Heat Transfer Engineering, 22:29–39, 2001 Copyright ° C 2001 Taylor & Francis 0145–7632/01 $12.00 + .00
Experimental Study on Closed-Loop Two-Phase Thermosyphon Devices for Cooling MCMs MIN-KYUN NA, JIN-SEOK JEON, and HO-YOUNG KWAK Mechanical Engineering Department, Chung-Ang University, Seoul, Korea
SANG-SIG NAM H/W Environment Section, Electronics and Telecommunication Research Institute, Taejon, Korea
Thermosyphon cooling modules, to cool multichip modules (MCMs), were designed and tested. The cooling module consists of a cold plate with micro nned channels and a plate-type integrated condenser. A separate ow model was employed to predict the mass ux and the pressure drop in the channel of the cold plate. The local and average convective boiling heat transfer coef cients and the corresponding wall superheat were calculated using the Chen’s correlation. Experiments were performed to nd out how the thermal performance of the cooling module was affected by the condenser size and the amount of charging uid. Great emphasis was placed on the transient characteristics of the cooling module. For an allowable temperature rise of 58°C on the surface of the heater, the cooling module can handle a heat ux of as much as 2.5 W/cm2 . No boiling retardation was observed inside the cold plate, which resulted in smooth transition from the transient state to the steady one. It was also found that the appropriate size of the condenser and the adequate amount of charging liquid are crucial factors affecting the performance of a closed two-phase thermosyphon device.
Switching systems in a broadband integrated services digital network (B-ISDN ) will use the asynchronous transfer mode (ATM ) and will have a throughput on the order of terabits per second. Correspondingly, the heat ux in a B-ISDN is one or two orders of magnitude higher than in conventional systems, and will reach 1 – 2 W/cm2 , which will necessitate new packaging systems and ingenious cooling technology [1, 2]. The multichip module (MCM ) has been considered [ 3]
an appropriate packaging design for B-ISDNs because its high packaging density will decrease signal propagation delay and reduce electromagnetic emission. The direct immersion cooling module [4] has been considered a promising method for such application because it removes a large amount of heat effectively. However, application of such a cooling method to communication systems is not an easy task because of dif culties in maintenance and reliability. As an alternative, ingeneous air-cooled thermosyphon modules [5, 6] and indirect liquid-cooling thermosyphons [7, 8] have been proposed for cooling high-density electronic packaging. However, it has been observed [9] that the
Address correspondence to Prof. Ho-Young Kwak, Mechanical Engineering Department, Chung-Ang University, 221, Huksuk-Dong, DongjakKu, Seoul 156-756, Korea. E-mail:
[email protected]
29
thermosyphon cooling module, which has a horizontal refrigerant path in the heat sink [5, 6], brings signi cant heat transfer crisis and consequent temperature rise due to the retardation of boiling inception inside the tube at a lower heat ux of about 1 W/cm2 . On the other hand, a thermosyphon action due to boiling of the refrigerant in a vertical channel completely eliminates the heat transfer crisis problem [ 9]. A design procedure for the cold plate of a two-phase thermosyphon cooling device to cool MCM modules was studied analytically. A separate ow model was employed to predict the mass ux and the pressure drop in the channel of the cold plate [10]. The local and average convective boiling heat transfer coef cients and the corresponding wall superheats were calculated using Chen’s correlation [ 11]. Thermosyphonic cooling modules, which have been designed based on the analysis, were fabricated and tested to cool a multichip module plugged into a planned packaging system. The cooling module, which consists of a cold plate having a vertical refrigerant path and an integrated plate condenser, is essentially a separate type of closed thermosyphon. Experiments were performed to nd out how the thermal performance of the cooling module was affected by the condenser size and the amount of charging liquid. Great emphasis was placed on the transient characteristics of the cooling module. No boiling retardation was observed inside the cold plate, which resulted in smooth transition from the transient state to the steady one. The measured values of the heat transfer coef cient are in good agreement with the mean values of calculated ones. It has also been found that appropriate size of the condenser and an adequate amount of charging liquid were crucial factors affecting the performance of the closed two-phase thermosyphon device. EXPERIMENTAL APPARATUS AND PROCEDURES A multichip module [3] has been proposed for improving the performance of communication systems. The high packaging density of an MCM will decrease signal propagation delay as well as electromagnetic emission. In this study, two 120 mm £ 100 mm £ 3.5 mm aluminum-covered plate heaters were used to simulate an MCM. The maximum allowable heat ow rate from this heater was about 345 W at the applied voltage of 220 V. The corresponding maximum heat ux achieved was about 2.54 W/cm2 when one side was insulated. The heaters were attached to the cold plate by using Omega Therm-101, a thermal paste. R-11 was chosen as a working uid in the thermosyphon device. A schematic of a thermosyphon cooling module tested is shown in Figure 1. The cooling module con30
heat transfer engineering
Figure 1 Schematic design of the TSCMF and monitoring positions of temperature and pressure.
