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

Table 2 FLOW REGIME GIVEN BY GREGORY ET AL. [30] WITH ITS EQUIVALENT SPEED

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Fig. 1. YASA derivation topology begins with (a) NS Torus-S topology, (b) exclude the stator yoke and (c) other design changes [6]. Conventional AFPM machines are comprised of yokes on a rotor and stator to support the field system and protect the armature of the motor. Innovatively, the YASA machine topology is able to increase the performance of AFPM by removing the stator yoke and implementing other design changes, as illustrated in Fig. 1. The derivation of YASA topology began with NS Torus-S topology (Fig. 1-a), then evolved by excluding the stator yoke (Fig. 1-b), as it did not seriously impact the electromagnetic behaviour of the system. In Fig. l-c, one can see that the pitch of the teeth is enlarged and individual windings feature a high strength square coil wrapped in a bonding material. 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, although the YASA machine achieves an improvement in peak performance. Therefore, optimisation of thermal design is required to improve the motor’s reliability and to further increase its performance. The current YASA machine design uses oil as a direct stator cooling method to counter the heat generated from iron and copper losses. This is accomplished by sealing the stator assembly from the rotors and al ows coolant oil to pass through between the windings. Camilleri et al. [16] have investigated this cooling method and mapped the temperature profile of YASA machine pole piece for different coolant flow rates. The study on this machine also considered the influence of stator coolant flow path geometry [17]. The findings provided sufficient thermal management for low and medium continuous power applications.
Fig. 1. YASA derivation topology begins with (a) NS Torus-S topology, (b) exclude the stator yoke and (c) other design changes [6]. Conventional AFPM machines are comprised of yokes on a rotor and stator to support the field system and protect the armature of the motor. Innovatively, the YASA machine topology is able to increase the performance of AFPM by removing the stator yoke and implementing other design changes, as illustrated in Fig. 1. The derivation of YASA topology began with NS Torus-S topology (Fig. 1-a), then evolved by excluding the stator yoke (Fig. 1-b), as it did not seriously impact the electromagnetic behaviour of the system. In Fig. l-c, one can see that the pitch of the teeth is enlarged and individual windings feature a high strength square coil wrapped in a bonding material. 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, although the YASA machine achieves an improvement in peak performance. Therefore, optimisation of thermal design is required to improve the motor’s reliability and to further increase its performance. The current YASA machine design uses oil as a direct stator cooling method to counter the heat generated from iron and copper losses. This is accomplished by sealing the stator assembly from the rotors and al ows coolant oil to pass through between the windings. Camilleri et al. [16] have investigated this cooling method and mapped the temperature profile of YASA machine pole piece for different coolant flow rates. The study on this machine also considered the influence of stator coolant flow path geometry [17]. The findings provided sufficient thermal management for low and medium continuous power applications.
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. 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].
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
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
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.
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
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. 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. 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. 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
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
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.
Fig. 11. Average air temperature within the rotating region. Fig. 10. Average air outlet temperature of each fan design.
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EngineeringMechanical EngineeringAutomotive EngineeringRotor CoolingFan Performance
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