Flow Boiling Heat Transfer of R134a in Multi Micro Channels

Proceedings of the World Congress on Mechanical, Chemical, and Material Engineering (MCM 2015) Barcelona, Spain – July 20 - 21, 2015 Paper No. 296

Flow Boiling Heat Transfer of R134a in Multi Micro Channels Ekhlas M. Fayyadh University of Technology, Department of Mechanical Engineering 10066 Alsina'a Street, Baghdad, Iraq [email protected]

Mohamed M. Mahmoud Brunel University London, College of Engineering, Design and Physical Sciences Kingston Lane, Uxbridge, Middlesex, London, UK UB8 3PH, Zagazig University, Zagazig, Egypt 44519 [email protected]; [email protected]

Tassos G. Karayiannis Brunel University London, College of Engineering, Design and Physical Sciences Kingston Lane, Uxbridge, Middlesex, London, UK UB8 3PH [email protected]

Abstract -Experiments were conducted to investigate flow boiling heat transfer of R134a in a multi micro channel heat sink. The heat sink consisted of 25 micro channels with nominal dimensions of 300 µm wide, 700 µm deep (Dh = 420 µm) and 200 µm fin thickness. The heat sink was made of oxygen free copper by CNC machining with 20 mm length and 15 mm width. Experimental operating conditions spanned the following ranges: heat flux 11.46 – 403.1 kW/m2, mass flux 50 – 300 kg/m2s and system pressure 6.5 bar. A high speed camera was used to capture the flow patterns simultaneously with heat transfer measurements. Three flow patterns were observed namely bubbly, slug and wavy-annular flow when the heat flux increased gradually. The heat transfer coefficient increased with heat flux and there was no mass flux effect. Assessing existing correlations indicated that the correlations of Mahmoud and Karayiannis (2013) and Cooper (1984) predicts the data very well with a mean absolute error less than 20 % compared to the other correlations. Keywords: Micro channel, Flow boiling, Heat transfer.

1. Introduction Two phase flow boiling micro channel heat sinks have emerged as one of the most effective solutions for cooling high and ultra-high heat flux devices such as electronics systems. Mudawar (2011) reported that the applications of two phase micro channel heat sinks should not be limited only to electronics cooling. He discussed and summarized other possible applications which include: (1) cooling turbine blades, (2) cooling fusion reactor blankets, (3) cooling the nozzles of rocket engines, (4) cooling power electronics in avionics and hybrid vehicles, (5) cooling hydrogen storage reservoirs, (6) refrigeration cooling, (7) thermal control in microgravity and capillary–pumped loops. Despite this wide range of applications, two phase micro channel heat sinks are still not commercially available. This arises from the fact that many fundamental issues in flow boiling at micro-scale are still unclear, e.g. flow instability, critical heat flux, lack of design equations, dominant mechanism(s). Research conducted in this area in the last two decades did not reach a comprehensive and universal conclusion about these issues. Table 1 summarizes some of the previous experimental studies that are relevant to mini/micro channels and the experimental conditions covered in these studies. They demonstrate that there is no common agreement on the dominant heat transfer mechanism(s). One group of researchers such as Chen and Garimella (2006), Bertsch et al. (2008), Bertsch et al. (2009), Madhour et al. (2011), and Leao et al. (2014) concluded that nucleate boiling is the dominant heat transfer mechanism. A second group such as Qu and Mudawar 296-1

(2003), Agostini et al. (2008), Mortada et al. (2012) reported that convective boiling is the dominant heat transfer mechanism. A third group such as Lee and Mudawar (2005), Harirchian and Garimella (2008), Lee and Garimella (2008), and Balasubramania et al. (2013) reported that nucleate boiling mechanism dominates at low vapor quality while convective boiling mechanism dominates at high qualities. These researches used the conventional criterion to infer the dominant heat transfer mechanism, i.e. heat transfer coefficient depends only on vapor quality and mass flux in convective boiling while it depends only on heat flux in nucleate boiling. This discrepancy among research groups may be attributed to the difference in experimental conditions, channel size, length, material and surface finish, see Karayannis et al. (2012). Table. 1. Summary of prior studies relevant to parallel rectangular multi mini/micro channels Author Agostini et al. (2008) Balasubramanian (2013)

et

al.

Bertsch et al. (2008) Bertsch et al. (2009)

Material

Dh [mm]/L[mm]/No. of channels

Fluid/ Tin [oC]/Tsat [oC]/ Pin [bar]

G [kg/m2 s]

q" [kW/m2]

Copper

0.336/20/67

R236fa /-/25/2.73

281-1501

36-2210

Copper

0.504-0.4887/25, 20/ 40,16

Water/90/-/-

88-751

Up to 4200

Copper

1.090 / 9.53 / 17

R134a/-/8.9-29/4-7.5

20.3--81

Up to 200

Copper

1.090-0.54 / 9.53 / 17, 33

R134a, R245fa/-/830/-

20-350

Up to220

Silicon

0.389 / 12.7 / 24

FC-77/ 71/-/-

267-458

40 – 800

Silicon

0.16-0.749/12.7/2-60

FC-77 /92/-/ -

250-1600

Up to 4000

Chen and Garimella (2006) Harirchian and Garimella (2008) Harirchian and Garimella (2009) Leao et al. (2014)

