Thrust bearing using a magnetic fluid lubricant under magnetic fields

Thrust bearing using a magnetic fluid lubricant under magnetic fields K. Nagaya, S. Takeda, A. Sato, S. Ikai, H. Sekiguchi and N. Saito

This paper describes a thrust bearing using a magnetic fluid lubricant under a magnetic field. The critical pressures of the bearing versus the magnitude of the magnetic flux densities have been investigated experimentally. It is shown that the critical pressures of the proposed bearing are larger than those of the normal lubricated bearing under high speeds. Keywords: machine elements, thrust bearings, magnetic fluids, critical pressures

Introduction A magnetic fluid is a colloidal suspension of ferromagnetic particles in a fluid. The fluid is non-conductive of electric current, but moves and changes its shape under a magnetic field. Many applications of this fluid have been reported 1-7. However, thrust bearings under magnetic field lubrication have not been investigated thoroughly. The present article experimentally details the critical pressure of the thrust bearing using a magnetic fluid lubricant under magnetic fields. Since the viscosity increases under the magnetic field, and there is also a sealing effect for the bearing with the fluid, the critical pressure increases compared with the normal thrust bearing under normal oil lubrication. When there is a magnetic seal along the circumferential side of the shaft, the magnetic fluid thrust bearing does not have leaks, and hence the bearing can be used without a supply of fluid. This implies that the bearing has advantages when compared with a normal thrust bearing in practical use.

Experimental results The thrust bearing proposed in the present article consists of an iron plate connected to a strong permanent magnet. The magnetic fluid is filled between the end surface of the shaft and the bearing, as mentioned above. The experimental results are first obtained under static load without shaft rotation, then results under the rotating load are obtained.

Fluid used in the experiment The magnetic fluid used in this experiment is Ferricolloid PA-40 with specific gravity = 1.24 (at 25°C), Department of Mechanical Engineering, Gunma University, Kiryu, Gunma 376, Japan

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saturation magnetic flux = 0.03 T, and viscosity = 170 mPa s (at 25°C). The ferrite particles of radius 100-200 ~ are suspended in the ot-oleic acid oil at 40 wt. %. To explain the effects of magnetic fluid, two kinds of normal oils are also chosen. One is the motor oil named SF-10W with viscosity 106 mPa s (at 25°C), and the other is a white spindle oil with viscosity 11 mPa s (at 25°C).

Effects of a magnetic field on the shape of a magnetic fluid on the bearing Figure 1 shows the geometry of the magnetic fluid thrust bearing for obtaining the critical pressure under static load. The bearing plate is connected to a permanent magnet. The end of the shaft is made of aluminium. It is supported by a thin film of magnetic fluid lubricant which lies on the surface of the bearing plate. The experimental tests are carried out for two surfaces of the bearing plate. One is the flat surface of diameter 46 mm as shown in Fig 2(a), where the bearing plate is made of iron, and, to avoid the fluid concentration at the edge, a thin circular copper ring of outer diameter 46 mm and inner diameter 38 mm lies along the edge of the bearing plate. The other surface has V-shaped grooves on the inner surface as shown in Fig 2(b). To promote fluid lubrication, there is a small step between the surface of the copper ring and the iron surface. There is a sealing effect for the magnetic fluid that is useful in supporting the shaft load. The sealing effect increases when there are grooves, while these grooves prevent the effect of fluid lubrication. Hence, using these two surfaces, we can investigate the effects of both on the crticial pressures. In the experimental apparatus in Fig 1 the magnetic flux densities vary when the electric current in the coil of the electromagnet, which is connected to the permanent magnet, is varied. The permanent magnet is made of samarium-cobalt metal named Smco-15. The number of turns of the coil wire is 840,

0301-679X/93/010011-05 © 1993 Butterworth-Heinemann Ltd

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K. Nagaya et aI--Thrust bearing using a magnetic fluid lubricant under magnetic fields Geometry of the groove

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the radius of the core of the electromagnet is 54 mm and its length is 84 mm. The magnetic flux densities on the surface of the grooveless bearing plate versus the distance from the centre are depicted in Fig 3, where the solid curve denotes the bearing without the permanent magnet, the chain curve the permanent magnet only, and the two-dot chain curve the bearing with both the permanent magnet and the electromagnet. The magnetic flux density increases with the distance from the centre. Since there is a copper ring of radii range 19 mm to 23 mm, the magnetic flux density has a maximum value near the edge of the iron surface (the edge of the copper ring) and then decreases in the radial direction. The shape of the magnetic field under the flux due to the electromagnet is shown in Figs 4 and 5, where 0 A to 4 A represent the values of the electric currents in the coil of the electromagnet. It can be seen that the magnetic fluid gathers near the edge in the case of the grooveless surface bearing (see

