Rotor Cooling Research Papers

The impact of thermal degradation on electrical machine winding insulation

2024, 2015 IEEE 10th International Symposium on Diagnostics for Electrical Machines, Power Electronics and Drives (SDEMPED)

Inter-turn stator short circuits can develop quickly leading to serious damage of an electric machine. However, degradation mechanisms of winding insulation are not yet fully understood. The main contribution of this article is analysis... more

descriptionView Paper arrow_downwardDownload

The impact of thermal degradation on electrical machine winding insulation

by Keith Burnham

2024, 2015 IEEE 10th International Symposium on Diagnostics for Electrical Machines, Power Electronics and Drives (SDEMPED)

Inter-turn stator short circuits can develop quickly leading to serious damage of an electric machine. However, degradation mechanisms of winding insulation are not yet fully understood. The main contribution of this article is analysis... more

descriptionView Paper arrow_downwardDownload

The impact of thermal degradation on electrical machine winding insulation

by Malgorzata Sumislawska

2024

Inter-turn stator short circuits can develop quickly leading to serious damage of an electric machine. However, degradation mechanisms of winding insulation are not yet fully understood. The main contribution of this article is analysis... more

descriptionView Paper arrow_downwardDownload

The impact of thermal degradation on electrical machine winding insulation

by Malgorzata Sumislawska

2024, 2015 IEEE 10th International Symposium on Diagnostics for Electrical Machines, Power Electronics and Drives (SDEMPED)

Inter-turn stator short circuits can develop quickly leading to serious damage of an electric machine. However, degradation mechanisms of winding insulation are not yet fully understood. The main contribution of this article is analysis... more

descriptionView Paper arrow_downwardDownload

High Potential of Magnet on the Performance of Dual Piezoelectric Fans in Electronics Cooling System

by Robiah Ahmad

2023, Indonesian Journal of Electrical Engineering and Computer Science

Recently, piezoelectric fan has gained attention as potential active cooling method for electronics devices. Even though the piezoelectric requires high voltage, there are findings to overcome the shortcomings. Adding on a magnet at the... more

Figure 6. Variation of Amplitude with Existence of Magnet

Figure 1. Piezoelectric Fan by www.piezo.com The amplitude of the piezoelectric fan can be calculated using Equation!

where K_pf and K_mag are the stiffness for piezoelectric fan and magnetic fans respectively. The stiffness of the magnetic fan and piezoelectric fan can be found in Equation 3 and Equation 4: where Bo, tinag, mag, X and jlo are the magnetic flux density, thickness of the magnet, radius of the magnet, distance between two magnets and permeability of the intervening medium respectively. At maximum amplitude, the integrated piezoelectric fan also encountered maximum air disturbance velocity which will affect the cooling rate of the cooling system. The velocity of the beam with respect to time is given in Equation 7 [18]:

Position of additional mass such as magnet is very important in order to obtain largest amplitude with same input voltage. Based on Equation | maximum amplitude can be achieved at point of moment of inertia of the beam, I, where illustration is in Figure 2. Moment of inertia is a measure of how difficult it is to rotate a particular body about a given axis. Greater the mass concentrated away from the axis, greater the moment of inertia [18]. In other words, moment of inertia is at the lowest value at the point on the axis of rotation to produce largest amplitude of the beam.

By using specifications in Table | and Figure 1, a function generator is used to power up the piezoelectric fan. The function generator is able to control the current flow thus able to oscillate the piezoelectric at its resonant frequency. The piezoelectric needs 115VAC thus a step down transformer has been used to drop the voltage from 240 VAC to 115 VAC. For a dual fans system, 2 casings are produced which are in different orientation, array and radial respectively. The schematics are shown in Figure 3.

Figure 4. Location of Magnet on Piezoelectric Fan The overall experiment is about to prove the significant of magnet on piezoelectric fan, optimum number of magnet located on each fans and the optimum distance between magnets on the piezoelectric fan and adjacent passive fan. The radius of the magnet is kept constant as the position and distance between magnets are varied. The best position of the magnet was tested at single fan only with single magnet and dual magnet at position I, II and III as shown in Figure 4. The position of magnet with largest amplitude is selected for next experiment. Table 2 shown different cases applied on piezoelectric fan and magnetic fans. Table 3. Different Case Applied on Piezoelectric Fan and Magnetic Fans

As shown in Figure 5, the highest amplitude at each magnet position occurs at different frequency. Position I, III and no magnet demonstrate quite similar amplitude but not at position II.

