HYDRODYNAMIC BEARING AND A METHOD FOR INTRODUCING LUBRICANT INTO THE BEARING

Abstract: A rotatable shaft/thrust plate combination is disposed within a sleeve to form a first clearance space between the shaft and the sleeve and a second clearance space between the thrust plate and the sleeve. The external faces of the thrust plate are exposed to air. The clearance spaces are filled with a liquid lubricant and the sleeve includes pressure equaliza­ tion ports connecting the first and second clearance spaces. Surface tension dynamic seals are provided between axially extending surfaces of the thrust plate and sleeve. The equalization ports balance the hydrodynamic pressures in the lubri­ cant to prevent the lubricant from being pumped through one of the dynamic seals. The resulting bearing provides high precision with low repetitive and nonrepetitive runouts. The bearing provides hydrodynamic support of both radial and axial loads and the bearing seal is relatively insensitive to orientation of the spindle and minimizes the generation of debris and contaminating particles.

Description: This invention related to precision hydrodynamic bearings. One important limitation to increasing track density of computer disk drives is spindle bearing performance. A disk drive whose spindle bearing has low runout can accommodate higher track densities which results in more data storage capacity per disk. The kinematics of the spin axis of a spindle bearing determine the precision of the bearing. As the journal spins relative to the sleeve, the spin axis may trace out a path or orbit. The motion of this axis typically has components that are synchronous with the spin and repetitive in nature. These motions are termed repetitive runout. Other components of spin axis motion may be asynchronous and nonrepetitive with respect to spin. These components are termed nonrepetitive runout. As a general rule, spindle bearing precision is increased as repetitive and nonrepetitive runouts are decreased. Ball bearing spindle systems make up the majority of prior art disk drives. The kinematics of the rolling elements in ball bearings result in relatively large nonrepetitive runout. This results from the fact that the lubricant film thicknesses in ball bearings are very thin providing little attenuation of geometric defects in the bearing. In addition, ball bearings produce forces on the disk drive structure to which it is attached which are of relatively high frequency and large amplitude. Hydrodynamic spindle bearing designs are also known. The Hewlett-Packard Model No. 9154A, 3.5 inch micro-Winchester disk drive incorporates a hybrid hydrodynamic-ball bearing spindle. The performance of this bearing is degraded by the incorporation of the ball bearings. The Phillips video 2000 videocassette recorder utilizes a hydrodynamic bearing which employs grease as the lubricant limiting operation to low speeds. Other known hydrodynamic sping bearings for disk drives employ a ferromagnetic fluid as the lubricant for the bearing. This fluid is retained or sealed in the bearing by magnetic fields set up in pole pieces at each end of the bearing. Unless the magnetic fields and clearances are very precisely matched at each end of the bearing, one seal will be stronger than the other and when the bearing heats up, the lubricant can be spilled. See U.S. Patent No. 4,526,484. U.S. Patent No 4,274,684 discloses a hydrodynamic bearing in accordance with the pre-characterised portions of claims 1 and 12. The present invention provides hydrodynamic bearing comprising a shaft disposed within a sleeve for relative rotation thereof, the shaft/sleeve combination including mating portions defining a first lubricant fillable clearance space for supporting radial loads, an external face of said shaft/sleeve combination being exposed to an outside environment, said shaft/sleeve combination having cooperating axially extending surfaces which diverge in a tapered configuration as they open out axially to the outside environment, said outwardly diverging surfaces being fillable with a liquid lubricant and being configured for forming with the liquid lubricant, at the lubricant/outside environment interface, an external surface tension dynamic seal for sealing said clearance space; said sleeve including pressure equalization ports; characterized in that thrust plates are provided on said shaft and that said mating portions further define a second lubricant fillable clearance space for supporting the axial loads, in that said outwardly diverging surfaces diverge towards the end of the bearing a straight taper at an angle sufficient to form the surface tension dynamic seal, in that the pressure equalization ports are coupled between said first and said second clearance spaces and in that said first and second clearance spaces are configured such that a differential in lubricant pressure in said clearance