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Mechanism of enhancement of heat transfer for plate-fin heat sinks with dual piezoelectric fans
International Journal of Heat and Mass Transfer 90 (2015) 454–465
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International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
Mechanism of enhancement of heat transfer for plate-fin heat sinks with dual piezoelectric fans Sheng-Lun Ma a, Jing-Wei Chen a, Hung-Yi Li b,1, Jing-Tang Yang a,⇑ a b
Department of Mechanical Engineering, National Taiwan University, Taipei 10617, Taiwan Department of Mechatronic Engineering, Huafan University, Shihtin, New Taipei 22301, Taiwan
a r t i c l e
i n f o
Article history: Received 16 November 2014 Received in revised form 20 February 2015 Accepted 15 March 2015
Keywords: Dual piezofans Plate-fin heat sink Heat transfer Particle image velocimetry
a b s t r a c t The transient flow structure and heat transfer of plate-fin heat sinks cooled with dual piezoelectric fans (piezofans) have been investigated experimentally. The analysis integrated with the measurement of thermal resistance, velocity, and vorticity was conducted via high-speed particle image velocimetry (PIV) and infrared thermometry. The mechanism of the enhancement of heat transfer was deduced on varying the spacing, height, phase and orientation of the fan. As the ratio of fan spacing to the width of the heat sink (D⁄) and the ratio of the fan height to the height of the heat sink (H ) decreased, the thermal resistance decreased. The thermal resistance for the fans vibrating out of phase is invariably superior to the fans vibrating in phase. The thermal resistance in the parallel orientation was 22–27% less than in the vertical orientation because of the unobstructed ventilation of ambient air, dual fans and the heat sink. In contrast, the cooling performance in the vertical orientation decreased because the hot flow in the heat sink was isolated with only a small amount of entrained cold flow. The velocity induced by the dual fans decreased 30–50% because the buoyancy effect caused by natural convection annihilated the cold flow to move downward into the heat sink, resulting in a decreased cooling performance. The revealed mechanisms of interaction between the dual fans and the plate-fin heat sink shed light on the concept of optimal design for electronic cooling systems. Ó 2015 Elsevier Ltd. All rights reserved.
1. Introduction Portable electronic equipment such as notebook computers, smart phones and portable point-of-sale (POS) terminals has become smaller and thinner with greatly increasing heat fluxes. Usable cooling methods are limited because there is little space to mount cooling devices. Ensuring that electronic products operate stably at large heat fluxes without a diminished life expectancy significantly requires a novel cooling device that combines improved cooling performance with small dimensions. The cooling method to dissipate heat from electrical devices is generally forced convection using air. Rotary-type fans have been widely used in conventional large-sized devices because of their efficiency and performance, but it is difficult to apply rotary fans for small electronics because of limitations of producing components such as rotor, bearings, motor and shaft smaller than a critical size. A piezoelectric fan is one technique proposed as an alternative device ⇑ Corresponding author. Tel.: +886 2 33669875; fax: +886 2 23634254. E-mail addresses: [email protected] (J.-T. Yang). 1 Tel.: +886 2 26632102x4017.
(H.-Y.
http://dx.doi.org/10.1016/j.ijheatmasstransfer.2015.03.050 0017-9310/Ó 2015 Elsevier Ltd. All rights reserved.
Li),
[email protected]
to generate air flow [1,2]. When AC power is applied to a piezoelectric fan causing an electrical potential difference across a piezoelectric material deposited on the plate, the plate vibrates back and forth and generates an air flow. Its advantageous features include no electromagnetic noise, light weight, minimum space requirement, low energy consumption and long life expectancy [3]. Much work has been conducted on the cooling performance and the flow characteristics of piezoelectric fans. Toda [4,5], an early researcher, investigated the effectiveness of cooling and proposed the essential models for the piezoelectric fan. Some research attempts to characterize the cooling performance with a small piezoelectric fan for electric devices were made with computational [2,6–10] and experimental methods [2,11]. The investigation of the effects of the piezoelectric fan height on the flow and heat transfer in microelectronic cooling horizontally oriented shows that the piezoelectric fan farthest from the heat source conferred the best heat transfer and best decreased the temperature of the heat source [2]. Tan et al. [10] assessed numerically the effects of a resonating mode and the opening inlet on the flow and heat-transfer characteristics, and concluded that the opening inlet has a weak effect on the flow and heat transfer for a large gap between a piezoelectric fan and a heated surface. The
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S.-L. Ma et al. / International Journal of Heat and Mass Transfer 90 (2015) 454–465 Blade Swing
64.0 mm 76.7 mm
Mylar Blade
Piezoelectric Bending Element
22.0 mm (h) 17.0 mm
25.4 mm
36.0 mm
12.7 mm
2.2 mm
(a)
(b)
Fig. 1. (a) Schematic diagram of the piezofan, and (b) plate-fin heat sink.
Table 1 Piezoelectric fan specifications. Input voltage Capacitance Power consumption Volume flow rate
115 VAC/60 Hz 15 nF 30 mW 2 CFM (0.9 l/s)
two-dimensional flow field generated from a piezoelectric plate was investigated. Vortex shedding each time the fan passed the nature position was observed with the maximum fluid velocity roughly four times the maximum tip velocity [12]. Li et al. [11] experimentally investigated the thermal-fluid characteristics of plate-fin heat sinks cooled with a piezoelectric fan with varied configuration, position and dimensions of the heat sink. With high-speed PIV, the effects of these parameters on the vortex formation, evolution and transport were analyzed. The complexity of the structural, fluid and heat-transfer behavior requires further investigation [13]. The feasibility of applying a single piezoelectric plate as a cooling device in electronic systems has also been investigated. Prototypes of a single piezoelectric plate built in an enclosure the size of a cellular telephone and a typical notebook computer were
used to demonstrate the cooling feasibility and performance [1]. To explore further the possibilities of use of a single piezoelectric fan coupled with the heat sink to enhance the cooling performance, investigations are in progress [3,14,15]. A single piezoelectric fan coupled with the heat sink has been considered in much previous work, but little research to understand better the thermal-fluid behavior of multiple piezoelectric fans has been conducted. Ihara and Watanabe [16] investigated the flow around a pair of piezoelectric fans oscillating in phase and out of phase with a large amplitude, but the vortex motion generated by the piezoelectric fans was not clearly observed because of insufficient spatial resolution of laser doppler velocimetry (LDV); their work also neglected heat-transfer characteristics. Kimber and Garimella [17] investigated the influence of factors including the fan length, vibrational amplitude, frequency offset and distance from the heat source on the arrays of piezoelectric fans. Under optimal conditions, the cooling performance was at resonance and oriented normal to the heat source, resulting in an enhancement 3.75 times in the heat transfer relative to natural convection alone. Abdullah et al. [18] presented an experimental analysis of the effects of the tip gap and the amplitude of vibration of piezoelectric fans on the heat-transfer characteristics of finned heat sinks and concluded that the best result occurred at large amplitudes and with a minimum tip gap. Choi et al. [19,20] investigated the unsteady flow fields around single and dual vibrating fans in a free stream, and concluded that the performance of dual piezoelectric fans vibrating in counter-phase is invariably superior to dual piezoelectric fans vibrating in phase unless the distance is too small. Discussion of thermal-fluid characteristics is, however, lacking in the literature on the complexity of interactive mechanisms between the heat source and dual fans with varied geometry, amplitude, phase angle and resonance frequency of the piezoelectric fan, the gap between the piezoelectric fan and the heat source, and the position of the piezoelectric fan. In this work hence the effects of various parameters including the spacing, height, phase and orientation of the fan on the transient flow structure and heat transfer were investigated experimentally using high-speed PIV and infrared thermometry. The cooling performance of the dual piezofans were determined based on the analysis of the mechanisms of interaction between dual piezofans and the heat sink.
