Experimental set-up for a multiple-vibrating fan cooling system

The only difference between the fans used in the multiple-vibrating fan cooling system and the piezoelectric fans used in all previous studies is the magnet on the tip of the fans. However, the magnet is also the key component, allowing the five fans to vibrate simultaneously. Fig.4-11 shows the multiple-vibrating fan cooling system. A hole in the sheet was drilled by a CNC machine, and the magnet was then placed on the tip without adhesive or glue. Further, the magnet can be easily removed and replaced by another magnet. The geometries of the vibrating fans and the magnets used in this experiment are shown in Table 6.

Each fan is fixed by a clamp that can be moved along a track. Thus, the fans can be adjusted to obtain a suitable magnet position. The multiple-vibrating fan system

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can be integrated with a heat sink to yield a direct cooling system. A 20-W heat source is attached to the backside of the aluminum heat sink. An insulation block is used to cover the heat source, and the cooling system is placed in a large plastic box to prevent perturbation from external convective currents. An alternating voltage from 50 V to 100 V is applied to the PZT plate. The fan amplitude was observed with a high-speed camera (JVC-HM550), and the temperature was detected by the five thermocouples attached to the surface of the heat sink.

CHAPTER 5 RESULTS AND DISCUSSION

An innovative multiple-vibrating fan cooling system has been developed by using both the piezoelectric force and the magnetic force. In this design, a piezoelectric plate can be used to simultaneously actuate five vibrating fans. Further, the multiple-vibrating fan system can be directly embedded into a heat sink. The performance of the system is demonstrated by the total amplitude of the system and the temperature drop of the heat sink. The results show its satisfactory performance, including low power consumption. However, the multiple-vibrating fan cooling system is based on a single piezoelectric fan cooling system. Thus, the performance and the arrangement of a single piezoelectric fan should be investigated first.

5.1 Performance of the single piezoelectric fan cooling system

To choose the best arrangement for the single piezoelectric fan cooling system, different orientations of the single piezoelectric fan cooling system were investigated.

The operating frequency was between 15 Hz and 40 Hz, while the operating voltage was between 50 V and 100 V. The simulation model was also used to examine the correlation between the flow field induced by the piezoelectric fan and the temperature drop on the inner fin surface. Additionally, the power consumption of the piezoelectric fan was investigated and the important dimensionless parameters were used to more efficiently analyze the experimental data.

5.1.1 The arrangement of the system

The arrangement of the piezoelectric fan may influence the generated flow field and the performance of the cooling system. To find the best arrangement of the piezoelectric fan, the vertical orientation (λ), the horizontal orientation ( Ê_¸¹º/Ë_ÌÍÎ), and the inclined angle of the cooling system (θ) were considered as shown in Fig.5-1.

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According to previous studies [20], the gap between the fan and the heated surface strongly affects the performance of the cooling system. In this study, three different vertical orientations, where λ was equal to 5 mm, 10 mm, and 15 mm, were tested under the conditions of X /L =0.5 and θ=0°. As shown in Fig.5-2, the results indicate that the temperature drop on both the bottom and the side fin surfaces increases obviously when the piezoelectric fan approaches the bottom fin surface.

Thus, λ=5 mm was chosen as the vertical orientation. Tests of the inclined angle θ (at 0°, 15°, 30°, 45°, 60°, 75°, and 90°) were completed under the conditions of X /L =0.5 and λ=5 mm. The performance of the cooling system with θ=0° was superior to those of other inclined angles.

Fig.5-3 shows the dimensionless PZT-convection number ?M versus the horizontal orientation X /L under the condition of λ=5 mm and θ=0°. The experimental data demonstrates that the heat dissipation rate reaches its peak value between X /L =0.5 and X /L =0.6. Thus, the optimal arrangement of the piezoelectric fan embedded into a finned heat sink can be determined according to the experimental results. Further, these results also can be applied to the arrangement of the magnetic fans in the multiple-vibrating fan cooling system.

5.1.2 The flow field of the system

A three-dimensional model was built to investigate the flow field of the cooling system. Fig.5-4 shows that the maximum velocity near the bottom surface and the side surface both occur at approximately X /L = 0.3-0.5. Theoretically, a higher velocity leads to a larger temperature drop. Fig.5-2 shows good agreement with the anticipated phenomenon. The maximum temperature drop occurs at the position of maximum velocity. The three-dimensional simulation model may help us to decide the position of the heat source and the arrangement of the piezoelectric fan.

