Enhancement of Electronic Cooling

ME412Team 6 – The Heat Removers

Sapan PatelDominic Waldorf

Robert Zuerlein

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Abstract The increasing demand of computing power leads to a higher demand of electric power running through microprocessors. To combat this, the project team was tasked with the design, analysis, building and testing of a device to minimize the operating temperature of an electrically heated plate. The temperature at the top center of the plate was to be tested after 6 minutes, while a box fan is turned on from a distance of 12 inches after the first three minutes of heating. The group tested various designs by adjusting fin arrays and plate thicknesses as well as by adding small computer fans in various positions. The advantages, disadvantages, and test results of each option were then compared to create the final product used during Design Day. The final product showcased a thicker plate with a pin-fin array and an operating temperature of 31.2°C, a mass of 169.2 grams, and a cost of $2.60 to create.

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Introduction In today’s world, the increasing integration of computers into our lives has led the demand of microprocessor computing speed to approximately double every two years. Microprocessors are in technology all around us, from the alarm clock that starts our day off to the cars which get us to work. This increased computing power leads to an increased electric power demand, and therefore high amounts of heat created through the microprocessors. The challenge at hand is to keep the temperature of these microprocessors at a reasonable operating temperature. The design of many microprocessors include a heat dissipation strategy, such as a heat sink, a continuously flowing fluid, or a powered fan. The ability to create effective heat dissipation devices at a low cost has a critical impact on the success of companies when manufacturing and selling at an industrial scale.The project team is given an electric 20 Watt Kapton Flexible Heater to simulate the challenge. The flexible heater is pressed between two thin aluminum plates which are each 2” by 3” and will be heated for 6 minutes in a horizontal orientation. To expedite the heat transfer, a box fan set a distance of 12” is turned on after the first 3 minutes. At 6 minutes, a thermocouple is placed on the top center of the plate to take a temperature reading.

Figure 1. Kapton Flexible Heater

The heat transfer device created by the group is required to fit within a box of dimensions 14” x 14” x 12”, and must be self-sustaining after the start of the competition timer. The competition takes into account most parameters which are critical to companies who have to solve a similar. The group is evaluated on the device weight, device cost, and most importantly the operating temperature of the device. The points regarding the operating temperature of the plate are given 2.5x the value of the points regarding device weight and cost, to emphasize the ultimate goal at hand.

The methods used to design, analyze, build, and test the device are outlined in the following sections of the report.

Design Analysis During the process of concept generation, each team member presented initial drawings which utilized basic methods of heat dissipation. These initial drawings presented an overall idea, without much attention given to specific details. Each of these basic concepts of design were then to be analyzed for advantages and disadvantages using a decision matrix.

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Figure 2. Design ideas, A (Top left), B (Top Right), C (Bottom Left), D (Bottom Right)

Design A was the most complex of the concepts imagined by the team. The design was based around the idea of a fluid running through a casing which surrounds the heated plate, throughout the 6 minutes of heating. With this design, the team could either incorporate a pump system which would bring the fluid back through, or have a reservoir filled at the beginning. With the reservoir approach, the flow rate over time must be calculated and there must be a place for the fluid to exit into. This would lead to a very slow flow rate to keep the fluid moving for 6 minutes. In addition, the manufacturing processes for this process would become very complicated.

Design B combined the idea of a heat sink with a fluid. In this design, the heat sink is suspending the plate from the reservoir. The fluid does not need to be circulated, which eliminates the need for increased volume or a pump. In the sketch, cones are used to minimize the contact with the base of the reservoir. Other shapes were also considered for the geometry of fins. Various fluids were also considered in place of water. The focus of this design, however, would be removing the heat from the bottom plate more than the top. There was discussion of adding a heat sink on the top in combination with this design, but the weight would significantly increase. Utilizing the air forced convection would be more advantageous than a resting fluid.