Laser Sheet
Piezo Fan Heat Sink
y
y
AC Power Supply
x
z Experimental Chamber
Laser
x z
COMS Camera
Laser Sheet
Opcal Lens PC & Control Soware
(a)
D
x
y
H
z
(b)
Fig. 2. (a) Experimental apparatus to investigate the interactive behavior of the flow field between plate-fin heat sinks and a piezofan; (b) parallel (upper) and vertical (lower) orientation of combination of piezofan and heat sink with denotation of fan height (H) and fan spacing (D).
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2. Experiments The employed AC piezofan (Piezo system, mass 2.8 g, with Mylar blade 36.0 mm, width 12.7 mm and total length 76.7 mm) is shown in Fig. 1(a). The blade swing is 25.4 mm with maximal air velocity up to 2 m/s. Table 1 shows the detailed dimensions and specifications of the piezofan. Fig. 1(b) shows the plate-fin heat sinks made of a cheap aluminum alloy (6061) that is easily machined and has large thermal conductivity. The heat sinks are coated with a flat black paint (radiation emissivity 0.9 ± 0.05) to increase the accuracy of temperature measurement. The dimensions of the heat sink are width (w) 30.0 mm, length (l) 30.0 mm, height (h) 22.0 mm, with thickness (b) 5.0 mm of the fin. The dimensions of the base of the heat sink are the same as the cross-sectional area of the underneath heating element, 30 30 mm for base and the fin thickness is 5.0 mm. An acrylic heat sink employed in the experiments without heating was fabricated to have the same size as that of aluminum alloy (6061). For the investigation on dynamic motions in flow field, an acrylic heat sink with high transparency is used for formation of light sheet in the PIV experiments. The flow mechanisms induced by dual fans are inherently complicated because the flow field is highly unsteady [9]. For such a complicated flow, the high-speed PIV was employed to provide a practical analysis of flow motions and structures, which clarifies and validates the flow interaction between the dual fans. Fig. 2(a) shows the experimental apparatus including the high-speed PIV system, the piezofans, the acrylic experimental chamber (dimensions 300 300 300 mm), the AC power supply and the heat sink. The vibrational phase of the piezofan (0° and 180°) was controlled on altering the electrode connected to an AC power supply. A single-axis motion controller was employed to adjust the position of the piezofan. Fig. 2(b) shows the orientations of parallel (upper) and vertical (lower) for the piezofan and heat sink in combination. For the orientation of parallel, the motion of piezofan is in the positive and negative x-directions perpendicular to the open of heat sink; for the orientation of vertical, the motion of piezofan is in the positive and negative z-directions parallel to the open of heat sink. Parameters H⁄ and D⁄ are dimensionless quantities defined as the ratio of fan height (H) to height of heat sink (h) and ratio of fan space (D) to width of heat sink (w), respectively. The fan height is defined as distance between the fan tip and the bottom of base plate; the fan space is defined as distance between the dual piezofans. The fan height and the fan space were varied by 2-axies stages precision translator-rotator with a fixture to adjust the position of dual piezofans. The conditions with variation of D⁄, H⁄, phase and orientation tested in this work are
Table 2 Eight conditions with variation of H⁄, D⁄, orientation, and phase. Case
H⁄
D⁄
Orientation
Phase
A a B b C c D d E e F f G g H h
1.10 1.10 1.10 1.10 1.10 1.10 1.10 1.10 0.50 0.50 0.80 0.80 1.10 1.10 1.10 1.10
0.67 0.67 0.80 0.80 0.93 0.93 1.06 1.06 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.00
Parallel Parallel Parallel Parallel Parallel Parallel Parallel Parallel Vertical Vertical Vertical Vertical Vertical Vertical Parallel Parallel
In phase Out of phase In phase Out of phase In phase Out of phase In phase Out of phase In phase Out of phase In phase Out of phase In phase Out of phase In phase Out of phase
Temperature Measurement Module
PC & Control Soware
IR Camera
Piezo fan
Heat Sink
y
Experimental Chamber x Heater
AC Power Supply
DC Power Supply
Fig. 3. Experimental setup for thermal image of plate-fin heat sink cooled by piezofan captured by infrared camera.
listed in Table 2. High-speed PIV was used to acquire the flow fields induced by the dual piezofans. For high-speed PIV measurement, a high-speed CMOS camera (Phantom V7.3, Vision Research) capable of capturing 190,000 frames s1 at 800 600 pixels was employed. The camera, equipped with a lens (Nikon Nikkor, 50 mm, f/1.2), was employed to acquire images with 640 480 pixels (spatial resolution 109.86 lm/pixel) at 3000 frames s1. Al2O3 particles of mean diameter about 1 lm were seeded as tracer particles into the transparent acrylic box. The diode-pumped solid-state laser (15 W, Lighthouse Photonics, wavelength 532 nm) served as a light source. The laser beam was formed into a light-sheet (thickness 1 mm) through a light-sheet forming system. The orientation of the plane of the light-sheet perpendicular to the camera depends on the orientation of the piezofan and the heat sink, as shown in Fig. 2(b). The raw images were calculated (Insight 3G, TSI; interrogation window size 32 32 pixels) with 50% overlap of the interrogation window [21]. The detailed procedure for the analysis of the flow field was similar to previous work [11,22]. To investigate the thermal performance of the plate-fin heat sink cooled by the piezofan, the infrared thermography was employed for capturing the thermal image, as shown in Fig. 3. The infrared camera (FLIR A325) used in the experiments has an uncooled focal-plane array detector (320 240 pixels) and can receive a wavelength from 7.5 lm to 13 lm. The range of temperature measurement from 20 °C to 350 °C is covered with accuracy ±2%. The field of view is 25° 19°/0.4 m with thermal sensitivity 0.05 °C at 30 °C. The images were recorded and analyzed with software (ThermaCAM Research Pro 2.9). The detailed description of the heating system has been described in previous work [11]. The heating system (LW-4671, Long Win Science and Technology Corporation) connected to the heat sink includes a heating element, temperature-measuring module and DC power supply, providing a heating power of 3-W (±5%) to the heat sink. The heating element was completely wrapped with insulation material except in areas in contact with the base of the heat sink. The heating power of the system is
Q¼
kal AðT l T u Þ d
ð1Þ
where kal is the thermal conductivity of aluminum alloy, A is the cross-sectional area of the heating element, T u and T l are the temperatures of the upper and lower measuring points inside the heating element, respectively, and d is the distance between the two measuring points. The temperature measurement is a measurement module. Two platinum thermally sensitive resistors (PT-100) are used to measure the temperatures inside the heating element. A T-type thermocouple is used to measure cooling fluid temperature. The thermal resistance is obtained from the resistance and voltage
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457
Fig. 4. Periodic velocity and vorticity distribution with, parallel orientation and in phase vibration (H⁄ 1.10 and D⁄ 0.67; Case A). (White arrow denoting the direction of main flow).
signals from the platinum thermally sensitive resistors and the thermocouple connected to the data acquisition device applying the equation
Rth ¼
T av e T 1 Q
ð2Þ
in which Q represents the heating power of the system and T av e and T 1 represent the average surface temperature of the heat-sink base plate and the temperature of the cooling fluid, respectively. The heating power and temperature monitoring are controlled with a computer. In the experiments with heat addition, the laser beam and laser sheet were orientated from top to bottom because of the heating system on the original optical path, resulting in a limited region of flow field obtained. 3. Results and discussion
x-direction. This cold flow was directed by the flow induced by the vibrating fans to form a jet down into the base plate (heat source) of the heat sink. The exchange between the accumulated flow in the heat sink and the downward jet of the cold flow occurred near the base plate of the heat sink; the mixed flow was directed upward to the ambient air. In contrast, for Case C with large D⁄ most dual fans induced entrainment flow shift from the left to the right side at the top of the heat sink, leading the decreased flow velocity into the base plate of the heat sink. When the cold flow from ambient air moved more into the heat sink, the thermal energy from the heat sink was removed better and the heat-transfer performance was improved [11]. The enhanced exchange of thermal energy between the hot flow and the cold flow for dual fans with D⁄ smaller than unity was more obvious than with large D⁄. As D⁄ decreased, the thermal resistance decreased, as shown in a subsequent section.