sists of a cold plate with ns (called a CPF ) and a platetype integrated condenser (called a PIC ). Such a cooling module is capable of cooling an array of chips twice as large as the IBM module or two arrays of MCMs. As shown in Figure 2a, eight micro ns were made on the surface of the ow channel in the cold plate. The length, height from root to tip, and width of the n are 150 mm, 7 mm, and 2.5 mm, respectively. The lateral distance between adjacent ns is about 9 mm. The effective heat transfer area in the cold plate is about 160 cm2 , and the residential volume for the refrigerant inside the cold plate is about 256 cm3 . The condenser, shown in Figure 2b, is made of an aluminum block whose dimensions are 55 mm £ 275 mm £ 100 mm so that ow channels for the vapor from the cold plate are close to the condensing surface to minimize the thermal resistance there. Eight horizontal channels for water were fabricated in the block. The channel diameters are 8 mm for water, with equal lengths of 10.0 cm for each channel. The condensing surface area is about 168 cm2 and the residential volume for the vapor inside the condenser is about 257 cm3 . In addition to the cold plate (CPF ) and the plate-type condenser (PIC ) mentioned above, another cold plate (called a CPFE) and plate-type condenser (called a PICE ), which have larger residential volume for the refrigerant or its vapor, were utilized. Combinations of cold plates and condensers employed in this experiment are shown in Table 1. The coolant (water ) was circulated to the condenser by a constant-temperature-bath circulator. The water ow rate to the condenser was xed at about vol. 22 no. 2 2001
returning condensed liquid had different paths, ooding phenomena did not occur in such a separated-type closed thermosyphon unit [5] which allows more heat to be transported. The temperatures of the cold plate and the condenser were measured using 20 T-type thermocouples mounted on the surfaces as shown in Figure 3. Acquisition of data obtained from the T-type thermocouples was done by a Yokogawa recorder (HR-1300 ) connected to a PC. The data collected from each thermocouple at 2-s intervals over 2 min were averaged separately to make a data set. Also, the vapor temperatures at the inlet of condenser, Tc,in, in and at the outlet of the cold plate, Te,out , and the liquid temperatures at the inlet of the cold plate, Te,in, and at the outlet of the condenser, Tc,out , were measured by T-type thermocouples whose monitoring positions are shown in Figure 1. Acquisition of data obtained from these thermocouples was done by a Yokogawa recorder (LR-8100 ). At a heat ux of about
Figure 2 Cold plate (a) and condenser (b) design for TSCMF (units in mm).
5.58 liters/min. The temperature of the water at the inlet to the condenser varied depending on the preset surface temperature of the condenser. The internal space of the cold plate was lled with a working uid. Charging of the working uid to the system was as follows; rst the system was evacuated up to 10 torr by a vacuum pump and then the liquid was charged to the cold plate. After the charge, the system pressure rises normally to 660 torr. The vapor generated at the cold plate owed to the condenser, where it was condensed by the circulating water of a xed ow rate. The liquid then returned to the evaporator by gravity. Since the vapor and Table 1 Combinations of cold plates and condensers and their residential volumes (inside volume of the connecting pipes is about 80 cm3 ) Combinations of cold Residential volume of cold System plate/condenser plate/condenser (cm3 ) volume (cm3 ) CPF/PIC (A-type) CPF/PICE (B-type) CPFE/PIC (C-type) CPFE/PICE (D-type)
256/257
593
256/331
667
350/257
687
350/331
761 Figure 3
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Thermocouple positions on cold plate (units in mm).
