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Flow boiling in microchannel with synthetic jet in cross-flow
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Flow boiling in microchannel with synthetic jet in cross-flow Ehsan Sourtiji∗, Yoav Peles Department of Mechanical and Aerospace Engineering, University of Central Florida, Orlando, FL 32828, USA
a r t i c l e
i n f o
Article history: Received 3 July 2019 Revised 23 October 2019 Accepted 8 November 2019 Available online xxx Keywords: Micro synthetic jet Microchannel Phase change Flow boiling Heat transfer Active enhancement
a b s t r a c t Flow boiling heat transfer in a microchannel with a micro synthetic jet powered by liquid/vapor phase change was studied. The synthetic jet had no moving parts and used the motion of the interfacial layer between vapor and liquid to propel liquid. Bubbles were generated on a micro heater in a chamber, 3.5 mm in radius and 220 μm tall, which was connected to the microchannel through a 300 μm nozzle opening. Periodic growth and collapse of bubbles in the chamber powered the synthetic jet. Using highspeed photography and a microscope, the sequence of bubble growth and collapse in the chamber, bubble nucleation in the microchannel, and their interaction with the synthetic jet were captured and analyzed. The jet velocity was estimated using image processing and its momentum coefficient was used to characterize the strength of the synthetic jet. The results showed that the synthetic jet enhanced nucleate flow boiling heat transfer in the microchannel by up to 20%. dry-out spots formed over the heated surface for high heat fluxes were mitigated by the synthetic jet as it assisted rewetting the surface and maintaining the integrity of the thin liquid film adjacent to the heated surface. © 2019 Elsevier Ltd. All rights reserved.
1. Introduction The demand for cooling high heat flux surfaces has been steadily growing for a range of applications, such as cooling of electronic devices. To address this issue, flow boiling heat transfer in diminishing length scales has been extensively studied due to its high heat transfer coefficient [1–5], high latent heat of vaporization, and its potential to enable direct chip-level cooling [6]. For instance, Kiyofumi et al. [7] observed that the heat transfer coefficient of flow boiling in microchannels is 3–20 times higher than that of a single-phase flow. In the last decade, a range of passive and active enhancement approaches for flow boiling heat transfer in diminishing length scales were studied, such as nano engineered surfaces [8–10], pin fins and microstructures [11–13], liquid phase separation [14], intricate microchannel geometries [15,16]. A unique passive heat transfer enhancement approach using auxiliary channels and multiple micronozzles embedded into a microchannel domain was extensively studied by Li et al. [17– 22]. Considerable flow boiling heat transfer enhancement was observed, which also suppressed two-phase flow instabilities. The enhancement was linked to a disruption of the boundary layer growth, flow mixing, and effective management of bubble confine-
∗
Corresponding author. E-mail addresses: [email protected] (E. Sourtiji), [email protected] (Y. Peles).
ments in the microchannel formed by a secondary flow of microjets emanated from the auxiliary channels. It was also deduced that the collapse of confined bubbles, due to direct condensation, enhanced the mixing and the associated heat transfer process. Active methods have also been studied. For instance, a transverse jet in crossflow with flow boiling in microchannel was studied by Vutha et al. [23] who observed considerable reduction in thermal oscillations amplitudes associated with the two-phase flow. This was attributed to the replenishment of liquid in dry-out regions, which subsequently suppressed large temperature raises. Xu et al. [24] used seed bubbles to improve heat transfer efficiency in a microchannel. The seed bubbles were generated using microheaters upstream in the microchannel and driven by pulse voltage signal. It was shown that introducing the seed bubbles, stabilize flow and heat transfer in micro-boiling, suppress the flow instability and decrease heating surface temperature. Fang and Khan [25] and Khan et al. [26] observed that a synthetic jet mitigated pressure and temperature fluctuations during flow boiling in microchannel. It was found that backflow due to bubble clogging and expansion were notably delayed by the synthetic jet, and that the thermal boundary layer was disturbed, and as a result, the heat transfer was enhanced. It was also inferred that the synthetic jet effectively broke elongated bubbles. Fabbri et al. [27] carried out an experimental study with microjet heat transfer and observed an increase in the heat transfer coefficient. Brun et al. [28] investigated the influence of confined jet array impingement on boil-
https://doi.org/10.1016/j.ijheatmasstransfer.2019.119023 0017-9310/© 2019 Elsevier Ltd. All rights reserved.
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2. Experimental apparatus and method Nomenclature 2.1. Micro-device Ah AN AC Cμ Cμ, Avg Dh Hc h I m˙ P qeff Q˙ loss R t t¯
τ
Ts Tsat V vavg vj, avg V¯b V¯ jet wc wN
surface area of the heater cross section area of the nozzle throat cross section area of the microchannel momentum coefficient average momentum coefficient hydraulic diameter microchannel height two-phase heat transfer coefficient electrical current mass flow rate applied electrical power effective heat flux heat loss rate electrical resistance time dimensionless time oscillation period surface temperature saturation temperature electrical voltage spatial average velocity in the microchannel crosssection area spatial average velocity in the nozzle throat crosssection area dimensionless bubble volume dimensionless jet velocity width of the microchannel opening width of the nozzle throat
Abbreviation HTC heat transfer coefficient nm nanometer ONB onset of nucleate boiling RTD resistance temperature detector μm micrometer via vertical interconnect access ∞ refers to the quantities in the microchannel flow Greek symbols P fluid density μ fluid dynamic viscosity ϕ oscillation phase angle
ing heat transfer. They found that local bubble activity, opposed to convective influences, was the dominant heat transfer mechanism. Hence, both the onset of nucleate boiling and the critical heat flux are strongly dependent on the jet velocity. In a previous study [29], a new approach whereas a synthetic jet powered by bubble growth and collapse inside a micro chamber was studied. This low-power approach does not require moving parts and has the essential necessities for operation in micro domains. The introduced method is an active approach which can be controlled using an input signal frequency and a prescribed duty cycle. The present study investigates the effect of this micro synthetic jet on flow boiling heat transfer in a microchannel. The effect of the microjet on the flow structures was analyzed through high-speed photography. Dominant mechanisms at different stages of the boiling process were identified and their interactions with the synthetic jet were illustrated and analyzed.
