The effect of heating area orientation on flow boiling performance in microchannels heat sink under subcooled condition

The effect of heating area orientation on flow boiling performance in microchannels heat sink under subcooled condition

International Journal of Heat and Mass Transfer 110 (2017) 276–293 Contents lists available at ScienceDirect International Journal of Heat and Mass ...

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International Journal of Heat and Mass Transfer 110 (2017) 276–293

Contents lists available at ScienceDirect

International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt

The effect of heating area orientation on flow boiling performance in microchannels heat sink under subcooled condition R. Ajith Krishnan, K.R. Balasubramanian ⇑, S. Suresh Department of Mechanical Engineering, National Institute of Technology, Tiruchirappalli, Tamil Nadu 620015, India

a r t i c l e

i n f o

Article history: Received 10 November 2016 Received in revised form 9 February 2017 Accepted 10 March 2017

Keywords: Subcooled flow boiling Microchannels Orientation Heat transfer coefficient Pressure drop

a b s t r a c t Subcooled flow boiling heat transfer experiments were conducted in this work in order to investigate the effect of the heating area orientation of microchannels heat sink on flow boiling heat transfer and pressure drop characteristics. The flow boiling heat transfer experiments were conducted in a 31 parallel ‘‘U” shaped microchannels (width 305 mm and depth 290 mm) heat sink with deionized water as the working fluid. The tests were conducted for different orientations such as, Horizontal upward facing (HU), Horizontal with heating area vertically aligned (HV), Vertical with up flow (VUF), Vertical with downflow (VDF) and Horizontal downward facing (HD) in a forced convection loop with volume flow rate of 50 ml/ min, 100 ml/min and 150 ml/min. From the experimental results, it was observed that the performances of the heat sink under all orientation conditions were found to be almost identical except the vertical downflow orientation (VDF). The critical heat flux values are found to be less in the case of vertical downflow orientation. In the case of vertical downflow orientation, the critical heat flux values corresponding to 50, 100 and 150 ml/min flow rate were 44.1 W/cm2, 74 W/cm2 and 99.3 W/cm2 respectively. For VDF orientation the buoyancy force acts on bubbles against the flow direction. Consequently, the bubbles were built up and merge each other due to the difficulty in draining and reversed flow was created with less heat flux input. The total pressure drop observed to be more for vertical downflow orientation compared to other orientations. Significant pressure fluctuations were observed during flow boiling in microchannels with VDF and HD orientations at low flow rates. The percentage reduction in effective heat flux value at the incipience of critical heat flux (CHF) in VDF orientation for flow rate of 50, 100 and 150 ml/min were 13%, 10.30% and 7.40%, and the corresponding percentage reduction in maximum outlet heat transfer coefficient was 30%, 23% and 19% respectively. Ó 2017 Elsevier Ltd. All rights reserved.

1. Introduction Heat flux from electronic devices is becoming a great challenge in the thermal management of electronic circuits. The heat removal is an important consideration in the design of many electronic systems such as microprocessors, IC chips, battery packs and inverters, laser diodes, avionic packages and several space systems. Flow boiling in mini/microchannels has been considered as the most promising cooling technique for a variety of high heat density applications. It was Tuckerman and Pease [1] demonstrated the importance of microchannels in cooling of electronic circuits. The first effective execution of microchannels in silicon devices was deemed by them. They were capable of removing the heat flux of 7.9 MW/m2. Numerous researchers had evaluated and reported the performance of heat sinks based on flow boiling in microchan⇑ Corresponding author. E-mail addresses: [email protected] (R. Ajith Krishnan), [email protected] (K.R. Balasubramanian), [email protected] (S. Suresh). http://dx.doi.org/10.1016/j.ijheatmasstransfer.2017.03.030 0017-9310/Ó 2017 Elsevier Ltd. All rights reserved.

nels. These include mainly the experimental observations characterising the steady state heat transfer and fluid flow behaviour, studies regarding numerical modelling and analysis of flow boiling, procedures for enhancing the critical heat flux and also the flow visualisation studies to understand the heat transfer mechanism of flow boiling in small channels. Even though a large number of studies were reported regarding the flow boiling heat transfer and heat transfer enhancement in the microchannels heat sink, only limited studies are published to identify the effect of the heating area orientation of heat sink on flow boiling performance. In the application point of view, the microchannels heat sink may be placed in any orientations with heating area positioned vertically, horizontally, downward facing and also inclined at a specific angle. The variation in the flow direction due to the difference in orientation of the heating area of the heat sink plays an important role in flow boiling condition due to the influence of gravity. The influence of gravity on two-phase fluid flow and heat transfer is mainly due to the drastic density variation among vapour and liquid phase. Even if it is assumed that in