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For the micro nned channel, the increase in the contact area between liquid and solid surface due to installation of ns was taken into account for the friction coef cient. For the two-phase multiplier, the Fridel correlation [13] which is applicable to any uid except when l l / l g > 1,000, was utilized. With the assumption of uniform heat ow to the cold plate, the quality at an arbitrary point in the ow passage can be obtained by the energy balance equation. That is,
1 W/cm2 , the velocity or the mass ux of the vapor to the condenser could be measured by monitoring the pressure difference between the two positions 50 cm apart from each other. A differential pressure transducer (Sensotec, P30P ) was utilized in this measurement, and the data were calibrated with a water lled U-tube. With an assumed surface roughness of 0.001 d, the mass uxes were estimated by using the Haaland and Hagen correlation for the friction factor [12] and the measured value of the pressure difference. The system pressure monitored at the inlet of the condenser was measured using a piezoelectric pressure transducer (Omega ).
dx q = dz Ghfg d
An empirical equation used to correlate void fraction data for round tubes [14] was employed in this study.
ANALYSIS OF COLD PLATE PERFORMANCE An analysis was done for the heat transfer performance of a cold plate cooled by convective boiling of R-11 in a ow passage having smooth and micro nned surfaces. The pressure drop in the ow passage was calculated using the separated ow model [ 10]. The pressure gradient to account for friction, acceleration, and gravitation components for vertical channel is given by ³ ´ (1 ¡ x )2 dP 4k d x2 2 +G + ¡ = (1 ¡ a )q l dz d dz a q g
+ [a q
g
+ (1 ¡ a )q l ]g
(1 )
The pressure gradient due to friction may be obtained using the two-phase multiplier, U l . 4k = d
³
dP ¡ dz
´
l
¢ U
2 l
dP ¡ dz
´
l
¡ 1/ 5
= 0.046Re, lo
(3 )
if Re, lo < 2 £ 104 if Re, lo > 2 £ 104
(4 )
where
X tt =
³
1¡ x x
´0.9³ q
g
q
l
´0.5 ³
l
l
l
g
´0.1
(7 )
A correct mass ow rate may be obtained when the pressure drop in the ow passage is equal to the hydrostatic head. For the assumed mass ux, G, one can calculate the pressure drop in the ow passage for a given heat ux. By an iteration procedure, one can nd G with given hydrostatic heads. For the ow passage with micro nned surface, Chen’s correlation with modi cation of the nucleate boiling contribution by ns was utilized [ 11]. This is
³
AP + g 2 AF = h l F + h FZ S AP + g AF
´
(8 )
where (9a )
and m =
s
2h CB kl t F
(9b )
The single-phase liquid convective heat transfer coef cient h l based on the total mass ow rate was calculated using the following McAdams correlation [15] rather than that of Dittus and Boelter [16]:
³
q kl h l = 0.023 Prl0.4 Re, 0.8 x io d
Gd Re, lo = l l 32
(6 )
g = tanh (m H F ) / (m H F )
For the friction coef cient, the following correlations [12] was used: ¡ 1/ 4
0.378
where X tt is the Martinelli parameter,
h CB
2C f G 2 (1 ¡ x )2 = d q l
C f = 0.079Re, lo
¡ ¢¡ = 1 + X tt0.8
a
(2 )
where ¡ (dp / dz )l is a single-phase frictional pressure gradient calculated at a liquid mass ux only, which is given by
³
(5 )
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vol. 22 no. 2 2001
l
q
¡ q g
g
´0.8
+1
(10 )
The nucleate boiling heat transfer coef cient h FZ may be calculated from the Forster-Zuber equation [17], which is ! kl0.79 C pl0.45 q 0.49 l 0.24 0.75 h FZ = 0.00122 0.5 0.29 0.24 0.24 D TSAT D PSAT r l l h fg q g
±
(11 )
In the calculation of the nucleate boiling heat transfer coef cient h FZ , the heat transfer area increase due to ns was also taken into account in the same way to obtain h l . Since h FZ depends on D TSAT , and g depend on h CB , the local wall superheat and heat transfer coef cient can be obtained by iteratively solving the following energy balance equation with Eq. (8 ): q A b = h CB ( A P + g A F )D TSAT
(12 )
