FLOW BOILING HEAT TRANSFER AND PRESSURE DROP OF R134A IN A MULTI MICROCHANNEL METALLIC EVAPORATOR Adil A. Mohammed,
[email protected] Al-Mustansiriya University, College of Engineering, Baghdad, Iraq Ekhlas M. Fayyadh,
[email protected] University of Technology, Department of Mechanical Engineering 10066 Alsina'a Street, Baghdad, Iraq Mohamed M. Mahmoud,
[email protected],
[email protected] Brunel University London, College of Engineering, Design and Physical Sciences, Kingston Lane, Uxbridge, Middlesex, London, UK UB8 3PH, Zagazig University, Zagazig, Egypt, 44519 Adnan A. Abdulrasool,
[email protected] Al-Mustansiriya University, College of Engineering, Baghdad, Iraq Tassos. G. Karayiannis,
[email protected] Brunel University London, College of Engineering, Design and Physical Sciences, Kingston Lane, Uxbridge, Middlesex, London, UK UB8 3PH
Abstract Flow boiling in multi microchannel evaporators is one of the most efficient methods for cooling electronics and high heat flux devices. However, several fundamental issues such as flow instability and lack of accurate prediction methods are not completely resolved. This paper presents experimental results of flow boiling pressure drop and heat transfer rate of R134a in a multi microchannel evaporator. The evaporator consisted of 25 micro channels with dimensions of 297 µm wide, 695 µm deep and 209 µm separating wall thickness. It was made of oxygen free copper by CNC machining and was 20 mm long and 15 mm wide. Experimental operating conditions spanned the following ranges: wall heat flux 5 – 120 kW/m2, mass flux 50 – 300 kg/m2s and system pressure 8.5 – 12.5 bar. The results demonstrated that the heat transfer coefficient increases with heat flux and vapour quality with an insignificant mass flux effect. The effect of system pressure depends on mass flux, i.e. no pressure effect was found at low mass flux while the heat transfer coefficient increased with pressure at the high mass flux values. The dominant heat transfer mechanism(s) is not clear in the present study although there is some features of nucleate boiling. Flow visualization revealed that flow reversal is evident at most operating conditions. The heat flux at which flow reversal occurs tends to increase as the mass flux increases. The pool boiling correlation of Cooper (1984) and the flow boiling correlation of Mahmoud and Karayianns (2013) predicted the experimental data reasonably well. The measured two phase flow pressure drop increases with vapour quality and mass flux but decreases with increasing system pressure. None of the examined pressure drop correlations could predict the experimental data with a reasonable accuracy.
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INTRODUCTION Flow boiling in microchannels has received a considerable attention by the research community in the last two decades. It was motivated by the capability of microchannels to achieve very high heat transfer rates due to the huge heat transfer surface area per unit volume. This makes microchannel heat exchangers very useful means for cooling high heat flux systems such as electronics equipment. The commercial wide-spread use of flow boiling microchannel heat exchangers is hindered by the lack of understanding several fundamental issues such as the dominant flow boiling heat transfer mechanism(s), flow instability and the lack of design correlations. Several experimental studies were conducted in the past in order to understand these issues and propose design correlations for the prediction of heat transfer coefficient and pressure drop in microchannels. However, there are still large discrepancies among the published experimental data and the reasons are not fully understood. Karayiannis et al. (2012) investigated experimentally the effect of heated length and inner surface characteristics on flow boiling heat transfer rates in vertical mini diameter tubes and reported significant effects on the magnitude and trend of the local heat transfer coefficient versus local vapour quality. Thus, they attributed the discrepancies in the published flow boiling heat transfer data to the variations in the length of the test section and surface characteristics from one laboratory to another. Consolini and Thome (2009) reported that flow instability could be another reason for explaining the discrepancies in the published heat transfer data. This section presents a brief review on the past experimental studies that focused on flow boiling of refrigerants in multi microchannel configurations, which is very relevant to the work presented in this paper. Lee and Mudawar (2005a) investigated flow boiling of R134a in a copper multi-microchannel evaporator consisting of 53 channels of 0.231 mm wide, 0.713 mm deep (AR = 3.1) and 25.3 mm long. The definition of the aspect ratio (AR) in the text of the present paper will be taken as the ratio of the channel height to the channel width. The experimental conditions included inlet quality range 0.001 – 0.25, outlet quality range 0.49 – superheated vapour, mass flux range 127 – 654 kg/m2s, base heat flux range 159 – 938 kW/m2 and system pressure range 1.44 – 6.6 bar. The experiments were conducted in a vapour compression refrigeration system (compressor was used rather than a pump). The inlet vapour quality was varied using a throttling valve while the heat flux was kept constant. It was found that the heat transfer coefficient increases with increasing heat flux and decreases with increasing vapour quality for x < 0.55 after which the effect of heat flux diminishes. The decrease of the heat transfer coefficient with vapour quality was considered as a feature of annular flow and therefore convective boiling was reported by them as a dominant heat transfer mechanism while nucleate boiling was reported to be limited to very low vapour qualities (< 0.05) or very low heat flux values. They did not report on the effect of mass flux. Agostini