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microchannels flow boiling the surface tension force and shear force govern the thermal and momentum transport, we cannot completely neglect the effect of gravitational forces, especially in low mass flow rates. The best part of the available data regarding the influence of the orientation on boiling was mainly for macro scale size channels. To the best of the author’s knowledge, only a few publications were carried out pertinent to the effect of heating area orientations on the performance of the microchannels heat sink [2–8]. Kandlikar et al. [2] investigated the influence of the orientation of the heat sink on flow boiling characteristics of water in a set of six parallel mini-channels each having a hydraulic diameter of 333 mm. The three different orientations such as horizontal, vertical downflow and vertical up flow experimented under matching operating conditions of heat and mass fluxes. The flow boiling performance of their mini-channel heat sink was found to be similar to the horizontal case where a gravity vector was absent. The authors conducted all experiments only at a single mass flux. The effect of mass flux on orientation effects was not reported. Moreover, the channel size used by them had a higher width and lower depth, which will provide sufficient room for the expanding bubble and reduces the pronounced effect of backflow. Miyata et al. [3] conducted flow boiling studies in a single copper tube of hydraulic diameter 1 mm using R410A refrigerant as the working fluid. The 320 mm length tube was tested vertically in both downward and upward flow directions. They observed an early changeover from the slug flow to the annular flow at the low mass flux situation in the vertical downward flow case. No noticeable difference in heat transfer between upward flow and downward flow was observed. The authors also reported that the pressure drop in the vertical downward flow direction was considerably larger than in the vertical upward flow direction. Their experiments were in the channel diameter range of 1 mm and with refrigerant, where the bubble departure diameter was very small. This small bubble departure diameter helps in the easy bubble removal and thus imparts fewer effects on the performance of heat sink under different orientation. Furthermore, the authors considered only upward and downward orientation. The effect of inclination of microchannels heats sinks on flow boiling using HFE-7100 as a coolant was studied by Wang et al. [4]. Multiple parallel rectangular microchannels with 0.825 mm hydraulic diameter were considered in the study and it was tested in orientations spans from vertically upward to vertically downward flow. The authors observed that for the vertically upward and horizontal flow the heat transfer coefficient was comparable at low mass velocities. The results also signify that the vertical downward orientations always depreciate the heat transfer performance. From their flow visualisation studies, they inferred that the vapour slug velocity was improved in the upward flow path due to buoyancy effect. In the first part of a two-part study related to microchannel evaporator for space applications, Lee et al. [5] addressed the effectiveness of two-phase micro-channels in negating the body force effects. For this, they conducted flow boiling experiments with FC-72 as working fluid with three different flow orientations such as horizontal, vertical up flow and vertical downflow. From the experimental studies conducted over broad ranges of mass velocity and heat flux, they summarised that the influence of the orientation on two-phase heat transfer was significant for low mass velocities. Beyond this, they demonstrate that by using a sufficiently high mass velocity, the flow boiling in microchannels was highly effectual in negating the influence of body force/buoyancy force in space system. The flow boiling heat transfer performance comparison of HFE7100 at a different inclination of the test surface was investigated experimentally by Hsu et al. [6]. The multiport micro-channel tested by them had a hydraulic diameter of 440 lm. Different incli-

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nations include 90, 45, 0, 45, and 90° with respect to the horizontal. Heat transfer coefficient for upward arrangements in their study is superior to that of downward arrangement which shows the effect of orientation on boiling performance in microchannels. Leao et al. [7] experimented the flow boiling of R245fa in 50 parallel rectangular microchannels heat sink with different orientations. The experimental trial with heating area horizontally positioned provides the higher overall heat transfer coefficient in their studies when compared to other orientations. The vertically positioned orientation with upward flow through the microchannels recorded the highest pressure drop. For most of the experimental conditions with different orientations bubbly and elongated bubble flow patterns was observed by the researchers. According to the author’s view, the vertically positioned upward flow orientation provides uniform flow pattern distribution among the channel. Moreover, the presence of reverse flow is rare for this orientation. Recently, Tamanna et al. [8] performed both heat transfer and flow visualisation studies in order to identify the effect of orientation of heat sink on flow boiling in silicon nanowire microchannels. They conducted experiments in a forced convection loop with deionized water as the working fluid. Flow boiling test was conducted in heat sink consisting of five parallel straight microchannels with silicon nanowire and heat sink without nanowire (plain channels). Only two orientations were chosen in their studies: upward facing (0° Orientation) and downward facing (180° Orientation). The authors reported that a strong influence of orientation was observed in boiling curves of plain wall microchannels heat sink, with a very meagre heat transfer performance showed by downward facing orientation. Meanwhile, the nanowire microchannel heat sink shows insensitivity to orientations at medium to high mass fluxes. For both nanowire and plain channel case, the pressure drop was found to be very little sensitive to orientation. As seen from the above-discussed literature, the presented results concerning the effect of orientation of the boiling surface on the performance of microchannels heat sink are not decisive. Besides, the previously published research papers were relevant only to vertical and horizontal orientations of the microchannels heat sinks, where the working fluid selection in most of the literature was either the refrigerant or the dielectric fluid. The bubble departure diameters in the case of the dielectric liquid and refrigerant flow boiling were much smaller when compared to water flow boiling [9]. The blockage caused by the bubbles departed from the surface in the case of water flow boiling was much intense when compared to the refrigerant and dielectric flow boiling. Hence it is very much needed for evaluating the flow boiling performance of microchannels heat sink with different orientations and flow rates with water as the working fluid, so as to study the effect of gravity on heat transfer, pressure drop and flow instabilities. Additionally, there was no organised information about the influence of mass flow rates and all possible orientations including horizontal upward facing, horizontal downward facing, vertical up flow, vertical downflow and horizontal with vertically aligned channels on the performance of microchannels heat sink. In this context, it is the main objective of our study to investigate the impact of heating area orientation on the heat transfer and total pressure drop of microchannels heat sink with different flow rates during subcooled flow boiling, with deionized water as the coolant.