3.1. Effect of fan spacing 3.2. Effect of fan height The investigations of the effect of fan spacing were focused on a parallel orientation. The performance of dual fans with varied distance between them showed no interaction between the dual fans too far apart and interference between the dual fans too close together [23]. In this work, thus the condition with interruption between dual fans through considering that the double input power decreases the amplitude and produces less thrust was dismissed. Figs. 4 and 5 show the periodic (1/60 s) velocity vectors and vorticity contour of Cases A and C. For Case A with small D⁄, dual fans were capable of producing a larger space near the blade in the negative x-direction and entraining more cold flow from the ambient air on vibrating to the maximum deflection in the positive
The investigations of the effect of fan height (Cases E and G) were focused on a vertical orientation. For Case E, H⁄ is smaller than for Case G, meaning that the distance between dual fans and the base plate of the heat sink of Case E is smaller than that of Case G. The smaller distance between the dual fans and the base plate (heat source) causes improved cooling performance [9] in both parallel and vertical orientations [11,14,24,25], but few investigators considered the flow and structure interaction between dual fans and the heat sink. The periodic velocity and vorticity distribution for vibrating conditions in phase of Cases E and G are shown in Figs. 6 and 7, respectively. The obvious discrepancy
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Fig. 5. Periodic velocity and vorticity distribution with parallel orientation and in phase vibration (H⁄ 1.10 and D⁄ 0.93; Case C).
between Cases E and G is the intense velocity distribution and vortex structures between the dual fans and the heat sink. For Case E with small H⁄, the smaller distance effectively results in large velocities and vortex structures near the blade tip of dual fans, causing more flow to strike the base plate of the heat sink and so to enhance the cooling performance [9,12]. In contrast, the large velocities and vortex structures distributed far from the base plate of the heat sink for Case G resulted in a limited influence on the thermal energy with a decreased velocity of entrained cold flow in the heat sink. The tendency of periodic velocity and vorticity distribution for the vibration out of phase is the same as that found for the vibration in phase, as shown in Figs. 8 and 9. The cooling performances for Cases E and e are indicated to be better than those of Cases G and g because of the varied interactive behavior between the dual fans and the heat sink. The thermal resistance decreased with decreased H⁄ as shown in a subsequent section.
other near the blade tip moved into the heat sink, resulting in the generation of numerous small vortex structures. This highly turbulent effect including a directional change of cold flow and the dissipation of the vortices further augmented the transfer of thermal energy. As the dual fans moved away from each other to the maximum deflection in the positive and negative z-directions, the mixed flow within the heat sink moved along with the propagations of clockwise and counter-clockwise vortices and went outside. A region of low pressure was hence formed in the empty space between the two blade tips so that ambient air was pulled through the tunnel between the dual fans, moving into the heat sink. In summary, because of the collision between clockwise and counter-clockwise vortices induced by dual fans, and the formation of a suction effect induced by a region of low pressure in the heat sink, the cooling performance for vibration out of phase was much better than vibration in phase, resembling characteristics reported in Ref. [23].
3.3. Effect of fan phase 3.4. Effect of fan orientation The velocity and vorticity distribution of dual fans vibrating in phase (0°) and out of phase (180°) are shown in Figs. 6–9. For the vibration in phase (Figs. 6 and 7), the cold flow is entrained every half period and disappeared after a strong jet flow formed near the blade tip. Another jet flow formed after a half period carried cold flow into the heat sink and then disappeared again. In contrast, for the vibration out of phase (Figs. 8 and 9), the flow behavior differed from the characteristic for the vibration in phase. As the dual fans vibrated toward the center line between the dual fans, the cold flow was entrained from the left and right sides and the flow between the dual fans was then pushed upward. The clockwise and counter-clockwise vortices that collided with each
The effects of the orientation of a piezofan for both parallel and vertical orientations on the cooling performance have been investigated [11,14,24,25], but there are few investigations of the effects of orientation for a combination of dual fans and heat sink on the cooling performance. Further understanding of the detailed interactions of the heat transfer and flow structures between the dual fans and the heat sink plays an important role in applications as devices to cool electronics. Fig. 10(a) and (b) show the thermal resistance of the dual fans vibrating in phase and out of phase for varied D⁄ in the parallel orientation, and varied H⁄ in the vertical orientation. The conditions involved in Fig. 10(a) include Cases
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459
Fig. 6. Periodic velocity and vorticity distribution with vertical orientation and in phase vibration (H⁄ 0.50 and D⁄ 1.00; Case E).
A; B; b; C; c; D and d; the conditions involved in Fig. 10(b) include Cases E, e; F; G and g. For the parallel orientation, the thermal resistance decreased with decreasing D⁄ for dual fans vibrating both in phase and out of phase. For the dual fans vibrating in phase the thermal resistance decreased about 5% as D⁄ decreased from 1.06 to 0.67; for the dual fans vibrating out of phase, the minimum thermal resistance appeared as D⁄ = 0.80 because of the limited vibrational amplitude of the dual fans. For the vertical orientation, the thermal resistance decreased with decreasing H⁄ for the dual fans vibrating both in phase and out of phase. The thermal resistance of Case E (H⁄ = 0.50) was 10% less than that of Case G (H⁄ = 1.10) for dual fans vibrating both in phase and out of phase. Overall, the thermal resistance of dual fans vibrating out of phase was about 2% less than that of vibrating in phase for the parallel orientation, whereas the thermal resistance of the dual fans vibrating out of phase was about 3% less than that of vibrating in phase for the vertical orientation. The thermal resistance for the parallel orientation was 15% less than that in the vertical orientation. For both parallel and vertical orientations, the cooling performance of the dual fans vibrating out of phase was better than that in phase. The velocity and vorticity distributions captured with high-speed PIV were analyzed to reveal the influence of the fan orientation on the mechanism of the enhancement of heat transfer with the dual fans. For the case of a parallel orientation as shown in Figs. 4 and 5, a side cold flow was entrained by the dual fans and was split into two flow directions. One cold flow struck directly the outer surface of the heat sink; the cold flow moved along the direction perpendicular to the plane, increasing the convective heat transfer of the heat sink. The strong vortices outside
the heat sink were evidence of a fan-induced flow colliding on the surface of the heat sink. Another cold flow went through a small gap between the dual fans and the heat sink into the heat sink as the dual fans vibrated toward the center line. In addition to the limitation by the left- and right-side fins of the heat sink, ambient air sucked through dual fans moved downward and pressed the hot flow in the heat sink to move along the direction perpendicular out of plane. These two flows thus greatly increased the heat transfer of the heat sink. The entire flow circulation between ambient air, the dual fans and the heat sink benefited the ventilation and so the cooling performance. In contrast, for a vertical orientation as shown in Figs. 8 and 9, the characteristics were the same as that in the parallel orientation with the cold flow entrained by the dual fans and the suction effect due to the region of low pressure between the dual fans and the heat sink, but the fin of the heat sink was then in the path of air moving out in the positive and negative x-directions, so these two flows were limited by an enclosure formed in the heat sink. The trapped hot flow smoothly moved from the heat sink until a gap was created by the dual fans vibrating outwardly. For both parallel and vertical orientations, the various interactions of the flow structures were more obvious when H⁄ decreased. For the parallel orientation, both inner and outer sides of the fin of the heat sink were affected by the dual-fan-induced vortices, but for the vertical orientation the hot flow trapped in the heat sink and the cold flow outside the heat sink were unable to mix easily and smoothly. The cooling performance for the parallel orientation was thus much better than that of the vertical orientation because the heat transfer and ventilation benefited from the large surface of the heat sink for convection and effective circulation.