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Fig 3 Magnetic flux densities in the radial direction at the surface of the bearing against the distance from the centre for a grooveless surface. (1) Permanent magnet and electromagnet with I = 4 A; (2) permanent magnet only; (3) electromagnet only with I = 4 A Fig 4) but, in the case of the grooved surface (see Fig 2(b)), the fluid is dispersed uniformly on the surface of the bearing plate (see Fig 5). Critical pressures of the bearing under the static load

As shown in Fig 1, as the load on the weight table increases, the gap between the shaft and the bearing plate decreases, and finally, the shaft contacts the bearing plate. The presure when the shaft contacts

1993 VOLUME 26 NUMBER 1

K. Nagaya et aI--Thrust bearing using a magnetic fluid lubricant under magnetic fields

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Fig 4 Shapes of the magnetic fluid on the surface of the grooveless bearing under the magnetic fields generated by the electromagnet. (a) 1 = 0 A; (b) I = 2 A; (c) 1 = 3 A; (d) I = 4 A the bearing is called here the 'critical pressure'. The critical pressure is calculated by measuring the electrical resistance between the shaft and the bearing plate because the magnetic fluid is non-conductive. The electrical resistance is very large (more than 5 kf~) when the shaft does not contact the bearing plate, due to the effects of the non-conductive magnetic fluid, but when the shaft contacts the bearing plate, the electrical resistance becomes significantly small (about 20 11). Then, the critical pressure can be calculated by measuring the electrical resistance when the shaft contacts the bearing plate. The side of the bearing is enclosed by a clear acrylic pipe, and the shaft is also made of an acrylic pipe having a short aluminium shaft at the end, as shown in Fig 1. The horizontal displacement is restrained by the shaft guide with guide rollers. The quantity of magnetic fluid is 3 g, the shaft diameter is 44.8 mm, and the inner radius of the acrylic pipe of the bearing is 46 mm. Figure 6 depicts the experimental data for the critical pressures against the electric current in the coil of the electromagnet. The marks (3, A, x and [] denote the experimental data, while the solid curve denotes the results obtained by the least squares method. The chain curve and the two-dot chain curve show the results for the motor oil and the spindle oil--these have no dependence on the magnetic flux density. The largest critical pressure for the bearing under the largest magnetic field, as shown in Fig 6, agrees with TRIBOLOGY

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Fig 5 Shapes of the magnetic fluid on the surface of the bearing with grooves under the magnetic fields generated by the electromagnet. (a) I = 0 A; (b) I = 2 A; (c) I = 3 A; (d) I = 4 A 6.0

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Fig 6 Critical pressures for the static load without rotation. X--grooved surface with permanent magnet and electromagnet; ©--grooveless surface with electromagnet only; []--grooved surface with electromagnet only; A--grooveless surface with electromagnet only; .... motor oil; - - - - - spindle oil

curve (1) in Fig 3. The pressure for the grooved surface is greater than that for the grooveless surface. This implies that the sealing effect is of importance in the case of static load without rotation. 13

K. Nagaya et aI--Thrust bearing using a magnetic fluid lubricant under magnetic fields Effects of temperature under the static load The viscosity of the lubricant decreases with temperature, consequently the critical pressure varies with oil temperature. To investigate this effect, this study also gives the results of critical pressure versus temperature using the experimental apparatus as shown in Fig 1, in which a thermocouple lies in the lubricant. Figure 7 shows the results. It shows that the critical pressure decreases linearly with temperature, but the decrease of lubricant viscosity is very non-linear. This implies that the sealing effect of the magnetic fluid has important roles in this phenomenon.

Critical pressures of the bearing under rotating thrust loads Figure 8 depicts the geometry of the experimental apparatus for measuring the critical pressure of the bearing under rotating thrust loads. The shaft (made of aluminium) is connected to a servometer. To avoid the shaft whirling, there are guides between the shaft and the guide rails. The load of the servomotor and the shaft is applied to the bearing by varying the tension in a wire which is connected to the winch, and the value of the applied load is measured by the load meter. If the critical load is greater than the gravity force of the motor and the shaft, the additional load is applied by pulling the load bar. The electrical resistance of the magnetic fluid between the shaft and the bearing is measured by the ohmmeter, in which the electric current is detected by the slip ring and brush. With the static load, since the electrical resistance decreases suddenly when the shaft contacts the bearing plate, the critical pressure could be obtained easily. However, for the rotating load, it is very difficult to decide the accurate critical load because of the existence of both boundary lubrication and the fluid lubrication when the oil film is significantly small. The electrical resistance of the shaft and the bearing, without the magnetic fluid, is only 20 1), but when there is a full magnetic fluid film, the resistance is