In order to improve total amplitude of the system, dual fans have been used and arranged in array and radial orientation. By placing two magnets on each fan, the amplitude of each piezoelectric fan and passive fan are recorded at different distance between magnets which are 15mm, 30mm and40mm. At the same distance between magnets, the results of different orientation are recorded as shown in Figure 7. Each case shows that radial orientation gives better result than array orientation. Distance between magnets should be selected in such a way that the amplitude produced by magnetic fan is not too far from the driving fan. Figure 8 shown the amplitude difference between piezoelectric fan and passive fan.

Figure 8. The Amplitude Difference between Piezoelectric Fan and Passive Fan

Figure 9. Total Amplitude Gained Different Fan Orientation This is to stabilize the repulsive force delivered between the magnets and the magnetic fan oscillates evenly with the piezoelectric fan. Therefore, the magnets should be placed at a range of distance 15mm- 30mm to generate maximum total amplitude for dual fans (Figure 9).

Figure 11. Peak to Peak Displacement vs Power Consumption at Different Number of Magnets This kind of system has a good enhancement in reducing power consumption while increasing thé total amplitude to reduce the heat in the area. As shown in Figure 11, single fan with dual magnet located i1 the middle of the blade consumed the least power while producing highest total amplitude. Dual fans wil definitely enhance the total amplitude without increasing the power consumption because no addec piezoelectric required but only passive fans with magnets.

Figure 13. Thermal Resistance Vs Power Consumption At Different Magnet Location Figure 12. Thermal Resistance vs Power Consumption at Different Magnet Location

Table 2. Different Case Applied on Piezoelectric Fan and Magnetic Fans

2.4. Performance Parameters convection coefficient since it is related to temperature change. The heat flux, q = — where equivalent tc

Table 4. Variety of Cases under Different Position of Magnet and Number of Magnet A piezoelectric fan oscillated with larger amplitude when additional mass is located on the blade. The position of the mass is varied to three positions and at each position, single and double magnets are applied. A table of cases is summarized as shown in Table 4.

descriptionView Paper arrow_downwardDownload

High Potential of Magnet on the Performance of Dual Piezoelectric Fans in Electronics Cooling System

by Robiah Ahmad

2023, Indonesian Journal of Electrical Engineering and Computer Science

Recently, piezoelectric fan has gained attention as potential active cooling method for electronics devices. Even though the piezoelectric requires high voltage, there are findings to overcome the shortcomings. Adding on a magnet at the... more

descriptionView Paper arrow_downwardDownload

Air inlet/outlet arrangement for rotor cooling application of axial flux PM machines

by Ahmad Syahid A. Fawzal

2022, Applied Thermal Engineering

Publisher: Elsevier NOTICE: this is the author's version of a work that was accepted for publication in Applied Thermal Engineering. Changes resulting from the publishing process, such as peer review, editing, corrections, structural... more

Fig. 1: The CFD model based on modular CAD design that combines all designs and: (a) Design 1 — inlet on top to outlet at bottom, (b) Design 2 — tangential inlet-outlet and (c) Design 3 — conventional centrifugal fan

Fig. 2: Detail drawing of the 3D CFD model including section A-A and section B-B. Fig. 2: Detail drawing of the 3D CFD model including section A-A and section

Fig. 3: The torque, rotor losses and magnet losses of a YASA sealed machine produced by FEA.

Fig. 4: Pressure drop of the studied designs.

Fig. 5: Windage losses of the studied designs

Fig. 6: Average air temperature on (a) the outlet and (b) within the rotational system. Fig. 8 presents the average temperature of the rotor and the magnet. The

Fig. 7: Flow streamlines of each design and its ISO surface of low velocity region (<10m/s) at 4000rpm in blue translucent patches. Fig. 7: Flow streamlines of each design and its ISO surface of low velocity

Fig. 8: Average solid temperature of the rotor and the magnets for (a) Design 1, (b) Design 2 and (c) Design 3.

Fig. 9: Average heat transfer coefficient of the studied designs

Fig. 10: Cooling Performance Index (CPI) curves of all tested designs.