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Fig. 7. Periodic velocity and vorticity distribution with vertical orientation and in phase vibration (H⁄ 1.10 and D⁄ 1.00; Case G).
Fig. 8. Periodic velocity and vorticity distribution with vertical orientation and out of phase vibration (H⁄ 0.50 and D⁄ 1.00; Case e).
S.-L. Ma et al. / International Journal of Heat and Mass Transfer 90 (2015) 454–465
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Fig. 9. Periodic velocity and vorticity distribution of vertical orientation and out of phase vibration (H⁄ 1.10 and D⁄ 1.00; Case g).
Fig. 10. Thermal resistance of dual fans vibrating in phase and out of phase for (a) D⁄ of the parallel orientation, and (b) H⁄ of the vertical orientation.
3.5. Effect of hot flow and cold flow There has been little investigation of the effects of hot and cold flows on the combination of dual fans and a heat sink. Because of limitations of experiments to analyze the interaction between conditions of hot and cold flows, the characteristics of cold flow, without a heat source, and hot flow, with heat source, are considered separately in most investigations [11,17]. The interaction between hot and cold flows might be strongly affected and altered because of the buoyancy effect induced by natural convection. The velocity and vorticity distribution of hot flow in phase and hot flow out of phase for Case H are shown in Figs. 11 and 12, respectively; the
velocity and vorticity distribution of cold flow in phase and cold flow out of phase for Case H are shown in Figs. 13 and 14, respectively. Fig. 15 and 16 shows a time series of the horizontal velocity component (u) and the vertical velocity component (v ) for heights 5, 10 and 30 mm above the base plate along the center line with dual fans vibrating in phase and out of phase in Case H. For the case in the hot-flow condition, the region of interest is focused on the center of the heat sink to reveal the interaction between the dual fans and the heat sink, because of the greatly altered characteristic flow structure influenced by the hot- flow. Figs. 11 and 13 show that the velocity in the hot-flow condition was about half that in the cold-flow. Figs. 12 and 14 show that
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Fig. 11. Periodic velocity and vorticity distribution of hot flow and in phase vibration (H⁄ 1.10 and D⁄ 1.00; Case H).
Fig. 12. Periodic velocity and vorticity distribution of hot flow and out of phase vibration (H⁄ 1.10 and D⁄ 1.00; Case H).
S.-L. Ma et al. / International Journal of Heat and Mass Transfer 90 (2015) 454–465
Fig. 13. Periodic velocity and vorticity distribution of cold flow and in phase vibration (H⁄ 1.10 and D⁄ 1.00; Case H).
Fig. 14. Periodic velocity and vorticity distribution of cold flow and out of phase vibration (H⁄ 1.10 and D⁄ 1.00; Case H).
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(a) D = 5 mm
(b) D = 10 mm 0.8
0.4 in phase_cold in phase_hot
0.6
2
in phase_cold in phase_hot
0.4
0.0
in phase_cold in phase_hot
1
0.2
u (m/s)
u (m/s)
0.2
u (m/s)
(C) D = 30 mm
0.0
0
-0.2 -1
-0.4
-0.2
-0.6 -2
(a)
0.02
0.04
0.06
0.08
0.10
0.12
t (s)
-0.8 0.00
(a)
0.04
0.06
0.08
0.10
0.12
t (s)
out of phase_cold out of phase_hot
0.6
u (m/s)
0.0
0.02
0.04
0.06
0.08
0.10
0.12
0.08
0.10
0.12
t(s)
2
out of phase_cold out of phase_hot
0.4
0.2
0.00
(a)
0.8
0.4
u (m/s)
0.02
out of phase_cold out of phase_hot
1
0.2 0.0
u (m/s)
u
-0.4 0.00
0.800000
0
-0.2
-1
-0.4
-0.2
-0.6
-2
-0.4 0.00
(b)
0.02
0.04
0.06
0.08
0.10
0.12
t (s)
-0.8 0.00
(b)
0.02
0.04
0.06
0.08
0.10
0.12
t (s)
0.00
(b)
0.02
0.04
0.06
t (s)
Fig. 15. Time series of the horizontal velocity component (u) for heights (a) 5, (b) 10 and (c) 30 mm in Case H.
(a) D = 5 mm
(b) D = 10 mm
(C) D = 30 mm
0.8
0.4 in phase_cold in phase_hot
0.6
2
in phase_cold in phase_hot
0.4
0.2
in phase_cold in phase_hot
1
0.0
v (m/s)
v (m/s)
v (m/s)
0.2 0.0
-0.2 -0.4
-0.2
0
-1
-0.6
(c)
0.02
0.04
0.06
0.08
0.10
0.12
t (s)
-0.8 0.00
(c)
0.4
0.04
0.06
0.08
0.10
0.12
t (s)
(c)
0.8
out of phase_cold out of phase_hot
0.6
0.2
0.0
0.00
0.02
0.04
0.06
0.08
0.10
0.12
0.08
0.10
0.12
t (s)
2
out of phase_cold out of phase_hot
0.4
v (m/s)
v (m/s)
-2 0.02
out of phase_cold out of phase_hot
1
0.2
v (m/s)
v
-0.4 0.00
0.0
0
-0.2
-0.2
-1
-0.4 -0.6
-0.4 0.00
(d)
0.02
0.04
0.06
t (s)
0.08
0.10
0.12
-0.8 0.00
(d)
-2 0.02
0.04
0.06
t (s)
0.08
0.10
0.12
0.00
(d)
0.02
0.04
0.06
t (s)
Fig. 16. Time series of the vertical velocity component (v) for heights (a) 5, (b) 10 and (c) 30 mm in Case H.
the velocity decreased about 30% for the dual fans vibrating out of phase in the hot-flow condition with a large difference in the flow motion compared with the cold-flow condition. Under the influence of a hot-flow, the downward main flow at 3/4 period became upward flow through natural convection as the dual fans vibrated outwardly, as shown in Fig. 12. In contrast, the main direction of the cold-flow was downward as the dual fans vibrated outwardly because of the suction induced by low-pressure induced suction, as shown in Fig. 14. The horizontal velocity component (u) in the
hot-flow condition was 10% and 20% larger than that in the cold-flow condition for heights 10 and 30 mm, respectively. The vertical velocity component (v) below height 30 mm of the hot-flow condition was smaller than the cold-flow condition because of the formation of a thermal boundary: the vertical velocity component (v) for both conditions of cold and hot flows increased as the height increased. At height 30 mm, the vertical velocity component (v ) of the dual fans vibrating out of phase in the hot flow was invariably positive, indicating that the flow on
S.-L. Ma et al. / International Journal of Heat and Mass Transfer 90 (2015) 454–465
top of the heat sink went upward. This characteristic agrees with the discrepancies of the flow structures between cold and hot flows found in Figs. 11–14. In contrast, for dual fans vibrating out of phase, the vortex structures induced by those dual fans at the natural position increased the heat transfer by forced convection, as the dual fans vibrating outwardly to the flow field within the heat sinks were influenced by the vortex structures induced by the dual fans and natural convection. The effect of natural convection was thus much stronger than that of forced convection in the transition period of the dual fans vibrating from the natural position to the maximum deflection. The discrepancy between hot-flow and cold-flow for the dual fans vibrating, in general, contributes to the natural convection induced by the temperature gradient, which was stronger than the forced convection in the transition period. The influence of natural convection on the vortex structures in the heat sink and the entrained cold flow was indicated to be inevitable. The influence of the thermal fluid on the vortex interaction between the dual fans and the heat sink must therefore be considered as the influence of natural convection was shown to be negligible. 4. Conclusion The mechanisms of enhanced cooling performance on applying dual fans on a plate-fin heat sink with various parameters including the spacing, height, phase and orientation of the fan, was investigated experimentally with high-speed PIV and infrared thermometry. For the effect of the fan spacing, the width of the tunnel between the dual fans was capable of entraining more cold flow with decreased D⁄. For the effect of the fan height, the formation of the turbulent flow structure near the heat source with decreased H⁄ strongly promoted the heat transfer. The thermal resistance, decreasing with decreased D⁄ (5%), and decreasing H⁄ (10%) of the fans vibrating out of phase was invariably superior to the fans vibrating in phase (3%). The thermal resistance in the parallel orientation was 22–27% less than in the vertical orientation because of the unobstructed ventilation, but decreased in the vertical orientation because the entrained cold flow into the heat sink was limited. The velocity induced by the dual fans decreased about 30–50% because the buoyancy effect caused by natural convection annihilated the cold flow to move downward into the heat sink. The effect of the thermal fluid on the interaction between the vortex and the heat sink must be considered. The revealed mechanisms of thermal effects and flow interaction for the combination of the dual piezofans and the heat sink shed light on the design and development of electronic cooling systems. Conflict of interest None declared. Acknowledgment National Science Council of the Taiwan (Republic of China) for financially supported this research under contract NSC 102-2815-C-002-047-E.