2. Experimental methods and data reduction 2.1. Experimental facility The schematic diagram and the actual photograph of the experimental facility of the present study are shown in the Figs. 1 and 2.

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It’s composed of a deionized water reservoir, micro gear pump, preheating unit, microchannels test section and a condenser unit. The test fluid, deionized water was pumped into the loop by a micro gear pump (Micropump-IDEX Corp-GA-T23) united with a voltage controller (HOLMARC) in order to provide the required liquid flow rate. The water in the reservoir was always maintained at atmospheric temperature. For different subcooled inlet temperature studies at the entry of the test section, the temperature of the water is controlled using a 1000 W Nichrome wire heater in the preheating unit which is connected to a 0–240 V AC autotransformer power supply. For the present study, the water inlet to the test section is maintained at 35 °C. 2.2. Test section The main component of the experimental facility is the channel test section. The Constructional details of the channel test section are shown in Fig. 3. The test section consists mainly of a copper testing block, a thermally insulating PEEK casing, a polycarbonate glass top plate, a heating block made of copper inside of which four-cartridge heaters each having a capacity of 400 W were inserted, and an insulating box made of high-quality glass wool packing. Photograph of the experimental test section is shown in Fig. 4. Detailed physical properties of the materials used in the current investigation is shown in Table.1. On the top of the copper testing block, 31 parallel ‘‘U” shaped microchannels of width(w) 305 mm, depth(d) 290 mm and length (l) 30 mm was machined using CNC wire cut EDM (CONCORD DK7725). The surface roughness was measured with a surface roughness tester, Mitutoyo SURFTEST SJ-410. The roughness was measured at 3 positions on the channel and the following average results were found: Ra = 1.730 mm, Rq = 2.177 mm and Rz = 10.535 mm. Due to the difficulty in measuring the roughness of the 305 mm width wire EDM machined microchannels, similar

dummy channels with more width was machined on the separate copper block with similar cutting parameters and the roughness was measured. The details of the copper testing block with machined microchannels are shown in Fig. 5. The copper test block was inserted into the PEEK housing and ensures a tight fit. During assembly, the top surface is levelled and covered by a transparent Polycarbonate (Lexan) glass sheet. A Nitrile rubber cord was inserted in between the PEEK housing and the polycarbonate sheet to make the test section leak proof. Inlet and outlet plenum to the parallel microchannels along with inlet and exit pressure tapping ports were machined in the PEEK housing. The deionized water was supplied to the inlet plenum of the heat sink test section (as shown in Fig. 3) and drained from the outlet plenum of the heat sink by drilling a 6 mm diameter ports on the top polycarbonate glass sheet. Two pressure transmitters (Delta Ohm, HD9408T 0–250 millibar) were connected to the inlet and exit pressure tapping ports in order to determine the pressure at both inlet and outlet of the microchannels heat sink. The copper testing block is heated using four-high watt density cartridge heaters (10 mm diameter, 50 mm length and 400 W) inserted into the heating copper block as shown in Fig. 3. A thick layer of silica aerogel insulation along with an insulating wooden box is provided around the heating copper block to avoid the heat loss to the surroundings. The top row and bottom row cartridge heaters are connected to separate autotransformers (0–240 V) in order to control the heating rate. The microchannels bottom surface temperature was measured by three K-type thermocouples (1 mm diameter probe) placed in holes drilled 3 mm below the copper testing block top surface as shown in Fig. 5. Another three thermocouples were placed at a distance of 13 mm from the copper testing block top surface in order to monitor the overheating of the block. Inlet and outlet plenum temperature were also measured by using two thermocouples inserted into the cavities. The electrical signals from the thermo-

Fig. 1. Schematic diagram of the experimental facility.

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Fig. 2. Photograph of flow boiling experimental facility.

Fig. 3. Constructional details of the channel test section.

couples and pressure transmitters were acquired, processed and stored by using a Keysight 34972A data acquisition system with 34901A control unit module and a computer. The performance of the microchannels test section was evaluated with different orientations of the footprint area as shown in Fig. 6. The direction of gravity vector is also shown. In Fig. 6(a) the footprint surface and microchannels are horizontally aligned with channels facing upward and flowing of working fluid through

the top of the channels, hence named hereafter as ‘horizontal upward facing’ (HU). In Fig. 6(b), the heating surface area of the microchannels is vertically placed and the parallel microchannel arrays are horizontally aligned, hence named as horizontal with a heating area vertically aligned (HV). In Fig. 6(c), the heating area is oriented inverse to the case of horizontal upward facing hence named as ‘horizontal downward facing’ (HD). The parallel microchannels are oriented vertically with vertical up flow and

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Fig. 4. Photograph of experimental test section.

