Piezoelectric Fan Cooling: A Novel High Reliability ... - IEEE Xplore

IEEE TRANSACTIONS ON INDUSTRIAL ELECTRONICS, VOL. 60, NO. 11, NOVEMBER 2013

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Piezoelectric Fan Cooling: A Novel High Reliability Electric Machine Thermal Management Solution Gareth M. Gilson, Stephen J. Pickering, David B. Hann, and Chris Gerada

Abstract—Electric machine thermal management is critical for the correct operation of high power density electrical machines. This is, however, challenging to achieve in safety-critical applications where the reliability of the cooling system needs to be significantly higher than conventional solutions. Piezoelectric fans are presented here as a novel fault-tolerant forced cooling convective system for electric machines. Particle image velocimetry techniques in conjunction with infrared thermal measurements were implemented to map and quantify the flow fields generated by such a cooling arrangement as well as to determine the effective cooling enhancement. This paper also outlines the main design variables for such a system and highlights the main considerations to be accounted for to optimize the cooling potential. For the specific machine presented in this paper, the optimal fin/fan geometry resulted in mean flows in excess of 2.41 m/s and turbulence values in excess of 2.38 m/s, which resulted in an average convective heat transfer coefficient enhancement of 364.8% on the fin base and a further enhancement of 53.6% on each of the fin side walls. This, in turn, led to a 61.8% reduction in the electric machine heat sink cooling mass. Index Terms—Fault-tolerant machines, fluid flow visualization/ measurement, heat transfer enhancement, machine cooling, piezoelectric fans, thermal management.

I. I NTRODUCTION

E

LECTRICAL MACHINES for high performance applications, such as in the aerospace industry, often operate close to the allowable thermal limits due to high power density requirements. The torque or power density which can be achieved in electrical machines is mainly thermally limited. Improving the cooling efficiency generally will lead to an improved machine performance. Different thermal management solutions for electrical machines have been proposed and implemented along the years [1]–[4]. Fluid cooled systems such as forced air cooling [5]–[10], water cooling [11]–[16], and oil cooling [17]–[19], to name a few, have proved to be superior

Manuscript received June 19, 2012; revised August 29, 2012; accepted September 24, 2012. Date of publication October 11, 2012; date of current version June 6, 2013. G. M. Gilson was with the Division of Energy and Sustainability, The University of Nottingham, Nottingham NG7 2RD, U.K. He is now with Cummins Generator Technologies, Stamford PE9 2NB, U.K. (e-mail: gareth.gilson@ cummins.com). S. J. Pickering and D. B. Hann are with the Division of Energy and Sustainability, The University of Nottingham, Nottingham NG7 2RD, U.K. (e-mail: [email protected]; [email protected]). C. Gerada is with the Division of Electrical Systems and Optics, The University of Nottingham, Nottingham NG7 2RD, U.K. (e-mail: chris.gerada@ nottingham.ac.uk). Color versions of one or more of the figures in this paper are available online at http://ieeexplore.ieee.org. Digital Object Identifier 10.1109/TIE.2012.2224081