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Contents lists available at ScienceDirect
International Journal of Heat and Mass Transfer journal homepage: www.elsevier.com/locate/ijhmt
Mechanism of enhancement of heat transfer for plate-fin heat sinks with dual piezoelectric fans Sheng-Lun Ma a, Jing-Wei Chen a, Hung-Yi Li b,1, Jing-Tang Yang a,⇑ a b
Department of Mechanical Engineering, National Taiwan University, Taipei 10617, Taiwan Department of Mechatronic Engineering, Huafan University, Shihtin, New Taipei 22301, Taiwan
a r t i c l e
i n f o
Article history: Received 16 November 2014 Received in revised form 20 February 2015 Accepted 15 March 2015
Keywords: Dual piezofans Plate-fin heat sink Heat transfer Particle image velocimetry
a b s t r a c t The transient flow structure and heat transfer of plate-fin heat sinks cooled with dual piezoelectric fans (piezofans) have been investigated experimentally. The analysis integrated with the measurement of thermal resistance, velocity, and vorticity was conducted via high-speed particle image velocimetry (PIV) and infrared thermometry. The mechanism of the enhancement of heat transfer was deduced on varying the spacing, height, phase and orientation of the fan. As the ratio of fan spacing to the width of the heat sink (D⁄) and the ratio of the fan height to the height of the heat sink (H ) decreased, the thermal resistance decreased. The thermal resistance for the fans vibrating out of phase is invariably superior to the fans vibrating in phase. The thermal resistance in the parallel orientation was 22–27% less than in the vertical orientation because of the unobstructed ventilation of ambient air, dual fans and the heat sink. In contrast, the cooling performance in the vertical orientation decreased because the hot flow in the heat sink was isolated with only a small amount of entrained cold flow. The velocity induced by the dual fans decreased 30–50% because the buoyancy effect caused by natural convection annihilated the cold flow to move downward into the heat sink, resulting in a decreased cooling performance. The revealed mechanisms of interaction between the dual fans and the plate-fin heat sink shed light on the concept of optimal design for electronic cooling systems. Ó 2015 Elsevier Ltd. All rights reserved.
1. Introduction Portable electronic equipment such as notebook computers, smart phones and portable point-of-sale (POS) terminals has become smaller and thinner with greatly increasing heat fluxes. Usable cooling methods are limited because there is little space to mount cooling devices. Ensuring that electronic products operate stably at large heat fluxes without a diminished life expectancy significantly requires a novel cooling device that combines improved cooling performance with small dimensions. The cooling method to dissipate heat from electrical devices is generally forced convection using air. Rotary-type fans have been widely used in conventional large-sized devices because of their efficiency and performance, but it is difficult to apply rotary fans for small electronics because of limitations of producing components such as rotor, bearings, motor and shaft smaller than a critical size. A piezoelectric fan is one technique proposed as an alternative device ⇑ Corresponding author. Tel.: +886 2 33669875; fax: +886 2 23634254. E-mail addresses: [email protected] (J.-T. Yang). 1 Tel.: +886 2 26632102x4017.
(H.-Y.
http://dx.doi.org/10.1016/j.ijheatmasstransfer.2015.03.050 0017-9310/Ó 2015 Elsevier Ltd. All rights reserved.
Li),
[email protected]
to generate air flow [1,2]. When AC power is applied to a piezoelectric fan causing an electrical potential difference across a piezoelectric material deposited on the plate, the plate vibrates back and forth and generates an air flow. Its advantageous features include no electromagnetic noise, light weight, minimum space requirement, low energy consumption and long life expectancy [3]. Much work has been conducted on the cooling performance and the flow characteristics of piezoelectric fans. Toda [4,5], an early researcher, investigated the effectiveness of cooling and proposed the essential models for the piezoelectric fan. Some research attempts to characterize the cooling performance with a small piezoelectric fan for electric devices were made with computational [2,6–10] and experimental methods [2,11]. The investigation of the effects of the piezoelectric fan height on the flow and heat transfer in microelectronic cooling horizontally oriented shows that the piezoelectric fan farthest from the heat source conferred the best heat transfer and best decreased the temperature of the heat source [2]. Tan et al. [10] assessed numerically the effects of a resonating mode and the opening inlet on the flow and heat-transfer characteristics, and concluded that the opening inlet has a weak effect on the flow and heat transfer for a large gap between a piezoelectric fan and a heated surface. The
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S.-L. Ma et al. / International Journal of Heat and Mass Transfer 90 (2015) 454–465 Blade Swing
64.0 mm 76.7 mm
Mylar Blade
Piezoelectric Bending Element
22.0 mm (h) 17.0 mm
25.4 mm
36.0 mm
12.7 mm
2.2 mm
(a)
(b)
Fig. 1. (a) Schematic diagram of the piezofan, and (b) plate-fin heat sink.
Table 1 Piezoelectric fan specifications. Input voltage Capacitance Power consumption Volume flow rate
115 VAC/60 Hz 15 nF 30 mW 2 CFM (0.9 l/s)
two-dimensional flow field generated from a piezoelectric plate was investigated. Vortex shedding each time the fan passed the nature position was observed with the maximum fluid velocity roughly four times the maximum tip velocity [12]. Li et al. [11] experimentally investigated the thermal-fluid characteristics of plate-fin heat sinks cooled with a piezoelectric fan with varied configuration, position and dimensions of the heat sink. With high-speed PIV, the effects of these parameters on the vortex formation, evolution and transport were analyzed. The complexity of the structural, fluid and heat-transfer behavior requires further investigation [13]. The feasibility of applying a single piezoelectric plate as a cooling device in electronic systems has also been investigated. Prototypes of a single piezoelectric plate built in an enclosure the size of a cellular telephone and a typical notebook computer were
used to demonstrate the cooling feasibility and performance [1]. To explore further the possibilities of use of a single piezoelectric fan coupled with the heat sink to enhance the cooling performance, investigations are in progress [3,14,15]. A single piezoelectric fan coupled with the heat sink has been considered in much previous work, but little research to understand better the thermal-fluid behavior of multiple piezoelectric fans has been conducted. Ihara and Watanabe [16] investigated the flow around a pair of piezoelectric fans oscillating in phase and out of phase with a large amplitude, but the vortex motion generated by the piezoelectric fans was not clearly observed because of insufficient spatial resolution of laser doppler velocimetry (LDV); their work also neglected heat-transfer characteristics. Kimber and Garimella [17] investigated the influence of factors including the fan length, vibrational amplitude, frequency offset and distance from the heat source on the arrays of piezoelectric fans. Under optimal conditions, the cooling performance was at resonance and oriented normal to the heat source, resulting in an enhancement 3.75 times in the heat transfer relative to natural convection alone. Abdullah et al. [18] presented an experimental analysis of the effects of the tip gap and the amplitude of vibration of piezoelectric fans on the heat-transfer characteristics of finned heat sinks and concluded that the best result occurred at large amplitudes and with a minimum tip gap. Choi et al. [19,20] investigated the unsteady flow fields around single and dual vibrating fans in a free stream, and concluded that the performance of dual piezoelectric fans vibrating in counter-phase is invariably superior to dual piezoelectric fans vibrating in phase unless the distance is too small. Discussion of thermal-fluid characteristics is, however, lacking in the literature on the complexity of interactive mechanisms between the heat source and dual fans with varied geometry, amplitude, phase angle and resonance frequency of the piezoelectric fan, the gap between the piezoelectric fan and the heat source, and the position of the piezoelectric fan. In this work hence the effects of various parameters including the spacing, height, phase and orientation of the fan on the transient flow structure and heat transfer were investigated experimentally using high-speed PIV and infrared thermometry. The cooling performance of the dual piezofans were determined based on the analysis of the mechanisms of interaction between dual piezofans and the heat sink.