spaces relative to pressure of the outside environment can be equalized by movement of the lubricant in the equalization ports and by movement of said seal along its associated taper until surface tension and pressure forces are in balance at least during said relative rotation. The present invention also provides hydrodynamic bearing comprising a shaft disposed within a sleeve for relative rotation thereof, the shaft/sleeve combination including mating portions defining a first lubricant fillable clearance space for supporting radial loads, an external face of said shaft/sleeve combination being exposed to an outside environment, said shaft/sleeve combination having cooperating axially extending surfaces which diverge in a tapered configuration as they open out axially to the outside environment, said outwardly diverging surfaces being fillable with a liquid lubricant and being configured for forming with the liquid lubricant, at the lubricant/outside environment interface, an external surface tension dynamic seal for sealing said clearance space; said sleeve including pressure equalization ports; characterized in that: said shaft/sleeve combination (102,112) further includes mating tapered portions which define a second lubricant fillable clearance space for also supporting axial loads; that said outwardly diverging surfaces diverge toward the end of the bearing in a straight taper at an angle sufficient to form said surface tension dynamic seal; and that said pressure equalization ports are coupled between said first and second clearance spaces, and that said clearance spaces are configured such that a differential in lubricant pressure in said clearance spaces relative to pressure of the outside environment can be equalized by movement of the lubricant in the equalization ports and by movement of said seal along its associated taper until surface tension and pressure forces are in balance at least during said relative rotation. The hydrodynamic bearing of the instant invention achieves lower levels of runout than ball bearings as a result of a thick film of lubricant which separates the sliding metal surfaces. This film provides a high degree of viscous damping which significantly attenuates nonrepetitive runout to levels which are less than state of the art rolling element bearings. In addition, the bearing generates forces on the structure attached to it which are low frequency and low amplitude relative to ball bearings. This reduction in the forcing function bandwidth and amplitude minimizes other vibrations in the disk drive and further improves tracking performance. The pressure equalization ports reduce pressure differentials which are caused by pumping actions inside the bearing. Because of the pressure balancing, the bearing does not tend to pump lubricant in a preferential manner through one seal or the other. Thus, only the external pressure differential across the bearing influences the position of the dynamic seal interfaces. The surface tension seals of the present invention do not leak nor do they generate solid debris. Brief Description of the Drawing Fig. 1 is a cross-sectional view of the bearing of the invention; Fig. 2 is an elevational view of the sleeve portion of the bearing; Fig. 3 is an expanded view of a portion of Fig. 1; Fig. 4 is an expanded view of a portion of Fig. 1; Fig. 5 is an expanded view of a portion of Fig. 4; Fig. 6 has been removed from the present specification; Fig. 7 is a cross-sectional view of a particularly preferred embodiment of the present invention; and Fig. 8 is a cross-sectional view of an embodiment of the invention utilizing a tapered shaft. Description of the Preferred Embodiment A hydrodynamic bearing 10 shown in Fig. 1 includes a sleeve 12 including a portion of smaller inside diameter 14. A journal or shaft 16 fits within the sleeve 12 forming a first clearance space 18. The journal 16 may include a recess 20. Thrust plates 22 and 24 rest on the journal 16 and are sealed by means of O-ring seals 26. The thrust plates 22 and 24 form second clearance spaces 28 with respect to radially extending surfaces of the smaller inside diameter portion 14 of the sleeve 12. The portion 14 of the sleeve 12 also includes axially extending passageways 30 and radially extending passageways 32. As shown in Fig. 2 the passageways 30 and 32 are arranged around the circumference of the sleeve 12. Four sets of passageways 30 and 32 are shown in Fig. 2 but more or fewer may be employed. Fig. 2 also shows spiral relief patterns 34. These relief patterns cooperate with patterns on the journal to generate radially directed inward hydrodynamic pressure. Relative rotation between the journal 16 and the sleeve 12 is provided for by the clearance spaces 18 and 28. Suitable dimensions for the clearance spaces 18 and 28 are 0.005 to 0.025 mm (0.0002 to 0.001 inches) and 0.0127 to 0.05 mm (0.0005 to 0.002 