in many applications; however, they usually imply a reliability and maintenance penalty [20]–[23]. Rotary fans are regularly utilized in industrial applications [10], [24]–[29] to increase the bulk fluidic motion and enhance heat transfer. However, the weight, space, cost, and power density, together with the low reliability (stalled scenario—w = 0 r/min), of such traditional fans may, however, not be feasible for safety-critical applications. Furthermore, in applications where both liquid and the forced air cooling are not possible due to the high reliability demand, naturally ventilated machines are often used [7], [30], [31]. This is typically the case with aerospace actuation motors where the required reliability level does not allow for conventional fans and any liquid cooling would carry too much of an overhead in terms of system weight and reliability. This paper introduces piezoelectric fans as a potential faulttolerant forced cooling convective system (FCCS), capable of enhancing electric machine housing heat transfer. Piezoelectric fans have been used in a number of applications, mainly in electronics cooling [32]–[37]; however, their application to electrical machine cooling has not been previously considered. This paper will look at the potential of adopting such fans in conjunction with machine housings to enhance heat transfer. The work presented will investigate the effects of the piezoelectric fan vibrational amplitude, separation distance, fin length, fin spacing, and fan orientation through experimental means. The application considered is an actuator motor for primary flight control, whereby, traditionally, such machines are naturally ventilated. Comprehensive particle image velocimetry (PIV) and thermal studies have been conducted to evaluate the effective cooling enhancement of the proposed FCCS. The experimental work in this paper considers a section of the machine casing, and the effects of the piezoelectric fan geometry, aspect ratio, fan orientation, and fin geometry are investigated on each of the heat sink surfaces independently: the fin base (FB) and the fin side walls (FSW). The next section will describe the concept of piezoelectric fans and their possible adoption for electrical machines. Section III will then describe the experimental setup adopted to characterize the flow and determine the effective heat transfer coefficients. Section IV will then investigate the flow characteristic excited by such fans to better understand the flow regimes and the eventual placement and geometry of these fins. Section IV will look at a study to determine the effective surface heat transfer coefficient for electrical machines adopting the proposed thermal management system. This will be followed by a case study and conclusions.

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Fig. 1. Radial finned electric machine housing.

Fig. 2. Operation of a piezoelectric fan.

II. P IEZOELECTRIC FANS FOR E LECTRICAL M ACHINE T HERMAL M ANAGEMENT This section presents the concept of adopting piezoelectric fans for electrical machines and reviews the main aspects to be considered in such a thermal management arrangement. A review of work associated with piezoelectric fan-based cooling systems will also be described. Conventional naturally cooled machines are usually of the form shown in Fig. 1 with a casing having extended surfaces, better known as fins. The implementation of piezoelectric fans as an alternative FCCS has great potential; however, as yet, they are still untested, and their application to electrical machine cooling has not been reported. In essence, a piezoelectric fan is fabricated by the bonding of a piezoelectric patch (typically PZT) to a lightweight cantilever beam (typically Mylar). When subjected to an electric field, the randomly oriented ions go into alignment, which, in turn, cause a slight deformation of the piezoelectric patch. As the generated displacement is small, piezoelectric patches are commonly compounded with lightweight cantilever beams to further amplify the displacement. Furthermore, on applying an alternating current, the blade vibrates back and forth with the same frequency as the supplied alternating current, thus generating an air flow (see Fig. 2). Papers highlighting the flow characterization of an oscillating piezoelectric fan are infrequent, while papers measuring the 2-D or 3-D flow velocities are virtually nonexistent. This is very important to identify the placement of the fins with respect to the electrical machine housing and the relative geometrical dimensions of the fan and housing. The first detailed flow visualization experiments to be performed were those by Purdue