et al. (2008) conducted an experimental study on flow boiling of R245fa and R236fa in a silicon multi-microchannel heat sink consisting of 67 parallel channels (AR = 3.05, Hch = 0.68 mm, Wch = 0.223 mm, L = 20 mm). They conducted the experiments at the following experimental conditions: base heat flux range 36 – 2210 kW/m2, mass flux range 281 – 1501 kg/m2s, exit vapour quality up to 0.75 and saturation temperature 25 0C. An orifice of 0.5 × 0.223 mm was created at each channel inlet in order to reduce the back flow and instabilities. The local temperature was measured at five axial locations using integrated sensors deposited on the back side of the substrate. The fluid entered the channels as a two phase mixture due to flashing through the inlet orifices. They reported a complex behaviour for the local heat transfer coefficient versus local vapour quality. The local heat transfer coefficient was found to increase with vapour quality at low heat flux values and low to medium mass flux values. There was little mass and heat flux effect. They attributed this behavior to the flashing effect, i.e. boiling did not start with a conventional nucleation process. At intermediate heat flux values, the heat transfer coefficient was found to increase with heat flux (h α qw0.67 ) with little dependence on vapour quality and mass flux. At high heat flux values, the heat transfer coefficient decreased with vapour quality, slightly decreased with heat flux and slightly increased with mass flux. This was attributed by Agostini et al. (2008) to the occurrence of intermittent dryout. Finally, they concluded that convective boiling regime does not seem to exist in microchannels. They did not conclude on the dominant heat transfer mechanism, although there is a strong heat flux effect and weaker mass flux and vapour quality effect in the intermediate heat flux range (possibly indicative of nucleate boiling). It is worth noting that the channel dimensions and aspect ratio are very similar to those examined by Lee and Mudawar (2005a) and the inlet conditions were also similar (two phase inlet rather than sub-cooled inlet). However, the heat transfer characteristics were completely different. The reasons for the differences are not clear, i.e. different fluids (R134a vs. R245fa and R236fa), channel material (copper vs. silicon), experimental system (vapour compression vs. pumped loop system) and temperature measurements (thermocouples vs. integrated sensors). All these parameters could result in a difference in the heat transfer results and consequently the conclusions on the dominant heat transfer mechanism(s). Bertsch et al. (2009) studied flow boiling of R134a and R245fa in two copper multi-microchannel heat sinks consisting of channels having the same aspect ratio (AR = 2.5). The first heat sink consisted of 17 channels (Hch = 1.9 mm, Wch = 0.762 mm, L = 9.53 mm) while the second one consisted of 33 channels (Hch = 0.953 mm, Wch = 0.381 mm, L = 9.35 mm). The channel surface roughness was < 0.5 – 0.6 µm. The mass flux ranged from
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20 to 350 kg/m2s, wall heat flux up to 220 kW/m2, vapour quality ranged from - 0.2 to 0.9 and saturation temperature 8 – 30 0C. The heat transfer coefficient was found to peak at vapour quality value of about 0.1 then stays approximately constant up to vapour quality value of 0.4 – 0.5 then it rapidly decreases with quality. At high heat fluxes, the plateau disappeared where the coefficient rapidly decreased with quality from its maximum value (inverted V-shape). Since the heat transfer coefficient showed strong dependence on heat flux while the mass flux effect was found to be small, nucleate boiling was concluded as a dominant heat transfer mechanism. Madhour et al. (2011) investigated flow boiling of R134a in a copper multi-microchannel evaporator. The evaporator consisted of 100 channels, 0.1 mm wide, 0.68 mm high (AR = 6.8), 0.072 mm separating wall thickness, 15 mm long. The experimental conditions included base heat flux range 25.7 – 1890 kW/m2, mass flux range 205 – 1000 kg/m2s and saturation temperature 63 0C (18.04 bar). The test section was heated using 35 micro-heaters and it included also sensors for the temperature measurements. The flow entered the evaporator through an inlet slit located on top of the evaporator at the mid plane such that the flow stream splits into two separate paths and then the fluid exits from two outlet slits. The footprint heat transfer coefficient (calculated based on the base heat flux) was plotted versus local vapour quality for different base heat flux values. At low mass flux value (205 kg/m2s), the heat transfer coefficient was found to increase rapidly with vapour quality for x < 0.4 then it remained constant for x > 0.4. At high mass flux value (1000 kg/m2s), the heat transfer coefficient remained almost constant with vapour quality except at very low heat flux values where it decreased with quality and at very high heat flux values where it increased with vapour quality. At very low heat flux values, the heat transfer coefficient was found to decrease with increasing heat flux and according to Madhour et al. (2011) the reason of this behaviour was difficult to be explained. They did not report on the dominant heat transfer mechanism(s). The total measured pressure drop was found to increase with increasing mass flux and exit vapour quality and the highest pressure drop value reached 94 kPa at G = 1000 kg/m2s and base heat flux value of 1880 kW/m2. It is worth mentioning that the measured pressure drop was negative at low mass and heat flux values. This was attributed by Madhour et al. (2011) to the possibility of pressure recovery at the outlet. Additionally, they did not report on the pressure drop signal along with the heat transfer results which could be useful to understand whether the flow was stable or not. Mortada et al. (2012) studied flow