Laser Sheet
Piezo Fan Heat Sink
y
y
AC Power Supply
x
z Experimental Chamber
Laser
x z
COMS Camera
Laser Sheet
Opcal Lens PC & Control Soware
(a)
D
x
y
H
z
(b)
Fig. 2. (a) Experimental apparatus to investigate the interactive behavior of the flow field between plate-fin heat sinks and a piezofan; (b) parallel (upper) and vertical (lower) orientation of combination of piezofan and heat sink with denotation of fan height (H) and fan spacing (D).
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2. Experiments The employed AC piezofan (Piezo system, mass 2.8 g, with Mylar blade 36.0 mm, width 12.7 mm and total length 76.7 mm) is shown in Fig. 1(a). The blade swing is 25.4 mm with maximal air velocity up to 2 m/s. Table 1 shows the detailed dimensions and specifications of the piezofan. Fig. 1(b) shows the plate-fin heat sinks made of a cheap aluminum alloy (6061) that is easily machined and has large thermal conductivity. The heat sinks are coated with a flat black paint (radiation emissivity 0.9 ± 0.05) to increase the accuracy of temperature measurement. The dimensions of the heat sink are width (w) 30.0 mm, length (l) 30.0 mm, height (h) 22.0 mm, with thickness (b) 5.0 mm of the fin. The dimensions of the base of the heat sink are the same as the cross-sectional area of the underneath heating element, 30 30 mm for base and the fin thickness is 5.0 mm. An acrylic heat sink employed in the experiments without heating was fabricated to have the same size as that of aluminum alloy (6061). For the investigation on dynamic motions in flow field, an acrylic heat sink with high transparency is used for formation of light sheet in the PIV experiments. The flow mechanisms induced by dual fans are inherently complicated because the flow field is highly unsteady [9]. For such a complicated flow, the high-speed PIV was employed to provide a practical analysis of flow motions and structures, which clarifies and validates the flow interaction between the dual fans. Fig. 2(a) shows the experimental apparatus including the high-speed PIV system, the piezofans, the acrylic experimental chamber (dimensions 300 300 300 mm), the AC power supply and the heat sink. The vibrational phase of the piezofan (0° and 180°) was controlled on altering the electrode connected to an AC power supply. A single-axis motion controller was employed to adjust the position of the piezofan. Fig. 2(b) shows the orientations of parallel (upper) and vertical (lower) for the piezofan and heat sink in combination. For the orientation of parallel, the motion of piezofan is in the positive and negative x-directions perpendicular to the open of heat sink; for the orientation of vertical, the motion of piezofan is in the positive and negative z-directions parallel to the open of heat sink. Parameters H⁄ and D⁄ are dimensionless quantities defined as the ratio of fan height (H) to height of heat sink (h) and ratio of fan space (D) to width of heat sink (w), respectively. The fan height is defined as distance between the fan tip and the bottom of base plate; the fan space is defined as distance between the dual piezofans. The fan height and the fan space were varied by 2-axies stages precision translator-rotator with a fixture to adjust the position of dual piezofans. The conditions with variation of D⁄, H⁄, phase and orientation tested in this work are
Table 2 Eight conditions with variation of H⁄, D⁄, orientation, and phase. Case
H⁄
D⁄
Orientation
Phase
A a B b C c D d E e F f G g H h
1.10 1.10 1.10 1.10 1.10 1.10 1.10 1.10 0.50 0.50 0.80 0.80 1.10 1.10 1.10 1.10
0.67 0.67 0.80 0.80 0.93 0.93 1.06 1.06 1.00 1.00 1.00 1.00 1.00 1.00 1.00 1.00
Parallel Parallel Parallel Parallel Parallel Parallel Parallel Parallel Vertical Vertical Vertical Vertical Vertical Vertical Parallel Parallel
In phase Out of phase In phase Out of phase In phase Out of phase In phase Out of phase In phase Out of phase In phase Out of phase In phase Out of phase In phase Out of phase
Temperature Measurement Module
PC & Control Soware
IR Camera
Piezo fan
Heat Sink
y
Experimental Chamber x Heater
AC Power Supply
DC Power Supply
Fig. 3. Experimental setup for thermal image of plate-fin heat sink cooled by piezofan captured by infrared camera.
listed in Table 2. High-speed PIV was used to acquire the flow fields induced by the dual piezofans. For high-speed PIV measurement, a high-speed CMOS camera (Phantom V7.3, Vision Research) capable of capturing 190,000 frames s1 at 800 600 pixels was employed. The camera, equipped with a lens (Nikon Nikkor, 50 mm, f/1.2), was employed to acquire images with 640 480 pixels (spatial resolution 109.86 lm/pixel) at 3000 frames s1. Al2O3 particles of mean diameter about 1 lm were seeded as tracer particles into the transparent acrylic box. The diode-pumped solid-state laser (15 W, Lighthouse Photonics, wavelength 532 nm) served as a light source. The laser beam was formed into a light-sheet (thickness 1 mm) through a light-sheet forming system. The orientation of the plane of the light-sheet perpendicular to the camera depends on the orientation of the piezofan and the heat sink, as shown in Fig. 2(b). The raw images were calculated (Insight 3G, TSI; interrogation window size 32 32 pixels) with 50% overlap of the interrogation window [21]. The detailed procedure for the analysis of the flow field was similar to previous work [11,22]. To investigate the thermal performance of the plate-fin heat sink cooled by the piezofan, the infrared thermography was employed for capturing the thermal image, as shown in Fig. 3. The infrared camera (FLIR A325) used in the experiments has an uncooled focal-plane array detector (320 240 pixels) and can receive a wavelength from 7.5 lm to 13 lm. The range of temperature measurement from 20 °C to 350 °C is covered with accuracy ±2%. The field of view is 25° 19°/0.4 m with thermal sensitivity 0.05 °C at 30 °C. The images were recorded and analyzed with software (ThermaCAM Research Pro 2.9). The detailed description of the heating system has been described in previous work [11]. The heating system (LW-4671, Long Win Science and Technology Corporation) connected to the heat sink includes a heating element, temperature-measuring module and DC power supply, providing a heating power of 3-W (±5%) to the heat sink. The heating element was completely wrapped with insulation material except in areas in contact with the base of the heat sink. The heating power of the system is
Q¼
kal AðT l T u Þ d
ð1Þ
where kal is the thermal conductivity of aluminum alloy, A is the cross-sectional area of the heating element, T u and T l are the temperatures of the upper and lower measuring points inside the heating element, respectively, and d is the distance between the two measuring points. The temperature measurement is a measurement module. Two platinum thermally sensitive resistors (PT-100) are used to measure the temperatures inside the heating element. A T-type thermocouple is used to measure cooling fluid temperature. The thermal resistance is obtained from the resistance and voltage
S.-L. Ma et al. / International Journal of Heat and Mass Transfer 90 (2015) 454–465
457
Fig. 4. Periodic velocity and vorticity distribution with, parallel orientation and in phase vibration (H⁄ 1.10 and D⁄ 0.67; Case A). (White arrow denoting the direction of main flow).