Silicon

0.16-0.749/12.7/2-60

FC-77 /92/-/-

225-1461

Up to 3500

Copper

0.167/15/50

R407C/ΔTsub =5-15 /-

400-1500

up to 310

Lee and Garimella (2008)

Silicon

0.16-0.538/12.7/10-60

Water/90-95.1/ -/-

46-126

100 – 3400

Lee and Mudawar (2005)

Copper

0.349/25.3/53

R134a/-/-/1.44-6.6

127-654

159 – 938

Madhour et al. (2011)

Copper

0.174./15/100

R134a/-/63/-

205-1000

25.7 – 1890

1.1/ 300/6

R1234yf, R134a/-//7.7

20-100

2 – 15

Mortada et al. (2012)

Aluminium

400 - 1300 879 - 2400 Dh is hydraulic diameter, L is channel length, Tin is fluid inlet temperature, Tsat is saturation temperature, Pin is inlet pressure, G is mass flux, q" is heat flux. Qu and Mudawar (2003)

Copper

0.349/44/21

Water/30-60/-/1.17

135-402

Some researchers conducted flow visualization simultaneously with heat transfer measurements in order to understand the prevailing heat transfer mechanism(s). For example, Chen and Garimella (2006) conducted a flow visualization study for flow boiling of FC-77 inside a silicon-based multi micro channel heat sink. The location of the high speed camera was at the centre of the heat sink. At low heat fluxes, small bubbles were observed to nucleate on the channel walls then grow, detach and move downstream with a little chance of coalescence. As heat flux was increased, bubble growth and coalescence rates increased resulting in confined bubbles and slug flow. The slug flow was sustained over a narrow range of heat flux. With further increase in heat flux, alternating churn and wispy-annular flows were observed and continued until dryout occurred. It is interesting to note that when the camera was moved to the upstream side near the channel inlet, flow reversal was observed at the heat flux value corresponding to the occurrence of alternating churn and wispy-annular flow. In other words, this alternating flow occurs due to flow reversal according to Chen and Garimella (2006). Harirchian and Garimella (2009) investigated the effect of channel width and mass flux on flow boiling patterns using FC-77 as the working fluid. They reported five flow patterns namely, bubbly, slug, churn, wispy-annular and annular flows. The flow patterns observed in channels of width 0.1 and 0.25 mm were found to be similar but different from those 296-2

observed in the channels of width ≥ 0.4 mm. The difference is that bubbly flow was supressed in the channels of small width and instead, slug flow developed immediately after boiling incipience. With increasing heat flux, intermittent churn/annular flow occurred. By contrast, in the channels of large width, bubbly and intermittent churn/wispy annular flows were observed. Accordingly, Harirchian and Garimella (2009) concluded that nucleate boiling dominates in channels of width ≥ 400µm while convective boiling dominates in channels of smaller width. The above review demonstrates that more research is still required in order to understand the flow boiling heat transfer characteristics and mechanism(s) in micro channels. Thus, the aim of the present work is to investigate experimentally flow boiling heat transfer characteristics of R134a in multi micro channels of 0.42 mm hydraulic diameter. High speed flow visualization was also conducted simultaneously with the heat transfer measurements. Additionally, the experimental data were used to assess some existing prediction methods.

2. Experimental Setup 2. 1. Flow Loop The experimental facility consists of two closed loops. The first one is the test loop using R134a refrigerant while the second is an auxiliary cooling loop using R404a. Figure 1 shows a schematic diagram of the test loop including a close-up photograph of the test section part. It consists of a reservoir, a gear pump, a sub-cooler, two Coriolis flow meters with an accuracy of ±0.1% (one for low flow rate (20 - 100 kg/hr) and the other for high flow rate (100 - 350 kg/hr)), a pre-heater, a test section and a condenser. The system includes two inline filters of 25µm size and three sight glasses in the upstream side of the test section, which were used to confirm that there is no boiling in the pre-heater. System pressure was kept constant by controlling the reservoir temperature using an immersion heater controlled by a PID controller. Flow meters Heater

Cooling coil

Fig. 1. Schematic drawing of the test loop and closed-up photograph of the test section

2. 2. Test Section Figure 2 depicts the details of the test section. It consists of an oxygen free copper block with micro channels on its top surface, polycarbonate housing, polycarbonate and quartz glass top cover plates (not shown in the figure) and cartridge heaters. Twenty-five rectangular micro channels were cut into the top surface of the oxygen free copper block using CNC machine with a feed rate of 550 mm/min. The copper block has overall dimensions of 15 mm width, 20 mm length, i.e. (20 x15 mm2 in the base area) and 74 296-3

mm height (64 + 10 mm, see Fig. 2(c)). The nominal dimensions of the micro channel, see Fig. 2(b), are 300 µm width (Wch), 700µm depth (Hch), 200 µm fin thickness (Wth) and 2000 µm length (L). These dimensions were measured using an electron microscope and the actual values are 297 µm for width, 695 µm for height and 209 µm for fin thickness. The surface roughness of the bottom wall was measured using a Zygo NewView 5000 surface profiler and the value of relative surface roughness was found to be 0.301 µm. Three cartridge heaters of 175 W heating power each were inserted horizontally at the bottom of the