boiling heat transfer and pressure drop of R134a and R1234yf in a multichannel heat sink consisting of 6 rectangular channels having hydraulic diameter 1.1 mm and was 300 mm long. Their experiments were conducted over a amass flux rang 20 – 100 kg/m2s, heat flux range 2 – 15 kW/m2 and vapour quality range 0 – 1. The heat sink was heated from the top and the bottom surface (no transparent top cover). It was found for R134a that there is no significant heat flux effect and the heat transfer coefficient increases slightly with vapour quality and mass flux up to vapour quality value of 0.4 then it starts decreasing with vapour quality due to dryout. They reported that this is an indication of the dominance of convective boiling. The behavior of R1234yf was reported to be similar to that of R134a except that the heat transfer coefficient of R1234yf was higher than that of R134a. Contrary to the discrepancy in the published heat transfer data, it seems that there is a common agreement on the effect of operating conditions on the measured two phase flow pressure drop. Many researchers such as Lee and Mudawar (2005b), Lee and Mudawar (2008), Mahmoud et al. (2014) and Pike-Wilson and Karayiannis (2014) measured two phase flow pressure drop of refrigerants in microchannels. They reported that the pressure drop decreases slightly with heat flux in the single phase region then it increases rapidly with heat and mass flux in the two phase region. The abovementioned disparity in the published heat transfer data calls for more research in this area in order to understand the fundamentals of flow boiling heat transfer in microchannels. The work presented here in this paper is a continuation to the work started by Fayyadh et al. (2015) to include higher inlet pressures (8.5 – 12.5 bar). Fayyadh et al. conducted flow boiling heat transfer the tests at an inlet pressure value of 6.5 bar and flow visualization was conducted at the mid location of the heat sink. In the present study, the heat transfer rate and pressure drop data are presented along with flow visualization conducted at the mid location near the channel inlet including part of the inlet manifold in order to investigate flow reversal. Additionally, the experimental data were used to assess existing pressure drop and heat transfer correlations in order to have an insight into recommendation for the design of microchannel heat exchangers. EXPERIMENTAL SETUP The detailed description of the experimental facility can be found in Fayyadh et al. (2015). It consisted of a reservoir, a gear pump, a sub-cooler, two Coriolis flow meters, a pre-heater, a test section and a condenser, see Fig. 1. The system pressure was adjusted at the required value by controlling the reservoir temperature. The test section depicted in Fig. 2 consisted of an oxygen free copper block, polycarbonate housing, polycarbonate and quartz glass top cover plates and cartridge heaters. Twenty five rectangular micro channels with the nominal dimensions of 0.3 mm wide (Wch), 0.7 mm deep (Hch), 0.2 mm separating wall thickness (Wth) and 20 mm length (L) were cut on top of the copper block using CNC machining. The surface roughness of the bottom wall was
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measured using a Zygo NewView 5000 surface profiler and the value of the average surface roughness was found to be 0.301 µm. Three cartridge heaters of 175 W per each were inserted horizontally at the bottom of the copper block to provide the heating power to the test section. The power was controlled by a variac and measured by a power meter Hameg HM8115-2 with accuracy of ± 0.4 % for current and voltage. Six T-type thermocouples were inserted vertically along the centerline of the copper block at 12 mm equidistance to help measure the heat flux. Three T-type thermocouples inserted at 1 mm distance below the channel surface and spaced by 8 mm equidistance along the axial direction were used to measure the local heat transfer coefficient along the channel. All thermocouples were calibrated with an accuracy of ± 0.3 K. The copper block was inserted into the polycarbonate housing and was sealed using two O-rings. The housing consisted of the inlet/outlet plenums and manifolds. The fluid inlet and outlet temperature were measured using T-type thermocouples of 1 mm diameter, which were calibrated with an accuracy of ± 0.5 K. The fluid inlet and outlet pressure were measured using absolute pressure transducers located immediately before and after the test section. These transducers were calibrated with an accuracy of ± 0.15 % and ± 0.32 % for inlet and outlet respectively. The pressure drop was measured directly across the test section using a calibrated differential pressure transducer (PX771A100DI) with an accuracy of ±0.081 %. The depth of the inlet/outlet manifold was the same as the depth of the micro channel. A number of holes (0.7 mm) were drilled into the housing to pass the thermocouples wires through them. A transparent layer of quartz glass with 8 mm thickness was sandwiched between the upper surface of the housing and the top polycarbonate cover plate. This layer was sealed using Oring embedded in the top surface of the housing. The top cover plate has a visualization window of similar dimension as the microchannels including the manifolds. Flow visualization was conducted using a high speed camera Phantom V.6 with 1000 f/s at full resolution 512 × 512 pixels and 32000 f/s at 256 × 256 pixels. The camera was integrated with a microscope for better flow visualization. The data were recorded using IMP35951 data acquisition and LabVIEW software after the system reaches steady state, i.e. constant readings with small oscillations. The experiments were conducted by keeping the flow rate constant and increasing the heating power gradually. The data were recorded for 2 min at a frequency of 1 Hz then were averaged to be used in the data reduction process.
