HIGH SPEED ROTOR ASSEMBLY

Abstract: A ring-shaped magnet (19) is bonded to an outer circumferential portion of a rotary body (21) supported on a rotary shaft (13) which forms a radial dynamic gas pressure bearing. A ring-shaped magnet (23) is provided with a predetermined distance maintained between itself and the ring-shaped magnet (19). Owing to the magnetic attractive force of these two ring-shaped magnets (19, 23), the rotary body (21) is restricted in its axial movement and held radially rotatable. Since the rotary body (21) is held in a 0.5 decentered state due to the difference in the magnetic balance between the ring-shaped magnets (19, 23), the radial dynamic gas pressure bearing generates a high dynamic pressure. A high dynamic pressure occurs due to herringbone dynamic gas generating grooves (15a, 15b) as well. Owing to the synergetic effect of these parts, a very high radial load capacity occurs, and the rotational accuracy is improved.

Description: The present invention relates to a high speed rotor assembly which is strongly required to stably rotate at a high accuracy, such as rotary bearing for polygon mirror. Polygon mirrors are used in, for example, laser printers, bar code readers, facsimiles and laser copy machines. Since the efficiency of the polygon mirror is enhanced as the number of rotations increases, the number of rotations of 20000 r.p.m. or more has been requested. As the bearing for such a high speed rotor assembly, for example, precise ball bearings have heretofore been used. However, it is difficult to keep a stable accuracy of rotation for an extended period of time due to problems such as wearing and the limit of the rotational speed is about 16000 r.p.m. Accordingly, grooved pneumatic dynamic pressure bearings for bearing such a rotor which is rotating at a high speed are used. The grooved pneumatic dynamic pressure bearings have a feature of low noise and low friction loss since they bear a rotor via a gas film in a non-contact manner but they are weak for external shocks and vibrations of the rotor are liable to occur since the load bearing ability of the bearing is low and it is difficult to obtain a stable accuracy of rotation. In order to overcome the above-mentioned problems, there has been proposed for example a high speed rotor assembly in which the bearing rigidity is increased to suppress the vibration of the rotor 52 by enlarging the bearing 51 as is disclosed in JP-A-60-98213 and as is shown in the sectional view of Fig. 8 and a high speed rotor assembly in which thrust direction pneumatic dynamic pressure bearings 62a and 62b are provided at opposite ends of the radial pneumatic dynamic pressure bearing 61 to suppress the vibrations of the rotor 63 by the pneumatic dynamic pressure which is generated in a thrust direction and the size of a rotary shaft 61 is reduced as is disclosed in JP-A- 63-266420 and as is shown in the sectional view of Fig. 9. In the high speed rotor assemblies which use axially long pneumatic dynamic pressure bearing having an increased radial rigidity for bearing the rotor as is disclosed in the cited references, the rotor is larger than that of the high speed rotor assembly using conventional ball bearings, resulting in an increase in the size of the assemblies. The pneumatic dynamic pressure bearings generally require precise machining in which the accuracy of shape such as degree of cylindricality, roundness and surface roughness should be in the order of 1 µm to provide excellent accuracy of rotation. Therefore, the fact that the rotary shaft is long in an axial direction makes it difficult to perform machining Although it is possible to perform machining, the manufacturing cost is so high that it is not practical. Since the area of contact between the rotor and the thrust bearing surface is large in the rotor assembly in which the radial pneumatic dynamic pressure bearing is provided at both ends thereof with the thrust dynamic pressure bearings, a high starting torque is required for starting the rotor. Ringing effect (adsorption effect on the smooth face) is liable to occur on the thrust bearing face so that activation of the rotor assembly may not be performed. When the inner peripheral side of a cylindrical work piece is machined, the inner diameter at opposite ends is liable to be less or higher than that at the midposition thereof. In other words, the inner peripheral side of the work piece may be machined into arcuate shape as shown in Figs. 2(1) and 2(2). Therefore, unstable dynamic pressure distribution is liable to occur at the both ends of the radial bearing due to the changes in the gap of the bearing and the resulting changes in the number of rotations. Vibrations in a conical mode may occur due to occurrence of unforeseen vibrations, amplification of the vibrations of the rotor at a specific vibration frequency of synchronized vibration and resonation of