(b)

Copper block

(c) dimensions in mm

(a) Fig. 2. (a) Test section, (b) microscopic picture of microchannel, and (c) dimensions of copper block

copper block to provide the heating power to the test section, see Fig. 2(a). 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 centreline of the copper block at 12 mm equidistance to measure the heat flux. Three T-type thermocouples were inserted along the axial direction of the channel at 8 mm equidistance and 1 mm from the channel bottom to help measure the local heat transfer coefficient along the channel. Another row of thermocouples parallel to these three thermocouples and spaced by 12 mm vertical distance were used to help verify that there is no heat flow in the horizontal direction (direction of flow), see Fig. 2(c). All thermocouples were 0.5 mm diameter and inserted at the center plane of the copper block. These 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 as seen in Fig. 2. The housing consists 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. 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 is 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 296-4

polycarbonate cover plate. This layer was sealed using O-ring embedded in the top surface of the housing. The top cover plate has a visualization window of similar dimension as the micro channels 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 recoded 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 1Hz then was averaged to be used in the data reduction process.

3. Data Reduction For single phase flow, the net pressure drop along the micro channel ∆Pch is given by: ∆Pch = ∆Pm − ∆Ploss

(1)

where ∆Pm is the total measured pressure drop and ∆Ploss is the total pressure losses due to the inlet manifold ΔPmi, outlet manifold ΔPmo, sudden contraction ∆Psc, and sudden expansion ∆Pex, see Eq. (2) below. ∆ploss = ∆pmi + ∆psc + ∆pex + ∆pmo

(2)

The components of the pressure losses in the above equation are given below, see Remsburg (2000). 1 2

∆𝑃𝑚𝑖 = [1 − 𝜎 2 + 𝐾𝑚𝑖 ] × 𝐺 2 𝑣𝑓 1

(3)

1

∆𝑃𝑚𝑜 = − [𝜎2 − 1 + 𝐾𝑚𝑜 ] × 2 𝐺 2 𝑣𝑓

(4)

The loss coefficients 𝐾𝑚𝑖 and 𝐾𝑚𝑜 depend on the manifold convergence and divergence angle 𝜃 and the values are summarized in the textbook given by Shaughnessy et al. (2005) in a table format as a function of the area ratio 𝜎 and angle 𝜃. The values are 0.134 for 𝐾𝑚𝑖 and 0.11 for 𝐾𝑚𝑜 for our design. In the above equations 𝜎 is the small to large cross sectional area ratio, 𝐺 is the mass flux and 𝑣𝑓 is the liquid specific volume. The sudden contraction and expansion losses in Eq. (2) are given by the following equations reported in Remsburg (2000): 1

1

∆𝑃𝑒𝑥 = − [𝜎2 − 1 + (1 − 𝜎)2 ] × 2 𝐺 2 𝑣𝑓

(5)

1

∆𝑃𝑠𝑐 = [1 − 𝜎 2 + 0.5(1 − 𝜎)] × 2 𝐺 2 𝑣𝑓

(6)

The experimentally determined single phase friction factor is then calculated as: ∆𝑃

𝐷

𝑓𝑒𝑥𝑝 = 2𝐿𝑣𝑐ℎ 𝐺ℎ2

(7)

𝑓

For two phase flow, the pressure loss due to sudden expansion ∆𝑃𝑒𝑥 and outlet manifold ∆𝑃𝑚𝑜 in Eq. (2) must be considered as two phase flow. The pressure loss due to the outlet manifold can be calculated using Eq. (8) below given by Fang and Eckhard (2008). ∆pmo = 0.425𝐺 2 [1 − 𝜎 2 ] [

2 𝑥𝑒𝑥𝑖𝑡 ρg

+

(1−𝑥𝑒𝑥𝑖𝑡 )2 𝜌𝑓

]

(8) 296-5

The pressure loss due to sudden expansion can be calculated using Eq. (9) given by Collier and Thome (1994). ∆𝑃𝑒𝑥 =

𝐺2 (1 − 2

𝜎)2 𝑣𝑓 [1 +

𝑣𝑔 −𝑣𝑓 𝑣𝑓

𝑥𝑒𝑥𝑖𝑡 ]

(9)

where 𝑣𝑔 is the gas specific volume. The exit vapour quality in the above equation can be calculated as: xexit = (𝑖𝑒𝑥𝑖𝑡 − if )⁄(ig − if )

(10)

where 𝑖𝑒𝑥𝑖𝑡 , 𝑖𝑓 , 𝑖𝑔 are the specific enthalpies of the fluid at the exit, saturated liquid and saturated vapour respectively. These parameters are estimated at the exit temperature and pressure. Since the flow enters the channel as a single phase liquid, the channel pressure drop is divided into single phase part, ∆𝑃𝑠𝑝 and a two phase part, ∆𝑃𝑡𝑝 . Thus, the net two phase pressure drop along the channel is calculated as: ∆𝑃𝑡𝑝 = ∆𝑃𝑐ℎ − ∆𝑃𝑠𝑝

(11)

The single phase pressure drop along the single phase region can be calculated from: ∆psp = G2 𝑣𝑓 (fsp ∗ Lsp )⁄2Dh