Table 1 Physical properties of materials used in the current investigation. Material

Physical properties

PEEK (polyether ether ketone)

Density: 1320 kg/m3 Thermal conductivity: 0.25 W/m K Melting point: 343 °C Glass transition temperature: 143 °C

Polycarbonate glass sheet

Density: 1200–1220 kg/m3 Thermal conductivity: 0.19–0.22 W/m K Melting point: 225 °C Glass transition temperature: 147 °C

Silica aerogel insulation

Density: 3–350 kg/m3 Thermal conductivity: <0.1 W/m K

Copper

Density: 8960 kg/m3 Thermal conductivity: 401 W/m K Melting point: 1084 °C

vertical downflow in Fig. 6(d) and (e). These two orientations are named as ‘vertical with up flow’ (VUF) and ‘vertical with downflow’ (VDF) respectively. 2.3. Experimental procedure Before starting the experiment, the deionized water was strenuously boiled for 20 min in order to remove the dissolved gases present and hence to obtain the accurate subcooled flow boiling data. A micro gear pump with a voltage controller was used to pump the water into the test section. The water inlet temperature at the test section was maintained at a constant temperature of 35 °C using the water pre-heater unit. The Instantaneous change of the water inlet temperature to test section for volume flow rate of 100 ml/min and HU orientation is shown in Fig. 7. The inlet temperature was kept at 35 °C during the start of the each experimental trial. The fluid after carrying the heat from surface passed through the chiller for maintaining the inlet temperature at 35 °C. From the time history of inlet temperature, it is clear that the inlet temperature varies within 35.2–36 °C. It is very difficult to maintain the inlet temperature exactly at a particular temperature since the fluid is recirculated in the experimental loop. The heat was supplied to the copper testing block in small increments

by cartridge heaters inserted in the heating copper block. After each increment in heat flux, the system was allowed to reach the steady state and the readings were recorded. The experiment was then repeated for different values of heat flux up to the incipience of the critical heat flux. Experimental results for the present study were based on the experimental conditions described in Table 2. The effective heat flux based on the footprint area is given by [9] as,

q00fp ¼

Q af Afp

ð1Þ

q00fp = Effective heat flux based on the heat sink footprint area, W/ cm2. Q af = Actual heat input to the test section after reducing the losses to the environment, W. Afp = Footprint area of the microchannel heat sink, cm2. The actual heat input to the test section is given by,

Q af ¼ Q thi  Q loss

ð2Þ

Q thi is calculated as the product of electrical current (I) and voltage (V) supplied to the cartridge heaters. The procedure to estimate the Qloss described in [10] was employed in the present study to estimate the heat loss for different orientations of the test section. The same procedure was also reported in flow boiling works done by Lee and Garimella [11], Bao et al. [12], Alam et al. [13], and Bogojevic et al.[14]. In order to find out the fraction of the total heat input that is dissipated via the flow boiling process, heat losses via other paths, i.e., natural convection, radiation, and conduction via the PEEK casing, an energy balance can be written as

Q thi ¼ Q af þ Q loss

ð3Þ

The actual heat input to the test section is dissipated by the working fluid during its heating. When the test section is completely drained of working fluid, then the Q af term can be eliminated from Eq. (3). The test section is then heated by supplying a small constant voltage to the cartridge heaters. The new energy balance equation now becomes as in Eq. (4)

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Fig. 5. Copper test block with machined microchannels.

Q thi ¼ Q loss ¼ htotal Aloss

ðT s;av g  T amb Þ

ð4Þ

T s;av g = Average surface temperature, °C. T amb = Ambient temperature, °C. Since the total heat transfer coefficient, htotal and the total heat loss area, Aloss is constant. The heat loss can be directly related to ðT s;av g  T amb; Þ as in Eq. (5)

Q loss

a ðT s;av g  T amb Þ

ð5Þ

Hence based on Eq. (5), for different values of T s;av g  T amb and supplied heat input linear relations can be made and heat loss at higher temperature can be calculated from those relations. The detailed procedure is provided in the next paragraph. Steady state temperature readings of the three thermocouples connected in the copper substrate and two thermocouples measuring the ambient temperature are recorded over a two-minute period and averaged. Heat loss tests are repeated for different levels of input power ranging from 1 W to 8 W and the average thermocouples temperature difference ðT s;av g  T amb Þ is then correlated to the input power supplied during the heat loss test. This is a linear relation which is used to find out the heat loss during flow boiling test. The linear relations to find out the heat loss during flow boiling under different orientations are shown in Fig. 8. The heat loss characteristic curves were almost same for all orientations indicating that the change in the orientation does not affect the heat loss, gratitude goes to the silica aero-gel insulation.

Based on the conditions at the channel outlet, the data analysis in this paper was done. Hence, the local heat transfer coefficient at the channel outlet was calculated as by [15] as,

ho ¼

q00fp T s;out  T f ;out

ð6Þ

T s;out = Outlet local wall temperature, °C. T f ;out = Fluid temperature at outlet, °C. The outlet local wall temperature T s;out was extrapolated from the thermocouple reading by assuming one-dimensional thermal conduction. The pressure ports are positioned at inlet and outlet plenum of the heat sink to measure the pressure upstream and downstream of the microchannels heat sink. The total pressure drop recorded during experimental trials represents the combined losses in the test section due to the frictional loss in the microchannels and minor losses across the bends. The sudden contraction from the inlet manifold to microchannels and expansion from microchannels to the outlet manifold also contribute to the minor losses. The measured pressure drop was corrected for the minor losses in order to estimate the experimental pressure drop along the channel based on the methods described in [16]. A total added pressure loss of about 6% of the total pressure drop was identified, and since the losses are small compared to total pressure drop it is ignored in our study. Hence, the pressure drop data reported in our experiments are directly obtained from the pressure sensors.

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Fig. 6. Schematic and 3D modelled images of different orientations of the microchannel heat sink.