the vibration with the magnetic force balance in a thrust direction. As a result, the bearing of the rotor may contact with the stationary shaft and the vibrations due to shock may not be dumped. JP-A-6-159 361 relates to a pneumatic and magnetic bearing type motor. This motor provides a hollow rotary member fitted rotatably on a stationary shaft, radial pneumatic bearings having dynamic pressure generating grooves formed on one of the stationary shaft and hollow rotary member, supporting the hollow rotary member in a radial direction with the aid of dynamic pressures generated, and a thrust magnetic bearing supporting the hollow rotary member in a thrust direction thereof. Therefore, the present invention was made in order to overcome the above-mentioned problems. It is an object of the present invention to provide a high speed rotor assembly which is compact in size and requires less starting torque, and has an excellent stable accuracy of rotation in a wide range from the beginning of rotation to a high speed rotation and a high shock resistance. This object is achieved with the features of the claims. The term "eccentricity" used herein is defined as the offset of the center of the rotor from the center of the radial bearing / the gap of radial bearing. The load bearing ability of a radial pneumatic dynamic pressure bearing of the invention for bearing a rotor may be larger than the radial load bearing ability of the magnetic bearing in a range of applicable numbers of rotations of said rotor (3000 - 40000 r.p.m.). The term "high speed rotor assembly" used herein generally refers to rotors having a dynamic pressure bearing mechanism. When the rotor is stopped to rotate by the magnetic bearing, the movement in an axial direction is restricted by the magnetic force so that the rotor is levitated in a given position in a thrust direction. A pneumatic dynamic pressure is generated in a gap between the stationary shaft and the rotary shaft by the herringbone grooves, when the number of rotations of the rotor increases. The rotor will have an increased bearing rigidity and damping characteristics in both radial and thrust directions against the external shock over a wide range from an initial rotation to high speed rotation due to the combined effect of the pneumatic dynamic pressure with the magnetic force of the ring-shaped magnets. The load bearing ability of the pneumatic dynamic pressure due to the herringbone grooves will become larger than that of the magnetic forces from the magnets as the number of the rotations of the rotor increases. This suppresses the whirl motion of the rotor ( precession of the rotor ) to provide a higher accuracy of rotation at a higher speed. By forming the bearing surfaces of said stationary and rotary shafts into smooth surfaces in the radial pneumatic bearing, the radial pneumatic bearing can be manufactured at a low cost. By forming the herringbone pneumatic dynamic pressure generating grooves on the bearing surface of the stationary shaft in the radial bearing, the stability can be enhanced. The annular grooves at the ends of the herringbone grooves formed on the bearing surface of said stationary shaft which is continuous thereto and has a larger depth prevent the distribution of dynamic pressure which occurs at the ends of the rotor. By making the stationary and rotary shafts of ceramic material, dead lock due to contact between stationary and rotary shafts made of metal which otherwise occurs can be prevented. In the following preferred embodiments of the present invention will be described in connection with the drawings, in which: Fig. 1 is a sectional view showing a high speed rotor assembly of a first embodiment of the present invention; Fig. 2 is a schematic view for illustrating the shape of a rotary shaft of the present invention; Fig. 3 is a schematic view for explaining the principle of a pneumatic dynamic pressure bearing; Fig. 4 is a schematic view for explaining the whirl movement which occurs in the rotor assembly using the pneumatic dynamic pressure bearing; Fig. 5 is a table showing a result of a simulation in which the performance of the prior art is compared with a high speed rotor assembly of the present invention; Fig. 6 is a sectional view showing a high speed rotor assembly of a second embodiment of the present invention; Fig. 7 is a table showing a result of simulation of the performance of the high speed rotor assembly of the second embodiment; Fig. 8 is a sectional view of a high speed rotor assembly for illustrating the prior art; and Fig. 9 is a sectional view of a high speed rotor assembly on which a miniaturized polygon mirror is mounted for illustrating the prior art. Now, preferred embodiments of the present invention will be described. Prior to the description of them, the principle of a pneumatic dynamic pressure bearing and whirl motion will be described. The pneumatic dynamic