For the calculations of the two-phase heat transfer coef cient multiplier F and the nucleate boiling suppression factor S, the curve- t equations recommended by Collier [18] were used. These are F =1
for
1 < 0.1 X tt
³
1 + 0.213 = 2.35 X tt S =
Rel =
´0.736
1 1 + 2.56 £ 10 ¡ 6 Re1.17 TP (1 ¡ x )Gd l l
for
1 > 0.1 X tt (13 )
(14 )
where ReTP = Rel F 1.25 and the calculation of the pressure drop and the heat transfer coef cient were accomplished by dividing the length of the ow channel in the cold plate into about 500 elements of equal size. The local pressure drop and the local heat transfer coef cient were calculated for each element. In turn, the local heat transfer coef cient was used to determine the local wall superheat. At a heat ux level between 0.5 and 3.5 W/cm2 , the values of the parameters such as mass ux, exit quality, average heat transfer coef cient, and base wall superheat for two different cold plate units are shown in Figures 4 and 6. The pressure head used in this calculation was obtained by observation. The higher exit quality for the cold plate with the micro nned channel is a direct consequence of the lower mass ux than that for the smooth surface-channel unit. It can be seen in heat transfer engineering
Figure 4 Calculated mass ux and exit quality depending on the heat ux with or without micro ns in ow channel.
Figure 4 that the cold plate unit with a smooth surface operates at a little higher mass ux level. The measured and calculated values of the mass uxes depending on the heat uxes applied are shown in Figure 5. Except for the cases of lower heat uxes, both results show a similar trend: the mass ux decreases as the heat ux increases. A clear demarcation point in the mass ux can be seen at the heat ux of 0.5 W/cm2 . Below a heat ux of 0.5 W/cm2 , pumping action due to the density gradient of the liquid coolant occurs, so that the mass ux increases as the applied heat ux increases. In fact, formation of a couple of bubbles can only be seen at the heat ux of 0.5 W/cm2 , as shown later in Figure 15a. Once boiling occurs at a higher heat ux, the thermosyphonic action due to the boiling of the refrigerant in the channel becomes signi cant. In this regime, the predicted value of the mass ux with a two-phase ow model yields reasonable agreement with the measurements. However, vapor trapped in the manometer line gave some dif culty to measuring the pressure drop in the two-phase ow regime, which may cause deviation between the experimental and theoretical values.
Figure 5 Calculated and observed mass uxes depending on the heat ux applied.
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Figure 6 Calculated and measured average heat transfer coef cient depending on heat uxes applied. The measured values were obtained from A-type (. ), C-type (j ), and D-type (² ) cooling modules.
The measured and calculated values of the average heat transfer coef cients depending on the heat uxes applied for the micro nned channel are shown in Figure 6. The measured data, denoted by . , j , ² , are the cases for employing the cooling module types A, C, and D as described in Table 1, respectively. The measured mean heat transfer coef cients, which vary linearly with the applied heat uxes, have a similar trend as the calculated values for the smooth surface rather than the micro nned one. It turns out that the micro nned channel results in only a small increase in the heat transfer area, consequently slight enhancement in the heat transfer coef cients. This may be because boiling does not occur on most of the surface area of the ns. The measured and calculated values of the mean values of the bare wall superheat to the cold plate for the micro nned channel unit are plotted as a function of the applied heat uxes in Figure 7. Again, the wall super-
Figure 7 Calculated and measured average wall superheats depending on heat uxes applied. The measured values were obtained from A-type (. ), C-type (j ), and D-type (² ) cooling modules.