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Design C was based around the idea of drawing the heat away from the plate to an outside copper casing. Copper’s low specific heat capacity and high thermal conductivity relative to cost would make it an efficient and lightweight material choice to draw the heat to. This would also allow the group to increase the volume of material around the outside of the plate to hold more heat. Finding raw copper materials for machining was difficult for the team, where aluminum was in abundance. In addition, the design was drawing heat away from the outsides of the plate, allowing the heat to still concentrate at the center of the heated plate.

Design D was to test various designs of heat sinks to the advantage of the air flow introduced at 3 minutes. With the horizontal flow, a heat fin array as shown in the figure above would help push the heat out if oriented parallel with the air flow. In addition, the textbook was referenced to understand the advantages of various fin array designs. If an axial flow is introduced to the system, then a pin-fin design may be advantageous. A heat sink could be applied to the bottom as well when the plate is elevated.

Table 1. Decision Matrix

Emphasis Design ScoringParameters A B C D A B C DFunction/Performance 5 4 3 4 5 20 15 20 25Product Cost 4 1 3 4 4 4 12 16 16Size 3 2 4 2 5 6 12 6 15Weight 4 1 2 2 5 4 8 8 20Manufacturability 3 1 3 3 4 3 9 9 12Score: 37 56 59 88

The decision matrix took into account the following parameters:

1. Function/Performance – How well will the device dissipate the heat from the top center? Will it function throughout the 6 minutes of testing reliably?

2. Product Cost – How much will each individual part cost? Parts such as batteries, pumps, and pre-manufactured nuts and bolts can add up and hurt the final score during the competition.

3. Size – Will the product fit within the 14” x 14” x 12” requirement? Do any of the designs require something large such as a reservoir, pump, or fan?

4. Weight – How heavy will everything add up to be? The addition of fluids or thick plates add to the weight significantly, hurting the final score.

5. Manufacturability – Is the design simple enough to allow the team to manufacture it with the tools available at machine shop?

The importance of each of these parameters were also added to place more emphasis on the parameters which will affect the final score the most. The operating temperature score was the most important, followed by the weight and cost. Each of these three scores will be compared for a final placing.

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The team gathered to compare the various design options based on the decision, keeping in mind many further options and alterations could be made to each basic design. Doing so allows the team to put a value to the quality of each approach. The scores showed that Design D, involving some type of heat fins with the possibility of adding more components, seemed to be the most beneficial by a substantial value.

There are many variations of fin arrays which the team could use over the small heated plate. The group began by looking into various heat sinks and applications. On a computer repair company’s website, the idea behind heat sinks is explained in an article called Heatsink 101. While many basic heat sinks for computers are simply fin arrays, many become more complicated. The design of these arrays plays to the advantage of the air flow which can be created. An increased air flow through the fins also helps the heat dissipate more efficiently. At this point, the team looked into increasing the air flow through the device. Adding an increased airflow through small fans could give the group a significant advantage over groups simply relying on the air flow from the box fan. In addition, there would not be a limitation on how close these fans are placed.

It was determined that a fan approximately the size of the heated plate is the desired dimensions, to strike a fair balance between weight and surface area which the air flow would cover. After searching through computer stores and online catalogs, the team decided on two 70 mm ball bearing case fans, which created a 21.33 CFM flow rate and 3500 RPM at a weight of only 35 grams each. These fans cost $10.96 each after tax.

Fan Out

Fan in

Plate Box Fan Box Fan PlateFan In

Fan Out

Figure 3. Device Placement

The group then had to decide on whether a fin array or a pin-fin array will be utilized with the fans. If only the box fan were being used, the rectangular fin array as shown in Design D would be the best option. From a top-down view, adding fans to either side of the plate with perpendicular air flow to the box fan’s would be ineffective as the air flow from the small fans would be blocked by the first fin plate for the fan in. Attempts to increase the air flow by placing the fans before and after the entrance to the flow would also be hindered as the box fan seemed to only slow down the velocities of the mini fans when placed behind them. Orientations in which the computer fans are placed on the top and bottom of the fin design were determined to have the biggest gain from the increased air flow the fans offer.