Fig. 11: The YASA machine assembly integrated with Design 2 and Design 3 inlet and outlet arrangement including dual mass flow sensors on the inlet. DE - disc side, before exiting to the ambient. Fig. 11: The YASA machine assembly integrated with Design 2 and Design 3 inlet and outlet arrangement including dual mass flow sensors on the inlet. Page 14 of 25

‘ig. 12: The new YASA machine topology including rotor cooling on both rotors. the NDE side until the simulations converge. study. In reality a thin layer of adhesive material is present between the

Fig. 13: The prototype of the new YASA machine topology in-cooperate with dual rotor cooling via the air transfer duct.

Fig. 14: Comparison between CFD and experimental data of the mass flow rate. Fig. 14: Comparison between CFD and experimental data of the mass flow

Fig. 15: Average air temperature on the outlet and within the rotational system of the design implementation. (Fig. 16-a) as compared to the initial study.

Fig. 17: Average solid temperature of the rotor and the magnets of the design implementation. Fig. 17: Average solid temperature of the rotor and the magnets of the design

4.3 Thermal Analysis machine will operate under speed less than 1000rpm for a long duration. significant amount of the heat out, thus its outlet temperature is similar to the development without stalling or negative pressure. However, further analysis

descriptionView Paper arrow_downwardDownload

Fan Performance Analysis for Rotor Cooling of Axial Flux Permanent Magnet Machines

by Ahmad Syahid A. Fawzal

2022, IEEE Transactions on Industry Applications

This document is the author's post-print version, incorporating any revisions agreed during the peer-review process. Some differences between the published version and this version may remain and you are advised to consult the published... more

Fig. 1. YASA derivation topology begins with (a) NS Torus-S topology, (b) exclude the stator yoke and (c) final design changes [6]. Conventional AFPM machines are comprised of yokes on the rotor and stator to support its field system and protect the armature of the motor. The YASA machine topology is able to improve the performance of AFPM by removing the stator yoke and implementing other design changes, as illustrated in Fig. 1. The derivation of the YASA topology began with the NS Torus-S topology (Fig. 1-a), then evolved by excluding the stator yoke as it did not seriously impact the electromagnetic behaviour of the system (Fig. 1-b). In Fig. l-c, the YASA topology was formed by the pitch of the teeth is enlarged and square windings are used. This new topology improved torque density by 20% and recorded a peak efficiency of 95% with 50% less iron in the stator compared to a conventional AFPM machine [6]. Nevertheless, the full potential of the machine is currently limited by its operating temperature. Therefore, optimisation of the thermal design is required to improve the motor’s reliability and to further increase its performance.

Fig. 2. The position of fan blade attached to the rotor allows introduction of cool air in the system. The initial concept of rotor cooling for the AFPM machine was first proposed by Vansompel [4]. The implementation of rotor cooling on the current YASA machine topology was accomplished by introducing an opening at the centre of the rotor and a single out et on the radial cover. This is adopted from the concept of a centrifugal fan, where the air will flow from the centre inlet t blade and exits throug hrough the system that drives by a fan h the side outlet. Fig. 2 shows that the fan blade component is attached directly to the machine rotor of non drive end side. While the machine rotates, the rotor with a fan blade will create pressure gradient inside the system, thus air flow is developed. This air flow is expected to absorb the heat from the rotor (by convection) and disperse it to the outlet. As this happens, heat from the magnet will flow to the rotor assembly through conductive heat transfer.

Fig. 3. The fan blade design selection with (a) backward-inclined aerofoil blade [20], (b) radial aerofoil blade [20] and (c) tear drop pillar blade [21]. Fig. 3-b, a radial aerofoil blade — from a typical fan engineering application [22] and Fig. 3-c, a tear drop pillar blade — from an automotive disc brake [23]. The backward- inclined aerofoil blade and radial aerofoil blade were selected as it is known that they provide good air flow as a pump or fan. Additionally, it is interesting to see how a tear drop pillar blade performs as this is the cooling technique used for heat removal in brake disc. For rotor cooling applications, the operation of a backward- inclined aerofoil blade is limited to an unidirectional machine, while the radial aerofoil blade and the tear drop pillar blade can be used in a machine with a bi-directional application. In this study, unidirectional operational was considered because current machine design is specific for automotive traction application with medium rotational speed in the intended directionand slow speed in the reverse.