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heat transfer engineering
Figure 8 Calculated and measured local heat transfer coef cient over the active length of the cold plate in smooth ow channel at two different heat uxes. The measured values for 1 W/cm2 (±) and 2.5 W/cm2 (² ) were obtained from the A-type cooling module.
heat measured in the various cooling modules of types A, C, and D as denoted by . , j , and ² , respectively, have a similar trend as the calculated results with the smooth surface approximation. Note that inside wall temperatures of the cold plate were calculated from the measured wall temperature of the cold plate by employing the one-dimensional, steady-state conduction equation. Figure 8 shows the calculated variation of the local heat transfer coef cient with downstream position for the smooth channel at two different heat uxes. The variation of the base wall superheat with downstream location for smooth cold plate is shown in Figure 9. The heat transfer coef cients and the wall superheat values measured at three different locations along the channel are also shown in Figures 8 and 9, respectively. Actually, the values were obtained by taking the average of the values measured at two lateral positions shown in
Figure 9 Calculated and measured local wall superheat over the active length of the cold plate in smooth ow channel at two different heat uxes. The measured values for 1 W/cm2 (±) and 2.5 W/cm2 (² ) were obtained from the A-type cooling module.
vol. 22 no. 2 2001
Figure 3. At the lower heat ux of 1 W/cm2 , the difference between measured and calculated values in the base wall superheats and the corresponding heat transfer coef cients is rather large at the bottom of the channel. This is because boiling hardly occurs in the lower part of channel at the heat ux of 1 W/cm2 , as will be clearly seen in Figure 15b. On the other hand, one can predict the wall superheats and the heat transfer coef cients correspondingly in the upper part of the channel reasonably when the higher heat ux of 2.5 W/cm2 is applied. The pumping action due to bubbles in the channel may enhance the heat transfer coef cients more than the ones predicted by Chen’s correlation. The wall superheats and the corresponding heat transfer coef cients measured at the middle of the channel are in good agreement with these calculated with the smooth surface approximation. The relative error in the system pressure measurement is about §0.25%, which yields an error of 0.08° C in the measurement of the saturated temperatures. The uncertainties in the temperature measurements are less than §0.1° C, so the calculated magnitudes of the wall superheat are accurate within §1.0%. Since the uncertainties in the heat ux measurements are approximately §3.0%, the uncertainty in the measurement of heat transfer coef cient was calculated to be within §4.0%. EXPERIMENTS FOR PERFORMANCE OF COOLING MODULE Experiments for the designed cooling module were performed without insulation of the condensing part, condenser wall, or tube connecting the condenser with an evaporator, in order to simulate more accurately the real conditions of electronic chip cooling. Experiments were performed to nd out how the thermal performance of the designed cooling module was affected by the condenser size and the amount of charging liquid. Great emphasis was put on the transient characteristics of the cooling module, which has only been studied previously by Nam et al. [9, 19]. The performance of the D-type cooling module at the heat ux of 1.0 W/cm2 is shown in Figure 10. It is shown by these results that the temperature of the simulated chip reaches steady state within 10 – 15 min. Smooth transition to the steady state without showing the heat transfer crisis can also be seen. The transient characteristics of the cooling modules are closely related to the boiling inception inside the cold plate. The temperature overshoot due to the retardation of boiling inception in the cold plate, which may occur in horizontal tubes in the heat sink [9], could damage the electronic chips. Overcharging of the coolant with a ll ratio of heat transfer engineering
Figure 10 (a)Transient characteristicsfor D-type cooling module at the heat ux of 1.0 W/cm2 with R-11 as coolant. (b) Temperature trend at various positions on evaporator of D-type cooling module at the heat ux of 1.0 W/cm2 with R-11 as coolant. The ll ratio of the working uid was 0.46.
0.73 for the A-type cooling module, the case shown in Figure 11, delays the steady-state operation of the cooling module. This is also true for the oversizing condenser, which is the B-type module shown later, in Figure 13. For both cases, the system experiences high pressure. Especially, as the system pressure increases to as high as 2.5 bar, so that operation of the cooling module was stopped for protection in the case of the overcharging case. This is because excess cooling at the
Figure 11 Transient characteristics for A-type cooling module with ll ratio of coolant 0.73 at the heat ux of 2.5 W/cm2 .