University [35], [38]. Acikalin et al. [33] indicated that the time-averaged flow field generated is highly turbulent but symmetric about the fan and that the maximum fluid rejection velocity recorded was approximately 0.3 m/s. On the other hand, Wait et al. [39] considered the performance of piezoelectric fans operating at higher resonance modes (i.e., operating above the resonance frequency). Flow visualization experiments were performed on a number of commercially available piezoelectric fans of varying length to better understand the transient and the steady-state fluidic motion generated by piezoelectric fans. However, it was concluded that certain advantages of the piezoelectric fan operating at higher resonance modes are offset by increased power consumption and decreased fluid flow. Thermal-oriented papers such as that of Acikalin et al. [32] and Kimber et al. [40], discussing the cooling capabilities of piezoelectric fans over natural convection, have also been published. It was reported that a convective heat transfer coefficient enhancement (%hInc. ) exceeding 375% and 444% respectively relative to a purely natural convection state is achievable on a vertical flat surface. Acikalin et al. [34] further performed comparative studies between an oscillating piezoelectric fan and two commercially available axial fans. Results indicate that, to achieve the same cooling performance, the tested axial fans consumed approximately ten times more power than the piezoelectric fan. This has been attributed to the mechanical losses present in the axial fans (bearing loss) being much larger than any losses present in the operation of the piezoelectric fan. Furthermore, it was shown that the heat source was cooled by more than 25 ◦ C with the piezoelectric fan when compared to a natural convection heat sink of the same volume as that of the tested piezoactuator system. However, no studies relating to the cooling enhancement generated by piezoelectric fans on electrical machines nor on general finned surfaces have been conducted. The suggested cooling mechanism in this paper is the combination of both a passive (fins) and an active (piezoelectric fan array) cooling system as shown in Fig. 3. III. E XPERIMENTAL S ETUP The piezoelectric fan used and tested throughout the duration of this study was obtained from Piezo Systems, INC., and has a total length of 76.7 mm, a width of 12.7 mm, a thickness of less than a millimeter, and a weight of 2.80 g. On applying the maximum input voltage (VP z = 115 V) and operated at the natural frequency (f = 60 Hz), the maximum vibrational envelope reaches 25.4 mm. As this paper details both the fluid flow characterization and the thermal enhancement capabilities of piezoelectric fans, two distinct experimental rigs were commissioned. A. Fluid Flow Rig The test section comprises a movable vertical straight-finned glass heat sink measuring 200 mm × 100 mm. The test section is considered as one of many repeating unit cells making up the electric machine finned housing. The literature [41] has

GILSON et al.: PIEZOELECTRIC FAN COOLING: ELECTRIC MACHINE THERMAL MANAGEMENT SOLUTION

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(VX , VY ), the resultant of the mean vectors (VMean ), and the RMS turbulent velocity (VRMS ), are given NI

VX =

VMean

VY

1

NI = (VX )2 + (VY )2 NI

VarX =

VX2

1

VarY = VRMS =

indicated that the electric motor housing curvature may be negated as results are similar to a straight-finned geometrical heat sink. A contact-free nonintrusive optical measurement technique better known as PIV capable of tracking tracer particles (being affected locally by the fluid flow) in a 2-D plane was set up (see Fig. 4). The utilized time-resolved PIV system comprised a dual pulsed Nd:YLF Litron laser with a 527-nm wavelength in conjunction with a SpeedSense 9060 Dantec 1 MP—1280 × 800 pixel resolution high speed camera with accompanying Tamron micro-lenses. Process control and initial data acquisition was provided by Dantec Dynamics v2.30 commercial PIV software, while image processing and data manipulation was performed in MATLAB by means of user-generated scripts. An adaptive correlation technique was implemented using an interrogation area measuring 32 × 32 pixels with a 50% overlap. No filters were applied however; a central difference technique with two refinement steps was adopted. A moving average validation was further applied so as to iron out any erroneous vectors caused by uncoupled particle pairs. Flow measurements were taken on two distinct planes, namely, the fin base and the fin side wall of a vertical straightfinned heat sink (see Fig. 4), and the effects of vibrational amplitude (AP z ), separation distance (G), fin length (L), and fan orientation were analyzed. The correlations utilized in the evaluation of the flow characteristics, particularly the x- and y-components of velocity

1

VY2

NI

(2) (3)

NI NI

Fig. 3. Artist impression of FCCS. (a) Horizontally mounted piezofan array. (b) Vertically mounted piezofan array.

(1)

NI NI

VY =

VX

1

− (VX )2

(4)

− (VY )2

(5)

(VarX ) + (VarY )

(6)

where NI = 500 and refers to the number of image pairs captured by the camera and “Var” refers to the variance. The above formulation may be explained further in the literature such as [42] and [43]. The maximum uncertainty in the attained 2-D mean and turbulent velocities was quoted at 5% by Westerweel [44], [45].