(12)

The length of the single phase region Lsp is calculated from an energy balance as: ṁ𝑐

Lsp = (Tsat −Tf,in ) ′′𝑝𝑓 𝑞 𝑊

(13)

where 𝑚̇, 𝑐𝑝𝑓 , Tsat, Tf,in, 𝑞 ′′, W are the total mass flow rate, liquid specific heat, saturation temperature at the location of zero vapour quality, fluid inlet temperature, base heat flux and width of the heat sink respectively. The single phase friction factor fsp in Eq. (12) is evaluated using the Shah and London (1978) equation for developing flow. The base heat flux 𝑞 ′′ is calculated from the measured temperature gradient as: 𝑑𝑇

𝑞 ′′ = 𝑘𝑐 𝑑𝑦

(14)

where kc is the thermal conductivity of copper and y is the vertical distance. The local heat transfer coefficient was calculated at the midpoint of the heat sink using the mid thermocouple Tw,mid after correcting the reading using the 1D heat conduction equation as follows: h=

𝑞′′ Wch (Tw,mid −Tf )(Wch +2ηHch )

(15)

If the midpoint is located in the single phase region, the fluid temperature Tf is calculated using Eq. (16) below. Tf = Tf,in +

𝑞′′ Wch Z ṁ𝑐𝑝𝑓

(16)

where z is the axial distance, η is the fin efficiency (𝜂 = tanh(𝑚𝐻𝑐ℎ )⁄𝑚𝐻𝑐ℎ ), m is the fin parameter √2ℎ⁄𝑘𝑐 𝑊𝑡ℎ . The local vapour quality is calculated as: 296-6

x(z) = (i(z) − if )⁄(ig − if )

(17)

If the midpoint is located in the two phase region, the saturation temperature is used. This saturation temperature is obtained assuming that pressure drop varies linearly along the two phase region. Experimental data were acquired over the following range of parameters: mass flux G = 100-500 kg/m2 s, heat flux 𝑞 ′′ = 11.46– 403.1 kW/m2, system pressure 6.5 bar. The propagated uncertainty analysis is conducted using the method explained in Coleman and Steel (1999) and the values were 2.5 - 2.39 %, 2.7 – 3 % and 14-40 % respectively for heat flux, heat transfer coefficient and fanning friction factor. The higher value of the uncertainty of the friction factor (40 %) is due to error in the flow and pressure drop measurement at low values at the lowest Reynold number.

3. Results and Discussions 3. 1. Single-Phase Results Single phase experiments were conducted before boiling experiments to validate the experimental system. Figure 3(a) depicts the measured friction factor compared with the Shah and London (1978) correlation for developing and fully developed flow. The figure demonstrates that there is a good agreement with the correlations and the deviation is within the experimental uncertainty. Figure 3(b) shows the experimental Nusselt number compared with the predictions from the correlations of Shah and London (1978) and Stephan and Preuber (1979). It is obvious that the experimental values show similar trend where Nusselt number increases with Reynolds number but the values are much higher than the predictions from the two correlations. However, the current experimental results agree very well with the experimental results of Agostini et al. (2008) for multi-microchannel configuration. 0.2 0.1

30 Shah and London (1978) ,Developing Flow Lam.

20

10

Nu (-)

f (-)

Stephan and PreuBer (1979)

0.01

5 Shah and London(1978)

Shah and London (1978), Fully developed Lam.

2

Present experimental w ork Experiment of Agostini et al.(2008)

Present w ork

0.001 200

1000

1 100

7000

Re (-)

1000

7000

Re (-)

(a) (b) Fig. 3. Single phase results (a) Fanning friction factor versus Reynolds number (b) Nusselt number versus Reynolds number

3. 2. Two Phase Results Figure 4 depicts the flow boiling patterns captured at the middle of the heat sink at a mass flux value of 50 kg/m2 s as a function of heat flux. At low heat flux (16.6 kW/m2), very small bubbles were observed in most channels with majority of bubbles nucleating near the channel corners, see Fig. 4(a). These bubbles detached from the wall and moved downstream without coalescence after they reach a certain size. As heat flux increased to 47.16 kW/m2 (Fig. 4(b)) the bubble growth rate increased and some channels exhibited confined bubble flow where bubble size is such that it spans the channel width, while bubble coalescence was observed in some other channels. Increasing the heat flux to 64.1kW/m2 (Fig. 4(c)) resulted in slug flow in most channels. In this regime, small nucleating bubbles are still observed in 296-7

the liquid slug and near the channel walls. With further increase in heat flux to 152.5 kW/m2 as seen in Fig. 4(d), the nucleation at the channel walls is not observed and the flow shows the features of wavyannular flow.