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( ~ Spring balancer Fig 8 Experimental apparatus for the bearing under the rotating thrust load more than 5 kl), as mentioned above. For the normal thrust bearing under normal oil lubrication, the critical pressure is often decided with the electrical resistance of the fluid being 100 1) (see Ref 8). The magnetic fluid is also non-conductive of electric current, as is the normal oil, so that the present study applies the resistance (100 1~) to calculate the critical pressure. Consequently, the actual critical pressure will be a little greater than its measured values. This implies that the measured critical pressure in this study gives the conservative value of the design. Figure 9 shows the critical pressures of the grooveless bearing versus the shaft rotating speeds for various kinds of fluid and various magnetic flux densities. The effects of grooves are depicted in Fig 10. These results are obtained under the constant temperature of 25°C. The figure shows that the critical pressure under the rotating load is significantly larger than that for the static load. This is due to the effects of fluid lubrication. The comparison between the critical pressure of a magnetic fluid lubricant and that of the normal oil lubricant suggests very interesting phenomena. The critical pressures for the normal oil lubrication under low speeds (<400 rpm) are greater than those 1993 V O L U M E 26 NUMBER 1

K. Nagaya et al--Thrust bearing using a magnetic fluid lubricant under magnetic fields 60

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However, at high speed (>400 rpm), the critical pressures for the magnetic fluid lubricant under strong magnetic fields become significantly larger than those for the normal oil lubricant. Therefore, the magnetic fluid lubrication has advantages for the bearing under high speed. From Fig 10, it can be also seen that the critical pressures for the grooved bearing are small compared with the grooveless bearing. This is the opposite tendency to that of the static load, and the reason for this is due to the fluid lubrication, because the grooves break the oil films. Therefore, the grooveless bearing has advantages for the thrust bearing.

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Fig 9 Critical pressures against the shaft rotating speeds for the grooveless bearing under the rotating thrust load. ×--motor oil; O--magnetic fluid without magnetic fields; 1--electromagnet with I = 4 A; A - permanent magnet; ~>--permanent magnet and electromagnet with I = 4 A

This article presents the thrust bearing using a magnetic fluid lubricant under magnetic fields. The results are summarized as follows. (1)

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Since the critical pressures for the magnetic fluid lubricant become larger than those of the normal oil lubricant at high speed, due to the effects of the magnetic fields, a magnetic fluid thrust bearing under strong magnetic fields has advantages compared with the normal oil thrust bearing. The grooveless bearing is better than the grooved bearing for a magnetic fluid thrust bearing at high speed.

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Acknowledgements The authors wish to thank the reviewers for their useful advice.

References

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1. Kamiyama S., Oyama T. and Htwe J. Basic study on the performance of magnetic fluid seals. Proceedings of the JSLE International Tribology Conference, Tokyo, Japan, 1985, 985-990

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2. Fertman V.E. Heat dissipation in high-speed magnetic fluid shaft seal. IEEE Trans. Magnetics, 1980, 16, (2), 352-357

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600 900 1200 Shaft rotationfrequency (rpm)

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Fig 10 Comparison between the critical pressures for the grooved bearing and the grooveless bearing under the fluxes due to the permanent magnet and electromagnet. O--with grooves for 1 - O; 1 - - w i t h grooves for I = 4 A; A--without grooves for I = O; ~ - - w i t h o u t grooves for 1 = 4 A for magnetic fluid lubrication. This implies that the lubrication effects of the motor oil on the critical pressure are larger than those of the magnetic fluid.

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3. Lutset M.O. and Starovoitov V.A. Experimental studies of highspeed cryogenic magnetic fluid seals. IEEE Trans. on Magnetics, 1980, 16, 343-346 4. Pinkus O. Model testing of magnetic-fluid seals, ASLE Reprint, No. 80-LC-2B-4., 1980 5. Moskowitz R. Dynamic sealing with magnetic fluids ASLE Trans. 1974, 18, (2), 135-143 6. Perry M.P. and Jones T.B. Dynamic loading of a single-stage ferromagnetic liquid seal. J. Appl. Phys., 1978, 49, (4), 2334-2338 7. Mikhalev Yu.O., Orlov D.V. and Stradomskii Yu.I. Ferrofluidic

seals. Magnetohydrodynamics, 1979, 15, 285-291 8. Bowden F.P. and Tabor D. The Friction and Lubrication of Solids, Oxford University Press, 1986, 247-260

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