Reduction of Mechanical Loss of Flywheel Energy Storage System with Spherical Spiral Groove Bearing

Transcription

1 Reduction of Mechanical Loss of Flywheel Energy Storage System with Spherical Spiral Groove Bearing Takeo Suuki, Takumi Masuda, Jun-ichi Itoh, Noboru Yamada Nagaoka University of Technology Nagaoka, Niigata, Japan IEEE PEDS 2017, Honolulu, USA December 2017 Abstract- Flywheel Energy Storage System (FESS) is known as a mechanical battery to store electricity. In a small-scale FESS, mechanical loss due to frictions of bearings must be reduced. In this study, a Spherical Spiral Groove Bearing (SSGB) is used to reduce the bearing loss. The bearing performance of SSGB is greatly affected by the groove shape, therefore the aim of this paper is to optimie the groove shape by means of computational fluid dynamics (CFD). CFD simulation results show that the groove width ratio Δ and groove depth h 0 have dominant effects on the bearing performance; and that the bearing loss can be reduced by 36.5% with Δ = 0.70 and h 0 = 0.02 mm as compared to the SSGB currently used. The prototype SSGB with the optimied shape was fabricated and the actual bearing loss was experimentally tested. As a result, it was confirmed that the bearing loss was reduced by 37.8 %. I. INTRODUCTION The number of renewable energy sources has increased, and a leveling method using energy storage systems is needed. Electrical energy can be stored in various forms, chemically, mechanically, thermally or magnetically. The flywheel energy storage system (FESS) stores electric energy as kinetic energy of a rotating flywheel and discharges energy by converting kinetic energy into electric energy via an electric motor and a generator. The FESS is a mechanical battery having advantageous characteristics such as lifetime and durability against temperature change, as compared with a chemical battery. A typical FESS consists of a flywheel rotor, a motor / generator, a casing and a bearing as shown in Fig. 1. During charging, the motor rotates the flywheel rotor, and at the time of discharge, the generator converts rotational energy into electrical energy. A comparison between FESS and other types of energy storage systems is shown in Table I. Compared with other energy storage systems, FESS is less prone to deterioration over time, and its output density is high due to high reaction speed. Various characteristics are obtained depending on the design of the flywheel rotor, so FESS can be used for a wide range of applications. However, in FESS, mechanical loss occurs due to friction loss caused by rotation of the rotor, and self discharge occurs. This mechanical loss consists mainly of windage loss and bearing loss. By bringing the inside of the casing close to a vacuum, windage loss can be relatively easily reduced [1]. On the other hand, reduction of bearing loss is difficult. Although a method of reducing friction loss by using high-temperature superconducting magnetic bearings (HTS) has been reported [2-4], the manufacturing and maintenance cost increases and control of magnetic bearings is difficult. In addition, the FESS using HTS becomes large and it is difficult to apply it to a small-scale system. In this research, for the purpose of developing a small scale high efficiency FESS, we tried a method to reduce cost without complicated control by using spherical spiral groove bearing (SSGB). The output and storage capacity of the prototype FESS is 1 kw / kwh, and the mass of the steel rotor is 380 kg. The advantages of SSGB are as follows. 1. Low friction loss than using multiple ball bearings, 2. Low TABLE I COMPARISON FESS WITH OTHER ENERGY STORAGE SYSTEMS Fig. 1. Schematic diagram of typical FESS components Storage method Flywheel Ni-Cd Battery Capacitor Rotation energy Chemical reaction Ionic migraton Charging time > 100 ms > 1000 ms > 10 ms Aged deterioration Energy and power density No Yes Yes Wh/kg W/kg Wh/kg W/kg Wh/kg ~100,000 W/kg /17/$ IEEE 355