As shown in Fig. 1(a), the microchannel (1.5 mm wide, 20 mm long, and 220 μm tall) contained the heater in the middle of the channel. An adjacent chamber (3.5 mm in radius and 220 μm tall) housed another micro-heater that powered the synthetic jet [29]. The microchannel was connected to the chamber through a 300 μm nozzle. The micro-device was fabricated using standard microfabrication technology and consisted of several substrates. The electrical circuits including the heater and the resistance temperature detectors (RTDs) were deposited on the bottom of a 500-μm thick Pyrex substrate. The microchannel, chamber, and the nozzle were subsequently formed in a patterned 220-μm thick SU8 layer on top of the bottom substrate. It was then bonded to another transparent 500-μm thick Pyrex substrate using a transparent 135-μm thick polyester sticker with a double-side adhesive-coating. The micro-heaters and RTDs were made of two layers consisting of Titanium (7-nm) to promote adhesion and Platinum (30-nm) through several sequences of depositions and etching steps (both dry and wet etches). Aluminum (700-nm) was also deposited to form electrical connections between the heater/RTDs and contact pads. The two heaters were first formed on the substrate and then a 1-μm thick layer of silicon dioxide was deposited to electrically insulate them form the RTDs, which were formed in a subsequent step above the heater. After fabricating the RTDs, another layer of silicon dioxide, 500-nm thick, was deposited to electrically insulate them from the working fluid. Finally, the protective silicon dioxide layer was selectively patterned and etched to provide access to the contact pads and to enable electrical connection between the micro-device and electrical spring loaded probe pins. Fig. 2 depicts a typical RTD and its vias taken by an optical microscope. After cleaning and dehydrating the wafer using acetone, nitrogen, and a hot plate, SU8 100 was dispensed and spun at 1500 RPM for 45 s to form a thickness of 220 μm. The wafer was then placed on a level hot plate for a soft bake to evaporate the SU8 solvent and densify the film. After cooling to room temperature, it was exposed to ultraviolet (UV) light for 120 s through a photomask. Subsequently, a post exposure bake over a hot plate at 60 °C for 10 h was performed and then cooled to room temperature. The sample was placed in a SU8 developer container for 3 h at room temperature. It was then rinsed and dried using isopropyl alcohol (IPA) and nitrogen. Fig. 3 illustrates a scanning electron microscope (SEM) image of the nozzle connected to the microchannel in the SU8 layer. The sticker was used to bond the bottom substrate, which carried the electrical circuits and the patterned SU8, to another transparent Pyrex substrate. Both the sticker and the upper Pyrex wafer were patterned with a laser-cutter (Hurricane 80 W) to create through-holes for the contact pads. Finally, the inlet and outlet manifolds were drilled (D = 1 mm) using Milltronic machine and a diamond drill through the bottom substrate. The transparent Pyrex substrate and the sticker allowed to visualize the flow during experiments. 2.2. Experimental set-up A closed fluid (HFE-70 0 0) loop was constructed and is shown in Fig. 4. Liquid was propelled through an electric pump (Micropump DB-380.A Brushless Drive) from a reservoir into the microchannel and the micro-chamber before returning to the reservoir. A calibrated OMEGA rotameter was used to measure flow rates, and a T-type thermocouple measured the inlet temperature. Two OMEGA pressure transducers were installed upstream and downstream of
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Fig. 1. (a) Schematic of the microfluidic device and (b) its different layers.
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Fig. 2. A typical RTD and its vias connectors. The RTD was made of two layers consisting of Tiatainum (7 nm) and Paltinum (30 nm). An additional 700 nm thick Aluminum was deposited to form the vias.
the device to measure fluid inlet and outlet pressures. The flow rate in the loop was controlled by adjusting the pump power. In addition, a needle valve was used to fine-adjustment the flow rate. Power was supplied to the thin-film heater by a GW Instek® DC power supply. The voltage across and the current through the heater were measured simultaneously using two Keysight digital multimeters. A LabVIEW program and a National InstrumentsTM data acquisition hardware were used with a personal computer (PC) to record the experimental measurements. Pulsating power in the micro-heater was generated by controlling the applied voltage of the DC power supply using an N-channel MOSFET, which functioned as a switch. The MOSFET was triggered by a GW Instek® electric function generator, which also allowed to change the signal’s frequency and duty cycle. The microfluidic device was placed in a custom made package (Fig. 5), precision machined from Delrin, to provide the fluid connection between the micro-device and the external fluid loop. The package was placed under an inverted microscope to visualize the flow in the microchannel and in the micro-chamber during operation. Images were captured using a high-speed camera connected to the microscope and were transferred to the PC for storage and subsequent analysis. The experimental apparatus was placed on an optical table with an anti-vibration system to minimize vibration and leveling effects. The system pressure was maintained at atmospheric pressure during all experiments, corresponding to a boiling temperature of 34 °C. The flow rate in the main microchannel was fixed at a mass flux of 63.18 kg/m2 s.
Fig. 3. SEM images of the nozzle and the microchannel fabricated in the SU8 layer.
Table 1 Experimental uncertainties. Measurements
Uncertainty
Mass flow rate, m˙ Inlet temperature, Tin Voltage on the heater, V Current on the heater, I Electrical power on the heater, P RTDs’ resistance Bubble volume Jet velocity Heat transfer coefficient, htp
±3% ±0.5 °C 0.10% 0.10% 0.12%±0.5 ±2%±4%±6.75% ~ 17%
2.3. Experimental procedure Prior to running the experiments, the RTDs were calibrated to obtain the resistance-temperature relations. The micro-device was placed in a copper plate package while a thermocouple was connected to the micro-device center. The package was then placed in an oven along with another thermocouple to measure the temperature inside the oven. By gradually ramping up the temperature and recording the resistance of the RTDs, the resistancetemperature curve for each individual sensor was established. These relations were then entered into the LabVIEW code and used during experiments. To estimate heat losses, experiments were performed after the calibration of the RTDs. Electrical power was applied to the heater located in the microchannel without flow and the temperature
changes were captured by the data acquisition unit and transferred to the PC. After reaching steady state, the power and the temperature difference between the ambient and the test section were recorded. The experiment was done for a wide range of input powers to establish the relation between the power and temperature (Fig. 6). This relation was used to estimate the heat loss (Q˙ loss ) during the experiment. Uncertainties associated with the experimental equipment are given in Table 1. Uncertainties of the measured values were obtained from the manufactures’ specification sheets, and the uncertainties of the derived parameters were calculated according to the standard propagation of uncertainty analysis [30].
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Fig. 4. Schematic of the experimental setup.
2.4. Data reduction
was expressed through the momentum coefficient [31,32], Cμ , defined as:
The average velocity in the microchannel was calculated using the following equation:
Vavg =
m˙
(1)
ρ (Hc wc )
where m˙ is the mass flow rate in the microchannel, ρ is the density, Hc and wc are the height and width of the microchannel, respectively. An image processing software, ImageJ, was used to calculate the bubble dimensions (i.e., area) obtained through high-speed imaging. Since the boundaries of the bubble were captured with the high-speed camera, the areas of the regions covered by the bubble were calculated. Considering that the bubbles were trapped between two flat plates, which were 220 μm apart, the areas obtained were multiplied by the distance between the two plates (i.e., the bubble height) to calculate the bubble volume and the associated liquid displacement as follow:
Vb = Abubble × hch
(2)
By calculating the bubble volume changes at different stages during the operation, the net flow rate at the nozzle section can be calculated subsequently, as follow:
V˙ jet =
Vb,
t2 − Vb, t2 − t1
t1
(3)
Uncertainties of the bubble volume and the nozzle net flow rate were estimated to be ±2% and ±4%, respectively. In order to characterize the synthetic jet and quantify its strength in comparison to the flow in the microchannel, its relative momentum (with respect to that of the microchannel flow),
ρ jV j,2 Avg AN ρ∞ VA2vg AC
Cμ =
(4)
where ρ ∞ is the fluid density, ρ j and Vj,Avg are the fluid density and mean velocity of the jet in the nozzle’s smallest cross section area corresponding to the location of maximum jet velocity. Since the fluid in the microchannel and the fluid emanating from the nozzle, were the same, ρ ∞ = ρ j . Additionally, the height of the microchannel and the nozzle were the same, thus the momentum coefficient is reduced to:
V j,2 Avg WN VA2vg WC
Cμ =
(5)
As the jet velocity varied periodically, the momentum coefficient varied with the strength of the jet. The average momentum coefficient is defined as follows:
Cμ, Avg =
τ
1
τ
∫ Cμ (t ) dt
t = 0
(6)
where τ is time corresponding to one full cycle of the synthetic jet (i.e., full cycle that includes the bubble growth, collapse, and incubation period). Dimensionless time, bubble volume and jet velocity are also defined as follow:
t¯ =
t
τ
, V¯b =
V jet Vb , V¯ jet = Vch Vc
(7)
where Vb , Vch , Vjet and Vc are the bubble volume, chamber volume, jet velocity at the nozzle throat and the average flow velocity in the microchannel respectively.