B. Thermal Rig Corresponding thermal measurements were taken on both the FB and FSW. A thermally insulated fin base heating element assembly unit, together with a similar fin side wall heating assembly arrangement and an adjustable infrared (IR) transparent dummy fin, made up the vertical finned heat sink unit cell (see Fig. 5). The FB and FSW heating element assembly units each consisted of a 0.021-mm-thick carbon fiber mat securely clamped between two vertical copper bus bars (one at each end of the mat). Both surfaces were independently fed from a DC voltage source. This assembly ensured a constant heat flux surface over the FB and FSW surfaces. A three-axis positioning system was further utilized to locate the piezoelectric fan in relation to the heat sink. The entire setup was placed inside a significantly sized enclosure so as to eradicate any atmospheric and stray fluctuations in the test section region. K-type thermocouples were used to measure the surrounding ambient air (T∞ ) together with the conduction loss from the back of the heat sink. On the other hand, 2-D local FB and FSW wall temperatures were measured through an IR camera, and the local convective heat transfer coefficient (h) and percentage convective heat transfer enhancement over natural convection (%hInc. ), together with the equivalent heat sink thermal resistance (REquivalent ), were calculated for the various tested fan/fin geometries according to (7)–(9), respectively.

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Fig. 5.

Graphical representation of thermal setup.

TABLE I P IEZOELECTRIC FAN VOLTAGE C ONVERSION M EASUREMENTS

Fig. 4. Flow measurement facility. (b) Fin base (FB) measurement plane. (c) Fin side wall (FSW) measurement plane.

A local heat flux balance (7) is used to determine the natural convective heat transfer coefficient (hNC ) as well as the forced (hFC ) convection heat transfer coefficient generated by the oscillating piezoelectric fan. The electrically generated heat flux over the heater surface is uniform and is dependent on the supply voltage (Vs ), current (Is ), and the heater surface area (As ). On the other hand, the radiation loss is a function of the surface emissivity (ε), Stefan–Boltzmann’s constant (σ), and the temperature difference, while the conduction loss is dependent on the local temperature difference and the RLoss , which is derived from the compounding insulation thickness. The equivalent heat sink thermal resistance REquivalent is dependent on the relative FB thermal resistance (rFB ) and the two relative FSW thermal resistances (rFSW ) and is evaluated on

the bases that these resistances are in parallel. This correlation is given in (9) and has units (K · m2 /W) ΔT VS IS 4 − εσ TS4 −T∞ − RLoss AS (7) h= ΔT hFC −hNC %hInc = ×100% (8) hNC rFB .rFSW . (9) REquivalent =AFB . rFSW +2rFB Evaluating the latter parameter ensures that the attained equivalent thermal resistance values are a function of the electric machine length (L) or, better, the exposed machine housing fin base area (AFB ). On utilizing the uncertainty error formulation established by Kline and McClintock [46], the maximum uncertainty in REquivalent was equivalent to 7.7%, while that for %hInc. was equivalent to 12%. IV. F LUID F LOW C HARACTERIZATION Preliminary tests ascertained that the fan amplitude (AP −P ) is linearly related to the applied voltage (VP z ). The Reynolds number (ReP z ) calculation is based on the equivalent hydraulic diameter, AP −P , and w and is described in further detail in [47] (see Table I).


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TABLE II P EAK V ELOCITY M EASUREMENTS: G = 10 mm. (a) FB. (b) FSW

Fig. 6. Three-dimensional flow characterization of the respective mean (LHS) and turbulent velocities (RHS) for VP z = 115 V. (a) G = 15 mm. (b) G = 10 mm. (c) G = 5 mm.

The selected input voltage measurement range ensures that a large Reynolds number range is investigated, which, in turn, should highlight the different flow regimes (laminar, transition, and turbulent). While there is no conclusive way in establishing the exact transition point, this transition might be visible in the detailed PIV imagery presented. Operating the fan at peak amplitude generates a flow disturbance in excess of 60 mm in the oscillation direction and a further 40 mm in the transverse direction (see Fig. 6). Furthermore, for all the applied fan voltages, the magnitude of the upward traveling flow is equivalent to that traveling downward, indicating flow symmetry. Table II(a) lists the velocity values attained on the FB, while Table II(b) lists the velocity values attained on the FSW when the fan was operated under different voltages.