Fig. 4. The experimental rig with mass flow sensor at the inlet.

Fig. 5. CFD model including fluid domain (blue), rotational fluid domain (red), rotor & ducting (grey) A complete domain of the non drive end was modelled due to the non-symmetrical design of the inlet and outlet positions. The final CFD model included the stator wall, the rotor assembly, the outlet extrusion and the inlet duct with dual mass flow sensors (Fig. 5). The contact interface between the Solid-Fluid regions and the Fluid-Fluid regions was created by imprinting the mating face to allow conformal mesh between the regions. Using the automated mesh generation in STAR-CCM+, an unstructured volume mesh (polyhedral cells) was used to discretise the solid and fluid domain. The prism layer was set based on the calculated dimensionless wall distance [24] of (1)-(2) with a local Reynolds number of 3.73e5 (3).

three fan blade designs. Fig. 6. Computational mesh of fluid domains for backward-inclined aerofoil blade at front cross section (top) and side cross section (bottom).

Fig. 7. Comparison CFD and experimental result of (circle) flow without mesh-filter on the outlet, (diamond) with mesh-filter on the outlet.

Fig. 8. a) Mass flow rate and b) pressure development of each fan design.

Fig. 10. Fan curve and power curve of (a) backward-inclined aerofoil blade, (b) radial aerofoil blade and (c) tear drop pillar blade. The contour plot of air temperature, Fig.1 , was recorded at Z +3mm above the front heated rotor surface for 3500rpm with the range between 26.85°C to 80°C. In Fig. 11-a, a hot air patch at 80°C can be observed on the backward -inclined lower aerofoil surface near to its trailing edge. This hot air then flows to the radial space of the machine while slightly reducing its temperature before going to the outlet. T temperature at the outlet of the backward-in he average air clined blade in

Fig. 11. Air temperature contour at z +3mm from the front heated rotor surface of (a) backward-inclined airfoil blade, (b) radial airfoil blade and (c) tear drof pillar blade.

Fig. 12. Average air outlet temperature of each fan design. On the other hand, no hot spot is present on the air temperature contour plot of the pillar type (Fig. 11-c). Nevertheless it shows a fair temperature gradient that collects at the radial space of the machine compared to the radial blade design. The average air outlet temperature of pillar type (Fig. 12) is marked between the two others fan blade designs with increasing value as the speed increases. Furthermore, the capability of each fan design in dissipating the rotor temperature can be analysed by plotting the average air temperature within the rotating region, as presented in Fig. 13. The hot air patch of the backward- inclined and the radial blade (Fig. 11) proves that both designs have high average temperatures within the system. However, for the radial blade, this hot air remains within the system as discussed before due to the flow circulation at the downwind region of the blade which made the average air outlet temperature (Fig. 12) of this design the lowest.

Fig. 14 is calculated by (4) where the heat flux of the rotor and magnet walls are divided by the temperature difference between the rotor and the system average temperature of Fig. 13.

Fig. 14. Average heat transfer coefficient of each fan design.

Fig. 15. Rotor Cooling Performance Index (RCPI) curves for this study.

LIST OF COEFFICIENT FOR ALL SELECTED TURBULENCE MODELS

BOUNDARY CONDITIONS USED FOR ALL SIMULATIONS TABLE I

BOUNDARY CONDITIONS USED FOR ALL SIMULATIONS

descriptionView Paper arrow_downwardDownload

The fan design impact on the rotor cooling of axial flux permanent magnet machines

by Ahmad Syahid A. Fawzal

2022, 2016 XXII International Conference on Electrical Machines (ICEM)

This document is the author's post-print version, incorporating any revisions agreed during the peer-review process. Some differences between the published version and this version may remain and you are advised to consult the published... more

Fig. 2. The fan blade design selection with (a) backward-inclined aerofoil blade [20], (b) radial aerofoil blade [20] and (c) tear drop pillar blade [21]. For this purpose, three different fan blade designs, adopted from other engineering applications, have been selected and the flow characteristic, power requirement and thermal characteristic of each fan design have been compared. Fig. 2 shows the selection of fan blade designs, which are: Fig. 2-a, a backward-inclined aerofoil blade — from a force-draft blower application [20], Fig. 2-b, a radial aerofoil blade — from a typical fan engineering application [20] and Fig. 2-c, a tear drop pillar blade — from an automotive disc brake [21].