Flow direction

(b) 𝒒𝒃 =47.2 kW/m2 (bubbly/confined bubble flow)

(a) 𝒒𝒃 =16.6 kW/m2 (bubbly flow)

(c) 𝒒𝒃 =64.1 kW/m2 (slug flow)

(d) 𝒒𝒃 =152.5 kW/m2 (wavy-annular flow)

Fig. 4. Flow boiling patterns as a function of heat flux at G = 50 kg/m2s

Figure 5 illustrates the effect of mass flux on the boiling curve plotted based on the temperature measured by the thermocouple located at the middle of the heat sink. As seen in the figure, there is no temperature undershoot for all mass fluxes and boiling started smoothly at very low wall superheat (0.8 – 1.2 K). The figure indicates that the heat flux increased almost linearly with wall superheat and the wall superheat at heat flux value of 250 kW/m2 is about 5.2 K. It is worth mentioning that this experiment was not planned to reach critical heat flux. Additionally, it is obvious from the figure that there is no clear mass flux effect on the boiling curve. Figure 6 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 at the middle of the heat sink. It is obvious from Fig. 6(a) that the heat transfer coefficient increases with heat flux and does not depend on mass flux. Figure 6(b) depicts the heat transfer coefficient versus vapour quality for different mass flux. The figure indicates that the heat transfer coefficient increases with vapour quality and mass flux, which contradicts the conclusion that one can reach if only Fig. 6(a) is assessed. The results of figure 6 makes inferring the dominant heat transfer mechanism using the conventional criteria very difficult. Figure 6(a) exhibits features of nucleate boiling while Fig. 6(b) shows features of convective boiling. However, flow visualization demonstrated that bubble nucleation was observed up to intermediated heat flux values. Accordingly, the dominant heat transfer mechanism in this study seems to be nucleate boiling but still it needs more investigation.

296-8

350 300

Heat flux ,q",(kW/m2)

250 200 150 100

G=50 kg/m2s G=100 kg/m2s G=200kg/m2s

50

G=300kg/m2s 0 0

1

2

3

4

5

6

7

Tw -Tsat (K)

Fig. 5. Boiling curve at four different mass fluxes 20

25000

20000

htp (W/m2 K)

htp (kW/m2 K)

15

10

G=50 kg/m 2s G=100 kg/m 2s G=200 kg/m 2s G=300 kg/m 2s

5

0 0

50

100

150

200

250

300

15000

mass flux(G)

10000


5000

0 0

350

0.2

0.4

0.6

0.8

1

xe,exit ( - )

2

Heat flux, q" (kW/m )

(b)

(a)

Fig. 6. (a) Heat transfer coefficient versus heat flux, (b) heat transfer coefficient versus vapor quality

4. Comparison With Exiting Prediction Methods The measured heat transfer coefficient in the present work is compared with the heat transfer coefficient predicted by the correlations summarized in the appendix. The accuracy of each correlation is estimated using the mean absolute error percentage (MAEP) as follows: 𝑀𝐴𝐸𝑃 =

ℎ −ℎ 1 ∑ | 𝑒𝑥𝑝 𝑝𝑟𝑒𝑑 | ∗ 100 𝑁 ℎ𝑒𝑥𝑝

(18)

As shown in Fig. 7, the macro scale correlations of Gungor and Winterton (1986 and 1987) and Kandlikar (1990) highly over predict the experimental data where the mean absolute error ranged from 61.2 to 154 % as seen in Fig, 7(a-c). The correlations of Lazarek and Black (1982) and Kew and Cornwell (1996) predicted the experimental data reasonably well at low heat and mass fluxes but the prediction gets worse as mass and heat flux increase. Tran et al. (1996) modified the correlation of Lazarek and Black (1982) by replacing the liquid Reynolds number with Weber number. This correlation slightly under predicted the experimental data – but still within the ±30% error band with a mean absolute error of 27.6 %. Warrier et al. (2002) proposed a heat transfer correlation for micro channels which includes the effect of boiling number and vapour quality. As seen in Fig. 6(g), this correlation shows scattered data with a mean absolute value of 53%. The correlations of Cooper (1984) and Mahmoud and Karayiannis 296-9

(2013) show the best predictions with a mean absolute error less than 20 %. It is worth mentioning that the correlation of Mahmoud and Karayiannis was developed for micro tubes using R134a data based on the model of Chen (1966). They modified the nucleate boiling suppression factor and the convective boiling enhancement factor and also used the Cooper (1984) correlation for the nucleate pool boiling part. This correlation takes into account also the effect of channel diameter which is embedded in the enhancement factor. The similarity in performance between this correlation and Cooper correlation and the agreement with the experimental data means that the correlation predicts the dominant mechanism very well. In other words, the correlation predicts that there is suppression for the convective boiling enhancement factor. The agreement with the Cooper correlation may be considered as an indication on the dominance of nucleate boiling heat transfer mechanism in this study. 50000

70000

50000

Gungor Wintrton,1986 MAEP=154%

Gungor and Winterton 1987

Kandlikar 1990 MAEP=94.9%

60000

MAEP=61.17%

+30%

-30%

20000 2

G=50 kg/m s

50000

-30% 40000

30000

20000

2

G=100 kg/m 2s G=200 kg/m 2s

10000

G=300 kg/m 2s 10000

20000

30000

40000

0

10000 20000 30000 40000 50000 60000 70000

Tran et al.(1996) MAEP=27.57%

Kew and Cornwell 1997

20000

G=50 kg/m 2s G=100 kg/m 2s G=200 kg/m 2s

10000

Predicted HTC(W/m 2K)

-30%

+30%

-30% 30000

20000

G=50 kg/m 2s G=100 kg/m 2s 10000

40000

50000

0

10000

Measured HTC (W/m2 K)