An increase in VP zs sees an increase in all velocity components, for both the FB and FSW. A reduction in G yields a significant increase in the principal velocity VY together with an increase in the FB coverage area. Numerically, the maximum VY increases from 1.430 m/s when G = 15 mm, to 1.960 m/s when G = 10 mm, and to 2.488 m/s when G = 5 mm (see Fig. 6). Flow symmetry is once again attained when monitoring the FSW flow characteristics. However, results indicate that the fan is not as efficient in cooling the FSW as the FB (see Table II). At minimal separation distances, very little air is available to the fan (due to the presence of the FSW’s) which, in turn, results in a lower fan effectiveness. However, with an increase in G, more air is readily available to the fan, which, in turn, results in an increased fan cooling effectiveness and, hence, in large mean velocities on the FSW’s. Numerically, the maximum VMean decreases from 0.892 m/s when G = 15 mm to 0.834 m/s when G = 5 mm (see Fig. 4). PIV studies have highlighted the complexity of the unsteady highly turbulent nature of the flow generated by an oscillating fan for the various fin/fan geometrical configurations. Mean flows in excess of 2.48 m/s and turbulence values in excess of 2 m/s have been recorded for the optimum fin/fan geometry.

V. T HERMAL C HARACTERIZATION Fig. 7 shows the 2-D cooling plots generated by the piezoelectric fan on the FB. For the considered finned setup, with a fin spacing of S = 40 mm, a central vertical elliptical symmetric pattern spanning in excess of 80 mm, highly oriented in the principal oscillation direction (VY ), is generated. This holds true for all tested amplitudes and separation distances. Results correlate directly to PIV tests and indicate that the best geometrical cooling configuration is that of the following: VP z = 115 V and G = 5 mm. From the presented results, it can be concluded that a reduction in the vibrational envelope (AP z ) and an increase in the separation distance (G) both lead to a reduction in the overall cooling capabilities of the oscillating

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TABLE III FAN C OOLING C APABILITIES: VP z = 115, G = 10 mm, AND S = 40 mm

Fig. 7. Evaluation of %hInc. on the FB surface: VP z = 115 V. (a) G = 15 mm. (b) G = 10 mm. (c) G = 5 mm.

Fig. 8. Evaluation of %hInc. on the FSW surface—VP z = 115 V. (a) G = 15 mm. (b) G = 10 mm. (c) G = 5 mm.

piezoelectric fan. This latter statement agrees with similar tests conducted on an unfinned surface together with those published in [47]. The series of 2-D contour plots presented in Fig. 8 reflect the cooling plots acquired on the FSW. The first noticeable feature in the contour maps is that the cooling magnitude (%hInc ) is significantly lower than that observed on the FB, indicating that a horizontally mounted piezoelectric fan is not as effective in cooling the FSW as the FB—a result which agrees with the conducted PIV tests. Furthermore, the cooling effect on the FSW is constrained in close proximity to the FB and runs along the entire length of the FB (i.e., ZP os = 0 mm) for large vibrational envelopes. Results indicate that the best FSW geometrical cooling configuration is that of the following: VP z = 115 V and G = 15 mm. That is, a large separation distance is seen to be beneficial when cooling the FSW. Furthermore, it is seen that both a reduction in the vibrational envelope (AP z ) and a decrease in the separation distance (G) lead to a reduction in the overall cooling capabilities of the piezoelectric fan. A direct comparative study is presented in Table III. However, as conflicting results seem to arise with regard to the optimal separation distance (G), (9) was implemented to establish the overall optimum fin/fan geometry, the results of which are shown in Figs. 9–13.

Fig. 9.

Fig. 10.

Effect of vibrational amplitude for a horizontal setup.

Effect of the separation distance (G): VP z = 115 V.

A. Constant Heat Flux Tests 1) Effect of Fan Amplitude (VP z ): As with any finned setup, an increase in LFin results in an increase in both the natural and forced convection components, represented by a decreasing REquivalent (see Fig. 9). Furthermore, introducing the proposed FCCS and applying a minimal fan voltage of VP z = 30 V result in a significant improvement in the cooling capabilities over natural convection. This is further evident with an increase in


Fig. 11. Effect of the fin spacing (S): VP z = 115 V.