pressure bearing makes use of the characteristic which gas possesses as an viscous fluid. As shown in Fig. 3, load bearing capability is obtained by two effects. A first effect which is shown in Fig. 3(1) is a dynamic pressure effect in which when a gas is entrained into a tapered gap 33 between two opposite surfaces 31 and 32 by the sliding movement thereof, a pressure of the gas is built up to bear a depressing force, that is, a load, so that a gas film is formed between the sliding surfaces. A second effect which is shown in Fig. 3(2) is caused when two opposite surfaces move toward each other at a given speed. At this time, the gas 36 between the two surfaces 34 and 35 should be expelled at a rate which is proportional to the approaching speed of the two surfaces 34 and 35. This expelling of the gas causes the pressure of the gas to increase against the depressing forces of the two surfaces. This effect is a dynamic pressure effect due to non-steady gas film formation and is also referred to as squeeze effect. Fig. 4 is a schematic view explaining the whirl motion of a radial bearing. For ease of explanation, a vertical axis is assumed and no effect of gravity is considered. If it is assumed that the center of a rotary shaft 41 which is rotating at an angular velocity ω is offset from the bearing center 43 by e, a pressure is generated in the gas lubricating film between a fixed bearing 42 and the rotary shaft 41 due to the above-mentioned dynamic pressure effect. A force Fe which serves to return the rotary shaft 41 to the original bearing center 43 is generated, as well as a force Fω which causes precession of the rotary shaft 41 around the bearing center 43 . When the rotary shaft 41 begins precession at an angular speed Ω, a pressure is generated due to the squeeze effect, causing a force FΩ which serves to suppress the precession of the shaft 41. The centrifugal force acting on the rotor 41 is represented by meΩ 2 where m is the mass of the rotor 41 which is borne by the bearing. The rotor 41 will perform the precession at such an angular velocity Ω that meΩ 2 = Fe. If Fω = FΩ at this time, there is no force which accelerates or decelerates the precession. Accordingly, the steady precession of the rotor with a given radius 44 will continue. If Fω < FΩ, the precession is decelerated so that the centrifugal force will gradually decrease. If Fω > FΩ, the precession motion will be accelerated so that the centrifugal force will surpass the centripetal force and gradually increase. Therefore, in order to suppress the whirl motion from occurring, the centripetal force component Fe of the reaction of the gas film should be increased. It is necessary to decrease the influence of the dynamic pressure on the changing component of the reaction of the gas film without decreasing the squeeze effect. In the first embodiment of the present invention, the whirl motion is suppressed by increasing the centripetal component Fe of the reaction of the gas film by making the rotary shaft eccentric within a given range with a dynamic pressure caused by the herringbone shaped grooves and a magnetic force from a magnetic bearing. Now, the first embodiment of the present invention will be described with reference to drawings. Fig. 1 shows the schematic sectional structure of a high speed rotor assembly having a polygon mirror of the first embodiment. A stationary shaft 12 made of a ceramic material is provided in the center of a housing 11 made of aluminum so that is erects thereon. The ceramic stationary shaft 12 is formed with dynamic pressure generating grooves, or herringbone dynamic pressure generating grooves 15a, 15b having a depth of 6 to 8 microns, which are inclined in opposite directions with respect to the axial direction so that a maximum pressure is generated in the center 14 of bearing surface of the ceramic rotary shaft 13. Annular grooves 16a and 16b having a depth of 10 microns are formed at the opposite ends of the herringbone dynamic pressure generating grooves 15a, 15b. The annular grooves 16a, 16b have a width of 2.0 mm and extend outwardly beyond the edge of the bearing surface of the rotary shaft 13 by one third of their width. A ring-shaped magnet 23 is bonded on or secured to the inner side of the housing 11 so that it is coaxial with the ceramic stationary shaft 12. A flange 17 which is made of aluminum is secured to the outer periphery of the ceramic rotary shaft 13 by shrinkage fit. The aluminum flange 17 is formed with recesses on the lower and outer sides thereof. Ring-shaped magnets 18 and 19 are fitted in the recesses. A polygon mirror 20 is secured to the upper side of the aluminum flange 17 by means of adhesive or bolts. These components constitute a rotor 21. The rotor 21 is fitted on the ceramic stationary shaft 12 so that a small gap or clearance 22 which is in the order of about 5 microns is formed between the inner side of