated printed circuit cards activated inside open and hermetic enclosures. The experimental study concludes that under all circumstances the temperature rise of devices can be expressed as a summation of two terms: temperature rise inside the thermal boundary of the device (the adiabatic temperature rise) and the temperature rise of the air at the device level. The air temperature rise is a function of the flow inside the enclosure, while the adiabatic temperature rise is practically independent of the cabinet geometry and is a very weak function of the local flow at the device level.

Study of the heat transfer at the device level shows that the transfer phenomena is dominated by the generation of the thermal plume around heated small bodies. The factors which generate and control the formation of the plume are affected very little by flow at the low Reynolds number which exist inside electronic cabinets. It is safe to say that any action which will spread or extract energy from the plume will enhance the heat dissipation from the electronic devices or from any other heated small body.

The paper compares some simple methodologies that can dissipate or extract energy from the thermal plume and investigates their effect in thermal transfer enhancement. Some practical recommendations for potential electronic equipment designers are made.

SYMBOLS

A = external surface area of the enclosure [sq m]

CR = correction factor for cross flow effects

h = convection heat transfer coefficient [W/C*sq m]

k = heat conductivity [w/m*C]

P = power dissipated inside the enclosure [W]

T = temperature [C]

Nu = Nusselt number

4th IEEE Workshop on Accelerated Stress Testing; Pasadena, CA; 9/22-24/1998Witzman 12

Pr = Prandtl number

Ra=Rayleigh number

Re = Reynolds number

Subscripts

Ad = adiabatic

AE = air inside the enclosure

D = device

DE = device inside enclosure

DFA= device in free air

MC = mixed convection

Tot= total (total power inside the enclosure)

INTRODUCTION

The natural convective flow and heat transfer which occur between a body and its enclosure has received in the past a great deal of attention [1,2]. Unfortunately, the knowledge accumulated through this research can seldom be applied in solving problems related to electronic instrumentation packaging. One of the reasons for this is the type of boundary conditions selected for both analytical and experimental research programs: single heated bodies or uniformly heated smooth plates.

Another major obstacle in using existing computation methodologies is related to the process of packaging design itself and the type of inputs the designer has at the time decisions are necessary. Most of the time, the device selection takes place far before the geometry of the enclosure is considered. For high power devices, most of the thermal resistance is inside the device thermal boundary layer. If design problems are not detected at this stage because predicting the flow inside enclosure is not yet possible predict, costly correction measures will be necessary at later stages of the product development.


Figure 1. Temperature field for an electronics carrying board.

a) Open Spaceb) Inside Enclosure

c) Forced Convection (Vertical Flow)

The situation with which the thermal packaging engineer is faced is not that of single heated bodies or uniformly heated smooth plates, but that of point heat sources connected to relatively non-conductive plates inside arbitrarily shaped enclosures. The literature is sorely lacking in analytical approaches to this problem and the numerical solutions are presently too time consuming to set-up and run. The present research program was therefore directed to provide simple correlations which would allow the designer to evaluate the feasibility and the associated cost of a proposed solution at any stage of the design iteration. By studying the phenomena related to the heat transfer between the device and its environment, a better understanding of the impact of different factors in the device cooling process was achieved. As a byproduct of this, ways to enhance the process were proposed and investigated.

EXPERIMENTAL PROGRAM FINDINGS

To help avoid confusion, the nomenclature used in the following sections will be defined here. The "device adiabatic temperature" is defined as the temperature that an element achieves when the convective heat transfer from the element to the fluid goes to zero. In the absence of conduction and radiation from other elements, the device adiabatic temperature is the temperature achieved when no power is applied to the element and the rest of the system is activated [3]. This is not to be confused with the "device adiabatic temperature rise", which is the temperature of the powered device above the device adiabatic temperature.

TD = TAd + TAd ( 1 )

Visualization and measurements of the thermal field of populated printed circuit boards were performed for boards cooled by forced convection and free convection inside fully open as well as sealed enclosures.4th IEEE Workshop on Accelerated Stress Testing; Pasadena, CA; 9/22-24/1998Witzman 14

The results (see Figure 1) show that the temperature field is similar for all these conditions, the temperature difference is identical at any considered point and it is a function of the flow passing through the enclosure or the size of the external surface area for sealed cabinets.