inches), respectively. These clearance spaces are filled with a lubricant such as oil which reduces wear between the journal and sleeve and provides a medium through which a hydrodynamic pressure field may be generated. Relative rotation or radial motion between the journal 16 and sleeve 12 is required to set up the hydrodynamic pressure field. The hydrodynamic bearing 10 supports loads by metal-to-metal contact when there is no relative motion. During normal operation, the spinning of the journal 16 sets up a steady pressure field around the clearance spaces which pushes the journal and sleeve apart and thus prevents metal-to-metal contact. The hydrodynamically pressurized film provides the stiffness needed to support the radial load of the disk, motor and associated hardware. Note that the hydrodynamic film stiffness is a measure of the resistance of the clearance space to change size under the influence of a load. Axial loads along the journal 16 spin axis are supported by the hydrodynamic pressure field in the clearance spaces 28 between the thrust plate faces and the sleeve portion 14. The amount of separation between the thrust plate faces and sleeve is controlled by the hydrodynamic film stiffness and the applied axial load (usually the weight of the entire rotating assembly). Pressure building geometries such as the relief pattern 34 shown in Fig. 2 are employed to generate film stiffness of sufficient magnitude. The sealing of the lubricant within the hydrdynamic bearing 10 will now be described in conjunction with Figs. 1, 4 and 5. There are two types of seals in the bearing 10, namely, static and dynamic seals. Static seals 26 which are preferably O-ring seals prevent lubricant leakage between the thrust plates 22 and 24 and the journal 16. They are called static seals in that there is no relative rotation or sliding between the thrust plates 22 and 24 and the journal 16. Dynamic sealing is required in the clearance space 36 between the thrust plates and the sleeve. These seals must not leak or generate solid debris. Sealing is provided by surface tension-capillary seals in which a lubricant-air interface 38 provides the surfaces forces. As shown in Fig. 5, two components, the liquid-gas (lubricant-air) interface 38 and the solid surfaces of the thrust plates and sleeve make up each seal. Surface tension forces directed axially away from each end of the bearing indicated by the arrows 40 balance the forces due to pressure differentials which may be applied across each interface as indicated by the arrows 42 and a force due to gravity. The magnitude of the axial surface tension forces depends on the wetted perimeter of the liquid-gas interface 38, the surface tension (a property of the liquid lubricant), the taper angle and the contact angle. The forces due to pressure differentials are dependent on the pressure differentials and the lubricant-air interface area. Since the solid boundaries of the seal are tapered, the wetted perimeter and area of the interface vary with the axial position of the interface. As a result, the axial position of the interface varies with pressure differences applied to the bearing until the surface tension forces and pressure forces balance. Stability of the interface is sensitive to the angle of taper. A taper angle of approximately 2° has been experimentally determined to be optimum for insuring interface stability. During bearing 10 operation, it is necessary that the pressures be nearly the same at the lubricant side of each lubricant-air interface 38. This pressure balance is provided by the pressure equalization ports 30 and 32 which connect the clearance spaces 18 and 28. Without the equalization ports, pumping actions inside the bearing may set up pressure differentials. For example, the thrust plates 22 and 24 produce an inwardly directed radial pumping action. The equalization ports tend to equalize the pressures. Furthermore, the passages should maintain a constant radial position in the neighborhood of the thrust plates. This requirement prevents large pressure gradients from developing in the passages due to the centrifugal pumping effects caused by the thrust plates. The bearing 10 is thus pressure balanced and does not tend to pump the lubricant in a preferential manner through one seal or the other. Only the external pressure differential across the bearing, therefore, influences the position of the interfaces. The equalization ports coupled with the surface tension dynamic seals result in a hydrodynamic bearing of higher precision with respect to runout relative to conventional bearing designs. Lubricant must be introduced into the bearing in such way that a minimal amount of air is trapped in the bearing. This is necessary because trapped air in the bearing expands as the bearing heats up and tends to push the lubricant out of the bearing. Fig. 7 is a particularly preferred embodiment of the invention