Fig. 1. Schematic drawing of the test loop DATA REDUCTION For single phase flow, the net pressure drop along the micro channels ∆Pch is given by:
∆pch = ∆pm − ∆ploss
(1)
The total pressure loss in equation 1 above is given as:
∆ploss = ∆pmi + ∆psc + ∆pex + ∆pmo
4
(2)
The symbols are defined in the Nomenclature section. The pressure loss in the inlet and outlet manifold are given by equation (3) and (4) below, Remsburg (2000).
[
]
1 ∆pmi = 1 − β 2 + K mi × G 2 v f 2
(3)
(b)
Copper block
(c) dimensions in mm
(a) Fig. 2 (a) Test section, (b) microscopic picture of microchannel, and (c) dimensions of copper block 1 1 ∆pmo = − 2 − 1 + K mo × G 2 v f β 2
The loss coefficients K mi and K mo depend on the manifold convergence and divergence angle
(4)
θ . These values,
which are functions of the area ratio β and angle θ are included in Shaughnessy et al. (2005) in a table format. In this case these are 0.11 and 0.134 for K mi and K mo respectively. Remsburg (2000) provides a relation for the sudden contraction and expansion losses in Eq. (2) as follows:
1 1 2 ∆pex = − 2 − 1 + (1 − β ) × G 2 v f β 2
(5)
1 ∆psc = 1 − β 2 + 0.5(1 − β ) × G 2 v f 2
(6)
[
]
The single phase friction factor is then calculated as:
f exp =
∆pch Dh 2 Lv f G 2
(7)
In two phase flow, the pressure loss due to the outlet manifold can be calculated using Eq. (8), Fang and Eckhard (2008).
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x2 (1 − xexit )2 ∆pmo = 0.425G 2 1 − β 2 exit + ρf ρ g
(
)
(8)
Collier and Thome (1994) provide Eq. (9) to calculate the pressure loss due to sudden expansion. ∆pex =
v −v G2 (1 − β )2 v f 1 + g f xexit 2 vf
(9)
The exit vapour quality can be calculated as:
xexit =
(iexit − i f )
(10)
(i g −i f )
The flow enters the channel as a single phase liquid. Therefore the channel pressure drop consists of a single phase, ∆psp and a two phase part, ∆ptp . The net two phase pressure drop along the channel is then calculated as: (11)
∆ptp = ∆pch − ∆psp
The single phase pressure drop can be calculated using:
∆psp = G 2 v f f sp Lsp 2 Dh
(12)
where Lsp, the length of the single phase region is calculated from an energy balance as: .
m c pf Lsp = (Tsat − T f ,in ) q ′′W
(13)
The single phase friction factor fsp in Eq. (12) is the apparent friction factor and is evaluated using Eq. (14) below given by Shah and London (1978) for developing flow.
f sp Re =
3.44 x+
+
K (∞) /( 4 x + ) + ( f Re) fd − 3.44 1 + C ′ ( x + ) 0.2
x+
(14)
+
In the above equation, x is the dimensionless axial distance ( L Re Dh ), ( f Re) fd , K (∞) and C ′ are constants that depend on the aspect ratio. For the geometry used in the present study, the values are 0.931, 0.000076 and 19.071 respectively for K (∞) , C ′ and ( f Re) fd . The base heat flux q′′ is calculated from the measured temperature gradient as:
q′′ = kc
dT dy
(15)
The local heat transfer coefficient was calculated using Eq. (16) below. The temperature of the wall was obtained using the three axial thermocouples after correcting the reading using the 1D heat conduction equation, as follows: h=
N (Tw,local
q′′Wch − T f )(Wch + 2ηH ch )
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(16)
If the thermocouple is located in the single phase region, the fluid temperature Tf is calculated using Eq. (17) below otherwise the saturation temperature is used. The pressure drop needed to obtain the saturation temperature was calculated assuming a linear pressure drop in the two phase region.