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condenser lowered the temperature of the condensate, which induced dif culty in boiling inception inside the cold plate. The ll ratio of coolant, f v , is de ned as the ratio of the charged coolant volume to the system volume of the cooling module. The vapor does not experience notable temperature variation in the tube from the cold plate to the condenser, which indicates that the vapor transport to the condenser is achieved adiabatically. On the other hand, the condensing uid receives much heat from an ambient, so the uid temperature at the inlet of the cold plate (15° C ) is higher than that at the outlet of the condenser (10° C ). The system pressure increases steadily to reach 1 bar at the steady-state condition. As can be seen in Figure 10b, the wall superheat difference between the bottom and top positions is about 4.6° C. This value is higher than the calculated one, which is about 2°C, as shown in Figure 9. The maximum temperature achieved on the surface of the chips was found to be less than 35° C at the heat ux of 1 W/cm2 . The transient characteristics of the D-type cooling module at 2.5 W/cm2 are shown in Figure 12. At such a
Figure 12 (a)Transient characteristicsfor D-type cooling module at the heat ux of 2.5 W/cm2 with R-11 as coolant. (b) Temperature trend at various positions on evaporator of D-type cooling module at the heat ux of 2.5 W/cm2 with R-11 as coolant. The ll ratio of the working uid was 0.55.
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heat transfer engineering
Figure 13 Average temperature of the cold plate and system pressure depending on ll ratio of coolant at the heat ux of 2.0 W/cm2 for various types of cooling modules tested.
high heat ux level, the average temperature of the cold plate turned out to be 52.5° C, and the base wall superheat varied from 6.4 to 11.5° C along with upstream position. These observed values are also in agreement with the calculated ones, which varied from 6.7 to 13.0° C as shown in Figure 9. Surprisingly, the wall superheat at bottom of the cold plate, Tc8– Tc9, has the lowest value, while the wall superheat at the position of Tc7 is the highest. At this high heat ux level, the system pressure reaches 1.60 bar at a steady state as shown Figure 12a, so that the saturation temperature becomes 37.3° C. However, the base wall superheat differences at the majority of the cold plate are in the range of 4°C. This is favorable for cooling of electronics because the temperature difference on the one unit of the chip is so small. Adequate charging coolant turns out to be an important factor affecting the performance of the cooling modules. For various types of thermosyphon devices employed, the average temperature of the cold plate and the system pressure, depending on the ll ratio of the coolant to the system volume, are shown in Figure 13.
Figure 14 Average temperature of cold plate and system pressure depending on heat ux for C-type cooling module with working uids of R-11 and HCFC-123.
vol. 22 no. 2 2001
Figure 15 Convective boiling in the cold plate ow channel with micro ns at the heat ux of (a) 0.5 W/cm2 , (b) 1 W/cm2 , and (c) 2.5 W/cm2 .
Higher ll ratios, over 50%, the C-type cooling module, yields high system pressure. Charging of the coolant with 90 – 110% of the residential volume of the cold plate produces the best performance among the devices. Instead of the working uid of R11, HCFC-123, an alternative to refrigerant R11, was also tested. The average temperature and the system pressure with HCFC-123 for the C-type module are shown in Figure 14. Similar results for the transient characteristics and the system performance were also obtained by using this coolant. However, the cooling module developed did not work with a coolant of FC-72 under the conditions tested. This may be due to the relatively higher boiling point of FC-72. It is better to use a working uid whose boiling point is close to R11 with the cooling module. A bit of uctuation in the temperature of the cold plate near the outlet (Tc1) was observed. This phenomenon is due to the change of dry patched area at this position from time to time, which was con rmed by visualization of the cold plate. Convective boiling in the con ned channel with micro ns is shown in Figure 15. At the heat ux of 1 W/cm2 , bubble formation can only be seen in the upper part of the channel, so two-phase ow approximation is no longer valid in the lower part of the channel, which is con rmed in Figures 7 and 8. Below the heat ux of 0.5 W/cm2 , forheat transfer engineering
mation of a couple of bubbles occurs on the surface of the cold plate. With the heat ux of 2.5 W/cm2 , the dry patched area at the top of the channel is clearly seen in this gure, which con rms that the maximum heat removed with this cooling module is about 2.5 W/cm2 . Such a heat load, which can be handled by the cooling module developed, is much higher than the maximum heat ux of 1.6 W/cm2 achieved by air-cooled closed thermosyphon devices [5]. It has been observed that the temperature uctuation ampli ed more when one charged liquid to the cold plate with a small amount, which is about 90% of the inner space of the cold plate. It was also found that the time required to reach the steady state is longer and the system’s performance is unstable when one uses an oversized condenser, the B-type cooling module case. Such results indicate that appropriate size of condenser is one of the crucial factors affecting the system’s performance. CONCLUSIONS Two-phase thermosyphon cooling modules with micro ns in the channel of cold plate (TSCMF ) were analyzed and tested to cool MCMs. The cooling module consists of a cold plate and a plate-type integrated condenser. Test results have revealed that the cooling vol. 22 no. 2 2001
37
module showed smooth transition from transient to steady state without heat transfer crisis at any heat ux level. The wall superheats on the cold plate were in the range of 5 – 12° C at the heat ux level of 2.5 W/cm2 , which is also in close agreement with analytical results for the smooth ow channel of the cold plate. The appropriate size of condenser and an adequate amount of charging liquid to the system were found to be crucial factors affecting the system’s performance.