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2.4. Experimental uncertainties Uncertainties related to the various sensors used in this flow boiling experimental study are listed in Table 3. Experimental uncertainty in the flow rate, heat flux, and pressure drop and heat transfer coefficient are calculated as by Kline and McClintock [17]. The maximum experimental uncertainty in the flow rate, effective heat flux, heat transfer coefficient and pressure drop calculations, data are 9.6%, 2%, 9.3% and 7.8%.

3. Experimental results and discussion 3.1. Validation and repeatability test

Fig. 7. Inlet water temperature history for 100 ml/min volume flow rate.

Table 2 Experimental conditions evaluated in this work. Orientation

Volume flow rate (ml/min)

Temperature inlet to heat sink (°C)

Horizontal Upward Facing (HU) Horizontal with Heating Area Vertically Aligned (HV) Horizontal Downward Facing (HD) Vertical with Up Flow (VUF) Vertical with Down flow (VDF)

50, 100, 150 50, 100, 150

35 35

50, 100, 150

35

50, 100, 150 50, 100, 150

35 35

Single-phase heat transfer tests were carried out to validate the experimental setup. The experimental outlet heat transfer coefficient and numerical results obtained through computations using commercial software package FLUENT were compared. Trial with 100 ml/min volume flow rate, heat flux range of 4–30 W/cm2 and inlet temperature (Tin) of 32 °C was used for both simulation and experiments. Obtained result is shown in Fig. 9. A fair agreement between the experimental results and the computed results by software package FLUENT is observed. Repeatability test was carried out for horizontal upward (HU) facing orientation of the heat sink at a flow rate of 100 ml/min, the inlet temperature of deionized water to test section was maintained at 35 °C. The repetitive boiling curves and total pressure drop curves for the trials are more or less overlapped as shown in Fig. 10. Minimum deviations of 0.3 °C and 0.2 millibar and maximum deviations of 3.1 °C and 2.12 millibar are observed for the T s;av g  T in and total pressure drop respectively. 3.2. Boiling curves The effects of footprint area orientation on subcooled flow boiling curve for the microchannels heat sink at volume flow rates of 50, 100 and 150 ml/min with inlet fluid temperature of 35 °C are plotted in Figs. 11–13. Five different orientations: HU, HD, HV, VUF and VDF were used to perform the flow boiling test. In our present study, the experiment was repeated for different values of heat flux till the CHF (critical heat flux) was encountered. Occurrence of CHF can be judged by the following phenomena,

Fig. 8. Heat loss characterization curve for different orientations.

Table 3 Uncertainty of various sensors used in this flow boiling experimental study. Variable

Uncertainty

Footprint area width and length Flow rate Temperature measurements Voltage Current Inlet and outlet pressure

±0.02 mm ±4.8 ml/min ±0.5 °C ±0.1 V ±0.01 A ±2.5 mbar

Fig. 9. Experimental and numerical modelling results for outlet heat transfer coefficient (100 ml/min HU trial).

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Fig. 10. Boiling curve (a) and total pressure drop (b); repeatability curve for 100 ml/min.

Fig. 12. Flow boiling curve at different orientations for 100 ml/min flow rate. Fig. 11. Flow boiling curve at different orientations for 50 ml/min flow rate.

(a) Backflow of vapour towards the inlet section, (b) Sudden increase in surface temperature due to dry out of the surface, (c) Fluctuation of inlet temperature due the back flow of vapour, (d) The occurrence of periodic dry out and rewetting of the channel. Triggering of any one of the above phenomena can be considered as the starting point of dry out or CHF. The entire power supply to the heater was reduced at this time to prevent overheating of the test cell, so as to avoid the damage. Post CHF data were not taken at any trials experimented in the present work. CHF occurrence in channels is shown in Fig. 14. In Fig. 14a nucleation at the wall of the channel is shown. In the next Fig. 14b, dry out in the left channels occurred. The white colour patches in the figure show the vapour streams. Complete dry out in the channels is shown in Fig. 14c. In this case, the entire channels are covered with vapour streams (seen as white colour in channels).The rewetting of the channels after dry out is shown in Fig. 14d. The movie S1 shows the CHF occurrence in the microchannels. In all orientations, the

occurrence of periodic dry out and rewetting of the channel was observed. A single-phase region, which corresponds to the linear portion of the boiling curve was noticed before the ONB (Onset of nucleate boiling). Depending on the typical behaviour of the boiling curve during pool or flow boiling, after ONB the above-mentioned linear behaviour should vanish and the characteristic curvature of the boiling curve must come along. Visualisation of ONB for volume flow rate of 50 ml/min trial with HU orientation at 38 W/cm2 effective heat flux is shown in Fig. 15. From the figure, it is clear that the bubble nucleation is started at the channel walls, which represent the starting of the nucleate boiling process. 3.2.1. Effect of mass flux Boiling curves for the flow rates 50, 100 and 150 ml/min with horizontal upward facing orientation (HU) of the heat sink is shown in Fig. 16. The effects of varying mass flux can be clearly observed. It is narrated from the boiling curves that the higher mass flow rate, supports a higher heat flux at a given wall superheats. For a given microchannel heat sink, the critical heat flux

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35 °C) the wall superheat for ONB observed as 0.2 °C. That is the first bubble on the microchannel wall appears at a wall temperature of 100.2 °C. This can happen in much higher subcooled conditions with very low volume flow rate since the fluid got enough time to reach the saturation temperature while passing through channel surfaces. In the subcooled flow boiling experimental results provided by Sarwar et al. [18], it is clear from the boiling curve that at low mass flux cases with high inlet subcooling (75 °C) the ONB started at near 100 °C wall temperature. Similarly, subcooled flow boiling studies presented by Sun et al. [15] also reported that with subcooling temperature increased from 30 to 70 K (inlet water temperature 30 °C), the corresponding increment in the incipient superheat was found less than 0.5 K.