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Figure 4. Computer Heatsink

Adding fans to the top or bottom of the plate with a heat sink now changes the horizontal flow to an axial flow. With axial flow, pin-fins are far more advantageous. Many computer fans add an axial fan facing upwards to draw the heat away from the plate. For manufacturability purposes and access to materials, cylindrical pin fins were selected. Attaching the pin fins to the plate mechanically would lead to a much higher product cost and weight, considering the parts would be pre-manufactured metals. Therefore, high heat adhesives were considered. Through researching various glues used to build computers, the team came across a low-cost thermal adhesive which can operate at temperatures upward of 150°C. This Arctic Silver adhesive is made with micronized metals, which would help fill small gaps between the pins and the plate. The direct contact between the fin and plate with the glue as a medium would serve as a better heat transfer medium than a fluid.

The pins were made from a long aluminum rod with a diameter of 0.25 inches. The rod was cut into larger pieces based on iterations of testing using a bandsaw. Once the finalized dimensions were selected, the pins were cut down using a clamp and hand-saw to achieve a better finish and accuracy.

A thicker plate was found using scrap materials at the machine shop. The top plate from the flexible heater was replaced with a plate approximately 3 mm thick, weighing in at approximately 39 grams. This increased mass was outweighed by its significant impact on the final temperature when running tests. The pins were attached and cured using the Arctic Silver adhesive and a room temperature cure. In order to create a hole on the top center of the plate after the pins were attached, a combination of a drill bit and flat head drill were used on a mill. This hole was created to approximately 2.5 mm diameter with a depth leaving 1 mm to allow a thermocouple to reach the location which would have been tested on the original plate.

The position of the plate relative to the box fan also needed to be considered. The original design of the plate elevated the design to approximately 4 inches off the ground. This elevation allowed significant airflow under the device to provide forced convection to both top and bottom plates. However, minor alterations to the horizontal and vertical positions showed different results during testing. Therefore, tests were further conducted to learn what the ideal horizontal and vertical position of the plate relative to the box fan is. Using a velocity sensor, the team found the design for the plate needed to be elevated such that the heated plate is 7 inches off of the table. The addition of elevation support would mean an increased weight

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value, and a lightweight material needed to be used. Balsa wood provides a very good combination of low weight and low cost.

The process behind selecting number of pins along with geometric properties are further discussed as an iterative process in the model building and testing sections of this report. Discussion about how it was also decided that the weight of the fans would not decrease the operating temperature by a significant enough value to add in the costs and weights is also included.

The final design showcases 38 offset pins at a diameter of 6.35 mm on the top of the thick plate inches. Each of these pins is 30 mm tall, provide a total of 0.023945419 m2 of increased surface area to the air passing through.

The final design used during the competition showed very low total cost, at approximately $2.60 per unit. These expenses included the thermal paste, aluminum stock, 4 screws, and balsa wood stock.

Table 2. Device Cost

Material Mass (in gm) Cost ($)copper   thermal paste 0.60211 $0.759brass   aluminum 135.37 $0.254

screws4 @

$0.069/screw $0.276

balsa wood28 inch @ $1.69/yd $1.314

Total   $2.602

Figure 5. Final Device

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Model Building

There are 3 main processes of heat transfer at work in this setup. Each of these must be evaluated in order to further understand the theoretical heat transfer occurring, and to help design an optimal heat transfer device.

The first is natural convection, in which flow is induced by buoyancy forces due to density differences caused by temperature variations in the fluid. In this situation, the slowest method of cooling the plate would be only by natural convection. Calculating the natural convection temperature gives the group a starting point in order to understand the effects of forced convection and transient heat conduction. Using energy balance, equation 1 compares the power applied to the plate to its heat transfer coefficient (h), temperatures (T), and area (A). Simple algebra further simplifies the Eqaution 2, which isolates the desired variable Ts (surface area).