having several advantages as compared to the embodiment of Fig. 1. A bearing 70 includes a shaft 72 with thrust plates 74 and 76. The shaft 72 with attached thrust plates 74 and 76 rotates within a sleeve 78. The sleeve 78 includes a portion having increased inside diameter to create an air space 80. The bearing 70 includes external surface tension seals 82 and internal surface tension seals 84. The external surface tension seals 82 and internal surface tension seals 84 are connected by pressure equalization ports 86. The surface tension seals 82 and 84 and the pressure equalization ports 86 are filled with a lubricant. As with the embodiment of Fig. 1, the surface tension seals are created by diverging, axially extending surfaces. The embodiment of Fig. 7 results in reduced evaporation rate of the lubricant from the seals. When the orientation of a bearing changes, the position of the surface tension seals along the spin axis also changes. In the case in which the oil-air interface moves into the bearing, a film of oil is left on the region of the metal which was previously covered by the lubricant of the seal. This film of oil is then exposed to air and has a large amount of surface area compared to the seal oil-air interface area. As a result of this increased surface area, the evaporation of the oil is increased and the life of the lubricant supply is reduced. When a bearing is not operating, the position of the seals is determined by the pressure difference between the two sealed regions of the bearing which are connected together by the pressure equalization port or balance tube. The internal fluid pressure difference is controlled by the elevation difference between the two regions of the bearing and the specific weight of the lubricant fluid. The external pressure differences due to variations in air pressure around the bearing are usually negligible. Thus seal position and the change in seal position are controlled primarily by the elevation changes in the bearing. Splitting the lubricated regions of the bearing of Fig. 7 into two separate and shorter zones reduces the range of possible elevation differences and also the resulting range of seal position changes. This design thus reduces the wetted area of the bearing and the evaporation rate. The bearing of Fig. 7 also provides higher moment stiffness. The higher stiffness results from the fact that the length of the bearing can be made longer relative to the bearing of Fig. 1. Moment stiffness is proportional to the length of the bearing squared when all of the other bearing characteristics are held constant. The bearing of Fig. 7 can be longer than the bearing of Fig. 1 because the seal areas are split into separate zones so that the central region of the bearing can be lengthened without affecting the behavior of the seals. Another advantage of the embodiment of Fig. 7 is faster thermal transient response of the lubricant. It is desirable to have the lubricant come up to temperature as fast as possible during start up. When the lubricant oil is warm, it has a lower viscosity than when it is cold and thus the torque requirements are less when the oil is warm. Accordingly, when the oil can be made to heat up quickly, a shorter period of high load on the driving motor results which is very desirable for some applications. The faster thermal response of the bearing of Fig. 7 results from the reduction of oil volume in this bearing design and the resulting increase in bearing power to oil volume ratio. Fig. 8 is yet another embodiment of the present invention. A bearing 100 includes a spindle shaft 102 which has tapered portions 104 and 106. These tapered portions mate with tapered bearing shells 108 and 110 which reside within a spindle housing or sleeve 112. The spaces between the tapered shaft and tapered bearing shells are filled with a liquid lubricant. The lubricant is sealed by external surface tension or capillary seals 114 and 116, and internal capillary seals 118 and 120. An equalization port 122 connects the seals 114 and 118, and an equalization port 124 connects the seals 116 and 120. Because of the tapered surfaces, both radial and axial loads are supported by the bearing. The spindle housing and shaft surfaces are a single contiguous unit without any parting line. No O-ring seals are required since no secondary leakage is possible with the tapered arrangement. The tapered portions of the bearing shaft or the tapered bearing shell surfaces include herringbone patterns which generate a net liquid flow due to machining tolerances. This net liquid flow in the bearing is compensated for by a flow in the opposite direction through the equalization ports 122 and 124. The bearing shells 108 and 110 have grooves on their outer surfaces. These bearing shells are shrink fitted into the spindle housing 112 and the grooves cooperate with the housing 112 to create the equalization ports.