20000

(d)

30000

40000

0

10000

20000

20000

G=50 kg/m 2s G=100 kg/m 2s

10000

2

G=200 kg/m s

MAEP=19.89% 40000

(f) MAEP=19.% +30%

30000

40000

-30%

20000

G=50 kg/m 2s G=100 kg/m 2s

10000

40000 2

Measured HTC (W/m K)

50000

-30% 30000

20000


10000

G=300 kg/m 2s 0

30000

+30%

G=200 kg/m 2s

G=300 kg/m 2s 0

50000

Mohamed and karayinnis 2013


Predicted HTC [W/m 2 k]

30000

40000

50000

+30%

-30%

30000


(e)

MAEP=53.14%

20000

G=50 kg/m 2s G=100 kg/m 2s G=200 kg/m 2s G=300kg/m 2s

Cooper 1984

40000

10000

20000

50000

50000

0

-30%


50000

Warrier et al. 2002

30000

0

0 30000

+30%

10000

G=200 kg/m 2s G=300 kg/m 2s

G=300 kg/m 2s 0

50000

40000


MAEP=31.83% 40000

30000

40000

50000

MAEP=32.66% +30%

30000

(c)

(b)

Lazarek and Black 1982

20000

20000


50000

10000

10000


(a)


G=300 kg/m 2s 0

0

50000

50000

0

G=50 kg/m 2s G=100 kg/m 2s G=200 kg/m 2s

G=300 kg/m 2s


40000

20000

10000

0

0 0

-30%

30000

G=50 kg/m s

G=100 kg/m 2s G=200kg/m 2s




+30% 30000

10000


+30%

40000

40000

0 0

10000

20000

30000

40000

Measured HTC [W/m2 K]

50000

0

10000

20000

30000

40000

50000


(h) (g) (i) Fig. 7. Comparison of the measured boiling heat transfer coefficient with some existing correlations

5. Conclusions Flow boing experiments in a copper multi microchannel heat sink using R134a were performed for a mass flux range 100-500 kg/m2 s and heat flux range 11.46– 403.1 kW/m2. Flow boiling patterns were

296-10

also studied using a high-speed high-resolution camera. The main concluding points can be summarized as follow: 1. Three flow patterns were observed when the heat flux was increased gradually in small steps. These patterns include bubbly, slug and wavy-annular flows. Bubble nucleation was observed up to slug flow and was not observed in wavy-annular flow. 2. The heat transfer results demonstrated that the heat transfer coefficient depends strongly on heat flux while it is a week function of mass flux. However, plotting the heat transfer coefficient versus vapour quality indicated dependence on vapour quality and mass flux. Based on that, it would be difficult to infer the dominant mechanism using the conventional criteria. 3. Based on flow visualization, nucleate boiling seems to be dominant at least up to intermediate heat fluxes (or slug flow). This point needs more investigation. 4. Comparison with some existing macro and micro scale correlations demonstrated that only the correlations of Mahmoud and Karayiannis (2013) and Cooper (1984) predicted the experimental data very well with a mean absolute error less than 20 %. 5. The correlation of Mahmoud and Karayiannis is recommended for design purposes, although further research is also advisable.

References 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, 5400-5414. Balasubramanian, K., Jagirdar, M., Lee, P.S., Teo, C.J., & Chou, S.K. (2013). Experimental Investigation Of Flow Boiling Heat Transfer And Instabilities In Straight Microchannels. Int. J. Heat and Mass Transfer, 66, 655-671. Bertsch, S.S., Groll, E. A., & Garimella, S.V. (2008). Refrigerant Flow Boiling Heat Transfer In Parallel Micro-Channels As A Function Of Local Vapour Quality. Int. J. Heat and Mass Transfer, 51, 4775-4787. Bertsch, S.S., Groll, E. A., & Garimella, S.V. (2009). Effect Of Heat, Mass Flux, Vapour Quality, And Saturation Temperature On Flow Boiling Heat Transfer In Microchannels. Int. J. Multiphase Flow, 35, 142-154. Chen, J.C. (1966). A Correlation For Flow Boiling Heat Transfer To Saturated Fluids In Convective Flow. Industrial and Engineering chemistry ,Process design and development, 5(3), 322-329, Chen, T., & Garimella, SV. (2006). Measurement And High Speed Visualization Of Flow Boiling Of A Dielectric Fluid In A Silicon Micro- Channel Heat Sink. Int. J. Multiphase Flow, 32(8), 957-971. Collier, J.G., & J.R. Thome. (1994). Convective Boiling and Condensation. Oxford University Press, Oxford, UK, third edition. Coleman, H. W., & W.G. Steele. (1999). Experimentation And Uncertainty Analysis For Engineers. John Wiley and Sons Inc. Second edition, New York. Cooper, M. G. (1984). Saturation Nucleate Pool Boiling. 1st UK National Conference on Heat Transfer, 2, 785-793. (Industrial and Chemical Engineering Symposium Series No. 86). Liu, F., Groll, E. A. (2008). Analysis Of A Two Phase Flow Ejector For The Transcritical CO2 Cycle. International Refrigeration and Air Conditioning Conference at Purdue. Gungor, K. E., & Winterton, R.H.S. (1986). A General Correlation For Flow Boiling In Tubes And Annuli. Int. J. Heat and Mass Transfer, 29(3), 351-358. Gungor, K. E., & Winterton, R.H.S. (1987). Simplified General Correlation For Saturated And Comparisons Of Correlations With Data. The Canadian Journal of Chemical Engineering, 65(1), 148 – 156. Harirchian, T., & Garimella, SV. (2008). Microchannel Size Effects On Local Flow Boiling Heat Transfer To A Dielectric Fluid. Int. J. Heat and Mass Transfer, 51, 3724-3733. Harirchian, T., & Garimella, SV. (2009). Effect Of Channel Dimension, Heat Flux, And Mass Flux On Flow Boiling Regimes In Microchannels. Int. J. Multiphase Flow, 35, 349-362. 296-11