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Fig. 5. The micro-device package used for the experiments along with the printed circuit board (PCB), ribbon cables for date acquisition, probe pins, terminal strip, and cover plate.
The heat flux was calculated using the voltage and current, which were recorded during experiments. The input electrical power, P, and the heater resistance, R, are related to the measured electrical voltage, V, and current, I, as follow:
P =V ×I
(8)
And
R=V / I
(9)
The RTD temperatures were calculated using the resistancetemperature relations that were obtained during the calibration process. By considering the heat loss, the effective heat flux, qeff , is defined as follow:
qeff =
P − Q˙ loss Ah
(10)
where, Ah , is the heater area. Using the local surface temperature, Ts , the flow boiling heat transfer coefficient, h, is defined as follow:
h=
q eff Ts − Tsat
(11)
3. Results and discussions Initial experiments were performed without the synthetic jet and served as a baseline for comparison. Subsequently, the synthetic jet was turned on and the modified characteristics of the flow boiling were examined. 3.1. Flow boiling without the synthetic jet Fig. 7 depicts a sequence of flow patterns in the microchannel established by gradually increasing the heat flux from the onset of nucleate boiling (ONB) to fully developed flow boiling. Boiling was first observed at a heat flux of 24.5 kW/m2 , near the exit of the microchannel where the mean fluid temperature was
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Fig. 6. Heat loss estimation. Electrical power was applied to the heater located in the microchannel without flow and the temperature changes were captured and are shown here (R2 is the coefficient of determination of the linear regression).
Fig. 7. Different boiling regimes in the microchannel.
Fig. 8. Bubble coalescence and elongated vapor bubbles formation.
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Fig. 9. Rapid expansion of the tracked bubble that blows off the liquid and expanded dry-out regions.
Fig. 10. Sequential stages of the bubble expansion and condensation during different phases of an operating cycle.
the highest due to heating of the flow and the established thermal boundary layer. As the heat flux increased to 34.1 kW/m2 and 41.6 kW/m2 , additional nucleation sites were activated, especially upstream of the ONB location. As a result, bubbles gradually expanded upstream and encompassed a larger region above the heater; in addition the average bubble diameter increased. At a heat flux of 53.5 kW/m2 , bubbles started coalescing where smaller bubbles merged and formed larger bubbles that eventually evolved into a single elongated vapor slug confined within the microchannel cross-section (Fig. 8). Nevertheless, the heater surface was still covered by a thin liquid layer, preventing spikes in the surface temperature. As expected, nucleate flow boiling was accompanied by enhanced heat transfer due to the phase change process and more rigorous mixing brought about by the bubble ebullition pro-
cesses. At a heat flux of 94 kW/m2 , dry spots on the heater surface appeared, which was intensified at an even higher heat flux of 135 kW/m2 . It was observed that a sudden bubble expansion amplified the appearance of dry-out spots. As shown in Fig. 9, bubbles formed in the upstream region of the heater, then moved downstream and at some point, explosively expanded covering the microchannel cross-section area. This process ruptured the thin liquid film adjacent to the heating surface, leading to increased local dryout regions. 3.2. Flow boiling with synthetic jets The progression of the flow pattern described above was modified when the synthetic jet was introduced. To power the synthetic jet the DC power supply was set to 3.4 Watt, and the op-
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Fig. 11. (a) Time dependent dimensionless bubbles volume in the chamber and (b) dimensionless average jet velocity at the nozzle throat during a typical operating cycle.
erating frequency and duty cycle of the step waveform created by the function generator were set to 0.7 Hz and 2.5%, respectively. Fig. 10 depicts the pumping chamber at different phases within an operating cycle. The bubbles shown in the chamber at ϕ = 0, correspond to the residual bubbles from a previous cycle. As the heater was turned on, the vapor pressure inside the bubbles increased, and as a result, they began to grow, occupying an increasing volume in the chamber. The continuous bubble expansion in the chamber, propelled the liquid inside the chamber through the nozzle, creating a jet that emanated into the microchannel. Following the evaporation period, the heater was turned off and the bubbles gradually condensed. Concurrently, the bubble contracted and liquid from the microchannel was pulled into the chamber. The sequence of heating and condensation in the chamber generates a series of ejections and suctions at the nozzle throat, creating the synthetic jet in the microchannel. It was observed that the bubbles reached their maximum size in less than 5 percent of the duty cycle (i.e., less than 70 ms). The changes of the dimensionless bubble volume in the chamber and the dimensionless average jet veloc-
ity at the nozzle throat during one typical cycle are quantitatively illustrated in Fig. 11. (The data was obtained using image processing and by analyzing the rate of the bubble volume changes.) During the bubble growth, the liquid in the chamber was forced out, establishing a net positive velocity as depicted in Fig. 11(b). Negative velocity values are associated with the bubble collapse phase where the liquid was pulled into the chamber from the microchannel. A maximum positive jet velocity of about 52 times the average velocity in the microchannel was observed at the beginning of the cycle, which was associated with the bubble growth phase, and a maximum negative velocity of about 7 times the average velocity in the microchannel was observed during the suction phase. In order to quantify the synthetic jet in comparison to the flow in the microchannel, the momentum coefficient was calculated using the obtained jet velocity profile. In doing so, the maximum jet momentum coefficient and the average momentum coefficient over the entire operating cycle were calculated to be 281.4 and 2.1, respectively. The average momentum coefficient is more than unity suggesting that the momentum of the synthetic jet was larger than
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Fig. 12. (a) Formation of dry-out spots over the heater surface which diminished the heat transfer coefficient, (b) Influence of the applied jet on mitigating the dryout regions and thin film formation.