T f = T f ,in +
q′′Wch z m c pf
(17)
The fin efficiency is obtained using Eq. (18) below
η = tanh(mH ch ) / mH ch
and
m = 2h / kcWth
(18)
The local vapour quality is calculated as: x( z ) = (i ( z ) − i f ) (ig − i f )
(19)
The experiments were conducted over the following ranges: mass flux G = 50-300 kg/m2 s, heat flux qw = 5– 120 kW/m2, system pressure 8.5 – 12.5 bar. The propagated uncertainty analysis was conducted using the method explained in Coleman and Steel (1999) and the values were 2.5 - 2.39 %, 2.7 – 3 % and 14-40 % for the heat flux, heat transfer coefficient and fanning friction factor respectively. The experimental system was validated using single phase experiments and the results were in a reasonable agreement with the conventional theory, see Fayyadh et al. (2015). RESULTS AND DISCUSSIONS Flow Reversal Flow visualization was conducted at the middle of the heat sink at a location that shows the channel inlet and a small part of the inlet manifold in order to detect whether there is flow reversal or not. The visualization results demonstrated that flow reversal is observed at most operating conditions. For all heat flux values, no flow reversal was observed for G = 50 kg/m2s and P = 8.5 bar. When the mass flux was increased to 100 kg/m2s, flow reversal was observed to occur at a wall heat flux value 40 kW/m2. The heat flux value at which flow reversal occurs increased to 90 kW/m2 when the mass flux was increased to 200 kg/m2s while no flow reversal was observed for G = 300 kg/m2s for the examined heat flux range (up to 110 kW/m2). Bearing in mind that the view of the camera shows only 7 channels at the middle of the heat sink, flow reversal could occur at the side channels and may not be detected by the camera. This means that, for G = 50 kg/m2s and P = 8.5 bar, there is a possibility of unobserved flow reversal in the side channels. On the contrary, significant flow reversal was observed for P = 10.5 bar and G = 50 kg/m2s at the lowest heat flux value of 16.7 kW/m2, see Fig. 3 for the sequence of pictures at these conditions. Starting from a reference time t = 0, the top left picture did not show any bubbles in the viewed channels. After 3 ms, a small bubble appeared in the second channel from the bottom (see the arrow) and this bubble was moving in a reversed direction towards the inlet manifold. After 14 ms, this bubble grew rapidly into an elongated bubble that reached to the inlet manifold. In the meantime, two bubbles were also observed moving in the third channel from the bottom towards the channel inlet, which also grew into an elongated bubble. The two elongated bubbles formed in the two neighbouring channels met in the inlet manifold and merged into a U-shape elongated bubble after about 20 ms. After about 88 ms, the elongated bubbles continued moving towards the inlet manifold and pushed some vapour into the first channel from the bottom (see arrow). After 114 ms, two new elongated bubbles were observed in the upper part of the picture moving from the downstream to the upstream direction. After 136 ms, the view of the camera showed that most of the channels is blocked with nearly stagnant elongated bubbles with part of the vapour clearly observed in the inlet manifold. Figure 3 demonstrates that it took about 466 ms for the bubbles and vapour to disappear from the pictures and single phase liquid flow appears. The presence of the vapour at the channels inlet for long time may cause flow redistribution such that more fluid flows into the other channels. It is worth mentioning, for the same mass flux, Fayyadh et al. (2015) observed similar behaviour at P = 6.5 bar for all heat flux values but the process took about 210 ms, which is about 50 % shorter. Besides, they found that as the mass flux increases, the heat flux at which flow reversal occurs increases which agrees with the results of the current study for P = 8.5 bar. This tendency is not clear for the high pressures (10.5 and 12.5 bar. It is not clear from Fig. 3 whether the bubbles that moved towards the channel inlet were resulting from nucleation, growth and expansion or it is only a vapour flowing back from the outlet manifold. Reviewing the results presented in Fayyadh et al. (2015) for P
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= 6.5 bar and the resent findings for 8.5 – 12.5 bar, in can be stated that the effect of mass flux on flow reversal is not clear and needs to be examined further. The same applies for the effect of system pressure. Inlet manifold
t = 0 ms
t = 5 ms
t = 14 ms
t = 20 ms
t = 88 ms
Flow direction
t = 114 ms
t = 136 ms
t = 156 ms
t = 466 ms
Fig. 3. The sequence of pictures for a flow reversal occurring at G = 50 kg/m2s and P = 10.5 bar, flow direction from right to left. Heat Transfer Results The effect of heat flux, mass flux, vapour quality and system pressure on the heat transfer coefficient calculated at a location near the channel exit is presented in this paper. Figure 4 illustrates the effect of mass flux and system pressure on the boiling. As seen in Fig. 4a for P = 8.5 bar, boiling commenced at low wall superheat with small temperature undershoot for some mass fluxes such as G = 200 kg/m2s. There is no clear mass flux effect indicated in the figure. For some mass fluxes such as G = 300 kg/m2s, the wall superheat decreased with increasing heat flux to the range 40 – 50 kW/m2. This could be due to the activation of more nucleation sites at this heat flux range. Figure 4b indicates for P = 12.5 bar that there is no mass flux effect and all curves merged into a single curve for all heat flux values. A small sudden increase in the wall superheat is observed at the highest heat flux for each mass flux indicating the onset of dryout. The same behavior was also observed for P = 10.5 bar. Figure 4c depicts that the system pressure does not have a significant effect on the boiling curve for G = 200 kg/m2s. It can be concluded from figure 4 that there is no clear mass flux or system pressure effect on the boiling curve.