two-phase multiplier used in Eq. (2 ) surface tension, N/m shear stress, N/m U r k
Subscripts CB FC FZ g
NOMENCLATURE Ab AF Ap Cf Cp d F fv g G HF h h fg kl L P Prl D PSAT q Re S tF TW TSAT D TSAT x X tt z
a d g l
q 38
base area of cold plate channel segment, m2 surface area of ns in channel segment, m2 prime surface area of channel section including base area of ns, m2 friction factor given in Eq. (4 ) speci c heat at constant pressure, kJ/kg K hydraulic diameter based on wetted perimeter used in Eq. (4 ), m macroscopic heat transfer coef cient multiplier used in Eq. (13 ) ll ratio of coolant gravitational constant, m/s2 mass ux, kg/m 2 s dimension of ns from root to tip, m heat transfer coef cient, W/m2 K latent heat of vaporization, kJ/kg uid thermal conductivity, W/m K effective length of the cold plate, m pressure, N/m2 liquid Prandtl number used in Eq. (10 ) saturation pressure difference corresponding to D T SAT used in Eq. (11 ) surface heat ux to base area of cold plate, W/cm2 Reynolds number nucleate boiling suppression factor used in Eq. (13 ) n thickness, m wall temperature of prime surface (base wall ) of channel, K saturation temperature of coolant, K wall superheat ( = TW ¡ TSAT ), K mass quality Martinelli parameter used in Eq. (7 ) downstream (vertical ) coordinate measured from inlet edge of active cold plate surface, m vapor void fraction gap size in the cold plate channel, m n ef ciency dynamic viscosity, N s/m 2 density, kg/m 3 heat transfer engineering
l lo TP
convective boiling forced convection Forster-Zuber vapor properties or corresponding to vapor ow alone in the channel liquid properties or corresponding to liquid ow alone in the channel liquid properties evaluated at the total mass ux two-phase ow
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cation for the Future, 9th Int. Display Research Conf., pp. 2 – 5, 1989. Toda, I., Innovation in Telecommunication Towards the 21st Century, NTT Review, vol. 1, pp. 160 – 164, 1990. Yamanaka, N., Kikuchi, S., Kon, T., and Ohsaki, T., Multichip 1.8Gb/s High Speed Space-Division Switching Module Using Copper Polymide Multilayer Substrate, 40th ECTC, Las Vegas, NV, pp. 562 – 570, 1990. Bar-Cohen, A., and Schweitzer, H., Thermosyphon Boiling in Vertical Channels, J. Heat Transfer, vol. 107, pp. 772 – 778, 1980. Kishimoto, T., and Harada, A., Two-Phase Thermosyphon Cooling for Telecom-multichip Modules, in Advances in Electronic Packaging (ASME), pp. 135 – 141, 1992. Budelman, S. A., High Ef ciency Heat Removal System for Electric Devices and the Like, U.S. Patent, 5,394,936, 1995. Ashiwake, N., Nakajima, T., Sasaki, S., Ohsone, Y., Harada, T., Iino, T., Kasai, K., and Idei, A., Apparatus for Cooling Semiconductor Device and Computer Having the Same, U.S. Patent 5,406,807, 1995. Pioro, I., Small Scale Two-phase Loop Thermosyphons for Cooling Telecommunication, Proc. 10th, Int. Heat Pipe Conf., Stuttgart, Germany, 1997. Choi, S., Nam, S., Kim, J., and Kwak, H., Thermal Characteristics of Two Phase Thermosyphon Cooling Module for Multichip Device, in Advances in Energy Ef ciency, Heat/Mass Transfer Enhancement (ASME), pp. 33– 43, 1996. Also in ETRI J., vol. 20, pp. 284 – 300, 1998. Chisholm, D., Two-Phase Flow in Pipelines and Heat Exchangers, George Godwin, London, 1983. Carey, V. P., Mandrusiak, G. D., and Roddy, T., Analysis of the Heat Transfer Performance of Offset Strip Fin Geometries in a Cold Plate Operating in a Two-Phase Thermosyphon, in Cooling Technology for Electronic Equipment, pp. 95– 112, Hemisphere, New York, 1988. White, F. M., Fluid Mechanics, 3d ed., McGraw-Hill, New York, 1994. Freidel, L., Improve Friction Pressure Drop Correlation for Horizontal and Vertical Two-Phase Flow, European TwoPhase Flow Group Meeting, Italy, 1979.