the momentum of the flow in the microchannel, and as a result expected to significantly affect the flow field and the associated heat transfer process. The synthetic jet mitigated the formation of dry-out spots during boiling in the microchannel as shown in Fig. 12. The left column correspond to flow boiling in the microchannel without the synthetic jet and the right column depicts the flow structures with the synthetic jet. Formation of dry-out regions over the heated surface is associated with a transition from a very effective heat transfer process to a very inefficient one leading to a rapid increase in the surface temperature. Mitigating these dry-out formations is a critical objective to maintain a desirable low surface temperature. As shown below, the synthetic jet assisted maintaining a thin liquid layer over the heater and mitigating hot spot areas. The effect of the synthetic jet on the two-phase heat transfer coefficient is illustrated in Fig. 13 for heat fluxes of q = 115 KW/m2 and 135 KW/m2 . dry-out formations was apparent under these heat fluxes. A step rise in the heat transfer coefficient (HTC) was observed with the activation of the synthetic jet. The increase in the HTC is related to the replenishment of dry
surface with a liquid film produced by the synthetic jet, thereby mitigating the increased surface temperature associated with dryout. Compared to the normal condition without the jet, an increase of up to 20 percent enhancement in the effective heat transfer coefficient was observed. Fig. 14 depicts the effective HTC at heat fluxes of 160 KW/m2 , corresponding to severe dry-out, and 95 KW/m2 with limited dry-out — a similar trend is observed. A more comprehensive result for the average HTC with and without the synthetic jet are shown in Fig. 15 for heat fluxes ranging from 27 KW/m2 to 160 KW/m2 . An initial raise in the HTC with heat flux is associated with the activation of nucleation sites and the associate increase in bubble formations, its latent heat, and mixing enhancement. Under these conditions, nucleate boiling was the dominate heat transfer mechanism. The detached bubbles in the channel also augment convection heat transfer by increasing the liquid superficial velocity. After the initial onset of nucleate boiling, further increase in the heat flux led to a decrease in the HTC. This is due to the coalescence of bubbles where they merged and occupied the entire cross section of the microchannel as a slug. As suggested in previous studies [33,34], this confinement effects suppressed bubble nucleation and induced lower HTC. At this stage, convective boiling and thin film evaporation prevailed and became the dominant heat transfer mechanisms. An evaporation momentum force repelled the liquid from the surface making it difficult for the liquid to spread over the heated area. As a result, the effective heat transfer surface area decreased, leading to lower HTC, which deteriorated with further increase in heat flux. It was observed that the synthetic jet increased the HTC for all the heat fluxes considered, although this enhancement appears to be more significant at higher heat fluxes. At lower heat fluxes where nucleate boiling was the dominated heat transfer mechanism, the HTC was more dependent on successive bubble growth and detachment from the heater surface. Applying the synthetic jet augments the HTC by promoting advection (i.e., flow mixing). For the slug flow, the contribution of the synthetic jet became more apparent as it tended to break the large elongated bubbles that suppressed nucleation. In higher heat fluxes where dryout spots began to form over the heater surface, the evaporation of the thin liquid film contributed significantly to the HTC. As shown in Fig. 12, the synthetic jet assisted spreading the thin liquid film over the heater area and mitigating dry-out spots, hence, augmenting the thin film evaporation mechanism and the overall HTC.
4. Conclusions Enhancement of nucleate boiling in microchannel using synthetic jet in cross-flow was experimentally studied. The synthetic jet was generated using bubble growth and collapse inside a chamber adjacent to the microchannel. The approach eliminated the need for mechanical moving parts to produce synthetic jet. Highspeed flow visualization and a microscope enabled to track the bubble ebullition processes in the micro-chamber during different boiling regimes in the microchannel. It also allowed the interactions between the synthetic jet and the flow in the microchannel. Results suggest that nucleate flow boiling was enhanced through the synthetic jet. For high heat fluxes where dryout regions formed, the synthetic jet assisted rewetting the heated surface by spreading a thin liquid film and mitigating the formation of hot spots. It was found that the enhancement of the two phase heat transfer coefficient was more noticeable at higher heat fluxes where thin film evaporation and bubble nucleation were augmented by the synthetic jet.
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Fig. 13. Heat transfer coefficient before and after the activation of the synthetic jet.
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Fig. 14. Heat transfer coefficient before and after the activation of the synthetic jet.
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Fig. 15. Time averaged heat transfer coefficient before and after the activation of the synthetic jet.
Declaration of Competing Interest The authors declare that there are no conflicts of interest. Acknowledgment This work was funded by the Office of Naval Research (ONR) with Dr. Mark Spector as program manager (Grant #N0 0 01415-1-2071). The authors would like to acknowledge the Cornell NanoScale Science & Technology Facility (CNF) for staff support. References [1] W. Qu, I. Mudawar, Measurement and correlation of critical heat flux in two-phase micro-channel heat sinks, Int. J. Heat Mass Transf. 47 (2004) 2045–2059. [2] A. Kos¸ ar, C.J. Kuo, Y. Peles, Suppression of boiling flow oscillations in parallel microchannels by inlet restrictors, J. Heat Transf. 128 (2005) 251–260. [3] S.G. Kandlikar, W.K. Kuan, D.A. Willistein, J. Borrelli, Stabilization of flow boiling in microchannels using pressure drop elements and fabricated nucleation sites, J. Heat Transf. 128 (2005) 389–396. [4] S.G. Kandlikar, advances History, and challenges in liquid flow and flow boiling heat transfer in microchannels: a critical review, ASME J. Heat Transf. 134 (2012) 034001. [5] M.C. Riofrío, N. Caney, J.A. Gruss, State of the art of efficient pumped two-phase flow cooling technologies, Appl. Therm. Eng. 104 (2016) 333–343. [6] Y. Peles, Advanced chip level liquid heat exchangers, Ann. Rev. Heat Transf. (2015) 155–198. [7] M. Kiyofumi, I. Akira, O. Hiroaki, The themohydraulic characteristics of two-phase flow in extremely narrow channels, Heat Transf. 21 (8) (1993) 823–837. [8] S. Shin, G. Choi, B.S. Kim, H.H. Cho, Flow boiling heat transfer on nanowire– coated surfaces with highly wetting liquid, Energy 76 (2014) 428–435. [9] A.K.M.M. Morshed, F. Yang, M.Y. Ali, J.A. Khan, C. Li, Enhanced flow boiling in a microchannel with integration of nanowires, Appl. Therm. Eng. 32 (2012) 68–75. [10] V. Khanikar, I. Mudawar T. Fisher, Effects of carbon nanotube coating on flow boiling in a micro-channel, Int. J. Heat Mass Transf. 52 (2009) 3805–3817. [11] D. Deng, W. Wan, Y. Qin, J. Zhang, X. Chu, Flow boiling enhancement of structured microchannels with micro pin fins, Int. J. Heat Mass Transf. 105 (2017) 338–349. [12] C. Woodcock, X. Yu, J. Plawsky, Y. Peles, Piranha pin fin (PPF) — advanced flow boiling microstructures with low surface tension dielectric fluids, Int. J. Heat Mass Transf. 90 (2015) 591–604. [13] S. Krishnamurthy, Y. Peles, Flow boiling on micropin fins entrenched inside a microchannel—flow patterns and bubble departure diameter and bubble frequency, J. Heat Transf. 132 (4) (2010) 041002. [14] W. Xing, J. Plawsky, C. Woodcock, X. Yu, A. Ullmann, N. Brauner, Y. Peles, Liquid-liquid phase separation heat transfer in advanced micro structure, Int. J. Heat Mass Transf. 127 (2018) 989–10 0 0.