Kandlikar, S. G. (1990). A General Correlation For Two-Phase Flow Boiling Heat Transfer Coefficient Inside Horizontal And Vertical Tubes. Int. J. Heat Transfer, 102, 219 – 228. Karayiannis, T.G., Mahmoud, M. M., & Kenning, D. B. R. (2012). A Study Of Discrepancies In Flow Boiling Results In Small To Microdiameter Metallic Tubes. Exp. Thermal and Fluid Science, 36, 126-142. Kew, P.A., & Cornwell, K. (1997). Correlations For The Prediction Of Boiling Heat Transfer In Small Diameter Channels. Applied Thermal Engineering, 17(8-10), 705 – 715. Lazarek, G.M., & Black, S.H. (1982). Evaporative Heat Transfer, Pressure Drop And Critical Heat Flux In A Small Tube With R113. Int. J. Heat Mass transfer, 25, 945 – 960. Leao, H. L. S., Nascimento, L., Do, F.J., & Ribatski, G. (2014). Flow Boiling Heat Transfer Of R407C In A Microchannels Based Heat Spreader. J. Experimental Thermal and Fluid Science, 59, 140151. Lee, J., & Mudawar, I. (2005). Two Phase Flow In High Heat Flux Micro-Channel Heat Sink For Refrigeration Cooling Applications: Part II-Heat Transfer Characteristics. Int. J. Heat and Mass Transfer, 48. 941-955. Lee, P. S., & Garimella, SV. (2008). Saturated Flow Boiling Heat Transfer And Pressure Drop In Silicon Microchannel Arrays. Int. J. Heat and Mass Transfer, 51, 789-806. Madhour, Y., Olivier, J., Patry, E.C., Paredes, S., Michle, B., & Thome, J.R. (2011). Flow Boiling Of R134a In A Multi-Micro Channel Heat Sink With Hotspot Heaters For Energy –Efficient Microelectronic CPU Cooling Application. IEEE Transactions on components, Packing and Manufacturing Technology, 1, 873-883. Mahmoud, M. M., & Karayiannis, T. G. (2013). Heat Transfer Correlation For Flow Boiling In Small To Micro Tubes. Int. J. of Heat and Mass Transfer, 66, 553-574. Mortada, S., Zoughaib, A., & Daurelle, C.A. (2012). Boiling Heat Transfer And Pressure Drop Of R134a And R-1234yf In Minichannels For Low Mass Fluxes. Int. J. Refrigeration, 35, 962-973. Mudawar, I. (2011). Two Phase Microchannel Heat Sinks: Theory, Applications And Limitations. J. of Electronic Packaging, 133, 041002:1 – 31. Qu, W., & Mudawar, I. (2003). Flow Boiling Heat Transfer In Two Phase Micro-Channel Heat SinkPart I. Experimental Investigation And Assessment Of Correlation Method. Int. J. Heat and Mass Transfer, 46, 2755-2771. Remsburg, R. (2000). Thermal Design of Electronic Equipment. CRC Press. Shah, R.K., & London, A. L. (1978). Laminar Flow Forced Convection In Ducts. Suppl. 1, Adv. Heat Transfer. Shaughnessy, E. J., Katz, I. M., & Schaffer, J. P. (2005). Introduction To Fluid Mechanics. New York Oxford, Oxford University Press. Stephan, K., & Preuber, P. (1979). Wӓrmeűbergang Und Maximale Wӓrmstromichte Beim Behӓltersieden Binӓrer Und Ternӓrer Flűssigkeitsgemische. Chem. Ing. Tech., 51, 37. Tran, T.N., Wambsganss, M.W., & France, D.M. (1996). Small Circular And Rectangular Channel Boiling With Two Refrigerants. Int. J. Multiphase Flow, 22, 485 – 493. Warrier, G.R., Dhir, V.K., Momoda, L.A. (2002). Heat Transfer And Pressure Drop In Narrow Rectangularchannels. Exp. Thermal Fluid Sci., 26, 53 – 64.