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[14] Wallis, G. B., One-Dimensional Two-Phase Flow, McGraw[15] [16] [17] [18] [19]
Hill, New York, 1969. Rohsenow, W. M., and Choi, H., Heat, Mass and Momentum Transfer, Prentice-Hall, Englewood Cliffs, NJ, 1961. Incropera, F. P., and Dewitt, D. P., Introduction to Heat Transfer, Wiley, New York, 1985. Forster, H. K., and Zuber, N., Dynamics of Vapor Bubble and Boiling Heat Transfer, AIChE J., vol. 1, pp. 531 – 535, 1995. Collier, J. G., Forced Convective Boiling, in Two Phase Flow and Heat Transfer in the Power and Process Industries, chap. 8, Hemisphere, New York, 1981. Nam, S., Kwak, H., and Kim, J., Multichip Module Cooling Apparatus, U.S. Patent 5,859,763, 1999.
Min-Kyun Na received B.S. and M.S. degrees in Mechanical Engineering from Chung-Ang University, Korea, in 1997 and 1999, respectively. Since 1999 he has been with the Research & Development Center, Halla Climate Control Cooperation, Taejon, Korea, where he is now a Researcher. His research interests are noise and vibration problems in cooling devices for cars.
Jin-Seok Jeon received B.S., M.S., and Ph.D. degrees in Mechanical Engineering from ChungAug University, Korea, in 1987, 1989, and 2000, respectively. Since 1992 he has been with the Research & Development Center, Korea Gas Corporation, Ansan, Korea, where he is now a Senior Researcher. His research interests are in sonoluminescence phenomena, boiling heat transfer,
heat transfer engineering
electronic equipment cooling and fuel cell power systems. He is a member of KSME. Ho-Young Kwak received a B.S. degree in Physics from Seoul National University in 1971 and an M.A. in Plasma Physics and a Ph.D. in Mechanical Engineering from the University of Texas at Austin. USA, in 1977 and 1981, respectively. He joined the Mechanical Engineering faculty, Chung-Ang University, Korea in 1981, and is currently a Professor there. He served as Dean of the College of Engineering from February 1997 to February 1999. His research interests are in bubble nucleation, bubble dynamics, sonoluminescence phenomena, boiling heat transfer, electronic equipment cooling, and exergoeconomic analysis for thermal systems. He is a member of KSME, ASME, KPS, APS, and ASA.
Sang-Sig Nam received B.S., M.S., and Ph.D. degrees in Electronic Engineering from Dankook University, Seoul, Korea, in 1981, 1983, and 1999, respectively. He joined ETRI in 1985, where was involved in developing the TDX-1A, 1B, and TDX-10 switching systems. Since 1994 he has participated in the ATM switching system development project, and he is now a Project Leader. His research interests are signal integrity engineering, system packing, and modeling and simulation of high-speed interconnection. He is a member of the KITE, KICS, and KEES of Korea.
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