[15] C.T. Lu, C. Pan, A highly stable microchannel heat sink for convective boiling, J. Micromech. Microeng. 19 (5) (2009) 055013. [16] D. Cooke, S.G. Kandlikar, Effect of open microchannel geometry on pool boiling enhancement, Int. J. Heat Mass Transf. 55 (2012) 1004–1013. [17] W. Li, F. Yang, T. Alam, X. Qu, B. Peng, J. Khan, C. Li, Enhanced flow boiling in microchannels using auxiliary channels and multiple micronozzles (I): characterizations of flow boiling heat transfer, Int. J. Heat Mass Transf. 116 (2018) 208–217. [18] W. Li, J. Ma, T. Alam, F. Yang, J. Khan, C. Li, Flow boiling of HFE-7100 in silicon microchannels integrated with multiple micro-nozzles and reentry micro-cavities, Int. J. Heat Mass Transf. 123 (2018) 354–366. [19] W. Li, T. Alam, F. Yang, X. Qu, B. Peng, J. Khan, C. Li, Enhanced flow boiling in microchannels using auxiliary channels and multiple micronozzles (II): enhanced CHF and reduced pressure drop, Int. J. Heat Mass Transf. 115 (2017) 264–272. [20] W. Li, X. Qu, T. Alam, F. Yang, W. Chang, J. Khan, C. Li, Enhanced flow boiling in microchannels through integrating multiple micro-nozzles and reentry microcavities, Appl. Phys. Lett. 110 (2017) 014104. [21] F. Yang, X. Dai, C.J. Kuo, Y. Peles, J. Khan, C. Li, Enhanced flow boiling in microchannels by self-sustained high frequency two-phase oscillations, Int. J. Heat Mass Transf. 58 (2013) 402–412. [22] X. Dai, F. Yang, R. Fang, T. Yemame, J.A. Khan, C. Li, Enhanced single-and two-phase transport phenomena using flow separation in a microgap with copper woven mesh coatings, Appl. Therm. Eng. 54 (2013) 281–288. [23] A.K. Vutha, S.R. Rao, F. Houshmand, Y. Peles, Active control of flow boiling oscillation amplitude and frequency using a transverse jet in crossflow, Appl. Phys. Lett. 108 (2016) 134104. [24] J. Xu, G. Liu, W. Zhang, Q. Li, B. Wang, Seed bubbles stabilize flow and heat transfer in parallel microchannels, Int. J. Multiph. Flow 35 (2009) 773–790. [25] R. Fang, J. Khan, Experimental study on the effect of synthetic jets on flow boiling heat transfer in microchannels, in: ASME 10th International Conference on Nanochannels, Microchannels, and Minichannels, 2012, pp. 81–89. [26] J.A. Khan, A.K.M.M.M Morshed, R. Fang, Towards ultra-compact high heat flux microchannel heat sink, Procedia Eng. 90 (2014) 11–24. [27] M. Fabbri, S. Jiang, V.K. Dhir, A comparative study of cooling of high power density electronics using sprays and microjets, J. Heat Transf. 127 (2005) 38–48. [28] C. de Brún, R. Jenkins, T.L. Lupton, R. Lupoi, R. Kempers, A.J. Robinson, Confined jet array impingement boiling, Exp. Therm. Fluid Sci. 86 (2017) 224–234. [29] E. Sourtiji, Y. Peles, A micro-synthetic jet in a microchannel using bubble growth and collapse, Appl. Therm. Eng. 160 (2019) 114084. [30] R.J. Moffat, Describing the uncertainties in experimental results, Exp. Therm. Fluid Sci. 1 (1988) 3–17. [31] M. Amitay, B.L. Smith, A. Glezer, Aerodynamic flow control using synthetic jet technology, AIAA 98 (1998) 0208. [32] J. Jung, C.J. Kuo, Y. Peles, M. Amitay, The flow field around a micropillar confined in a microchannel, Int. J. Heat Fluid Flow 36 (2012) 118–132. [33] J. Ma, W. Li, C. Ren, J.A. Khan, C. Li, Realizing highly coordinated, rapid and sustainable nucleate boiling in microchannels on HFE-7100, Int. J. Heat Mass Transf. 133 (2019) 1219–1229. [34] D. Deng, Y. Xie, Q. Huang, W. Wan, On the flow boiling enhancement in interconnected reentrant microchannels, Int. J. Heat Mass Transf. 108 (2017) 453–467.
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Contents lists available at ScienceDirect
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Flow boiling in microchannel with synthetic jet in cross-flow Ehsan Sourtiji∗, Yoav Peles Department of Mechanical and Aerospace Engineering, University of Central Florida, Orlando, FL 32828, USA
a r t i c l e
i n f o
Article history: Received 3 July 2019 Revised 23 October 2019 Accepted 8 November 2019 Available online xxx Keywords: Micro synthetic jet Microchannel Phase change Flow boiling Heat transfer Active enhancement
a b s t r a c t Flow boiling heat transfer in a microchannel with a micro synthetic jet powered by liquid/vapor phase change was studied. The synthetic jet had no moving parts and used the motion of the interfacial layer between vapor and liquid to propel liquid. Bubbles were generated on a micro heater in a chamber, 3.5 mm in radius and 220 μm tall, which was connected to the microchannel through a 300 μm nozzle opening. Periodic growth and collapse of bubbles in the chamber powered the synthetic jet. Using highspeed photography and a microscope, the sequence of bubble growth and collapse in the chamber, bubble nucleation in the microchannel, and their interaction with the synthetic jet were captured and analyzed. The jet velocity was estimated using image processing and its momentum coefficient was used to characterize the strength of the synthetic jet. The results showed that the synthetic jet enhanced nucleate flow boiling heat transfer in the microchannel by up to 20%. dry-out spots formed over the heated surface for high heat fluxes were mitigated by the synthetic jet as it assisted rewetting the surface and maintaining the integrity of the thin liquid film adjacent to the heated surface. © 2019 Elsevier Ltd. All rights reserved.
1. Introduction The demand for cooling high heat flux surfaces has been steadily growing for a range of applications, such as cooling of electronic devices. To address this issue, flow boiling heat transfer in diminishing length scales has been extensively studied due to its high heat transfer coefficient [1–5], high latent heat of vaporization, and its potential to enable direct chip-level cooling [6]. For instance, Kiyofumi et al. [7] observed that the heat transfer coefficient of flow boiling in microchannels is 3–20 times higher than that of a single-phase flow. In the last decade, a range of passive and active enhancement approaches for flow boiling heat transfer in diminishing length scales were studied, such as nano engineered surfaces [8–10], pin fins and microstructures [11–13], liquid phase separation [14], intricate microchannel geometries [15,16]. A unique passive heat transfer enhancement approach using auxiliary channels and multiple micronozzles embedded into a microchannel domain was extensively studied by Li et al. [17– 22]. Considerable flow boiling heat transfer enhancement was observed, which also suppressed two-phase flow instabilities. The enhancement was linked to a disruption of the boundary layer growth, flow mixing, and effective management of bubble confine-
∗
Corresponding author. E-mail addresses: [email protected] (E. Sourtiji), [email protected] (Y. Peles).
ments in the microchannel formed by a secondary flow of microjets emanated from the auxiliary channels. It was also deduced that the collapse of confined bubbles, due to direct condensation, enhanced the mixing and the associated heat transfer process. Active methods have also been studied. For instance, a transverse jet in crossflow with flow boiling in microchannel was studied by Vutha et al. [23] who observed considerable reduction in thermal oscillations amplitudes associated with the two-phase flow. This was attributed to the replenishment of liquid in dry-out regions, which subsequently suppressed large temperature raises. Xu et al. [24] used seed bubbles to improve heat transfer efficiency in a microchannel. The seed bubbles were generated using microheaters upstream in the microchannel and driven by pulse voltage signal. It was shown that introducing the seed bubbles, stabilize flow and heat transfer in micro-boiling, suppress the flow instability and decrease heating surface temperature. Fang and Khan [25] and Khan et al. [26] observed that a synthetic jet mitigated pressure and temperature fluctuations during flow boiling in microchannel. It was found that backflow due to bubble clogging and expansion were notably delayed by the synthetic jet, and that the thermal boundary layer was disturbed, and as a result, the heat transfer was enhanced. It was also inferred that the synthetic jet effectively broke elongated bubbles. Fabbri et al. [27] carried out an experimental study with microjet heat transfer and observed an increase in the heat transfer coefficient. Brun et al. [28] investigated the influence of confined jet array impingement on boil-
https://doi.org/10.1016/j.ijheatmasstransfer.2019.119023 0017-9310/© 2019 Elsevier Ltd. All rights reserved.