Fig. 12. Effect of piezofan orientation: VP z = 115 V.

Fig. 13. Comparative study: Constant heat flux versus constant temperature tests—VP z = 115 V.

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fan voltage. The largest cooling occurs when VP z = 115 V. Numerically, at the FB (LFin = 0 mm), REquivalent is reduced from 0.116 m2 · K/W (VP z = 0 V) to 0.027 m2 · K/W (VP z = 115 V)—a 76.7% reduction in the heat sink unit-cell thermal resistance. However, with an increase in the LFin , the effect of the FCCS is seen to decrease. This is attributed to the fact that a larger fin length implies a larger fin area (AF S ) which, in turn, implies that natural convection becomes the dominant cooling mechanism. Numerically, at LFin = 65 mm, REquivalent is reduced from 0.024 m2 · K/W (VP z = 0 V) to 0.013 m2 · K/W (VP z = 115 V)—a 45.8% reduction. Similar trends have been attained for a vertically mounted fan. Results indicate a 72.2% reduction when LFin = 0 mm and a 44% reduction when LFin = 65 mm—a direct comparison of which is presented in Fig. 12. Similar results shine true for the other tested fin spacings (S) and separation distances (G). 2) Effect of Separation Distance (G): The effect of separation distance on the resulting heat sink unit-cell thermal resistance (REquivalent ) is further investigated. Fig. 10 shows the results attained for one such test when operating the FCCS at the peak voltage and S = 40 mm. In the previous section, it was established that a minimal G is beneficial to cool the fin base, while a maximum G is beneficial to cool the fin side walls. However, on superimposing the results (i.e., to maximize the cooling of the entire heat sink unit cell), the overall effect of separation distance G has negligible effect (within the experimental uncertainty) on REquivalent (see Fig. 10). The same trends hold true for all other tested VP z and S. That is, experimental results indicate that, for both fan orientations, the separation distance G has negligible effect on the overall cooling enhancement of the heat sink unit cell and is thus not a design parameter that should be considered in the design of the FCCS. 3) Effect of Fin Spacing (S): The effect of fin spacing on the resulting heat sink unit-cell thermal resistance (REquivalent ) has also been investigated, the results of which are presented hereunder. Fig. 11 graphically indicates that a decrease in the fin spacing results in a significant reduction in the heat sink equivalent thermal resistance. The latter statement holds true for both the natural convection (Fan Off) plots as well as when the FCCS is in operation (Fan On). At the fin base (LFin = 0 mm), REquivalent is reduced from 0.113 m2 · K/W (VP z = 0 V and S = 30 mm) to 0.026 m2 · K/W (VP z = 115 V and S = 30 mm)—a 77% reduction. On the other hand, at LFin = 65 mm, REquivalent is reduced from 0.019 m2 · K/W (VP z = 0 V) to 0.010 m2 · K/W (VP z = 115 V)—a 47.4% reduction. Experimental results (see Fig. 11) indicate that the fin spacing is a critical design parameter that should be considered in the design of the FCCS. Moreover, a fin spacing of S = 30 mm results in the best cooling geometry. 4) Effect of Fan Orientation: Piezoelectric fan orientation (horizontal and vertical) is another critical design issue which was investigated. A direct comparison between the two fan orientations is presented in Fig. 12. Experimental results indicate that, for all tested fin spacings, a horizontally mounted fan outclasses a vertically mounted fan.