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(a)
(b)
(c) Fig. 4. Effect of mass flux and system pressure on the boiling curve plotted for the thermocouple near the channel exit; (a) P = 8.5 bar, (b) P = 12.5 bar, (c) system pressure effect at G = 200 kg/m2s
Figure 5 illustrates the effect of heat flux, mass flux and vapour quality on the two phase heat transfer coefficient. The heat transfer coefficient in this figure is calculated at one location close to the channel exit. It is obvious from Fig. 5a that the heat transfer coefficient increases with heat flux and does not depend on mass flux for all mass flux values. The same can be concluded at 12.5 bar, see Fig. 5, with a steeper HTC-heat flux gradient. Figure 5c indicates that for P = 8.5 bar, the heat transfer coefficient increases with vapour quality with an obvious mass flux effect. Additionally, the effect of vapour quality tends to be weak as the mass flux decreases. For example, for G = 50 kg/m2s, the heat transfer coefficient increased rapidly up to x ≈ 0.2 then it increased at a very small rate. The same behaviour was observed for P = 12.5 bar, as seen in Fig. 5d, where the heat transfer coefficient increased with vapour quality and mass flux. According to the conventional criteria for nucleate boiling to be dominant, the heat transfer coefficient also should depend on heat flux, be independent of the vapour quality and mass flux and it should increase with increasing system pressure. In the present study the heat transfer coefficient exhibits strong dependence on the vapour quality, while the effect of mass flux is not clear. Additionally, the effect of system pressure is not conclusive, see Fig. 6. The figure demonstrates for G = 100 kg/m2s that, there is no pressure effect while increasing the pressure from 8.5 to 10.5 bar at G = 300 kg/m2s resulted in an increase in the heat transfer coefficient. However, further increase in system pressure to 12.5 bar resulted in insignificant effect on the heat transfer coefficient. Accordingly, it is very difficult in this study to conclude on the dominant heat transfer mechanism. It seems that both mechanisms contribute to the heat transfer and it is difficult to segregate their separate contributions.
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(b)
(a)
(c) (d) Fig.5. (a) effect of heat and mass flux on the heat transfer coefficient at P = 8.5 bar, (b) effect of heat and mass flux on the heat transfer coefficient at P = 12.5 bar, (c) effect of mass flux and vapour quality on the heat transfer coefficient at P = 8.5 bar, (d) effect of mass flux and vapour quality on the heat transfer coefficient at P = 12.5 bar
(a) (b) Fig. 6. (a) Effect of system pressure on the heat transfer coefficient at (a) G = 100 kg/m2s, (b) G = 300 kg/m2s
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Comparison with Heat Transfer Correlations The measured heat transfer coefficient is compared to that predicted from existing correlations. Seven correlations proposed by Copper (1984), Gungor and Winterton (1986, 1987), Kandlikar (1990), Yu et al (2002), Li and Wu (2010) and Mahmoud and Karayiannis (2013) were assessed in the present study. The accuracy of each correlation is estimated using the mean absolute error percentage (MAEP) given by Eq. (20) below and the percentage of data within the ± 30 % error band (β). Table 1 summaries the performance of all examined correlations using 263 experimental data points. The table indicates that the best examined correlation is the one proposed by Copper (1984) where it predicted 74.78 % of the data within the 30 % error bands. Also, the correlation given by Mahmoud and Karayiannis (2013) predicted 40.1 % of the data at a mean absolute error value of 36.66%. Figure 7 depicts the global comparison with the correlations of Cooper (1984) and Mahmoud and Karayiannis (2013).