Fig. 13. Flow boiling curve at different orientations for 150 ml/min flow rate.

increased with the mass flux. Moreover, the average surface temperature (Ts,avg) required for ONB changes with mass flux. In the case of HU orientation, the Ts, avg at which ONB begins are 100.2 °C, 102.3 °C and 104.8 °C, for flow rates of 50, 100 and 150 ml/min respectively. Increasing coolant flow rate increase the convective heat transfer and at the same time increase the wall superheats needed to initiate the bubble nuclei, consequently postponing the ONB. In our experiments with a very small volume flow rate of 50 ml/min with inlet subcooling of 65 °C (i.e. inlet temperature of

3.2.2. Effect of orientation Based on the flow boiling curves shown in Figs. 11–13, for five different orientations at different flow rates, it is clear that the effect of heating area orientation on flow boiling is negligible except for the vertical downflow (VDF) orientation. In the case of the VDF orientation, the commencement of CHF occurs earlier than other orientations. The opposing buoyancy force as shown in Fig. 17 cause vigorous interactions between the flowing liquid and the elongating vapour phases causes severe flow reversal and premature CHF. This similar case was observed in works reported by Kandlikar et al. [2] and Wang et al. [4] with vertical downflow orientation. In the case of flow boiling in microchannels, the bubbles nucleate near the edges of the channel wall as shown in Fig. 18. The bubble then grows to fill the entire channel width leading to a slug flow. Once the bubble filled the entire channel width, it starts to expand by pushing the liquid in both upstream and downstream direction. This expansion of the bubble leads to the formation of vapour slug, and the vapour slug can even grow to fill the entire

Fig. 14. CHF occurrence in microchannels for 100 ml/min, VUF orientation, 82.6 W/cm2.

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Fig. 15. Visualisation of ONB for volume flow rate of 50 ml/min, HU orientation, 38 W/cm2.

Fig. 16. Flow boiling curve for HU orientation with flow rates of 50 ml/min, 100 ml/ min and 150 ml/min.

channel length causing dry out. In the case of HU, HD, VUF and HV orientations the individual bubbles nucleated from nucleation sites grew into vapour slugs individually. Merging of bubbles to form the vapour slug was rarely observed in these orientations. But the condition seems to be different in the case of VDF orientation, as shown in Fig. 19. Rapid evaporation coupled with the frequent merging of nucleated bubbles and vapour slugs leads to frequent churn flow in the case of VDF orientation (Fig. 19b and c). Even though backflow was seen in every orientation trials, it is much pronounced in the VDF orientation. The rapid evaporation at the surface aided by gravity causes the vapour slug to elongate and flow back to the upstream section. The vapour backflow prolongs all the way into the inlet plenum for all the trials of VDF orientation explored in this work, causing flow maldistribution in the channels. In the case of VUF orientation merging of bubbles and slugs was rarely observed. The bubble itself during its path convert to slug, elongated slug and fill the entire channel with the annular flow as shown in Fig. 20. The movie S2 provided shows the flow boiling in VDF orientation in microchannels and movie S3 shows the flow boiling in HV orientation.   The Effective heat flux q00fp value for different orientations under different flow rates at the incipience of CHF is shown in

Fig. 17. Illustration shows the buoyancy force in bubbles during VUF and VDF orientation.

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Fig. 18. Bubble, slug and annular flow pattern in microchannel.

Fig. 19. Flow boiling images for VDF orientation, 50 ml/min, 43 W/cm2.

Table 4. For the VDF orientation the percentage reduction in effective heat flux value at the incipience of CHF for volume flow rates of 50, 100 and 150 ml/min are 13%, 10.30% and 7.40%. The effective heat flux value at incipience of CHF of Horizontal upward orientation at different flow rates are taken as the reference for calculating the percentage reduction in effective heat flux at incipience of CHF. The premature CHF in VDF orientation is due to the gravity aided severe back flow and pronounced reversed vapour slug flow into the inlet plenum as discussed earlier. Occurrence of periodic dry out and rewetting of the channel was intense in case of VDF at the incipience of CHF when compared to other orientations. Even though it is considered that only surface force is significant in microchannels flow boiling, the orientation of channel with respect to gravity cannot be completely neglected. In the case of small hydraulic diameter channels, the effect of gravitational forces normal to the flow direction is not expected to be significant. But the orientation of channel length with respect to gravity vector matters. During the VUF orientation, the flow inertia is coexisting with the buoyancy force and thereby enhances the easy increase

of bubble velocity along the channel length [14]. The inverse of this, the flow inertia is in opposition to the buoyant force direction for VDF orientation and for this basis, the size of the vapour bubble elongated upstream and the heat transfer is deteriorated due to interactions between the flowing liquid and the vapour bubbles. The strong influence of horizontal downward facing orientation on boiling curve reported by Tamanna et al. [8] is not observed in our experiments. They observed a reduced heat transfer performance in the case of plain wall downward facing orientation. CHF and dry out occurs at a very low heat flux situation for plain wall downward facing microchannel, due to vapour stagnation, flooding at the upstream end and blockage of liquid renewal. In our experiments, the performance of HU and HD orientation seems to be almost similar. This proves that in the case of microchannels the variation of gravitational forces normal to flow direction does not affect the two-phase heat transfer performance. Also due to the higher surface area ratio, the influence of the gravitational force is expected to be small in microchannels. In the case of conventional channels (hydraulic diameter greater than or equal to

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Fig. 20. Flow boiling images for VUF orientation, 50 ml/min, 46 W/cm2.