the rotor 21 (rotary shaft 13) and the outer side of the stationary shaft 12 and a gap which is 1.5 mm is formed between the ring-shaped magnet 19 fitted on the outer side of the rotor 21 and the inner side of the housing 11. The rotor 21 is prevented from moving in an axial direction by the magnetic attracting forces of the two ring-shaped magnets 19 and 23 and is levitated in a given position in a thrust direction. The two ring-shaped magnets 19 and 23 are positioned substantially coaxially with respect to the center of the bearing while the rotor 21 is eccentric relative to the center of the bearing by 2.5 µm (eccentricity = 0.5) due to the fact that the magnetic balance is slightly different between the magnets 19 and 23. Accordingly, a large dynamic pressure can be generated even at not higher than 10000 r. p. m. and shock resistance can be enhanced. It is preferable that the gaps 24 and 22 between the magnetized portion along the outer side of the rotor and the ring-shaped magnet and between the rotor and the ceramic bearing be 0.05 to 5 mm and 1 to 30 µm, respectively. If the gap 24 is not larger than 0.05 mm, the magnetic attracting force exceeds the radial load bearing capability. If the gap 24 exceeds 5 mm, the rotor can not be borne in a thrust direction and the centrifugal force becomes smaller. Accordingly, if the gap is not within this range, it is hard to obtain the eccentricity of 0.3 to 0.7 on rotation at a high speed. If the gap 22 is not larger than 1µm, the load bearing ability becomes too high, resulting in unstable rotation. The friction loss on steady rotation is high and the power consumption becomes high. The amount of generated heat becomes high so that various functions including rotation stability and damage due to contact occur by the changes in dimension of mechanical parts due to thermal expansion. If the gap 22 is larger than 30 µm, the load bearing ability is lowered, resulting also in less rotation stability. By adjusting the gaps 22 and 24 so that they fall within these ranges, the eccentricity can be stably kept to 0.3 to 0.7 on high speed rotation by the combination of pneumatic bearing and magnetic bearing. In the present invention, it is preferable that the magnetized portion be formed on the outer peripheral side of the rotor and in a plane including the centre of gravity of the rotor, which is normal to the axis of the rotor. This is the reason why the conical vibrations can be suppressed by the existence of the magnetized portion on the plane including the gravity center. The rotor 21 is rotated by means of a motor comprising a coil 25 secured to the housing 11 and a magnet 18 secured to the aluminum flange 17 and a rotation control unit (not shown). The rotor 21 starts to rotate by means of the motor. As the number of rotations of the rotor 21 increases, a pressure is generated in the bearing gap 22 by the herringbone pneumatic dynamic pressure generating grooves 15a, 15b so that the radial load bearing ability is obtained. The increase in the radial load bearing ability enhances the shock resistance of the bearing and suppresses the whirl motion of the rotor 21 so that the radius of the whirl motion becomes small to provide increased precision of the rotation. In other words, a pneumatic dynamic pressure is generated in a gap between the stationary shaft 12 and the rotary shaft 13 by the herringbone grooves 15a, 15b as the number of rotations of the rotor 21 increases. The rotor 21 will have an increased bearing rigidity and damping characteristics in both radial and thrust directions against the external shock over a wide range from an initial rotation to high speed rotation due to the combined effect of the pneumatic dynamic pressure with the magnetic force of the ring-shaped magnets 19, 23. The load bearing ability of the pneumatic dynamic pressure due to the herringbone grooves 15a and 15b will become larger than that of the magnetic forces from the magnets 19 and 23 as the number of rotations of the rotor 21 increases. This suppresses the whirl motion of the rotor and provides a higher accuracy of rotation at a higher speed. Unstable dynamic pressure distribution which occurs at both ends of the rotary shaft 13 is suppressed and contact between both ends of the rotary shaft 13 and the stationary shaft 12 is prevented from occurring due to the fact that the annular grooves 16a, 16b which are continuous with the herringbone grooves 15a, 15b has a larger depth than that of the herringbone grooves 15a, 15b and extend outwardly beyond the edge of surface of the radial bearing of the rotary shaft 13 by one third of its width. It is considered that the unstable dynamic pressure distribution can be suppressed by the fact that the annular grooves 16a, 16b are offset outwardly by 1/3 to 2/3 of their width relative to the edge of the radial bearing surface of the rotary shaft 13. By making the stationary and