MAEP =
1 N
∑
hexp − h pred hexp
× 100
(20)
Table 1 assessment of the heat transfer correlations
Author Copper (1984) Gungor and Winterton (1986) Gungor and Winterton (1987) Kandlikar (1990) Yu et al. (2002) Li and Wu (2010) Mahmoud and Karayiannis (2013)
MAEP, % 24.28 250 98.98 54.9 80.5 44.57 36.66
β, % of data within ± 30 % 74.78 0.3 8.61 25.7 0.58 29.37 40.1
Pressure Drop Results and Comparison with Correlations Figure 8 illustrates the effect of mass flux and system pressure on the two phase pressure drop along the channels. Figures 8a and b indicate that the pressure drop increases rapidly with vapour quality in the low quality region then the rate of increase decreases as the vapour quality increases. The pressure drop decreased suddenly at vapour quality value of about 0.4 for G = 100 kg/m2s and P = 8.5 bar due to the occurrence of dryout. Additionally, the effect of mass flux on pressure drop is small for the mass flux range 50 – 150 kg/m2s. On the contrary, increasing the mass flux from 150 to 200 and 250 kg/m2s resulted in a higher mass flux effect while the effect became very small when the mass flux increased from 250 to 300 kg/m2s for P=8.5 bar. Figure 8c demonstrates that the pressure drop decreases slightly when the pressure was increased from 8.5 to 10.5 bar while it decreased significantly when the pressure increased further to 12.5 bar. The pressure drop results of the present study agree reasonably with the results reported by Mahmoud et al. (2014) and Pike-Wilson and Karayiannis (2014). The measured two phase pressure drop data (247 data points) are compared with the correlations of Mishima and Hibiki (1996), Lee and Lee (2001), Warrier et al (2002), Yu et al (2002), Qu and Mudawar (2003), Lee and Mudawar (2005b), Lee and Garimella (2008). The results of the assessment are summarized in Table 2. The table demonstrates that all examined correlations failed to predict the two phase pressure drop with a reasonable accuracy. The best correlation (Lee and Mudawar (2005b)) predicted only 35.1 % of the data within the ± 30 % error bands at a mean absolute error percentage of 47.2 %.
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(a)
(b)
Figure 7 The global comparison with the best examined correlations, (a) Cooper (1984), (b) Mohamed and Karayiannis(2013)
(a)
(b)
(c) Fig. 8. (a) effect of mass flux and vapour quality on the two phase pressure drop for P = 8.5 bar, (b) effect of mass flux and vapour quality on the two phase pressure drop for P = 12.5 bar, (c) effect of system pressure on the measured two phase pressure drop.
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Table 2 assessment of the pressure drop correlations
Author Mishima and Hibiki (1996) Lee and Lee (2001) Warrier et al. (2002) Yu et al. (2002) Qu and Mudawar (2003) Lee and Mudawar (2005b) Lee and Garimella (2008)
MAEP, % 75.9 61.8 48.2 71.3 59.2 47.2 49.9
β, % of data within ± 30 % 2.5 12.1 28.9 19.6 21.8 35.1 28.8
CONCLUSIONS Experiments have been performed for flow boiling heat transfer and pressure drop of R134a in a copper multi-channel heat sink for mass flux range 50 – 300 kg/m2s and pressure range 8.5 – 12.5 bar. The influence of operating parameters such as mass flux, heat flux and system pressure on heat transfer and pressure drop characteristics was discussed in the paper. The main concluding points are summarized as follows: a. The flow visualization revealed that flow reversal is evident for most operating conditions. The heat flux at which flow reversal occurs tends to increase as the mass flux increases. When it happens, the vapour could stay for more than 0.4s in the inlet manifold and block the inlet to the channels, particularly at low mass fluxes. This may result in dryout in these channels due to the temporary blockage of the channels and thus the overall performance of the heat sink might be influenced. The effect of pressure and mass flux on flow reversal needs further investigation. b. The two phase heat transfer coefficient was found to increase with increasing heat flux and system pressure. The effect of the mass flux is not clear. It is worth mentioning that the transparent top cover was not sealing the top of the channels very well. Therefore, overflow was observed with several nucleating bubbles on the top of the separating walls among the channels even at very high heat flux values. However, the strong dependence of the heat transfer