Table 4 Effective heat flux value at the incipience of CHF for different orientations tested with different flow rates. Flow rate (ml/min)

50 100 150

  Effective heat flux q00fp value at incipience of CHF (W/cm2) HU

HD

HV

VUF

VDF

50.7 82.56 107.23

50.39 82.8 107.21

50.57 83.1 107.05

50.23 82.65 106.32

44.10 74.0 99.29

3 mm) for the HD orientation and with small volume flow rates, the buoyancy causes the vapour to stratify above the liquid and adjacent to the heated wall. This stratification regime blocks liquid access to the heated wall and resulting in minor cooling of the heated wall from liquid ligaments, as shown in Fig. 21. The increasing of volume flow rate considerably offset the influence of buoyancy force on VDF orientation. The percentage reduction in effective heat flux at the incipience of CHF reduced with an increase in volume flow rate for vertical downflow orientation. Also at higher flow rates, a smaller difference in heat transfer coefficients was observed among various inclinations. As the flow rate increases the drag force become more rampant causing the generated bubble to take apart from the heated surface in advance and

move parallel to or slide along the heated surface. Because of this the bubble departure diameter decreases and bubble depart early from the nucleation cavity with the increase of the coolant flow rate. Thus, providing less chance for merging and occurrence of reversed vapour flow to inlet plenum. 3.3. Outlet heat transfer coefficient Local heat transfer coefficient at the channel outlet calculated as per Sun et al. [15] and Sujith et al. [19] for different orientations under different flow rates of 50, 100 and 150 ml/min is shown in the Figs. 22–24. The outlet heat transfer coefficient is plotted against the effective heat flux. It is clear from the figure that the

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Fig. 21. Horizontal downward facing (HD) orientation flow boiling in conventional and narrow channels.

Fig. 22. Outlet heat transfer coefficient for different orientations at a flow rate of 50 ml/min.

Fig. 24. Outlet heat transfer coefficient for different orientations at a flow rate of 150 ml/min.

Fig. 23. Outlet heat transfer coefficient for different orientations at a flow rate of 100 ml/min.

heat transfer coefficient increases with increasing heat flux in the boiling regime. While comparing the results for different flow rates, the heat transfer coefficient is found to increase with the increase in the coolant flow rate. This is due to the increase in the heat carrying capacity of the working fluid with increases in the mass flux [19].

Fig. 25. Total pressure drop variation with effective heat flux during flow boiling with a flow rate of 50 ml/min.

In considering the effect of the heating surface orientation of the microchannel heat sink on heat transfer coefficient, the vertical downflow (VDF) orientation only shows some variation in the gen-

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eral trend. The backflow in the vertical downflow case was more rigorous, with the evidently reversed flow of vapour into the inlet manifold. The consequential increase in flow maldistribution plays the key role in the heat transfer degradation in the vertical downflow case. As cited in the literature, Kandlikar and Balasubramanian [2] reported that the boiling performance and the heat transfer coefficients for horizontal and upward vertical flow orientations were worth comparing. When juxtaposing with vertical upward flow orientation the heat transfer coefficient for the vertical downward orientation in their experiments was impaired by 30–40%. For the present experimental study, the effect of heating surface orientation on the heat transfer coefficient is insignificant for HU, HV, HD and VUF cases, but in case of VDF orientation the maximum outlet heat transfer coefficient is degraded by 30, 23 and 19% for flow rates of 50, 100 and 150 ml/min. The gradual reduction in percentage degradation of outlet heat transfer coefficient with an increase in flow rate reveals the fact of diminishing the influence of gravity on boiling with an increase in coolant flow rate.

Fig. 27. Total pressure drop variation with effective heat flux during flow boiling with a flow rate of 150 ml/min.

3.4. Pressure drop The effect of footprint area orientation and heat flux on the averaged total flow boiling pressure drop at different flow rates in the microchannels heat sink are shown in Figs. 25–27. The average pressure drop is obtained based on the arithmetic averages of the transient results of total pressure drop after reaching a stable flow boiling state. It was clear from the previously discussed Fig. 10 that, during single phase flow the pressure drop decreases with the increase in heat flux due to the decrease of viscosity of the fluid with an increase in temperature. The total pressure drop across the heat sink just before the starting of ONB (Onset of Nucleate Boiling) and till the observation of incipience of CHF is shown in Figs. 25–27. After the ONB (Onset of Nucleate Boiling) the pressure drop increases and there was a fluctuating nature in pressure drop until CHF (Critical Heat Flux). Rigorous flow oscillations were observed during incipience of CHF with large pressure fluctuations. The total pressure drop likewise found to increase with an increase in coolant flow rate. This nature is as of now expected and is general because of the reality that, with increasing the two-phase flow velocity the frictional pressure also increases. Another important observation is that, the effect of orientation on the total pressure drop. The VUF orientation had the lowest

Fig. 28. Transient outlet pressure profile of the microchannels heat sink during flow boiling at a flow rate of 50 ml/min.