Fig. 3. The experimental rig with mass flow sensor at the inlet. speed range was selected, based on the limitation of the mass flow sensor (limited to 600litres/min), while the nominal speed of the machine is 2100rpm. The setup was set to run until the system reached a steady state condition for mass flow rate measurement for 30seconds on every 100rpm increment with a 10rpm/s ramp rate.

A complete domain of non-drive-end side was modelled, due to non-symmetrical design of inlet and outlet positions. The final CFD model included stator wall, rotor assembly, inlet duct with dual mass flow sensors and outlet extrusion (Fig. 4). Contact interface between Solid-Fluid regions and Fluid-Fluid regions was created by imprinting the mating face to allow conformal mesh between the regions. Using automated mesh generation of STAR-CCM+, an unstructured volume mesh (polyhedral cells) was used to discretise the solid and fluid domain. The prism layer was set based on the calculated dimensionless wall distance [22] of

Fig. 5. Comparison CFD and experimental result of (circle) flow without mesh-filter on the outlet, (diamond) with mesh-filter on the outlet. The error percentage in all turbulence models reduced at higher rotational speeds, as the flow changed to fully developed turbulent flow. The pattern of mass flow plot by CFD started to change after 2500rpm, due to the change from transitional flow and fully developed turbulent flow Table II presents an equivalent speed study for transitiona flow regime by Gregory et al. on their generic rotor-stato1 study [30].

Fig. 6. a) Mass flow rate and b) pressure development of each fan design. The mass flow rate and pressure development, or pressure rise, for the three fan blades are presented in Fig. 6. A radial blade design provides the highest mass flow and pressure development compared to the other designs. The mass flow rate steadily increases, while the pressure development sharply rises as the rotational speed increases. On the other hand, the pillar type and the backward-inclined show similar characteristics on driving the mass flow and pressure development with the same relative difference across the speed range.

Fig. 9. Air temperature contour at z +3mm from the front heated rotor surface of (a) backward-inclined airfoil blade, (b) radial airfoil blade and (c) tear dro pillar blade.

Fig. 8. Fan curve and power curve of (a) backward-inclined aerofoil blade, (b) radial aerofoil blade and (c) tear drop pillar blade. The contour plot of air temperature, Fig.9, was recorded at Z +3mm above the front heated rotor surface for 3500rpm with the range between 26.85°C to 80°C. In Fig. 9-a, hot air patches at 80°C can be observed on the backward-inclined Furthermore, the capability of each fan design in dissipating t the average presented in Fig. 11. The hot air patche inclined and he rotor temperature can be ana radial blade (Fig. 9) proves t ysed by plotting air temperature within the rotating region, as d of backward- hat both designs have high average temperature within the system. However, for radial blade, this hot air remains within t flow circula ion at the downwind region of made the average air outlet temperature o he system due to the blade, which f this design the

mass flow rate with reasona T a b In Fig. 8-a, the backward-inclined type provides a fair amount of pressure development across the production of ackward-inclined, this ro fi ow rate. ble amount of windage losses. he fan curve of radial blade in Fig. 8-b increases sharply and is closely followed by its However in Fig. 8-c, althoug power curve or windage losses. h the pillar type blade design is ble to provide a higher amount of mass flow, compared to or design shows significant windage losses that increase drastically with increasing mass

Fig. 11. Average air temperature within the rotating region. Fig. 10. Average air outlet temperature of each fan design.

BOUNDARY CONDITIONS USED FOR ALL SIMULATIONS A Moving Reference Frame (MRF) approach was applied to the rotational region to mimic rotational implicitly. This approach allowed the prediction of the rotational flow in the steady state condition and a faster solution compared to ransient simulation, e.g. [23]. The rotational speed for validation was varied similarly to the experiment with a backward-inclined aerofoil blade attached. Three selections of turbulence models, namely SST (shear-stress transport) k- omega [24], Realizable k-epsilon [25] and v2-f model [26]- [29] were compared during validation. Further, the urbulence model of SST k-omega was then chosen for the comparison of the three fan blade designs. TABLEI

FLOW REGIME GIVEN BY GREGORY ET AL. [30] WITH ITS EQUIVALENT SPEED

descriptionView Paper arrow_downwardDownload