2 cost than magnetic bearing, 3. Simple structure, 4. Can be used permanently. However, there have been no reports that SSGB has been adopted for FESS so far, the loss characteristics of SSGB and the optimum design method have not been established. Therefore, in this study, we evaluate the loss characteristics of SSGB by Computational Fluid Dynamics (CFD) and basic experiment and aim at optimiing groove shape which can reduce friction loss without sacrificing levitation force. II. SSGB FOR FESS SSGB is one of fluid bearings. As shown in Fig. 2, in the SSGB, the gap between the rotating shaft part and the bottom part is filled with the lubricating oil. A spiral groove is machined on the spherical surface of the tip of the shaft part. The shape of this groove is defined by the parameters shown in Table II and Fig. 2. When this shaft part is rotated in a counterclockwise direction, the lubricating oil is stirred by the grooves of the surface, and the shaft is floated up by hydrodynamic pressure. SSGB has an interesting ability to hold both radial load and axial load at the same time. In addition, it can be used permanently under proper operating conditions. In 1966 Muijerman et al proposed the concept of helical groove gearing [5]. Bootsma conducted research on SSGB theory and simulated load capacity [6, 7]. Sato et al. clarified the relationship between fluid inertia effect and load capacity in SSGB [8]. Many reports on planar thrust bearings and radial bearings such as herringbone grooved bearings have been reported [9, 10]. On the other hand, SSGB has so far not been studied because it is expensive to manufacture and no suitable application was available. Unlike other bearings, the SSGB has to be machined the thread groove of about 0.01 mm in depth to a perfect sphere and prepare the bottom part in addition to the shaft. Although these sophisticated processes increase the cost, in recent years it has become possible to manufacture with relatively low cost through development of processing techniques. The SSGB reported in this paper was manufactured using a 5-axis machining center. The shape of SSGB currently in use is R = 12.5 mm, α = 14 degree, Δ = 0.046, h 0 = 0.1 mm. For FESS, SSGB with high thrust load capacity and low friction moment is required. These characteristics change depending on multiple conditions such as rotation speed, lubrication temperature, groove shape and so on. In order to evaluate them in a unified way, we use the non-dimensional load capacity and moment defined by (1) and (2). C 12 R 2 r F 4 f (1) C 2 R r M 4 T (2) Fig. 2. Schematic diagram of SSGB used for kwh FESS TABLE II PARAMETERS OF GROOVE SHAPE R Sphere radius, mm α Groove angle, deg Δ Ratio of the width of groove to the width of ridge Φ 1 Equatorial angle, deg C r Radial clearance, mm Φ 2 Zenith angle, deg h 0 Groove depth, mm III. OPTIMIZATION OF SSGB BY CFD Since FESS repeats charge and discharge, lubrication temperature and rotation speed can not be kept constant. Therefore, at various lubrication temperatures and rotational speeds, the groove shape where the non-dimensional thrust load capacity is maximied and the non-dimensional moment is minimied was determined. For this optimiation, we used the CFD software SCRYU / Tetra (Software Cradle Co., Ltd). Shaft side Lubricant (a) Bottom side Fig. 3. Schematic diagram of SSGB and groove parameters (a) View at bearing shaft side surface Enlarged cross-sectional view of groove 356

3 (a) comparison result of the non-dimensional number. Since conditions of high oil temperature and low rotation are conditions that are severe for SSGB, analysis was carried out at oil temperature 80 C. and rotation speed 2,000 rpm. In consideration of actual machining accuracy, the increments of Δ and h 0 were set to 0.05 mm and 0.01 mm, respectively. Figure 4 shows the simulation results. It is found that the load capacity in the radial direction is greatly affected by the groove depth and the groove width ratio affects the thrust load capacity and the load moment. The grooves on the surface serve to send the lubricant downward and act like a pump. If the flow passage is too small, lubricating oil can not be supplied to the lower side, whereas if the flow passage is too large, the supply force becomes weak and pressure hardly occurs. The contact area has a large influence on the frictional force. By increasing the groove width ratio, the area to be a hill becomes smaller and the area of contact with the bottom surface becomes smaller. This is considered to be the cause of decreasing the non-dimensional moment. Comparison was made with h 0 = 0.10 mm, Δ = 0.05, which is the specification of SSGB currently used. Non-dimensional radial load capacity needs to be approximately the same range ( ) as the current SSGB. Also, the non-dimensional thrust load capacity needs to be higher than the rotor load under all operating conditions. When the rotor weight is 3,723 N (380 kgf), the rotation speed is 2,000 rpm, the kinematic viscosity is Pa s, the shaft radius is 12.5 mm, and the radial clearance is 0.01 mm, the required nondimensional thrust load capacity is It is necessary to minimie the non-dimensional moment while satisfying these conditions. As a result, h 0 = 0.02 mm, Δ = 0.70 was determined as an optimum shape parameter. In this case, the non-dimensional moment decreases by 36.5%, and the nondimensional thrust load capacity becomes about 3.6 times the reference. In the next section, experiments are conducted by prototyping this groove shaped SSGB. (c) Fig. 4. Simulation results for various groove width ratio and groove depth (a) Non-dimensional radial load capacity Non-dimensional thrust load capacity (c) Non-dimensional moment Fluid momentum and pressure were calculated by CFD and the performance of SSGB was obtained. The simulation model of SSGB is shown in Fig. 3. The thrust load capacity of the SSGB should be larger than the rotor weight, and at the same time a low friction moment is required. We changed the groove width ratio and the groove depth among the parameters that determine the SSGB shape and searched for the optimum shape based on the IV. EXPERIMENTAL RESULTS We conducted basic experiments to compare with CFD simulation and verify it. An outline of the experimental apparatus is shown in Fig. 5 and Table III. We used the two SSGBs, i.e., the current SSGB (Old-SSGB) and SSGB determined by CFD simulation (New-SSGB) for the experiment. The groove shape parameters of these SSGBs are shown in Table IV. While maintaining the applied load at 981 N (100 kgf) and the lubricating oil temperature at 35 C, the friction torque was measured when the rotational speed was changed. Temperature control of lubricating oil was carried out by passing cooling water around the bottom side of the SSGB. Using Krytox 143AD (DuPont) as the lubricating oil, the experiment was conducted in the range of 1,000 rpm to 2,000 rpm. The radial clearance C r can vary during the experiment because the thrust load capacity changes with rotational speed, thus the non-dimensional numbers were 357