Figure 2. Temperature field on a board inside a baffled enclosure (300 W/board, uniform distributed, 188 mm inlet opening, 23.6 deg angle baffle).

AIR COOLED OPEN CABINETS

Forced and natural convection studies of flow inside cabinets [4,5] performed at BNR showed that the dominant factor determining the flow characteristics is the channel inlet geometry and the inlet flow conditions. However, measurements of the temperature field on the boards showed little correlation with the channel flow pattern variation. In other words, the device adiabatic temperature rise could not be correlated to the flow. For the same substrate, the results also indicated that the device adiabatic temperature rise was almost independent of the device position in the array, even though a strong interdependency between device temperature and the substrate thermal conductivity was found. Additionally, the heat transfer from the device surface to the air surrounding that device (adiabatic heat transfer coefficient as defined by [3]), was found to be almost independent of the position of the device in the array as long as the orientation of the device was kept constant (i.e., it was not rotated about its z-axis). The exception to this was for the situation in which the device was mounted near the air inlet.

These results are surprising and remarkable when one considers the large number of variables involved: board lay-out, device shape and material conductivity as well as large variations in the flow patterns. They indicate that the heat transfer from a device to the adiabatic air temperature is a localised phenomenon, for the most part independent of the upstream or downstream conditions.


Based on the experimental findings for open enclosures, the device adiabatic temperature rise (the temperature rise inside the device thermal boundary layer) can be computed by considering the devices as isothermal bodies mounted on infinite fins (assuming the thermal conductivity of the printed circuit board material is of the order of 0.7 to 2.0 W/m*C) [5,10]. An equation that can be used for determining the local convection heat transfer coefficient was developed and it is presented in this paper (Equation 5). The terminology "local" is used in order to specify that the heat transfer coefficient (integrated over the surface area of the device) determined by this equation is valid for a specific device and its immediate vicinity. For high conductivity boards (k>2 W/m*C) this assumption is not always valid and experimental correction factors should be considered.

In some cases the impact of radiation is significant and a correction for the heat losses by radiation should be considered. An example of computation for the adiabatic device temperature rise is shown in [5]. The prediction of the device adiabatic temperature is more difficult and is still an open subject. Fortunately, for most practical applications (forced convection and unobstructed flow natural buoyancy enclosures) the adiabatic device temperature can be approximated as the summation of the bulk air temperature rise at the device level and the upstream device thermal wake [3,5]. The errors associated with this approximation are usually small (under 5 deg C).

In some cases, for example partially open free convection enclosures in which baffles are used at inlet or outlet in order to divert the flow, the prediction of the adiabatic device temperature might become a difficult task. Figure 2 shows the temperature profile of a board carrying a uniform heated device array inside a partially open (baffled) enclosure. The methodology in which this picture was produced is described in [6]. From this picture it appears that the prediction of the device adiabatic temperature under such conditions is not a simple task.

DEVICE TEMPERATURE RISE INSIDE SEALED ENCLOSURES

Predictions of temperature rise for devices inside sealed enclosures (air recirculation) is one of the most challenging tasks confronting the thermal management engineers in the electronic industry. Complex three dimensional numerical models are able to predict the flow patterns developed inside enclosures, but they generally fail to predict the device temperature within an acceptable degree of accuracy. Other mathematical approaches (for example flow inside porous media), have a high degree of complexity and relative difficulty, but can produce reliable data. However, they must be supported by a large library of experimental values and physical constants which usually are not readily available to the designer at the time these predictions are required.

Experimental measurements of the temperature rise on single, populated boards revealed that the thermal fields were similar for boards powered inside enclosures and in free space (see Figure 1). The temperature at any considered point on the board differs from the ambient temperature (just outside the boundary layer) at the same point in free space by the same value. This value was


found to be a function of a single variable - the external area of the enclosure - for all the cases were the devices were at least 15 mm from the enclosure wall.