Fig. 29. Transient microchannels heat sink surface temperature at outlet during flow boiling at a flow rate of 50 ml/min.

Fig. 26. Total pressure drop variation with effective heat flux during flow boiling with a flow rate of 100 ml/min.

total pressure drop and it is due to the fact that, the buoyancy force assisted the bubble movement which in time favoured the easy fluid flow in the microchannel heat sink. Meanwhile, the VDF orientation recorded the maximum total pressure drop during boiling. The opposing buoyancy force causes the bubble to elongate towards the inlet side which results in increased pressure drop

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Fig. 30. Detailed transient outlet pressure profile plot during flow boiling under 50 ml/min.

and severe backflow. The pressure drop curve shows very modest sensitivity to HU, HD and HV orientation for the tested experimental trials. Even though the HD orientation shows a slightly higher pressure drop at low flow rate trials, (50 ml/min and 100 ml/ min) and this may be due to the bubble stagnation and less prominent vapour acceleration effect in the downward facing boiling orientation. The transient outlet pressure profile of the microchannels heats sink during flow boiling at a flow rate of 50 ml/min and an effective heat flux of 47.8 W/cm2 for HU, HD, HV, and VUF orientation and 42.2 W/cm2 for VDF orientation is shown in Fig. 28. The corresponding transient response of the surface temperature near at the outlet of the microchannels heat-sink is shown in Fig. 29. It takes around 250–300 s for the test cell to reach the steady state. From the figure, it is clear that the outlet pressure oscillations in the VDF orientation were intense compared to other orientations. The outlet pressure continuously fluctuated in the range of 11–27 millibar in the case of VDF orientation; therefore, the flow was considered more chaotic in that orientation. Detailed transient outlet pressure profile plot for each orientation in the boiling regime is shown in Fig. 30. The oscillating nature of the outlet pressure is visible for VDF, HV and HD orientations, while the intensity is higher for VDF orientation compared to others. The VUF and HU orientation show only a marginal fluctuation in outlet pressure, and the flow was considered less chaotic in these orientations. The visible outlet pressure fluctuations in case of HD orientation are mainly due to vapour stagnation and blockage of liquid renewal. In the case of HV orientation, since the heating surface oriented vertically and the channels oriented horizontally, there is a chance for severe non-uniformity of the two-phase flow distribution among the microchannels channels. This non-uniformity in flow causes pressure fluctuations in HV orientation at low flow rates. The transient outlet pressure profile of the microchannels heat sink during flow boiling at a flow rate of 150 ml/min and an effec-

tive heat flux of 100.8 W/cm2 for HU, HD, HV, and VUF orientation and 97.5 W/cm2 for VDF orientation are shown in Fig. 31. The pressure fluctuations for VDF orientation is found to be less intense when compared to the pressure fluctuations in the case of 50 ml/ min flow rate. This inference highly promotes the conclusion of the negation of orientation effects in microchannels flow boiling at higher flow rates. From the detailed transient outlet pressure profile for flow boiling under 150 ml/min flow rate as shown in Fig. 32, it is clear that the fluctuations are looking similar for VDF, HV, HD and HU orientations. Moreover, the VUF orientation has fewer fluctuations when compared to others. The higher flow rate increases the drag force which accelerates the bubble departure and further flushes the bubble to the outlet plenum, hence the pressure fluctuations are found to be less at higher flow rates and the orientation effect can be completely neglected.

Fig. 31. The transient outlet pressure profile of the microchannels heat sink during flow boiling at a flow rate of 150 ml/min.

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Fig. 32. Detailed transient outlet pressure profile plot during flow boiling under 150 ml/min.

4. Conclusions The effect of heating area orientation on heat transfer and pressure drop during subcooled flow boiling of water in microchannels heat sink was experimentally investigated in this study. Five different orientations such as Horizontal upward facing (HU), Horizontal with heating area vertically aligned (HV), Vertical with up flow (VUF), Vertical with downflow (VDF) and Horizontal downward facing (HD) were investigated under different flow rates of 50, 100 and 150 ml/min. From the experimental study, the following key conclusions can be drawn:  The experimental results showed that the effect of orientation, especially the VDF orientation has an important role in the performance of microchannel heat sink at low flow rates. The VDF orientation shows degraded heat transfer performance, higher total pressure drop and severe outlet pressure oscillations during flow boiling at the flow rate of 50 ml/min.  A percentage reduction of 13% in CHF and 30% reduction in the outlet heat transfer coefficient were noted in the case of VDF orientation during boiling at low flow rates.  Higher flow rate causes a significant reduction in the influence of orientation on CHF and outlet pressure oscillations. This conduct can be explained by the way that, the higher flow rate increases the magnitude of drag and shear forces compared to buoyancy.  In the case of HV orientation pressure fluctuations are observed during the incipience of CHF due to the non-uniform two-phase flow distribution within the microchannels heat sink. The pressure oscillations in HD orientation during boiling at low flow rates are mainly due to vapour stagnation and liquid blockage on channel surface.

 The VUF orientation is found to be the best choice for flow boiling in microchannels due to the reduced pressure drop, higher CHF and fewer pressure fluctuations. The buoyancy assisted bubble movement helps to clear the channel before getting it clogged with vapour bubbles.

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