(1)

(2)

Two separate heat transfer coefficients must be found, since the value varies from top plate to bottom plate. The heat transfer coefficient (Equation 3) is found by relating the Nusselt number (NuL). The Nusselt number must be calculated using varying equations depending on the plate and the range of the Rayleigh number (RaL) and characteristic length value, which are found by using Equations 4 & 5.

(3)

(4)

(5)

The Nusselt number can then be determined using the equations provided in Equations 6-8. This value is plugged into the respective heat transfer coefficient in Equation 2.

(6-8)

Using the set-up provided, surface temperature (Ts), is actually found using an iterative loop. The iterative loop requires a guessed surface temperature to start, which can then be used to

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evaluate properties, Rayleigh number, Nusselt numbers, heat transfer coefficients, and finally the surface temperature again. When the guess converges with the guess, the surface temperature through natural convection is found.

Figure 6. Iteration Process for Natural Convection

Starting the guess for surface temperature at approximately 490K, the values were each calculated using an iterative table in Excel. After increasing the K value by 0.5 each iteration, the temperature guess and calculated values converged at after 26 iterations, at approximately 502.5K, or 229.35°C.

Table 3. Iteration Results for Natural Convection

Iteration TS Ra NuTop NuBottom hTop hBottom TS Convergence26 502.5 93926.01 9.4534 4.72 14.476 7.238 502.281 -0.219

The next heat transfer process at work during this setup is forced convection. This process is analyzed for after the 3-minute mark, when natural convection transitions to forced transition due to the fan being turned on. Forced convection is a type of transport in which fluid motion is generated by an external source such as a pump, fan, or suction device. In forced convection, the convective heat transfer coefficient is considered, which is the same for both top plate and bottom plate. Again, the energy balance shown in Equation 1 allows us to compare the power to the surface temperature (Ts), convective coefficient (h), areas of the plate (ATop and ABottom). The surface temperature is found by rearranging the energy balance to Equation 2.

(9)

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(10)

Next, a Reynolds (ReL) and Prandtl (Pr) number are needed (Equations 11 & 12) to calculate the Nusselt number (NuL) as shown in Equations 13 & 14. The fluid velocity, u∞, is measured using a velocity sensor.

(11)

(12)

(13-14)

Using an average of several measured velocities, the characteristic length, and properties of air defined from natural convection, Reynolds and Nusselt numbers are calculated.

Table 4. Velocity of Box Fan, Plate Area, and Characteristic Length

u (Avg) 3.1 m/s

Area0.00455327

6 m^2L 0.065 m

The heat transfer coefficient was calculated by using the thermal conductivity, k, the Nusselt number, and the characteristic length, giving a value of 36.76 W/(m2*K). The forced convection leads to a calculated surface value of 359.75K, or 86.6°C.

Table 5. Forced Convection Results

Re Nu h Ts12672.9

690.845

936.75765

1359.7486

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The next part of forced convection analysis is the model building of fins. Fins are one of the most effective methods for dissipating heat, because they provide more material for the heat to transfer to as well as more surface area in contact with the fluid flowing through. This was critical to the team to understand what to manufacture for the pin fins. The initial model must be built before variables can be input to minimize the theoretical temperature.

Even though fins are added on from the beginning, the convective heat transfer will continue from the fins as well as from the part of the plate still exposed. This equation is shown in Equation 15.

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(15)

The area of the exposed plate is represented by Aplate while the surface area exposed by the fins is Afins. The efficiency of fins efficiency is given by Equation 16, where the m value is found through Equation 17.

(16)

(17)

In Equation 17, AS is given by the surface area of a fin, and AC is the cross-sectional area. Using the above equations as well as geometrical equations to calculate areas, a mathematical model was created on Excel. This model allows for the input of various fin geometries which then are used to calculate a fin efficiency and Ts value. In addition, the mathematical model shows the weight of the fins used, which will play into the final cost of the product.