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Thus, from the current experimental results, it has been established that the best FCCS geometry to implement is a horizontally mounted fan, operated at VP z = 115 V and G = 5 mm with a fin spacing of S = 30 mm. B. Constant Temperature Tests While the aforementioned presented results all pertain to a constant heat flux scenario, further experiments were performed to evaluate the change in REquivalent when a constant surface temperature (TS − T∞ ) is maintained (rather than a constant heat flux). These tests were undertaken to cater for the common electrical machine design scenario where maximum allowable operating temperatures (TM , TEW dg ) are the constraint rather than running the machine at the same power density but at lower operating temperatures. Maintaining a constant surface temperature (TS ) inherently results in an increase in heater input power. This, in turn, implies that lower thermal resistance values arise (see Fig. 13). Fig. 13 compares a constant heat flux test against a constant temperature test for both fan orientations when VP z = 115 V and G = 10 mm and at a fin spacing of S = 30 mm. For both piezofan orientations, a lower REquivalent is obtained on conducting the constant temperature tests when compared to the constant heat flux tests. While this difference is minute for large LFin , the difference in REquivalent is seen to increase with a decrease in LFin . Furthermore, similar trends to those attained in a constant heat flux environment resulted, once again indicating that a horizontally mounted fan still outperforms a vertically mounted one. This result ensured that further constant temperature tests were irrelevant. VI. C ASE S TUDY: A EROSPACE ACTUATOR FAULT-T OLERANT M OTOR While the heat transfer enhancement advantages of implementing the suggested FCCS are highlighted in the previous section, a study on the weight savings capabilities of piezoelectric fans is now presented. A representative supporting structure design was created and optimized through FEA studies (see Fig. 14), and a prototype was fabricated and tested. The designed supporting piezoelectric fan structure was constructed from carbon fibre reinforced plastic (CFRP), which resulted in a structure mass of 3.80 g, which, in turn, implied a total FCCS mass of 6.60 g. The NC cooling mass (mass of FSW’s) and the FCCS cooling mass (combined mass of FSW’s, piezoelectric fan, and piezoelectric fan CFRP supporting structure) for the different fin spacings are presented in Fig. 15. Fig. 15 has no engineering relevance if considered as a separate entity. However, if superimposed onto Fig. 11, both the cooling enhancement and the weight saving capabilities of the designed FCCS may be established. The authors would like to highlight that no single optimum FCCS configuration exists; as the optimum FCCS configuration is dependent on the imposed constraints: REquivalent , m, and LFin (or cooling volume). The cooling mass (m) units are taken

Fig. 14. Von Mises stress analysis of FCCS.

Fig. 15. Comparative study between the NC and the FCCS mass for different fin spacings.

to be kg/m as the analysis carried out pertains to a unit cell. The results of this study can thus be scaled up depending on the machine length resulting in greater flexibility. To demonstrate this, each constraint shall be taken in turn, and the benefits of introducing the FCCS will be investigated for each of the three different constraints. A. Thermal Resistance Constraint (REquivalent ) If a thermal resistance constraint of REquivalent = 0.020 m2 · K/W is set up, by referring to Fig. 16, the following results are obtained: NC − S = 30 mm, LFin = 60 mm, m = 1.10 kg/m FCCS − S = 30 mm, LFin = 11 mm, m = 0.42 kg/m. The above results indicate an 81.7% reduction in fin length (which is directly related to the cooling volume) and a 61.8% reduction in cooling mass. On the other hand, applying a thermal resistance constraint of REquivalent = 0.065 m2 · K/W yields the following results: NC − S = 30 mm, LFin = 11 mm, m = 0.20 kg/m FCCS − S = 50 mm, LFin = 0 mm, m = 0.132 kg/m.


Fig. 16. Thermal resistance versus fin length versus cooling mass for various fin spacings: Horizontal fan—VP z = 115 V.

This latter scenario results in the complete elimination of all fin side walls (100% reduction), which, in turn, results in a 34% reduction in cooling mass. Irrespective of the required electric machine housing thermal resistance, the implementation of the suggested FCCS leads to an overall reduction in both the cooling volume and cooling mass.