signals from the platinum thermally sensitive resistors and the thermocouple connected to the data acquisition device applying the equation
Rth ¼
T av e T 1 Q
ð2Þ
in which Q represents the heating power of the system and T av e and T 1 represent the average surface temperature of the heat-sink base plate and the temperature of the cooling fluid, respectively. The heating power and temperature monitoring are controlled with a computer. In the experiments with heat addition, the laser beam and laser sheet were orientated from top to bottom because of the heating system on the original optical path, resulting in a limited region of flow field obtained. 3. Results and discussion
x-direction. This cold flow was directed by the flow induced by the vibrating fans to form a jet down into the base plate (heat source) of the heat sink. The exchange between the accumulated flow in the heat sink and the downward jet of the cold flow occurred near the base plate of the heat sink; the mixed flow was directed upward to the ambient air. In contrast, for Case C with large D⁄ most dual fans induced entrainment flow shift from the left to the right side at the top of the heat sink, leading the decreased flow velocity into the base plate of the heat sink. When the cold flow from ambient air moved more into the heat sink, the thermal energy from the heat sink was removed better and the heat-transfer performance was improved [11]. The enhanced exchange of thermal energy between the hot flow and the cold flow for dual fans with D⁄ smaller than unity was more obvious than with large D⁄. As D⁄ decreased, the thermal resistance decreased, as shown in a subsequent section.
3.1. Effect of fan spacing 3.2. Effect of fan height The investigations of the effect of fan spacing were focused on a parallel orientation. The performance of dual fans with varied distance between them showed no interaction between the dual fans too far apart and interference between the dual fans too close together [23]. In this work, thus the condition with interruption between dual fans through considering that the double input power decreases the amplitude and produces less thrust was dismissed. Figs. 4 and 5 show the periodic (1/60 s) velocity vectors and vorticity contour of Cases A and C. For Case A with small D⁄, dual fans were capable of producing a larger space near the blade in the negative x-direction and entraining more cold flow from the ambient air on vibrating to the maximum deflection in the positive
The investigations of the effect of fan height (Cases E and G) were focused on a vertical orientation. For Case E, H⁄ is smaller than for Case G, meaning that the distance between dual fans and the base plate of the heat sink of Case E is smaller than that of Case G. The smaller distance between the dual fans and the base plate (heat source) causes improved cooling performance [9] in both parallel and vertical orientations [11,14,24,25], but few investigators considered the flow and structure interaction between dual fans and the heat sink. The periodic velocity and vorticity distribution for vibrating conditions in phase of Cases E and G are shown in Figs. 6 and 7, respectively. The obvious discrepancy
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Fig. 5. Periodic velocity and vorticity distribution with parallel orientation and in phase vibration (H⁄ 1.10 and D⁄ 0.93; Case C).
between Cases E and G is the intense velocity distribution and vortex structures between the dual fans and the heat sink. For Case E with small H⁄, the smaller distance effectively results in large velocities and vortex structures near the blade tip of dual fans, causing more flow to strike the base plate of the heat sink and so to enhance the cooling performance [9,12]. In contrast, the large velocities and vortex structures distributed far from the base plate of the heat sink for Case G resulted in a limited influence on the thermal energy with a decreased velocity of entrained cold flow in the heat sink. The tendency of periodic velocity and vorticity distribution for the vibration out of phase is the same as that found for the vibration in phase, as shown in Figs. 8 and 9. The cooling performances for Cases E and e are indicated to be better than those of Cases G and g because of the varied interactive behavior between the dual fans and the heat sink. The thermal resistance decreased with decreased H⁄ as shown in a subsequent section.
other near the blade tip moved into the heat sink, resulting in the generation of numerous small vortex structures. This highly turbulent effect including a directional change of cold flow and the dissipation of the vortices further augmented the transfer of thermal energy. As the dual fans moved away from each other to the maximum deflection in the positive and negative z-directions, the mixed flow within the heat sink moved along with the propagations of clockwise and counter-clockwise vortices and went outside. A region of low pressure was hence formed in the empty space between the two blade tips so that ambient air was pulled through the tunnel between the dual fans, moving into the heat sink. In summary, because of the collision between clockwise and counter-clockwise vortices induced by dual fans, and the formation of a suction effect induced by a region of low pressure in the heat sink, the cooling performance for vibration out of phase was much better than vibration in phase, resembling characteristics reported in Ref. [23].
3.3. Effect of fan phase 3.4. Effect of fan orientation The velocity and vorticity distribution of dual fans vibrating in phase (0°) and out of phase (180°) are shown in Figs. 6–9. For the vibration in phase (Figs. 6 and 7), the cold flow is entrained every half period and disappeared after a strong jet flow formed near the blade tip. Another jet flow formed after a half period carried cold flow into the heat sink and then disappeared again. In contrast, for the vibration out of phase (Figs. 8 and 9), the flow behavior differed from the characteristic for the vibration in phase. As the dual fans vibrated toward the center line between the dual fans, the cold flow was entrained from the left and right sides and the flow between the dual fans was then pushed upward. The clockwise and counter-clockwise vortices that collided with each
The effects of the orientation of a piezofan for both parallel and vertical orientations on the cooling performance have been investigated [11,14,24,25], but there are few investigations of the effects of orientation for a combination of dual fans and heat sink on the cooling performance. Further understanding of the detailed interactions of the heat transfer and flow structures between the dual fans and the heat sink plays an important role in applications as devices to cool electronics. Fig. 10(a) and (b) show the thermal resistance of the dual fans vibrating in phase and out of phase for varied D⁄ in the parallel orientation, and varied H⁄ in the vertical orientation. The conditions involved in Fig. 10(a) include Cases
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459
Fig. 6. Periodic velocity and vorticity distribution with vertical orientation and in phase vibration (H⁄ 0.50 and D⁄ 1.00; Case E).