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2. Experimental apparatus and method Nomenclature 2.1. Micro-device Ah AN AC Cμ Cμ, Avg Dh Hc h I m˙ P qeff Q˙ loss R t t¯
τ
Ts Tsat V vavg vj, avg V¯b V¯ jet wc wN
surface area of the heater cross section area of the nozzle throat cross section area of the microchannel momentum coefficient average momentum coefficient hydraulic diameter microchannel height two-phase heat transfer coefficient electrical current mass flow rate applied electrical power effective heat flux heat loss rate electrical resistance time dimensionless time oscillation period surface temperature saturation temperature electrical voltage spatial average velocity in the microchannel crosssection area spatial average velocity in the nozzle throat crosssection area dimensionless bubble volume dimensionless jet velocity width of the microchannel opening width of the nozzle throat
Abbreviation HTC heat transfer coefficient nm nanometer ONB onset of nucleate boiling RTD resistance temperature detector μm micrometer via vertical interconnect access ∞ refers to the quantities in the microchannel flow Greek symbols P fluid density μ fluid dynamic viscosity ϕ oscillation phase angle
ing heat transfer. They found that local bubble activity, opposed to convective influences, was the dominant heat transfer mechanism. Hence, both the onset of nucleate boiling and the critical heat flux are strongly dependent on the jet velocity. In a previous study [29], a new approach whereas a synthetic jet powered by bubble growth and collapse inside a micro chamber was studied. This low-power approach does not require moving parts and has the essential necessities for operation in micro domains. The introduced method is an active approach which can be controlled using an input signal frequency and a prescribed duty cycle. The present study investigates the effect of this micro synthetic jet on flow boiling heat transfer in a microchannel. The effect of the microjet on the flow structures was analyzed through high-speed photography. Dominant mechanisms at different stages of the boiling process were identified and their interactions with the synthetic jet were illustrated and analyzed.
As shown in Fig. 1(a), the microchannel (1.5 mm wide, 20 mm long, and 220 μm tall) contained the heater in the middle of the channel. An adjacent chamber (3.5 mm in radius and 220 μm tall) housed another micro-heater that powered the synthetic jet [29]. The microchannel was connected to the chamber through a 300 μm nozzle. The micro-device was fabricated using standard microfabrication technology and consisted of several substrates. The electrical circuits including the heater and the resistance temperature detectors (RTDs) were deposited on the bottom of a 500-μm thick Pyrex substrate. The microchannel, chamber, and the nozzle were subsequently formed in a patterned 220-μm thick SU8 layer on top of the bottom substrate. It was then bonded to another transparent 500-μm thick Pyrex substrate using a transparent 135-μm thick polyester sticker with a double-side adhesive-coating. The micro-heaters and RTDs were made of two layers consisting of Titanium (7-nm) to promote adhesion and Platinum (30-nm) through several sequences of depositions and etching steps (both dry and wet etches). Aluminum (700-nm) was also deposited to form electrical connections between the heater/RTDs and contact pads. The two heaters were first formed on the substrate and then a 1-μm thick layer of silicon dioxide was deposited to electrically insulate them form the RTDs, which were formed in a subsequent step above the heater. After fabricating the RTDs, another layer of silicon dioxide, 500-nm thick, was deposited to electrically insulate them from the working fluid. Finally, the protective silicon dioxide layer was selectively patterned and etched to provide access to the contact pads and to enable electrical connection between the micro-device and electrical spring loaded probe pins. Fig. 2 depicts a typical RTD and its vias taken by an optical microscope. After cleaning and dehydrating the wafer using acetone, nitrogen, and a hot plate, SU8 100 was dispensed and spun at 1500 RPM for 45 s to form a thickness of 220 μm. The wafer was then placed on a level hot plate for a soft bake to evaporate the SU8 solvent and densify the film. After cooling to room temperature, it was exposed to ultraviolet (UV) light for 120 s through a photomask. Subsequently, a post exposure bake over a hot plate at 60 °C for 10 h was performed and then cooled to room temperature. The sample was placed in a SU8 developer container for 3 h at room temperature. It was then rinsed and dried using isopropyl alcohol (IPA) and nitrogen. Fig. 3 illustrates a scanning electron microscope (SEM) image of the nozzle connected to the microchannel in the SU8 layer. The sticker was used to bond the bottom substrate, which carried the electrical circuits and the patterned SU8, to another transparent Pyrex substrate. Both the sticker and the upper Pyrex wafer were patterned with a laser-cutter (Hurricane 80 W) to create through-holes for the contact pads. Finally, the inlet and outlet manifolds were drilled (D = 1 mm) using Milltronic machine and a diamond drill through the bottom substrate. The transparent Pyrex substrate and the sticker allowed to visualize the flow during experiments. 2.2. Experimental set-up A closed fluid (HFE-70 0 0) loop was constructed and is shown in Fig. 4. Liquid was propelled through an electric pump (Micropump DB-380.A Brushless Drive) from a reservoir into the microchannel and the micro-chamber before returning to the reservoir. A calibrated OMEGA rotameter was used to measure flow rates, and a T-type thermocouple measured the inlet temperature. Two OMEGA pressure transducers were installed upstream and downstream of
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Fig. 1. (a) Schematic of the microfluidic device and (b) its different layers.
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Fig. 2. A typical RTD and its vias connectors. The RTD was made of two layers consisting of Tiatainum (7 nm) and Paltinum (30 nm). An additional 700 nm thick Aluminum was deposited to form the vias.
the device to measure fluid inlet and outlet pressures. The flow rate in the loop was controlled by adjusting the pump power. In addition, a needle valve was used to fine-adjustment the flow rate. Power was supplied to the thin-film heater by a GW Instek® DC power supply. The voltage across and the current through the heater were measured simultaneously using two Keysight digital multimeters. A LabVIEW program and a National InstrumentsTM data acquisition hardware were used with a personal computer (PC) to record the experimental measurements. Pulsating power in the micro-heater was generated by controlling the applied voltage of the DC power supply using an N-channel MOSFET, which functioned as a switch. The MOSFET was triggered by a GW Instek® electric function generator, which also allowed to change the signal’s frequency and duty cycle. The microfluidic device was placed in a custom made package (Fig. 5), precision machined from Delrin, to provide the fluid connection between the micro-device and the external fluid loop. The package was placed under an inverted microscope to visualize the flow in the microchannel and in the micro-chamber during operation. Images were captured using a high-speed camera connected to the microscope and were transferred to the PC for storage and subsequent analysis. The experimental apparatus was placed on an optical table with an anti-vibration system to minimize vibration and leveling effects. The system pressure was maintained at atmospheric pressure during all experiments, corresponding to a boiling temperature of 34 °C. The flow rate in the main microchannel was fixed at a mass flux of 63.18 kg/m2 s.