Under the assumption that the adiabatic temperature rise for a device mounted inside a sealed enclosure is identical to the adiabatic temperature rise for a device mounted inside a ventilated enclosure, it follows that the device adiabatic temperature will be related to the temperature rise of the air inside the enclosure. Considering the average enclosure temperature as equivalent to the bulk air temperature for a device mounted inside a ventilated enclosure, the device temperature rise can be computed from the equation:

TDE= TDFA+ TAE (2)

The average temperature rise of the enclosure can be computed from a relation as:

TAE=PTot/h*A (3)

The experimental study produced sufficient data to create an empirical correlation for the value of the heat convection coefficient within small sealed enclosures:

h=12.696-16.774*A (4)

This correlation is valid for enclosures with a surface area between 0.05 to 0.250 square meters and it cannot be extrapolated safely outside this interval. A comparison between the correlation and the experimental results is shown in Figure 3. Ongoing research intended to produce more general equations is taking place in our laboratory.

Figure 3. Variation of the convection heat transfer coefficient for enclosures as a function of the enclosure external area.


Figure 4. Heated plate to ambient thermal resistance, as a function of the distance between the plate an isotheral surface.

As shown in Figure 4, the thermal resistance between two isothermal plates immersed in a fluid is a function of the distance between them, if the distance between the plates is under a certain critical value. The overall result is a combination between the conduction and the convection that takes place inside the fluid gap. It is expected that electronic devices mounted on the board and brought close to the enclosure wall will show a similar behaviour, from the heat transfer through the gap point of view. When the distance between the device and the wall is decreased below the 15 mm limit, the device temperature increases slightly initially, before decreasing sharply when the separation drops below 5 mm. This phenomenon can be associated to a sharp drop in the adiabatic temperature rise when conductive bodies are inserted inside the device thermal plume. Alternatively, the device temperature rise when the distance between the heated wall and the isothermal surface is under 15 mm, can be explained by the increase of the internal air temperature on the internal side of the inclosure wall (reduced convection).

The experimental program shows that under any studied situation, the influence of the upstream (or downstream) channel conditions on the device temperature can be superimposed on the temperature rise in the boundary layer of the device via the device adiabatic temperature [3,5,6]. This temperature represents in fact the temperature of the environment in which the device is located. Due to the relatively large distances between the heat sources - the silicon chips - very little interaction takes place between the thermal boundary layers of different devices mounted on the same board. This interaction is mainly in the form of a thermal wake produced by the high power devices mounted upstream from the considered device. The results are similar for forced and free convection cooled electronic cabinets, and, to a certain extent (distance between device to the outside enclosure larger than 15 mm) they are valid for hermetic (re-circulation) enclosures.

These conclusions justify studying the thermal Phenomena in the close proximity of the device, rather than considering the heat transfer and fluid flow inside the entire channel.


VISUALISATION OF THE THERMAL FIELD GENERATED AROUND A SMALL HEATED BODY

Although a large number of studies have been done to provide information about the fluid flow and velocity field inside the enclosures or around individual heated bodies, very little research related to monitoring the temperature field inside the cooling fluid itself has been done. Some tentative work of this kind implies the use of interferometry techniques used to visualize the thermal boundary layer around simple geometries [8] or use of the thermal print of heated fluids on solids covered with liquid crystals [9].

It was decided that the use of interferometric techniques was too expensive and not suitable for our purposes. The use of liquid crystal techniques is laborious, very slow and lacks the accuracy necessary for producing useful empirical equations. It was decided to use infrared measurement techniques which can produce fast and accurate thermal field measurements. The results can be stored and analyzed numerically. Obviously, the infrared thermometry cannot be used to directly monitor the temperature field inside the cooling fluid, especially for the case of air cooled objects. What can be monitored is the thermal print that hot gases leave on thin, nonconductive membranes immersed inside the hot fluid.

In order to study the thermal field created in the proximity of a small heated three dimensional body, a set up similar to the one shown in Figure 5 was created. The experimental apparatus consisted of a very thin foil of thermally non conductive material (polyethylene) with a window in which a cubical body of copper is suspended. The foil did not touch the body which could be heated by a resistor mounted inside. Due to the high thermal conductivity of the material (copper), the surface of the body was considered isothermal. No direct conduction was considered between the body and the foil. Due to the thickness and low thermal conductivity of the foil, the heat spreading inside the foil was minimal. The body and the foil were painted black to ensure a thermal emissivity of 0.92. Under such conditions it was considered that the thermal print on the foil, monitored by an infrared camera, represented the cross-section of the thermal boundary painted black to ensure a thermal emissivity of 0.92. Under such conditions it was considered that the thermal print on the foil, monitored by an infrared camera, represented the cross-section of the thermal boundary layer created around the heated body. Samples of the results are shown in Figure 6.