In the table shown, the actual geometry used for the fins shows a theoretical temperature of over 46°C. Various other geometries were entered to the mathematical model (see Appendix). Although many showed a higher fin efficiency at smaller fin lengths, tests showed that the tops of these pins were becoming too hot, showing that more material could be added on. In addition, using thicker pins showed a lower temperature as well as a higher efficiency, but the stock material for such geometries was unavailable to the team, the costs were higher, and manufacturing as many pins on a lathe was not allowed because of the high demand for machines in the shop.

Table 6. Fin Forced Convection Mathematical Model

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Fin effi ciency Fin S A Fin CS Aη fin m h As k Ac L q

0.96743882 10.62771941 36.7576514 0.000598473 3.16692E-05 0.03 20(Taken from forced) 0.000630143

38 Pins

Aluminum Properties Plate Finsρ 2702 kg/m^3 l 0.06476 0.003175 rN 38 w 0.07031 0.03 hV 3.61029E-05 m^3 Aplate 0.004553276cp 1013 J/(kg-K)k 205 W/(m-K)α 2.25E-05 m^2/sCost 1.47 $/kg Mass fins 0.097550057

Mass plate 0.038754295Weight Fins

0.097550057 kg Area Fins 0.0239454190.143398584 $ Cost

K CTs 319.6292913 46.47929131

The third and final mathematical model to be considered assumes the use of fins, which applies to this group’s design. This analysis is conducted to understand the effect of the 3-minute heating period on the device, and whether the predicted steady state temperature from the fin analysis will be reached by the device. Transient analysis is found by assuming a lumped system model, as shown in Equation 18.

(18)

In this situation, the qloss is set to zero due to the small amounts of energy lost by natural convection as the device is heating up. The equivalent heat capacity of the device is given by (mcP)equiv, and is found by Equation 19.

(19)

Equation 18 can be solved for TH. The cooling period after the 3-minute mark changes the equation, and is represented by Equation 20.

(20)

The above differential equation is solved for T, which is the temperature of the device (Equation 21). This equation can then be algebraically manipulated to give tss, which is the amount of

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time the device will take after the 3-minute mark to reach its predicted steady state temperature (Tss) from the fin analysis (Equation 22).

(21)

t ss=−(mC p ) equiv

h (Aplate+η finA fins )ln { 0 .2

TH−T∞−qheater

h (A plate+η finA fins ) } (22)

Using the predicted Tss value from the fin analysis, it was found that the steady state would be reached from the fins after approximately 5 minutes.

Table 7. Steady State Results from Transient Analysis

tss final 305.2655 (s) 5.087759 (min)

Testing and VerificationBefore making any alterations to the plate or doing any designs, it was necessary for the team to run initial tests to understand the range of temperatures currently at work. Initial tests with the plate resting on the table and no alterations done to it show value ranges around approximately 115°C after 3 minutes, and 150°C after 6 minutes when the fan is not run at all. With the fan running after 3 minutes, the values are significantly lower.

Table 8. Initial Plate Tests without Box Fan

Without Fan (°C)Trial 3 Min 6 Min

1 125.9 154.12 130 1743 107 1584 115.5 144.45 113.9 150.4

Table 9. Initial Plate Tests with Box Fan

With FanTrial 3 Min 6 Min

1 114 66.52 111.4 65.7

Initial tests were run using fins of 40 mm length, which is 10 mm longer than the final design used. The group saw that without the fan, the top few millimeters of the fins were at room

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temperature or below. This indicated to the group that the mathematical model must be consulted again to find a shorter length in which as much of the material applied as possible heats up.

With the two computer fans, the group believed theoretically it would work best if a fan was placed on top pointed upwards from the fans. However, the direct air blowing back into the center may have a positive impact on the thermocouple reading. To test these questions as well as many other mini fan orientations, the group kept the positioning of the plate and the fin array as a constant.