rotary shafts of ceramic material, dead lock due to contact between stationary and rotary shafts made of metal which otherwise occurs can be prevented. In the above-mentioned first embodiment of the high speed rotor assembly, the rotor 21 is eccentric with respect to the center of the bearing at 0.5 by the difference in magnetic balance between the ring-shaped magnets 19 and 23 as mentioned above. By making the rotary shaft 13 eccentric, a higher dynamic pressure is generated to enhance the accuracy of the rotation. It is found from the result of simulation that enhancement in the precision of the rotation due to the eccentricity of the rotor 21 provides a remarkable result when the eccentricity is not less than 0.3 on steady rotation. On the other hand, when the eccentricity is 0.7 or higher, the risk of contact between the stationary shaft 12 and the rotary shaft 13 occurs on application of a external force such as shock, etc. Therefore, the eccentricity is preferably 0.3 to 0.7. The results of comparison between the prior art shown in Fig. 9 and the simulation of the precision of rotation of the high speed polygon formed as mentioned above when rotated at 30000 r. p. m. is shown in Fig. 5. In the high speed rotor assembly of the first embodiment, the rotor 21 is levitated in an axial direction by the attracting forces of the ring-shaped magnets 19 and 23 as mentioned above. Therefore, the friction loss which occurs before levitation is eliminated so that the rise-up time until steady rotation is shorter and less starting torque is necessary. Particularly, in the first embodiment, a higher dynamic pressure can be generated since the rotor 21 is eccentric relative to the center of the bearing by the difference in magnetic balance between the magnets 19 and 23. A higher dynamic pressure can also be generated by the herringbone pneumatic dynamic pressure generating grooves 15a, 15b. It is deemed that these effects combine to generate a very high load bearing ability and to enhance the shock resistance of the bearing and the whirl motion of the rotor 21 is suppressed or at least minimized so that the radius of the whirl motion becomes small to enhance the accuracy of the rotation. It is considered that the mechanism of thrust dynamic pressure bearing is not necessary so that the power consumption is reduced since the rotor 21 is levitated in an axial direction by the above-mentioned magnets 19 and 23. It is understood from Fig. 5 that the high speed rotor assembly of the first embodiment has very excellent rotation characteristics. Now, a high speed rotor assembly of the second embodiment of the present invention will be described with reference to Fig. 6. Parts which are substantially identical with those of the first embodiment are denoted by like reference numerals and description thereof is omitted. In the second embodiment, the stationary shaft 112 of ceramic material is not provided on the surface thereof with herringbone pneumatic dynamic pressure generating grooves 15a, 15b, but its surface is smoothed. Similarly with the first embodiment, the rotor 21 is eccentric relative to the center of the bearing by 2.5 µm (eccentricity = 0.5) by the difference in magnetic balance between the ring-shaped magnets 19 and 23. In other words, by making the rotary shaft 13 eccentric, a high dynamic pressure is generated to prevent the rotor 21 from performing the whirl motion to provide a high precision of rotation. The result of simulation of the performance of the high speed rotor assembly of the second embodiment will be described with reference to Fig. 7. Fig. 7 shows the bearing number, stability characteristics, stability limit weight (kg) for the radial clearances of 0.0020, 0.0025, 00030 (mm) when the number of rotations is set to 8000 (r.p.m.) and the eccentricity is 0.1 to 0.8. It is found from the simulation result that a high stability limit weight can be obtained by making the eccentricity 0.3 or more. On the other hand, there occurs the risk of contact between the stationary shaft 12 and the rotary shaft 13 due to shocks, etc. as with the above-mentioned first embodiment when the eccentricity is 0.7 or more. Accordingly, it is preferable that the eccentricity be 0.3 to 0.7. The second embodiment has an advantage in that a high speed rotary polygon mirror can be provided at a low cost by omitting the herringbone pneumatic dynamic pressure grooves 15a, 15b which are difficult to be provided on the surface of a stationary shaft 112 of a ceramic material. As mentioned above, the high speed rotor assembly of the present invention can be used for a bearing for a polygon mirror used in the laser printer, bar code readers, facsimiles and laser copy machines. The rotor is also suitable for recording apparatus, machine tools, measuring instruments in which a member which is rotating at a high speed is borne by a thrust pneumatic dynamic pressure bearing.