Figure 5. Experimental set-up for determination of the Nu-Re correlations.


Figure 6a

Figure 6b

Figure 6c

Figure 6d


Figure 6. Visualisation of the thermal boundary layer, for a 20 mm size cube, 2 W power dissipation.

The results of the visualisation of the thermal boundary layer of a single heated body in a large enclosure (the room itself) were similar to the results achieved in the past by using other visualisation techniques [9] for the plume produced by single small bodies in free convection. More impressive was the visualisation of the plume when cross-flow of increasing speed was passed over the heated body, parallel to the visualization membrane direction (see Figure 6). For very low speed (low Re number flow), the plume rotated according to flow direction but decreased very little in size and temperature. Once the flow velocity was high enough that the plume was in the same direction as the flow, the plume decreases slowly in size, until the flow reaches very high local velocities of over 3 m/sec.

The temperature and velocity data for the above experiment have been non-dimensioned and are presented in Figure 7. In this diagram the characteristic dimension is the square root of the entire area of the heated body. The results were corrected to account for the radiation between the body and the environment. The local convection heat coefficient can be computed from Equation 5 [7].

NuMC=NuD+NuFC+NuNC+CR (5)

were :

NuD = diffusive limit

NuD=2√ (6)

Figure 7. Variation of the convection heat transfer coefficient for a suspended small heated body, as a function of the approaching air velocity.

NuFC = pure forced convection limit

NuFC=0.825*Re0.5*Pr0.33 (7)


NuNC= pure natural convection limit

NuNC=0.694*Ra0.25 (8)

CR = mixed convection, cross flow correction

CR=0.86 -3.54 * (Ra/Re2) 0 25 (9)

Considering the size and the temperature interval in which the electronic devices are operated, the value of this correction parameter is very small, approaching 0 when the air velocity exceeds l m/sec.

This equation covers the entire interval between the completely static air to the beginning of the turbulent flow regime. The characteristic length is the square root of the external surface area of the device.

The results obtained coincide with the results found at the board level: electronic devices are very high power small heat sources which create independent local heat plumes. Even though the board intersecting the plume spreads it to a certain degree, due to the relative low conductivity of the board and the very high local concentration of heat, we can still consider that very little interference occurs between the thermal boundary layer of the independent devices. The relatively low Re Number flow (under 2000 when the considered characteristic dimension is the square root of the external area of the heated body) has very little impact on the device thermal boundary layer. The plume or wake from the upstream devices will affect the adiabatic temperature of the device, without modifying the adiabatic heat transfer coefficient.

METHODOLOGIES OF ENHANCING THE HEAT TRANSFER FROM ELECTRONIC DEVICES

Upon accepting the idea that the device heat dissipation is controlled by the formation of the thermal plume around the device, it is reasonable to consider that any solution that will produce the dispersion of this plume or the spreading of it to a solid, will reduce the device thermal resistance. Simple solutions for achieving this result are the use of air movement (forced convection and air impingement jets), the use of conductive substrates and extraction of heat from the plume (non contact heat sink). Even though these solutions do not produce impressive results when compared with immersion or liquid cooling (in most cases they are less efficient than conventional heat spreaders), they are easy and inexpensive to implement. In most cases they can produce a reduction of 5 to 10 C in junction temperature, which might be the difference between success or failure of a specific project.

Increasing Local Air Velocity

It has been shown that the heat transfer from the device is a local phenomenon, practically independent of the channel flow conditions. The characteristic


dimension which best illustrates the local heat transfer conditions is the square root of the external area of the device. For suspended heated bodies, the sensitivity of Nu number with the increasing air velocity is quite high: it is proportional with the Re number to the power 0.5, when the local Re number varies between 0 to 3000. Unfortunately, for devices mounted on a substrate with thermal conductivity between 0.7 to 2 W/m*C as in the case for printed circuit boards, the efficiency of the board which acts as an infinite cooling fin, decreases rapidly with the increase of the local heat transfer coefficient [5,10]. As an overall result, the sensitivity of the local Nu number with the increase in air velocity is reduced to the power 0.3 (see Figure 8). Considering the silicon temperature, devices showing large internal thermal resistances will show an even lower sensitivity to the increase in local air velocity. An example of variation of silicon temperature with the increase of approaching air velocity is shown in Figure 9.