Fig. 3. SEM images of the nozzle and the microchannel fabricated in the SU8 layer.
Table 1 Experimental uncertainties. Measurements
Uncertainty
Mass flow rate, m˙ Inlet temperature, Tin Voltage on the heater, V Current on the heater, I Electrical power on the heater, P RTDs’ resistance Bubble volume Jet velocity Heat transfer coefficient, htp
±3% ±0.5 °C 0.10% 0.10% 0.12%±0.5 ±2%±4%±6.75% ~ 17%
2.3. Experimental procedure Prior to running the experiments, the RTDs were calibrated to obtain the resistance-temperature relations. The micro-device was placed in a copper plate package while a thermocouple was connected to the micro-device center. The package was then placed in an oven along with another thermocouple to measure the temperature inside the oven. By gradually ramping up the temperature and recording the resistance of the RTDs, the resistancetemperature curve for each individual sensor was established. These relations were then entered into the LabVIEW code and used during experiments. To estimate heat losses, experiments were performed after the calibration of the RTDs. Electrical power was applied to the heater located in the microchannel without flow and the temperature
changes were captured by the data acquisition unit and transferred to the PC. After reaching steady state, the power and the temperature difference between the ambient and the test section were recorded. The experiment was done for a wide range of input powers to establish the relation between the power and temperature (Fig. 6). This relation was used to estimate the heat loss (Q˙ loss ) during the experiment. Uncertainties associated with the experimental equipment are given in Table 1. Uncertainties of the measured values were obtained from the manufactures’ specification sheets, and the uncertainties of the derived parameters were calculated according to the standard propagation of uncertainty analysis [30].
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Fig. 4. Schematic of the experimental setup.
2.4. Data reduction
was expressed through the momentum coefficient [31,32], Cμ , defined as:
The average velocity in the microchannel was calculated using the following equation:
Vavg =
m˙
(1)
ρ (Hc wc )
where m˙ is the mass flow rate in the microchannel, ρ is the density, Hc and wc are the height and width of the microchannel, respectively. An image processing software, ImageJ, was used to calculate the bubble dimensions (i.e., area) obtained through high-speed imaging. Since the boundaries of the bubble were captured with the high-speed camera, the areas of the regions covered by the bubble were calculated. Considering that the bubbles were trapped between two flat plates, which were 220 μm apart, the areas obtained were multiplied by the distance between the two plates (i.e., the bubble height) to calculate the bubble volume and the associated liquid displacement as follow:
Vb = Abubble × hch
(2)
By calculating the bubble volume changes at different stages during the operation, the net flow rate at the nozzle section can be calculated subsequently, as follow:
V˙ jet =
Vb,
t2 − Vb, t2 − t1
t1
(3)
Uncertainties of the bubble volume and the nozzle net flow rate were estimated to be ±2% and ±4%, respectively. In order to characterize the synthetic jet and quantify its strength in comparison to the flow in the microchannel, its relative momentum (with respect to that of the microchannel flow),
ρ jV j,2 Avg AN ρ∞ VA2vg AC
Cμ =
(4)
where ρ ∞ is the fluid density, ρ j and Vj,Avg are the fluid density and mean velocity of the jet in the nozzle’s smallest cross section area corresponding to the location of maximum jet velocity. Since the fluid in the microchannel and the fluid emanating from the nozzle, were the same, ρ ∞ = ρ j . Additionally, the height of the microchannel and the nozzle were the same, thus the momentum coefficient is reduced to:
V j,2 Avg WN VA2vg WC
Cμ =
(5)
As the jet velocity varied periodically, the momentum coefficient varied with the strength of the jet. The average momentum coefficient is defined as follows:
Cμ, Avg =
τ
1
τ
∫ Cμ (t ) dt
t = 0
(6)
where τ is time corresponding to one full cycle of the synthetic jet (i.e., full cycle that includes the bubble growth, collapse, and incubation period). Dimensionless time, bubble volume and jet velocity are also defined as follow:
t¯ =
t
τ
, V¯b =
V jet Vb , V¯ jet = Vch Vc
(7)
where Vb , Vch , Vjet and Vc are the bubble volume, chamber volume, jet velocity at the nozzle throat and the average flow velocity in the microchannel respectively.
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Fig. 5. The micro-device package used for the experiments along with the printed circuit board (PCB), ribbon cables for date acquisition, probe pins, terminal strip, and cover plate.
The heat flux was calculated using the voltage and current, which were recorded during experiments. The input electrical power, P, and the heater resistance, R, are related to the measured electrical voltage, V, and current, I, as follow:
P =V ×I
(8)
And
R=V / I
(9)
The RTD temperatures were calculated using the resistancetemperature relations that were obtained during the calibration process. By considering the heat loss, the effective heat flux, qeff , is defined as follow:
qeff =
P − Q˙ loss Ah
(10)
where, Ah , is the heater area. Using the local surface temperature, Ts , the flow boiling heat transfer coefficient, h, is defined as follow:
h=
q eff Ts − Tsat
(11)
3. Results and discussions Initial experiments were performed without the synthetic jet and served as a baseline for comparison. Subsequently, the synthetic jet was turned on and the modified characteristics of the flow boiling were examined. 3.1. Flow boiling without the synthetic jet Fig. 7 depicts a sequence of flow patterns in the microchannel established by gradually increasing the heat flux from the onset of nucleate boiling (ONB) to fully developed flow boiling. Boiling was first observed at a heat flux of 24.5 kW/m2 , near the exit of the microchannel where the mean fluid temperature was
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Fig. 6. Heat loss estimation. Electrical power was applied to the heater located in the microchannel without flow and the temperature changes were captured and are shown here (R2 is the coefficient of determination of the linear regression).
Fig. 7. Different boiling regimes in the microchannel.
Fig. 8. Bubble coalescence and elongated vapor bubbles formation.
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Fig. 9. Rapid expansion of the tracked bubble that blows off the liquid and expanded dry-out regions.
Fig. 10. Sequential stages of the bubble expansion and condensation during different phases of an operating cycle.
the highest due to heating of the flow and the established thermal boundary layer. As the heat flux increased to 34.1 kW/m2 and 41.6 kW/m2 , additional nucleation sites were activated, especially upstream of the ONB location. As a result, bubbles gradually expanded upstream and encompassed a larger region above the heater; in addition the average bubble diameter increased. At a heat flux of 53.5 kW/m2 , bubbles started coalescing where smaller bubbles merged and formed larger bubbles that eventually evolved into a single elongated vapor slug confined within the microchannel cross-section (Fig. 8). Nevertheless, the heater surface was still covered by a thin liquid layer, preventing spikes in the surface temperature. As expected, nucleate flow boiling was accompanied by enhanced heat transfer due to the phase change process and more rigorous mixing brought about by the bubble ebullition pro-
cesses. At a heat flux of 94 kW/m2 , dry spots on the heater surface appeared, which was intensified at an even higher heat flux of 135 kW/m2 . It was observed that a sudden bubble expansion amplified the appearance of dry-out spots. As shown in Fig. 9, bubbles formed in the upstream region of the heater, then moved downstream and at some point, explosively expanded covering the microchannel cross-section area. This process ruptured the thin liquid film adjacent to the heating surface, leading to increased local dryout regions. 3.2. Flow boiling with synthetic jets The progression of the flow pattern described above was modified when the synthetic jet was introduced. To power the synthetic jet the DC power supply was set to 3.4 Watt, and the op-
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Fig. 11. (a) Time dependent dimensionless bubbles volume in the chamber and (b) dimensionless average jet velocity at the nozzle throat during a typical operating cycle.