5.1.3 The power consumption of the system

In the fan cooling system, the power consumption of the PZT fan is also an important issue. Fig.5-5 shows the relationship between the power consumption and M for different vibrational frequencies of the piezoelectric fan. The power consumption is 0.511 W and M is 1.4 when the operating frequency is 24 Hz.

However, the power consumption decreases to 0.022 W while M increases to 1.53 at an operating frequency of 30 Hz, because 30 Hz is the resonant frequency of the fan, allowing the PZT fan to vibrate at its maximum amplitude [16]. According to this result, the power consumption should be considered first when selecting the operating frequency. An inappropriate operating frequency will increase the power consumption by a factor of twenty times.

5.1.4 The amplitude of the system

The fan amplitude plays an important role in the cooling ability of a piezoelectric fan. However, the operating frequency is also a factor that influences the cooling ability. Fig.5-6 [29] shows how M depends on the fan amplitude at different operating frequencies. In this figure, M increases as the amplitude and the operating frequency are increased. According to the figure, the influence of the amplitude on M is clearly larger than that of the operating frequency, especially at low fan amplitude operating conditions. The M can be increased from 1.05 to 1.3 at 25 Hz by increasing the amplitude from 4 mm to 9 mm. However, the M is only increased from 1.05 to 1.08 at a 5-mm amplitude by increasing the operating frequency from 25 Hz to 30 Hz.

Because a higher operating frequency only increases the cooling ability slightly when the fan amplitude is low, it is unnecessary to increase the operating frequency.

Thus, the resonant frequency of the piezoelectric fans was designed in a low range

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between 20 Hz and 35 Hz.

5.1.5 The correlation between ÏÐ_ÑÒÓ and μ_ÑÒÓ of the system

According to Equation (3.8), the operating frequency and the fan amplitude are directly proportional to the Re . The two parameters can be easily observed, and they also play important roles in enhancing the Nu . However, the Nu cannot be calculated directly from the frequency and the fan amplitude, so the correlation between the Re and Nu may help in the estimation of the Nu conveniently. Fig.5-7 shows Nu and Re of the experimental data at different amplitudes and frequencies. When the Re is below 700, the Nu number increases almost linearly with the increasing Re number. Further, an approximate curve can be derived from these data. Thus, the Nu of the single piezoelectric fan cooling system can be inferred by its operating frequency and the fan amplitude.

5.1.6 The relationship between »_Ñ and Ri of the system

The dimensionless parameter is defined as Ri ?Gr/Re ^W) and represents the importance of natural convection relative to forced convection. When Ri<1, natural convection is negligible. However, when Ri>10, forced convection is negligible [45].

Fig.5-8 shows the dimensionless analysis of the experimental data at different frequencies and amplitudes. In Fig.5-8, two approximate curves are drawn according to the experimental data. M can be correlated with Ri and then generalized as the following equations. These equations mainly aid in judging which term should be considered seriously by a specific number. The equations may be used to analyze and assess the performance of the single piezoelectric fan cooling system more conveniently.

5.2 Performance of the multiple-vibrating fan cooling system cooling ability and usability when using the multiple-vibrating fan cooling system.

5.2.1 Effect of fan geometries on the resonant frequency

The thickness of the magnetic fan has a large influence on the vibrating amplitude, as well as the resonant frequency. As shown in Fig.5-9 and Fig. 5-10, the resonant frequency of case 1 (see Table. 6) is 36.4 Hz, which is higher than that of case 2 (21.8 Hz). With the same length, width, and material, the resonant frequency of the thick magnetic fan is higher than that of the thin one. This higher vibrational frequency is beneficial for improving the amount of induced flow. However, the thin magnetic fan may perform better at larger amplitudes, because it can be bent more easily, which can also improve the amount of induced flow. On the other hand, the length of the magnetic fan also affects the resonant frequency. Fig.5-10 and Fig.5-11 give a comparison between case 2 and case 3. The resonant frequency of case 2 is 21.8 Hz, whereas the resonant frequency of case 3 is 30 Hz. The results show a tendency toward Equation (2.7) and Equation (2.8). The resonant frequency of the fan decreases as the length of the fan is increased.