Table 10. Prototype Testing with Various Orientations

TemperatureTest Type 3 min 6 min

Plates Only - Test 1 110°C 64.0°CPlates Only - Test 2 114.9°C 64.8°CPlates Only - Test 3 110.5°C 61.7°C

Box Fan Only 49°C 37.4°CMini Fan Facing Up Only 53°C 36.6°C

Mini Fan Facing Down Only - 35.7°CBoth Mini Fans w/ One Facing Down - 36°C

Elevated 6 inches w/ Mini Fan Facing Down 52.7°C 35.2°C

The lowest temperature we were able to achieve was 34.7°C. This was one mini fan pointed downwards with the pin fin array. The next lowest temperature we were able to achieve was with the mini fan facing downwards when the plate was elevated. In many of these situations, it was seen that the batteries were failing on the fans before the end of the test, and results were not reliable. The increased cost added by the fans and supporting batteries would increase out estimated price by about 10x its original value, and add over approximately 44.2 grams per fan used. In addition, more weight would be needed to suspend the fans at a 7-inch height where the box fan showed the highest velocity. Seeing the small difference made by adding in the computer fans, the group decided the disadvantages in weight and cost were not worth the small change in degrees. The device could still perform better by lowering the pin size.

In adjusting the pins, more tests were conducted between various length sizes. Due to a limited amount of material, the group was unable to use smaller increments of testing, and had to rely on the mathematical model to verify lengths of pins. It was also seen using fewer pins than the amount the group was able to fit led to higher temperature readings as well.

Table 11. Testing Results with Various Pin Lengths

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Fin Length/Quantity Temperature @ 6 minutes40 mm, 31 pins 38.9°C40 mm, 38 pins 38.4°C35 mm, 38 pins 37.3°C30 mm, 38 pins 36.9°C25 mm, 38 pins 38.0°C

The next tests performed were to compare the temperature readings done at the regular 4 inches to the readings done at where the velocity was highest for the box fan. All orientations at this given height showed readings at the lowest of position settings, around 32-34°C.

Final testing was done to determine whether it’s better to have the air flow from the box fan flowing through the staggered geometry, or parallel with the straight lines.

Table 12. Testing Results with Different Fin Arrangements

OrientationTemperature @ 6

MinutesParallel, straight flow 32.9°CStaggered, wavy flow 35.1°C

With the competition day approaching, the team was satisfied with an operating temperature only 10°C above room temperature, and extremely low weights and costs.

In comparison to the mathematical models, the team saw fairly large discrepancies. With natural convection, the plate was never allowed to heat for a long enough time period to reach that temperature. When looking at the natural convection, the plate is estimated to reach around 230°C, while at the 6-minute mark it only reaches a maximum of 174°C from tests. Forced convection values for both without fins and with fins show a similar trend of being significantly higher than values in test results. Without the fins, the expected temperature is to be 86.6°C, while the tested values are slightly over 20°C less than that value. With fins, the tested value comes slightly closer, approximately 15°C lower than the predicted value of 47°C from the mathematical model.

With mathematical models, there are many explanations for increased values when compared to testing. Air properties and are taken to be ideal, and a room temperature is estimated using the thermocouple. While the room temperature can vary day by day, the model sticks with one average room temperature. This throws off the fluid property values. The aluminum is also considered to be a pure aluminum. If taken from the machine shop, the group has no way of knowing whether it is a 6061 alloy, or pure aluminum. By assuming the wrong kind, the properties of the plate and fins change immediately. In addition, the aluminum may have impurities in it. The addition of fins is not perfect, where there may be many gaps between the two materials as well as different shapes and sizes of the fins. The high temperature adhesive

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may also absorb and hold plenty of heat, which is not considered in the analysis. Fan power may vary from box fan to box fan, and other drafts within the test setting room are also ignored. There is, however, a similar trend with the pattern of experimental vs. theoretical values. When the theoretical decreases significantly, the experimental value does the same, showing that trends are recognized with the mathematical model.

Due to the various calculations, uncertainty errors needed to be taken into account. The uncertainty values calculated showed a smaller range than the actual variations between testing and theoretical values. The same method used in the Error Estimation Experiment was used in the derivation and calculation of the uncertainty values. The equations are as follows, where dq is equal to 0.5 watts and q is equal to 20 watts.