Even though increasing the air velocity is not an efficient way of improving the local heat transfer, in some cases (for example when the power density per board exceeds 90 W/sq m) large quantities of air must be circulated through the system in order to reduce the cabinet air temperature and the adiabatic temperature of the device. Impingement cooling, a method of blowing high air velocities over a device, has the advantage of using cool air from outside the cabinet instead of upstream heated air. The preliminary results with air jets directed at the top of the devices show an improvement of at least 15% in the local heat transfer coefficient when compared with channel flow of the same velocity. More research is necessary in order to confirm these very preliminary results. Unfortunately, the mechanical complexity of jet impingement cooling limits the use of this option to very limited applications.

Figure 8. Actual (effective) heat transfer coefficient, including conjugate phenomena, as a function of air flow


Figure 9. Variation of the silicon temperature with the approaching air velocity

Conductive Substrate

In very limited cases the conductivity of the substrate can be high enough to remove the heat from the device to the ultimate heat sink solely by conduction. For sparse arrays, the most common benefit of the conductive substrate is to spread the heat plume created by the device to a larger area from which it can be more readily transferred to the passing air. In this way, the substrate can be considered as a cooling fin, to which the rules for conjugate convection-conduction can be applied [5,10]. Experimental as well as analytical research indicate that for sparse arrays, the device external thermal resistance decreases with the square root of the increase in thermal resistance [10]. This rule of thumb is applicable for arrays in which the distance between devices dissipating heat is larger than the size of the device and the board conductivity varies between 0.7 to 2 W/m*C. For densely populated boards and highly conductive substrates, the impact of increasing board conductivity is less obvious and sometimes controversial.

Extraction of the Heat from the Thermal Plume (Non Contact Heat Spreader)

Most of the case to ambient thermal resistance is concentrated inside the thermal plume surrounding the device. It is logical to assume that introducing a conducting body inside the plume will reduce the overall resistance by conducting the heat away from the device. Experimental results have already indicated that this plume is not really affected by the low reflows existing inside the electronic enclosures. It is expected that conducting the heat from the boundary layer with a non-contact thermal spreader will be more efficient than spreading the plume inside the incoming or enclosure air flow.

Based on temperature distribution within the boundary layer and the law of heat conduction, one might think that two factors of equal importance should be considered in the design of a non-contact heat sink: the distance from the heated body and the surface of the heat sink. If an extended surface is used so that fins are within close proximity of the device, some points of this surface will be further from the heated surface than others. The efficiency of the more distant


points in extracting heat from the plume is expected to be lower than the efficiency of the points in close proximity to the heated body. It was expected that an optimum geometry or configuration might be achieved, in terms of the length to width ratio of the fins for a given thermal conductivity.

Figure 10. Non contact heat spreader testing apparatus

An experimental program in which different isothermal expanded surfaces were introduced inside the thermal plume was designed an carried out in order to find this optimum configuration. The non-contact heat sinks were maintained at a constant temperature by circulating water inside channels made into their bodies. The materials used for the heat sinks were aluminum and copper. A sketch of the experimental set-up is shown in Figure 10. Surprisingly, the flat plate having the same size as the device top surface outperformed all other geometries (for the condition of identical minimum distance between the device and the heat sink). This result indicates that diffusion is the dominant phenomena controlling the device thermal resistance. Some fluid circulation was identified inside the external layers of the plume. But, as a rule, most of the thermal resistance is concentrated in the first millimetre of the device thermal boundary layer.

CONCLUSIONS

The experimental results collected over the last five years converge towards the idea that the heat transfer from a device to its environment is a local phenomenon controlled by the thermal plume surrounding the device. It is practically independent of the upstream/downstream conditions and it is not a function of the flow pattern developed inside the channel. Under all the cases under study (free or forced convection inside open or sealed enclosures) the final temperature can be expressed as a summation of the device adiabatic temperature and the adiabatic temperature rise. The adiabatic temperature rise is independent of the device position inside the array.