A; B; b; C; c; D and d; the conditions involved in Fig. 10(b) include Cases E, e; F; G and g. For the parallel orientation, the thermal resistance decreased with decreasing D⁄ for dual fans vibrating both in phase and out of phase. For the dual fans vibrating in phase the thermal resistance decreased about 5% as D⁄ decreased from 1.06 to 0.67; for the dual fans vibrating out of phase, the minimum thermal resistance appeared as D⁄ = 0.80 because of the limited vibrational amplitude of the dual fans. For the vertical orientation, the thermal resistance decreased with decreasing H⁄ for the dual fans vibrating both in phase and out of phase. The thermal resistance of Case E (H⁄ = 0.50) was 10% less than that of Case G (H⁄ = 1.10) for dual fans vibrating both in phase and out of phase. Overall, the thermal resistance of dual fans vibrating out of phase was about 2% less than that of vibrating in phase for the parallel orientation, whereas the thermal resistance of the dual fans vibrating out of phase was about 3% less than that of vibrating in phase for the vertical orientation. The thermal resistance for the parallel orientation was 15% less than that in the vertical orientation. For both parallel and vertical orientations, the cooling performance of the dual fans vibrating out of phase was better than that in phase. The velocity and vorticity distributions captured with high-speed PIV were analyzed to reveal the influence of the fan orientation on the mechanism of the enhancement of heat transfer with the dual fans. For the case of a parallel orientation as shown in Figs. 4 and 5, a side cold flow was entrained by the dual fans and was split into two flow directions. One cold flow struck directly the outer surface of the heat sink; the cold flow moved along the direction perpendicular to the plane, increasing the convective heat transfer of the heat sink. The strong vortices outside
the heat sink were evidence of a fan-induced flow colliding on the surface of the heat sink. Another cold flow went through a small gap between the dual fans and the heat sink into the heat sink as the dual fans vibrated toward the center line. In addition to the limitation by the left- and right-side fins of the heat sink, ambient air sucked through dual fans moved downward and pressed the hot flow in the heat sink to move along the direction perpendicular out of plane. These two flows thus greatly increased the heat transfer of the heat sink. The entire flow circulation between ambient air, the dual fans and the heat sink benefited the ventilation and so the cooling performance. In contrast, for a vertical orientation as shown in Figs. 8 and 9, the characteristics were the same as that in the parallel orientation with the cold flow entrained by the dual fans and the suction effect due to the region of low pressure between the dual fans and the heat sink, but the fin of the heat sink was then in the path of air moving out in the positive and negative x-directions, so these two flows were limited by an enclosure formed in the heat sink. The trapped hot flow smoothly moved from the heat sink until a gap was created by the dual fans vibrating outwardly. For both parallel and vertical orientations, the various interactions of the flow structures were more obvious when H⁄ decreased. For the parallel orientation, both inner and outer sides of the fin of the heat sink were affected by the dual-fan-induced vortices, but for the vertical orientation the hot flow trapped in the heat sink and the cold flow outside the heat sink were unable to mix easily and smoothly. The cooling performance for the parallel orientation was thus much better than that of the vertical orientation because the heat transfer and ventilation benefited from the large surface of the heat sink for convection and effective circulation.
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Fig. 7. Periodic velocity and vorticity distribution with vertical orientation and in phase vibration (H⁄ 1.10 and D⁄ 1.00; Case G).
Fig. 8. Periodic velocity and vorticity distribution with vertical orientation and out of phase vibration (H⁄ 0.50 and D⁄ 1.00; Case e).
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Fig. 9. Periodic velocity and vorticity distribution of vertical orientation and out of phase vibration (H⁄ 1.10 and D⁄ 1.00; Case g).
Fig. 10. Thermal resistance of dual fans vibrating in phase and out of phase for (a) D⁄ of the parallel orientation, and (b) H⁄ of the vertical orientation.
3.5. Effect of hot flow and cold flow There has been little investigation of the effects of hot and cold flows on the combination of dual fans and a heat sink. Because of limitations of experiments to analyze the interaction between conditions of hot and cold flows, the characteristics of cold flow, without a heat source, and hot flow, with heat source, are considered separately in most investigations [11,17]. The interaction between hot and cold flows might be strongly affected and altered because of the buoyancy effect induced by natural convection. The velocity and vorticity distribution of hot flow in phase and hot flow out of phase for Case H are shown in Figs. 11 and 12, respectively; the
velocity and vorticity distribution of cold flow in phase and cold flow out of phase for Case H are shown in Figs. 13 and 14, respectively. Fig. 15 and 16 shows a time series of the horizontal velocity component (u) and the vertical velocity component (v ) for heights 5, 10 and 30 mm above the base plate along the center line with dual fans vibrating in phase and out of phase in Case H. For the case in the hot-flow condition, the region of interest is focused on the center of the heat sink to reveal the interaction between the dual fans and the heat sink, because of the greatly altered characteristic flow structure influenced by the hot- flow. Figs. 11 and 13 show that the velocity in the hot-flow condition was about half that in the cold-flow. Figs. 12 and 14 show that
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Fig. 11. Periodic velocity and vorticity distribution of hot flow and in phase vibration (H⁄ 1.10 and D⁄ 1.00; Case H).
Fig. 12. Periodic velocity and vorticity distribution of hot flow and out of phase vibration (H⁄ 1.10 and D⁄ 1.00; Case H).
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Fig. 13. Periodic velocity and vorticity distribution of cold flow and in phase vibration (H⁄ 1.10 and D⁄ 1.00; Case H).
Fig. 14. Periodic velocity and vorticity distribution of cold flow and out of phase vibration (H⁄ 1.10 and D⁄ 1.00; Case H).
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(a) D = 5 mm
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Fig. 15. Time series of the horizontal velocity component (u) for heights (a) 5, (b) 10 and (c) 30 mm in Case H.
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Fig. 16. Time series of the vertical velocity component (v) for heights (a) 5, (b) 10 and (c) 30 mm in Case H.
the velocity decreased about 30% for the dual fans vibrating out of phase in the hot-flow condition with a large difference in the flow motion compared with the cold-flow condition. Under the influence of a hot-flow, the downward main flow at 3/4 period became upward flow through natural convection as the dual fans vibrated outwardly, as shown in Fig. 12. In contrast, the main direction of the cold-flow was downward as the dual fans vibrated outwardly because of the suction induced by low-pressure induced suction, as shown in Fig. 14. The horizontal velocity component (u) in the
hot-flow condition was 10% and 20% larger than that in the cold-flow condition for heights 10 and 30 mm, respectively. The vertical velocity component (v) below height 30 mm of the hot-flow condition was smaller than the cold-flow condition because of the formation of a thermal boundary: the vertical velocity component (v) for both conditions of cold and hot flows increased as the height increased. At height 30 mm, the vertical velocity component (v ) of the dual fans vibrating out of phase in the hot flow was invariably positive, indicating that the flow on
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top of the heat sink went upward. This characteristic agrees with the discrepancies of the flow structures between cold and hot flows found in Figs. 11–14. In contrast, for dual fans vibrating out of phase, the vortex structures induced by those dual fans at the natural position increased the heat transfer by forced convection, as the dual fans vibrating outwardly to the flow field within the heat sinks were influenced by the vortex structures induced by the dual fans and natural convection. The effect of natural convection was thus much stronger than that of forced convection in the transition period of the dual fans vibrating from the natural position to the maximum deflection. The discrepancy between hot-flow and cold-flow for the dual fans vibrating, in general, contributes to the natural convection induced by the temperature gradient, which was stronger than the forced convection in the transition period. The influence of natural convection on the vortex structures in the heat sink and the entrained cold flow was indicated to be inevitable. The influence of the thermal fluid on the vortex interaction between the dual fans and the heat sink must therefore be considered as the influence of natural convection was shown to be negligible. 4. Conclusion The mechanisms of enhanced cooling performance on applying dual fans on a plate-fin heat sink with various parameters including the spacing, height, phase and orientation of the fan, was investigated experimentally with high-speed PIV and infrared thermometry. For the effect of the fan spacing, the width of the tunnel between the dual fans was capable of entraining more cold flow with decreased D⁄. For the effect of the fan height, the formation of the turbulent flow structure near the heat source with decreased H⁄ strongly promoted the heat transfer. The thermal resistance, decreasing with decreased D⁄ (5%), and decreasing H⁄ (10%) of the fans vibrating out of phase was invariably superior to the fans vibrating in phase (3%). The thermal resistance in the parallel orientation was 22–27% less than in the vertical orientation because of the unobstructed ventilation, but decreased in the vertical orientation because the entrained cold flow into the heat sink was limited. The velocity induced by the dual fans decreased about 30–50% because the buoyancy effect caused by natural convection annihilated the cold flow to move downward into the heat sink. The effect of the thermal fluid on the interaction between the vortex and the heat sink must be considered. The revealed mechanisms of thermal effects and flow interaction for the combination of the dual piezofans and the heat sink shed light on the design and development of electronic cooling systems. Conflict of interest None declared. Acknowledgment National Science Council of the Taiwan (Republic of China) for financially supported this research under contract NSC 102-2815-C-002-047-E.
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