coefficient on vapour quality makes inferring the dominant heat transfer mechanism very difficult. The strong dependence of the heat transfer coefficient on heat flux and vapour quality could be unique for parallel microchannels compared to single channel configurations. c. The comparison with some existing heat transfer correlations demonstrated that the correlation of Cooper (1984) predicted the experimental data reasonably well and better than the examined flow boiling correlations. The second best correlation was the one proposed by Mahmoud and Karayiannis (2013). d. The measured two phase flow pressure drop increases with vapour quality and mass flux but decreases with increasing system pressure. The comparison of the experimental data with some existing pressure drop correlations revealed that all correlations predicted poorly the experimental data. The best examined correlation was that proposed by Lee and Mudawar (2005) which predicted only 35.1 % of the data within the ± 30 % error bands at a mean absolute error percentage 47.2 %. NOMENCLATURE AR cpf
C′
Dh fsp G h Hch if ig iexit K mi K mo
kc K (∞)
L
Aspect ratio, Hch/Wch, [-] Liquid specific heat, [J/kg] Constant in Eq. (14) Hydraulic diameter, [m] Single phase Fanning friction factor, [-] Mass flux, [kg/m2s] Heat transfer coefficient, [W/m2K] Channel height, [m] Specific enthalpy of saturated liquid, [kJ/kg] Specific enthalpy of saturated vapour, [kJ/kg] Specific enthalpy at the exit pressure and temperature, [kJ/kg] Inlet manifold loss coefficient, [-] Outlet manifold loss coefficient, [-] Copper thermal conductivity, [W/m K] Constant in Eq. (14) Channel length, [m]
13
Lsp
m
m N Pi Po ∆P ∆pch ∆psc ∆pex ∆ploss
∆pmi ∆pmo ∆pm
∆psp ∆ptp
qw q′′
Tfi Tf Ti To Tsat Tw, mid vf vg W Wch Wth x xexit
x+
y z
Single phase length, [m] Mass flow rate, [kg/s] Fin parameter, see Eq. (17) Number of channels, [-] Inlet pressure, [Pa] Outlet pressure, [Pa] Pressure drop, [pa] Channel pressure drop, [pa] Pressure drop due to sudden contraction, [pa] Pressure drop due to sudden expansion, [pa] Pressure losses, pa] Pressure drop in the inlet manifold, [pa] Pressure drop in the outlet manifold, [pa] Measured pressure drop, [pa] Single phase pressure drop, [pa] Two phase pressure drop, [pa] Wall heat flux, [W/m2] Base heat flux, [W/m2] Fluid inlet temperature, [K] Fluid temperature, [K] Inlet temperature, [K] Outlet temperature, [K] Saturation temperature, [K] Mid wall temperature, [K] Specific volume of saturated liquid, [m3/kg] Specific volume of saturated vapour, [m3/kg] Heat sink width, [m] Channel width, [m] Channel separating wall thickness, [m] Vapour quality, [-] Exit quality, [-] Dimensionless axial distance, L/ReDh, [-] Vertical distance, [m] Axial distance, [m]
Greek Symbols β Small to large cross sectional area ratio, ± 30% error band Diffuser/nozzle Conical angle, [0] Ɵ η Fin efficiency, [-] ρf Liquid density, [kg/m3] Ρg Vapour density, [kg/m3] REFRENCES Agostini, B., Thome J.R., Fabbri M., Michel B., Calm D., Kloter U., 2008, High Heat Flux Flow Boiling in Silicon Multi-microchannels- Part I: Heat Transfer Characteristics of Refrigerant R236fa, Int. J. Heat and Mass Transfer 51, pp. 5400-5414. Bertsch S.S., Groll E. A., Garimella SV., 2009, Effect of Heat, Mass Flux, Vapour Quality, and Saturation Temperature on Flow Boiling Heat Transfer in Microchannels, Int. J. Multiphase Flow, 35, pp. 142-154. Coleman, H. W., W.G. Steele., 1999, “Experimentation and Uncertainty Analysis for Engineers”, John Wiley and Sons Inc. Second edition, New York, 1999. Collier J.G., Thome J.R., 1994, Convective Boiling and Condensation”. Oxford University Press, Oxford, UK, third edition. Consolini, L., and Thome, J.R., 2009, Micro-channel Flow Boiling Heat Transfer of R134a, R236fa and R245fa, Microfluidics and Nanofluidics, vol. 6, pp. 731-746.
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Cooper M. G., 1984, Saturation Nucleate Pool Boiling. 1st UK National Conference on Heat Transfer, 2, pp.785-793. (Industrial and Chemical Engineering Symposium Series No. 86). Fayyadh, E. M., M. M. Mahmoud and T. G. Karayiannis, 2015, Flow Boiling Heat Transfer of R134a in Multi Micro Channels, Proceedings of the 2nd Int. Conf. on Heat Transfer and Fluid Flow, Barcelona, Spain, 20 – 21 July. Fang, L., and Eckhard, A. G., 2008, Analysis of a Two Phase Flow Ejector for the Transcritical CO2
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