Claims: A high speed rotor assembly including : a radial pneumatic dynamic pressure bearing having a housing (11), a stationary shaft (12) which erects on and within said housing (11) and a rotary shaft (13) surrounding said stationary shaft (12), which is rotatably borne; and a magnetic bearing having a rotor (21) which is secured to said rotary shaft (13) and which is provided along its outer peripheral side with a magnetized portion (19), a ring-shaped magnet (23) secured to said housing (11), which is positioned in the same plane as said magnetized portion (19) for bearing said rotor (21) in a radial and thrust directions so that a gap (24) is formed between said rotary and stationary shafts (13, 12) in a radial direction, wherein the radial load of said rotor (21) is borne by the combination of the load bearing ability of said radial pneumatic dynamic pressure bearing and the radial load bearing ability of said magnetic bearing, characterized in that said rotor (21) is borne at an eccentricity of 0.3 to 0.7, the eccentricity being defined as the offset of the center of the rotor (21) from the center of the radial bearing divided by the gap (22) of the radial bearing between the inner side of the rotor (21) and the outer side of the stationary shaft (12). A high speed rotor assembly as defined in Claim 1 in which the load bearing ability of said radial pneumatic dynamic pressure bearing for bearing said rotor (21) will become larger than the radial load bearing ability of said magnetic bearing as the number of rotations of said rotor (21) increases. A high speed rotor assembly as defined in Claim 1 or 2 in which the bearing surfaces of said stationary and rotary shafts (12, 13) in said radial pneumatic bearing are formed into smooth surfaces. A high speed rotor assembly as defined in Claim 1 or 2 in which said stationary shaft (12) of said radial pneumatic bearing is formed on the bearing surface thereof with herringbone pneumatic dynamic pressure generating grooves (15a, 15b) and annular grooves (16a, 16b) which are contiguous to the both ends of said herringbone pneumatic dynamic pressure generating grooves (15a, 15b). A high speed rotor assembly as defined in Claim 4 in which said annular grooves (16a, 16b) formed on the bearing surface of said stationary shaft (12) have a depth which is larger than that of said herringbone pneumatic dynamic pressure generating grooves (15a, 15b). A high speed rotor assembly as defined in claim 4 or 5 in which said annular grooves (16a, 16b) extend beyond the edge of the radial bearing surface of said rotary shaft (13) by 1/3 to 2/3 of its width outwardly in an axial direction. A high speed rotor assembly as defined in any of Claims 1 to 6 in which said stationary and rotary shafts (12, 13) are made of ceramic material. A high speed rotor assembly as defined in any of Claims 1 to 7 in which the gap (24) between said magnetized portion (19) and said ring-shaped magnet (23) is 0.05 to 5 mm and a gap (22) between said stationary shaft (12) and said rotary shaft (13) is 1 to 30 µm. A high speed rotor assembly as defined in any of Claims 1 to 8 in which said magnetized portion (19) is formed on the outer peripheral side of said rotor (21) and on a plane including the gravity center of said rotor (21) and which is normal to the axis of said rotor (21). A high speed rotor assembly as defined in claim 1 in which said stationary shaft (12) of said radial pneumatic bearing is formed on the bearing surface thereof with herringbone pneumatic dynamic pressure generating grooves (15a, 15b). A high speed rotor assembly as defined in Claim 10 in which said stationary shaft (12) of said radial pneumatic bearing is formed at the ends of the bearing surface thereof with annular grooves (16a, 16b).

Inventors: IZUME HIROYUKI, JINBO NAOYUKI, KUWAYAMA YOICHI, OHNO KAZUSHIGE, IZUME, HIROYUKI, JINBO, NAOYUKI, KUWAYAMA, YOICHI, OHNO, KAZUSHIGE, IZUME, Hiroyuki, C/O Ogaki-kita-kojou, Ibiden KK, JINBO, Naoyuki, C/O Ogaki-kita-kojou, Ibiden KK, KUWAYAMA, Yoichi, C/O Ogaki-kita-kojou, Ibiden KK, OHNO, Kazushige, C/O Ogaki-kita-kojou, Ibiden KK
Applicants: FUJI XEROX CO LTD, IBIDEN CO LTD, FUJI XEROX CO., LTD., IBIDEN CO, LTD.
Application year: 1995
Main classification: F16C32
IPC codes: F16C17, H02K7, G02B26, F16C33, H02K5