Claims: Hydrodynamic bearing comprising a shaft (16) disposed within a sleeve (12) for relative rotation thereof, the shaft/sleeve combination (16,12) including mating portions defining a first lubricant fillable clearance space (18) for supporting radial loads, an external face of said shaft/sleeve combination (16,12) being exposed to an outside environment, said shaft/sleeve combination (16,12) having cooperating axially extending surfaces which diverge in a tapered configuration as they open out axially to the outside environment, said outwardly diverging surfaces being fillable with a liquid lubricant and being configured for forming with the liquid lubricant, at the lubricant/outside environment interface, an external surface tension dynamic seal for sealing said clearance space (18); said sleeve (12) including pressure equalization ports: characterized in that thrust plates (22,24) cooperating with said sleeve (12) are provided on said shaft (16) such that said mating portions further define a second lubricant fillable clearance space (28) for supporting the axial loads, in that said outwardly diverging surfaces diverge towards the end of the bearing a straight taper at an angle sufficient to form the surface tension dynamic seal, in that the pressure equalization ports (30,32) are coupled between said first and said second clearance spaces and in that said first and second clearance spaces (18,28) are configured such that a differential in lubricant pressure in said clearance spaces (18,28) relative to pressure of the outside environment can be equalized by movement of the lubricant in the equalization ports and by movement of said seal (36) along its associated taper until surface tension (40) and pressure forces (42) are in balance at least during said relative rotation. The bearing of claim 1 characterised in that the pressure equalization ports (30,32) include axially extending passageways (30) in communication with radially extending passageways (32). The bearing of claim 2 wherein the radially extending passageways (32) are located near the center of the bearing. The bearing of claim 1 characterised in that it includes four equalization ports (30,32). The bearing of claim 2 characterised in that the axially extending passageways are equally spaced around the sleeve. The bearing of claim 1 further including static seals (26) between the shaft and the thrust plates. The bearing of claim 1 further including relief patterns (34) in opposed sleeve/thrust plate faces to generate inwardly directed radial forces. The bearing of claim 6 wherein the static seals are O-ring seals. The bearing of claim 1 characterised in that each end of one bearing includes also internal surface tension dynamic seals with an air space disposed between the ends. The bearing of claim 9 wherein the pressure equalization ports connect the internal and external surface tension seals. The bearing of claim 10 wherein the pressure equalization ports are disposed diagonally. Hydrodynamic bearing comprising a shaft (102) disposed within a sleeve (112) for relative rotation thereof, the shaft/sleeve combination (102,112) including mating portions defining a first lubricant fillable clearance space for supporting radial loads, an external face of said shaft/sleeve combination (102,112) being exposed to an outside environment, said shaft/sleeve combination (102,112) having cooperating axially extending surfaces which diverge in a tapered configuration as they open out axially to the outside environment, said outwardly diverging surfaces being fillable with a liquid lubricant and being configured for forming with the liquid lubricant, at the lubricant/outside environment interface, an external surface tension dynamic seal for sealing said clearance space; said sleeve (112) including pressure equalization ports; characterized in that: said shaft/sleeve combination (102,112) further includes mating tapered portions which define a second lubricant fillable clearance space for also supporting axial loads; that said outwardly diverging surfaces diverge toward the end of the bearing in a straight taper at an angle sufficient to form said surface tension dynamic seal; and that said pressure equalization ports (122,124) are coupled between said first and second clearance spaces, and that said clearance spaces are configured such that a differential in lubricant pressure in said clearance spaces relative to pressure of the outside environment can be equalized by movement of the lubricant in the equalization ports (122,124) and by movement of said seal (36) along its associated taper until surface tension (40) and pressure forces (42) are in balance at least during said relative rotation. The bearing of claim 12 wherein the shaft and sleeve are each a continuous unit with no parting line. The bearing of claim 12 wherein the sleeve includes separate bearing shells having the tapered surfaces. The bearing of claim 14 wherein the bearing shells include grooves which cooperate with the sleeve to form the equalization ports. The Bearing of claims 1 and 12, wherein the straight taper is approximately 2°.

Inventors: CODROVA JACKIE, SCHAULE MAX WERNER, TITCOMB FORREST, CODROVA, JACKIE, SCHAULE, MAX WERNER, TITCOMB, FORREST
Applicants: DIGITAL EQUIPMENT CORP, DIGITAL EQUIPMENT CORPORATION
Application year: 1988
Main classification: F16C17
IPC codes: F16C33, F16C17, G11B19