B. Cooling Mass Constraint (m) The following constraint commonly results when the minimization of mass is the key. On setting up a cooling mass constraint of m = 1 kg/m and referring to Fig. 16, the following results: NC − S = 30 mm, REquivalent = 0.023 m2 · K/W, LFin = 55 mm FCCS − S = 30 mm, REquivalent = 0.011 m2 · K/W, LFin = 42.5 mm. The above results indicate a 52.2% reduction in the heat sink unit-cell thermal resistance, together with a further 22.7% reduction in the cooling volume. VII. C ONCLUSION The cooling capabilities of a novel electric machine housing forced cooling technique were evaluated for numerous

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different fin/fan geometrical configurations. Implemented PIV techniques highlight the complexity of the highly turbulent unsteady flow fields (with numerous vortices being shed in each swing) generated by one such oscillating piezoelectric fan. It was further deduced that the flow comprises a bulk mean flow superimposed onto a highly turbulent fluctuating component—indicating that the convective cooling enhancement is dependent on %hInc. = f (VMean , VRMS ). Peak velocities in excess of 2.41 m/s and turbulent flows reaching 2.38 m/s have been recorded. Furthermore, flow measurement results agree with thermal results and indicate that an average convective heat transfer coefficient enhancement of 364.8% on the fin base and an enhancement of 53.6% on each of the fin side walls is attainable. Experimental results further indicate that a horizontally mounted fan, operating at the natural frequency, with maximum vibrational envelope, minimal separation distance, and minimum fin spacing results in the minimal heat sink thermal resistance REquivalent , i.e., optimum FCCS geometry. The net improvement in adopting this thermal management methodology compared to natural convection is application dependent; however, for the case study considered, this resulted inasmuch as a 61.8% reduction in the electric machine heat sink cooling mass relative to natural convection cooling and the complete elimination of the housing fin side walls. These results, together with the high reliability, minimal weight, and minimal power requirements, make piezoelectric fans potentially suited for the thermal management of electric machines in safety-critical applications.

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Gareth M. Gilson received the B.Eng. degree in mechanical engineering from the University of Malta, Msida, Malta, in 2008 and the Ph.D. degree in energy and sustainability from The University of Nottingham, Nottingham, U.K., with a dissertation titled “Cooling of Advanced Aircraft Actuation Systems.” He is currently a Thermal Design Engineer with Cummins Generator Technologies, Stamford, U.K. His core research interests are the cooling of aircraft actuation systems for the More Electric Aircraft and the thermal management of current and future electric machines.

Stephen J. Pickering received the B.Sc. and Ph.D. degrees in mechanical engineering from The University of Nottingham, Nottingham, U.K., in 1979 and 1984, respectively. Since 1988, he has been a Lecturer with The University of Nottingham, where he is currently an Associate Professor and Reader in the Faculty of Engineering. He has extensive research experience in thermofluids and has undertaken research into the cooling of electric machines for over ten years.


David B. Hann received the B.Sc.(Hons.) degree in physics and the Ph.D. degree from Edinburgh University in 1990 and 1994, respectively. He subsequently worked at King’s College London, London, U.K., and Trinity College Dublin before taking a lectureship position at The University of Nottingham, Nottingham, U.K., in 2005. He is currently a Lecturer with the Department of Mechanical, Materials and Manufacturing Engineering, The University of Nottingham. His current research interests are acoustics and optical measuring techniques as well as thermofluids.

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Chris Gerada received the Ph.D. degree in numerical modeling of electrical machines from The University of Nottingham, Nottingham, U.K., in 2005. He subsequently worked as a Researcher at The University of Nottingham on high performance electrical drives and on the design and modeling of electromagnetic actuators for aerospace applications. In 2008, he was appointed as Lecturer in Electrical Machines and is now an Associate Professor within the power electronics, machines and controls research group at The University of Nottingham. His core research interests include the design and modeling of high performance electric drives and machines. Dr. Gerada has been the Project Manager of the GE Aviation Strategic Partnership since 2006 and, in 2011, was awarded a Royal Academy of Engineering Senior Research Fellowship supported by Cummins. He is also an Associate Editor for the IEEE T RANSACTIONS ON I NDUSTRY A PPLICATIONS and an executive member of the management board of the U.K. Magnetic Society and the Institute of Engineering and Technology Aerospace Technical and Professional Network.