Uncertainty Calculations

dTs=Ts∗( dqq ) (23)

dtss=tss∗( dqq ) (24)

Table 13. Uncertainty Error Values

Uncertainty Analysistss final 7.631 sTs (w=.005) 9.443 KTs(w=.01) 9.440 KTs(final design) 7.991 K

Recommendations / Conclusions –

The team’s final project placed 19th out of 40 teams, many of which used various versions of the design concepts discussed earlier in this report. For the three parameters, the placings and values are as follows:

1) Temperature – 31.2°C, 28th place2) Weight – 169.2 grams, 3rd place3) Cost - $2.60 – 15th place

The pin fin array design was manufactured based on the idea of using an axial air flow. The team overlooked the idea that air flows in different directions would hinder the performance of each fan. However, the pin fins were already manufactured when the team was close to the competition day, and switching to a rectangular fin array would lead to a risk. For future designs, the team would consider using the rectangular fins if only the box fan were being used. Other recommendations for the pin fin array would include thicker pins at a fewer number, at a shorter length. From the mathematical model, these thicker pins would be able to draw more

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heat away from the plate itself. In addition, the pins could have been manufactured with a tighter tolerance and cut with smoother finishes. The screws used that went into the balsa would could have been smaller. If cost were not as important, advanced heat sinks for computers utilize a combination of copper and aluminum. Hollow copper tubing of the right diameter could be tested with axial flow. The hollow piping could almost double the surface area increased by pins. Orientations between staggered pins and grid-like square patterns could also be tested.

References

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1. C.J. Paul, ME 412 Enhancement of Electronic Cooling. Michigan State University, 2016.

2. C.J. Paul, ME 412 Forced Convection Analysis. Michigan State University, 2016.

3. C.J. Paul, ME 412 Natural Convection Analysis. Michigan State University, 2016.

4. C.J. Paul, ME 412 Transient Analysis. Michigan State University, 2016.

5. "Cooling 101: The Basics of Heat Transfer." Koolance. Koolance, Inc. Web. 02 May 2016.

6. "Heatsink 101." Geek Easy Computers. Geek Easy Computers, 29 May 2013. Web. 02 May 2016.

7. Incropera, Frank P., and David P. DeWitt. Introduction to Heat Transfer. 6th ed. John Wiley &

Sons, New York. Print. 2011.

8. "Replacement 70mm TX3 Dual Ball Bearing CPU Cooler Fan." StarTech.com. StarTech.com. Web.

02 May 2016.

Appendix A – Box Fan Data

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Appendix B – Computer Fan Data

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PerformanceAir Flow Rate 21.33 CFMFan RPM 3500Noise Level 33 dBAPower

Input Current 0.25Input Voltage 12 DC

Power Consumption 2.16W

Physical Characteristics

Color BlackEnclosure Type PlasticProduct Height 2.8 in [70 mm]Product Length 2.8 in [70 mm]Product Weight 1.2 oz [35 g]Product Width 0.4 in [10 mm]

Appendix C – Air Fluid Properties

Appendix D – Properties of Materials Used

Balsa Wood Properties

Compressive Strength 7 MpaDensity 0.13 g/cm^3Specific Heat Capacity 1700 J/kg-K

Aluminum Properties

ρ 2702 kg/m^3N 38 V 0 m^3cp 1013 J/(kg-K)k 205 W/(m-K)α 2.25E-05 m^2/sCost 1.47 $/kg

Arctic Silver Adhesive Properties

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T (K) μ(Pa-s)

k(W/m-K) Pr Cp

(kJ/kg-K)ν

(m2/s)α

(m2/s)β

(1/K)300.00 1.842E-05 2.58E-02 0.707 1.007 1.515E-05 2.229E-05 0.00333333

Form ViscousColor SilverMelting Point >999°CBoiling Point >204°CDensity 2.28 g/cm^3

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