5.2.2 Effect of the distance between the magnets on the resonant frequency In this system, the repulsive magnetic force can be represented as a spring in

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compression, which induces an additional stiffness (K _' that influences the resonant frequency of the fans. To investigate the effect of this force, a PZT fan and a magnetic fan were set as shown in Fig.5-12(a). According to Equation (2.11), the amount of additional stiffness depends on the distance between the two magnets. The total stiffness (K_"++) should be increased when the distance is decreased. Thus, the resonant frequency of the magnetic fan may increase due to the increased total stiffness. Fig.5-13 shows that the resonant frequency of the magnetic fan is increased as the distance between the magnets is decreased. Fig.5-12(b) shows the resonant frequency of the middle magnetic fan affected by the two magnets, and Fig.5-14 shows the experimental results of the system. The two magnets have a larger influence on the resonant frequency of the system. The result proves that Equation (2.16) is not only applicable for a vibrating fan affected by one magnet but also for a vibrating fan affected by two magnets. Thus, Equation (2.16) can be used to assess the resonant frequency of the multiple-vibrating fan system.

5.2.3 Performance of the multiple-vibrating fan system

To demonstrate the performance of the multiple-vibrating fan system, the “total amplitude” was defined as the sum of the amplitude of each vibrating fan. A comparison of the total amplitude of the multiple-vibrating fan system (case 4) versus a single PZT fan under different voltages is shown in Fig.5-15. The amplitude of the single PZT fan was 8 mm at an operating frequency of 36.5 Hz and a voltage of 50 V.

However, the total amplitude of the multiple-vibrating fan system was 31 mm under an operating frequency of 33.4 Hz and a voltage of 50 V. The total amplitude of the multiple-vibrating fan system was approximately 400% more than that of the single PZT fan. Further, because the operating frequency of the multiple-vibrating fan system was decreased from 36.5 Hz to 33.4 Hz, the power consumption also

decreased. Fig.5-16 shows that the power consumption of the multiple-vibrating fan system is only 0.03 W under operating conditions of 50 V and 33.4 Hz, while the power consumption of a single PZT fan is 0.033 W under operating conditions of 50 V and 36.5 Hz.

5.2.4 Cooling ability of the multiple-vibrating fan system

The thermal performance of the multiple-vibrating fan cooling system can be demonstrated by checking the temperature drop of the thermocouples adhered to the fin surface, as shown in Fig.5-17. The temperature drop of the thermocouples is defined as the difference between the average temperature of the five thermocouples at the initial condition and the average temperature of the five thermocouples at steady state. Fig.5-18 shows the cooling ability of the multiple-vibrating fan cooling system.

The multiple-vibrating fan cooling system is turned on at t=0 sec. The system was operated under the conditions of 36.4 Hz and 50 V. When t=0 sec, the average temperature of the five thermocouples was 67 ℃ with a 20 ℃ room temperature.

When λ was 8 mm, the average temperature dropped from 67℃ to 55 ℃. However, when λ was 2 mm, the temperature dropped from 67 ℃ to 50 ℃. The results show that the temperature drop reaches 17 ℃ under the condition of 36.4 Hz and 50 V with 0.03 W of power consumption.

5.3 Improvement and application of the multiple-vibrating fan system

The methods used to further improve the performance of both the multiple-vibrating fan system and the cooling module combined with the multiple-vibrating fan system with a finned heat sink will be proposed in this section.

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5.3.1 Frequency modulation of a magnetic fan

The operation of the multiple-vibrating fan system completely takes advantage of the resonant effect such that the most important topic is the approach to adjusting the resonant frequencies of the fans to be uniform. According to our experiments, the difference between the resonant frequencies of the magnetic fans is very small (about 1~2Hz), hence the influence of the resonant frequencies on amplitude is not self-evident. For example, in Fig.5-9, the amplitudes of the No.2 fan and No.3 fan are nearly equivalent. However, in this figure, the amplitude of the No.1 fan is very small because it is operated at a non-resonant frequency. Thus, the use of a frequency frequency modulator weighing 0.715 g. The resonant frequency of the magnetic fan is 35.4 Hz when the graduation is 0 mm. However, the resonant frequency of the magnetic fan is decreased to 29.6 Hz when the graduation is 40 mm. Using the frequency modulator, the resonant frequency of a magnetic fan and a piezoelectric fan can be adjusted to the desired value. Further, the resonant frequencies of these fans can be adjusted to the same value. Thus, the maximum performance of the multiple-vibrating fan system can be reached.