erating frequency and duty cycle of the step waveform created by the function generator were set to 0.7 Hz and 2.5%, respectively. Fig. 10 depicts the pumping chamber at different phases within an operating cycle. The bubbles shown in the chamber at ϕ = 0, correspond to the residual bubbles from a previous cycle. As the heater was turned on, the vapor pressure inside the bubbles increased, and as a result, they began to grow, occupying an increasing volume in the chamber. The continuous bubble expansion in the chamber, propelled the liquid inside the chamber through the nozzle, creating a jet that emanated into the microchannel. Following the evaporation period, the heater was turned off and the bubbles gradually condensed. Concurrently, the bubble contracted and liquid from the microchannel was pulled into the chamber. The sequence of heating and condensation in the chamber generates a series of ejections and suctions at the nozzle throat, creating the synthetic jet in the microchannel. It was observed that the bubbles reached their maximum size in less than 5 percent of the duty cycle (i.e., less than 70 ms). The changes of the dimensionless bubble volume in the chamber and the dimensionless average jet veloc-
ity at the nozzle throat during one typical cycle are quantitatively illustrated in Fig. 11. (The data was obtained using image processing and by analyzing the rate of the bubble volume changes.) During the bubble growth, the liquid in the chamber was forced out, establishing a net positive velocity as depicted in Fig. 11(b). Negative velocity values are associated with the bubble collapse phase where the liquid was pulled into the chamber from the microchannel. A maximum positive jet velocity of about 52 times the average velocity in the microchannel was observed at the beginning of the cycle, which was associated with the bubble growth phase, and a maximum negative velocity of about 7 times the average velocity in the microchannel was observed during the suction phase. In order to quantify the synthetic jet in comparison to the flow in the microchannel, the momentum coefficient was calculated using the obtained jet velocity profile. In doing so, the maximum jet momentum coefficient and the average momentum coefficient over the entire operating cycle were calculated to be 281.4 and 2.1, respectively. The average momentum coefficient is more than unity suggesting that the momentum of the synthetic jet was larger than
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Fig. 12. (a) Formation of dry-out spots over the heater surface which diminished the heat transfer coefficient, (b) Influence of the applied jet on mitigating the dryout regions and thin film formation.
the momentum of the flow in the microchannel, and as a result expected to significantly affect the flow field and the associated heat transfer process. The synthetic jet mitigated the formation of dry-out spots during boiling in the microchannel as shown in Fig. 12. The left column correspond to flow boiling in the microchannel without the synthetic jet and the right column depicts the flow structures with the synthetic jet. Formation of dry-out regions over the heated surface is associated with a transition from a very effective heat transfer process to a very inefficient one leading to a rapid increase in the surface temperature. Mitigating these dry-out formations is a critical objective to maintain a desirable low surface temperature. As shown below, the synthetic jet assisted maintaining a thin liquid layer over the heater and mitigating hot spot areas. The effect of the synthetic jet on the two-phase heat transfer coefficient is illustrated in Fig. 13 for heat fluxes of q = 115 KW/m2 and 135 KW/m2 . dry-out formations was apparent under these heat fluxes. A step rise in the heat transfer coefficient (HTC) was observed with the activation of the synthetic jet. The increase in the HTC is related to the replenishment of dry
surface with a liquid film produced by the synthetic jet, thereby mitigating the increased surface temperature associated with dryout. Compared to the normal condition without the jet, an increase of up to 20 percent enhancement in the effective heat transfer coefficient was observed. Fig. 14 depicts the effective HTC at heat fluxes of 160 KW/m2 , corresponding to severe dry-out, and 95 KW/m2 with limited dry-out — a similar trend is observed. A more comprehensive result for the average HTC with and without the synthetic jet are shown in Fig. 15 for heat fluxes ranging from 27 KW/m2 to 160 KW/m2 . An initial raise in the HTC with heat flux is associated with the activation of nucleation sites and the associate increase in bubble formations, its latent heat, and mixing enhancement. Under these conditions, nucleate boiling was the dominate heat transfer mechanism. The detached bubbles in the channel also augment convection heat transfer by increasing the liquid superficial velocity. After the initial onset of nucleate boiling, further increase in the heat flux led to a decrease in the HTC. This is due to the coalescence of bubbles where they merged and occupied the entire cross section of the microchannel as a slug. As suggested in previous studies [33,34], this confinement effects suppressed bubble nucleation and induced lower HTC. At this stage, convective boiling and thin film evaporation prevailed and became the dominant heat transfer mechanisms. An evaporation momentum force repelled the liquid from the surface making it difficult for the liquid to spread over the heated area. As a result, the effective heat transfer surface area decreased, leading to lower HTC, which deteriorated with further increase in heat flux. It was observed that the synthetic jet increased the HTC for all the heat fluxes considered, although this enhancement appears to be more significant at higher heat fluxes. At lower heat fluxes where nucleate boiling was the dominated heat transfer mechanism, the HTC was more dependent on successive bubble growth and detachment from the heater surface. Applying the synthetic jet augments the HTC by promoting advection (i.e., flow mixing). For the slug flow, the contribution of the synthetic jet became more apparent as it tended to break the large elongated bubbles that suppressed nucleation. In higher heat fluxes where dryout spots began to form over the heater surface, the evaporation of the thin liquid film contributed significantly to the HTC. As shown in Fig. 12, the synthetic jet assisted spreading the thin liquid film over the heater area and mitigating dry-out spots, hence, augmenting the thin film evaporation mechanism and the overall HTC.
4. Conclusions Enhancement of nucleate boiling in microchannel using synthetic jet in cross-flow was experimentally studied. The synthetic jet was generated using bubble growth and collapse inside a chamber adjacent to the microchannel. The approach eliminated the need for mechanical moving parts to produce synthetic jet. Highspeed flow visualization and a microscope enabled to track the bubble ebullition processes in the micro-chamber during different boiling regimes in the microchannel. It also allowed the interactions between the synthetic jet and the flow in the microchannel. Results suggest that nucleate flow boiling was enhanced through the synthetic jet. For high heat fluxes where dryout regions formed, the synthetic jet assisted rewetting the heated surface by spreading a thin liquid film and mitigating the formation of hot spots. It was found that the enhancement of the two phase heat transfer coefficient was more noticeable at higher heat fluxes where thin film evaporation and bubble nucleation were augmented by the synthetic jet.
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Fig. 13. Heat transfer coefficient before and after the activation of the synthetic jet.
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Fig. 14. Heat transfer coefficient before and after the activation of the synthetic jet.
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Fig. 15. Time averaged heat transfer coefficient before and after the activation of the synthetic jet.
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Please cite this article as: E. Sourtiji and Y. Peles, Flow boiling in microchannel with synthetic jet in cross-flow, International Journal of Heat and Mass Transfer, https://doi.org/10.1016/j.ijheatmasstransfer.2019.119023