The characteristic dimension which best describes the thermal phenomena at the device level is the square root of the device external surface area.

Thermally conductive substrates help spread the heat concentrated inside the thermal plume by acting as cooling fins for devices. The film thermal resistance


(the adiabatic thermal resistance),for most practical applications, is reduced proportionally with the square root of the increase of the substrate thermal conductivity.

Spreading the thermal plume by low Re number flow, similar to the flows existing inside electronic enclosures, is inefficient and costly to achieve. The sensitivity of the device adiabatic thermal resistance with the local airvelocity is proportional to velocity raised to power 0.3. This makes forced convection cooling a very inefficient solution if air velocities higher than 2 m/sec are required.

Introduction of conductive bodies (non-contact heat sinks) inside the device thermal plume is a viable cost effective solution which can reduce the device adiabatic free convection thermal resistance to about a half. This solution is particulary advantageous when the product specification requires sealed enclosures or unpackaged chips on multichip modules. Although this solution is less effective than the conventional heat spreaders mounted on the device surface which can reduce the adiabatic thermal resistance by a factor higher than five, non contact heat spreaders can be used for cooling large surface mounted devices due to the fact that they do not produce any mechanical stress on the device leads.

ACKNOWLEDGEMENTS

The authors will like to acknowledge the work of the following students that help performing the experimental work included in this paper: N. Sharp, B. Van Stralen from the University of Waterloo, S. Delroy from the Queen's University, Y. Maciel from the McGill University and M. Cooper from the Western University.

REFERENCES

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[2] W. Aung, Cooling Technology for Electronic Equipment, New York, Hemisphere Publishing Corporation, 1988.

[3] A. Ortega, R. Moffat, "Heat Transfer from an Array of Simulated Electronic Components: Experimental Results for Free Convection With and Without a Shrouding Wall", Proc. 23rd AICHE/ASME National Heat Transfer Conference, Denver, Colorado, pp 5-15, 1985.

[4] K. Graham, D. Newport, T. Nlcoletta, S. Witzman, Pressure DroD Across Racked PCB's Under Forced Convection Coolina, BNR IR 386, 1986 (internal report).

[5] S. Witzman, D. Newport, T. Nicoletta, "An Alternative Methodoloay for Computing the Temperature Rise Or an Electronic Device Inside a Direct Air-


Cooled Electronic Cabinet", Proc.The Ninth International Heat Transfer Conference, Jerusalem, Israel, Volume 2, pp 313-319, 1990.

[6] S. Witzman, D. Newport, T. Nicoletta, "Free Convection Alr Cooling of the Electronic Eauipment Still a Miracle Waiting to be Explored and Explained", Proc. 1989 SEMITHERM, San Diego, California, pp 98-103, 1989.

[7] M. Yovanovich, C. Vanoverbeke, "Combined Natural and Forced Convection from Isothermal Spheresn, Proc. 1988 AIAA Thermophysics Plasma Dynamics and Lasers Conference, AIAA-88-2618, San Antonio, Texas, 1988.

[8] K. Kihm, D. Kastell, L. Fletcher, "Study of Thermal Boundary Layers Occurring Around the Leading Edge of a Vertical Isothermal Wall Vsing Speckle Photography", proc. 1991 ASME JSME Thermal Engineering Joint Conference, Reno, Nevada, pp 25-34, 1991.

[9] R. Moffat, "Experimental Heat Transfer" Proc. The Ninth International Heat Transfer Conference, Jerusalem, Israel, pp 187-206, 1990.

[10] K. Graham, S. Witzman, ~Analytical Correlation of Thermal Design of Electronic Packages", Cooling Technoloav for Electronic Equipment, pp 249-264, Hemisphere Publishing Corporation, New York, 1988.

[11] M. Yovanovich, "General Expression for Forced Convection Heat and Mass Transfer From Isopotential Spheroids", AIAA 26th AerosDace Science Meetina, AIAA 88-0743, Reno, Nevada, 1988.


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