5.3.2 Double-sided magnetic fan

According to Fig.1-11 and Fig.1-12, the air flow generated by a piezoelectric fan only results near the tip of the piezoelectric fan. However, the overall surface of a heat

sink must be cooled in a cooling system. A double-sided magnetic fan is able to generate air flow at the fan tip and the fan bottom simultaneously. Fig.5-21 shows an actual view of a double-sided magnetic fan. The resonant frequencies of the fan and its bottom beam are the same. Thus, both parts are able to vibrate at the same operating frequency. By using this fan, the air flow can be generated at the top side and the bottom side at the same time, as shown in Fig.5-22. This fan also can be used for the multiple-vibrating fan system, as shown in Fig.5-23.

5.3.3 An application of a multiple-vibrating fan cooling system

Recently, the required heat dissipation of CPUs has been increasingly reduced.

Table 10 demonstrates the specifications of the INTEL Pentium 4 [46] and the INTEL i7 series. The CPU-TDP means that the maximum wasted heat should be removed from the CPU. The CPU-TDP of Pentium 4 HT-672 is on the order of 115 W.

However, the CPU-TDP of i7-3770K is on the order of 75 W, hence the rotary fan for high heat dissipation from the CPU may no longer be used in the future. Thus, a low power consumption cooling system for the new CPUs should be investigated as a substitute for the rotary fan cooling system. Additionally, a rotary fan module increases the volume of the heat sink by approximately 20% to 30%, as shown in Fig.5-24.

The multiple-vibrating fan cooling system can be used for CPU cooling by directly inserting it into the heat sink. Fig.5-25 shows a plate that can be used to fix the different vibrating fans. The clamp end is fixed by a screw such that the type of vibrating fan can be changed easily. Further, the plate can be embedded into the heat sink directly, as shown in Fig.5-26. Fig.5-27 shows the multiple-vibrating fan cooling system as it was applied to cool a CPU.

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CHAPTER 6 CONCLUSIONS

An innovative multiple-vibrating fan cooling system was developed by using both the piezoelectric force and the magnetic force. The system can be placed directly into a heat sink without occupying extra space. The effects of the vibrating frequency, the operating voltage, the geometries of the fans, the distance between the fans and the heat sink, and the power consumption have been analyzed to assess the performance of the cooling system. Further, a method for evaluating the performance of a single piezoelectric fan cooling system also has been presented. The method can be used to assess the performance of the multiple-vibrating fan cooling system. The major conclusions of this study are summarized below:

(1) Using a three-dimensional, transitional model to account for the flow field induced by the piezoelectric fan, the maximum velocity was found to usually occur at the position 20 mm before the fan tip; the results show sufficient agreement between the temperature drop and the flow field.

(2) The orientation of the piezoelectric fan strongly affects the M . The M of the system, which operates at a frequency of 30 Hz and an amplitude of 18 mm, reaches 2.3 (X /L =0.5, λ=5 mm, θ=0°).

(3) Although the operating frequency does not affect the M as much as the amplitude does in this cooling system, it still regulates the power consumption of the PZT fan. The use of a 6 Hz fan frequency, which is offset from the resonant frequency, increases the power consumption by twenty times. Thus, the resonant frequency should be chosen for all conditions.

(4) In the case of ( X /L =0.5, λ=5 mm, θ=0°), the Nu almost linearly increases with an increasing Re . According to this correlation, the Nu can be derived from the amplitude and the operating frequency that determine the Re .

Nu J 2.9297 V 0.0387Re when Re × 700

(5) The M can be correlated with the dimensionless ratio of the natural convection to the forced convection (Ri) and then generalized as the following equations:

M J 1.389 X 0.2762 ln Ri Ri × 1 M J 1.268 X 0.05663 ln Ri Ri Ú 10

The major findings of the multiple-vibrating fan cooling system are given as follows:

(6) The fan geometries have a large influence on the resonant frequency of the magnetic fan. The resonant frequency can be decreased from 36.4 Hz to 21.8 Hz by reducing the thickness of the magnetic fan from 0.3 mm to 0.2 mm. However, the additional stiffness gained by the addition of the magnet only affects the resonant frequency of the magnetic fan slightly when the value of the distance between the two magnets (x) is larger than 15 mm. For example, the resonant frequency of the magnetic fan (case 1) decreased from 33.8 Hz to 30.4 Hz when X was increased from 13.7 mm to 25.1 mm.

(7) In the same operating conditions, the multiple-vibrating fan system does not consume more power than a single PZT fan does. However, the total amplitude of the multiple-vibrating fan system reaches 56 mm, whereas the amplitude of the

(7) In the same operating conditions, the multiple-vibrating fan system does not consume more power than a single PZT fan does. However, the total amplitude of the multiple-vibrating fan system reaches 56 mm, whereas the amplitude of the