296-12

Appendix Lazarek and Black (1982) Cooper (1984)

ℎ𝑡𝑝 = 30

𝑘𝑙 𝐷ℎ

𝑅𝑒𝑙0.857 𝐵𝑜 0.714 ,

(0.12−0.2 log10 𝑅𝑃 )

ℎ𝑡𝑝 = 55 𝑃𝑅 𝑁𝑢3

𝐵𝑜 = 𝑞′′ ⁄𝐺ℎ𝑓𝑔

𝑅𝑒𝑙 =

(− log10 𝑃𝑅 )−0.55 𝑀−0.5 𝑞′′0.67 ℎ𝑙 = 0.023 𝑅𝑒𝑙0.8 𝑃𝑟𝑙0.4

𝑘𝑙

If 𝐹𝑟𝑙 ≤ 0.05 replace 𝐸 by

Gungor and Winterton (1986)

ℎ𝑡𝑝 =

Gungor and Winterton (1987)

ℎ𝑡𝑝 = (𝑆𝑆2 + 𝐹𝐹2 ) ℎ𝑙 , 𝑆 = 1 + 3000 𝐵𝑜0.86 , 𝐹 = 1.12 [

𝑁𝑢4

(𝐸ℎ1 + 𝑆ℎ𝑛𝑏 ),

𝐷ℎ 𝐺 (1 − 𝑥𝑒) 𝜇𝑙

𝐷ℎ

0.1−2𝐹𝑟

(0.12)

(− log10 𝑃𝑅 )−0.55 𝑀−0.5 𝑞′′0.67 𝐸 = 1 + 24000 𝐵𝑜1.16 + 1.37 (1⁄𝑋𝑡𝑡 )0.86 , 𝑆 = (1 + 1.15 ∗ 10−6 𝐸 2 𝑅𝑒𝑙1.17 )−1 ℎ𝑛𝑝 = 55 𝑃𝑅

𝑥𝑒

1−𝑥𝑒

0.1−2𝐹𝑟𝑙 {𝐹𝑟𝑙 }

For horizontal and 𝐹𝑟𝑙 < 0.05 , 𝑆2 =

and 𝐹2 =

0.75

]

𝜌

[ 𝑙]

0.41

𝜌𝑔

⁄ {𝐹𝑟𝑙1 2 }

𝑙 𝐸𝐹𝑟𝑙 and 𝑆 by 𝑆𝐹𝑟𝑙0.5 Where

𝐵𝑜 = [

𝑞"

𝐺ℎ𝑓𝑔

otherwise

], 𝐹𝑟𝑙 = [

𝐺2

𝜌𝑙2 𝑔𝐷ℎ

𝑆2 = 1 Kandlikar (1990)

ℎ1 = 0.023 𝑅𝑒𝑙0.8 𝑃𝑟𝑙0.4 1−𝑥𝑒 0.8

𝐶𝑜 = [

𝑥𝑒

]

𝜌𝑔 0.5

[ ] 𝜌𝑙

𝑘𝑙 𝐷ℎ

ℎ𝑡𝑝

,

ℎ𝑙

= 𝐶1 𝐶𝑜 𝐶2 (25𝐹𝑟𝑙 )𝐶5 + 𝐶3 𝐵𝑜𝐶4 𝐹𝑓𝑙

Saturated flow boiling for 𝑅134𝑎: 𝐹𝑓𝑙 = 1.63

, for Co < 0.65: C1= 1.136, C2= -0.9, C3= 667.2, C4 0.7, C5= 0.3 for Co > 0.65: C1= 0.6683, C2= -0.2, C3= 1058, C4 0.7, C5= 0.3

Tran et al. (1996) Kew and Cornwell (1997) Warrier et al. (2002) Mahmoud and Karayiannis (2013)

−0.4

𝜌𝑙 ℎ𝑡𝑝 = 840000 (𝐵𝑜2 𝑊𝑒𝑙 )0.3 ( ) 𝜌𝑔 𝑘𝑙 ℎ𝑡𝑝 = 30 𝑅𝑒𝑙0.857 𝐵𝑜 0.714 (1 − 𝑥𝑒)−0.143 𝐷ℎ ℎ𝑡𝑝 = 𝐸ℎ𝑙 ,

ℎ𝑙 = 𝑁𝑢4

𝑘𝑙 𝐷ℎ

,

𝐸 = 1 + 6𝐵𝑜 1⁄16 − 5.3 (1 − 855𝐵𝑜)𝑥𝑒 0.65

𝑁𝑢4 = 8.235 (1 − 2.042𝛽 + 3.767𝛽2 − 2.477𝛽3 + 5.361𝛽4 − 0.1865𝛽5 ) ℎ𝑡𝑝 = (𝐸ℎ𝑙 + 𝑆ℎ𝑛𝑏 ), ℎ𝑙 = 0.023 𝑅𝑒𝑙0.8 𝑃𝑟𝑙0.4 𝑆=

ℎ𝑙 = 4.36 𝑘𝑙 𝐷ℎ

𝑘𝑙 𝐷ℎ

𝑓𝑜𝑟 𝑅𝑒𝑙 < 2000

𝑓𝑜𝑟 𝑅𝑒𝑙 > 3000,

1 1+2.56∗10−6 (𝑅𝑒𝑙 𝐸 1.25 )1.17

296-13

𝐸 = [1 +

𝐴 0.64 𝑋𝑡𝑡

]

𝐷ℎ 𝐺 (1 − 𝑥𝑒) 𝜇𝑙 𝐴 = 2.812 𝐶𝑜−0.408 𝑅𝑒𝑙 =

,

]