4 (a) Fig. 5. Schematic diagram of experimental setup for SSGB test TABLE III MEASURING INSTRUMENTS USED IN EXPERIMENT EQUIPMENT (a) PM motor (e) Ball bearing unit Air cylinder (f) Load cell (c) Torque meter (g) SSGB (d) Laser displacement meter TABLE IV PARAMETERS OF GROOVE SHAPE OF SSGB USED IN EXPERIMENT R 12.5 mm Φ 1 20 deg C r 0.01 mm Φ 2 90 deg α 14 deg Δ h (Old-SSGB) (New-SSGB) 0.10 (Old-SSGB) 0.02 (New-SSGB) - mm Fig. 6. Experimental result of the optimied New-SSGB and the currently used Old-SSGB (a) Non-dimensional thrust load capacity Non-dimensional moment calculated by applying the measured levitation height as C r to (1) and (2). Fig. 6 shows a graph obtained by linearly approximating the averaged measured values, and the error bar shows the maximum value and the minimum value. The result of the New-SSGB shows a reasonable agreement with the simulation result, and the measured non-dimensional moment of the New-SSGB was smaller than that of the Old-SSGB. The non-dimensional moment of the New-SSGB at 2,000 rpm decreases by 66.3% relative to the Old-SSGB. The measured non-dimensional thrust load capacity of the New- SSGB was on average, which is sufficiently larger than the required value (0.042). On the other hand, the measured non-dimensional numbers of the Old-SSGB were much larger than the simulation results. This may be because that in the simulation the non-dimensional numbers were calculated by assuming a constant value, C r = 0.01 mm, while in the Fig. 7. Measurement result of the bearing loss and levitation length at the 2,000 rpm of the optimied New-SSGB and the currently used Old-SSGB experiment, the measured levitation length was used. Further discussion is needed to clarify this matter. Fig. 7 shows the experimental results of levitation height and bearing loss, indicating that the bearing loss was successfully reduced by 37.8% (32 W). The levitation length was also decreased; however, it is within an allowable range for the present condition. 358

5 V. CONCLUSION In this paper, the performance of the SSGB used for the small-scale FESS was simulated by CFD, and the groove shape of the SSGB was optimied. Then, the experiments with the fabricated SSGBs were conducted to investigate the validity of the optimiation. The optimied SSGB showed a lower bearing loss than that of the unoptimied SSGB, which agreed with the simulation result. The present result indicates that the charts of non-dimensional numbers obtained by CFD is useful for the shape optimiation of SSGB. Nomenclature C r Radial clearance, mm f i Load capacity, N F i Non-dimensional load capacity h 0 Groove depth, mm L Length of groove and hill, mm M Non-dimensional moment R Radius, mm T Friction torque, N m u i Flow velocity of x i direction, m/s α Groove angle, deg Δ Groove width ratio μ Viscosity of lubrication, Pa s ρ Density of lubrication, kg/m 3 Φ 1 Equatorial angle, deg Φ 2 Zenith angle, deg ω Angular velocity, rad/s References [1] J. Itoh, K. Tanaka, Y. Saiki, N. Yamada, and K. Kato, Design and Experimental Evaluation of the Flywheel System for Power Leveling, in Industrial Electronics Society, IECON th Annual Conference of the IEEE, [2] M. Strasik et al., Design, fabrication, and test of a 5-kWh/100-kW flywheel energy storage utiliing a high-temperature superconducting bearing, IEEE Trans. Appl. Supercond., vol. 17, no. 2, pp , [3] S. Mukoyama et al., Test of REBCO HTS Magnet of Magnetic Bearing for Large Capacity Flywheel Energy Storage System, Phys. Procedia, vol. 65, no. 3, pp , [4] Y. Miyaaki et al., Development of superconducting magnetic bearing for flywheel energy storage system, Cryogenics (Guildf)., vol. 80, pp , [5] E. A. Muijderman, Ed., Spiral groove bearings.pdf. N. V. Philips, [6] J. Bootsma, Spherical and Conical Spiral Groove Bearings Part I: Theory, Trans. ASME J. Lubr. Technol., vol. 97, no. 2, pp , [7] J. Bootsma, Spherical and Conical Spiral Groove Bearings Part II: Load Capacity and Stability, Trans. ASME J. Lubr. Technol., vol. 97, no. 2, pp , [8] Y. Sato, Fluid Inertia Effects on the Load Capacity of Spherical Spiral Groove Bearings.pdf, Trans. ASME J. Tribol., vol. 108, no. 1, pp , [9] T. Hirayama, N. Yamaguchi, S. Sakai, N. Hishida, T. Matsuoka, and H. Yabe, Optimiation of groove dimensions in herringbonegrooved journal bearings for improved repeatable run-out characteristics, Tribol. Int., vol. 42, no. 5, pp , [10] M. Sahu, M. Sarangi, and B. C. Majumdar, Thermohydrodynamic analysis of herringbone grooved journal bearings, Tribol. Int., vol. 39, no. 11, pp ,