Abstract:
The shaft may be supported for rotation by a conical bearing rotating within a sleeve. To prevent misalignment of the rotor and stator as the motor heats up and fluid viscosity changes, a magnetic preload is established; in a preferred embodiment, the magnetic preload is achieved using a magnetic back iron aligned with the stator magnet, the magnetic back iron being supported from the base. The shaft may further include a lower journal bearing for maintaining radial alignment and/or stiffness. 
     A shaft may be supported for rotation relative to a sleeve by a combination of journal bearing and thrust bearing whose gaps are connected and grooved to cooperate. The bearing system includes a magnetic preload at the end of the shaft distal from the journal bearing/thrust bearing combination, the magnetic force balancing the spiral groove thrust bearing to maintain the bearing support for the shaft and the load (including hub and disc) that it supports. Further, the journal bearing balances against the thrust bearing so that as fluid is drawn further into the thrust bearing, it is withdrawn from the journal bearing to reduce the working area of the journal bearing. A reservoir terminating in a capillary seal also provided on the far side of the thrust bearing from the journal bearing. This design allows the journal bearing to drain itself as the thrust bearing spins up and its pressure increases so that the pressure of the journal bearing matches the thrust bearing.

Description:
CROSS REFERENCE TO A RELATED APPLICATION 
     This application claims priority to two provisional applications, Ser. No. 60/363,986 filed Mar. 12, 2002, entitled CONSTANT PRESSURE MAGNETICALLY PRELOADED FDB MOTOR invented by Norbert Steven Parsoneault, Troy Michael Herndon and Jim-Po Wang, and provisional application Ser. No. 60/368,675, filed on Mar. 29, 2002, entitled CONSTANT PRESSURE CONICAL FDB invented by Norbert Steven Parsoneault and Troy Michael Herndon and incorporated herein by reference. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates to the field of fluid dynamic bearings, and more particularly to fluid bearings which are less temperature dependent. 
     BACKGROUND OF THE INVENTION 
     Disc drive memory systems have been used in computers for many years for storage of digital information. Information is recorded on concentric tracks of a magnetic disc medium, the actual information being stored in the forward magnetic transitions within the medium. The discs themselves are rotatably mounted on a spindle, while the information is accessed by read/write has generally located on a pivoting arm which moves radially over the surface of the rotating disc. The read/write heads or transducers must be accurately aligned with the storage tracks on the disk to ensure proper reading and writing of information. 
     During operation, the discs are rotated at very high speeds within an enclosed housing using an electric motor generally located inside the hub or below the discs. Such known spindle motors typically have had a spindle mounted by two ball bearing systems to a motor shaft disposed in the center of the hub. The bearings are spaced apart, with one located near the top of the spindle and the other spaced a distance away. These bearings allow support the spindle or hub about the shaft, and allow for a stable rotational relative movement between the shaft and the spindle or hub while maintaining accurate alignment of the spindle and shaft. The bearings themselves are normally lubricated by highly refined grease or oil. 
     The conventional ball bearing system described above is prone to several shortcomings. First is the problem of vibration generated by the balls rolling on the bearing raceways. This is one of the conditions that generally guarantee physical contact between raceways and balls, in spite of the lubrication provided by the bearing oil or grease. Hence, bearing balls running on the generally even and smooth, but microscopically uneven and rough raceways, transmit the rough surface structure as well as their imperfections in sphericity in the vibration of the rotating disc. This vibration results in misalignment between the data tracks and the read/write transducer. This source of vibration limits the data track density and the overall performance of the disc drive system. Vibration results in misalignment between the data tracks and the read/write transducer. Vibration also limits the data track density and the overall performance of the disc drive system. 
     Further, mechanical bearings are not always scalable to smaller dimensions. This is a significant drawback, since the tendency in the disc drive industry has been to continually shrink the physical dimensions of the disc drive unit. 
     As an alternative to conventional ball bearing spindle systems, much effort has been focused on developing a fluid dynamic bearing. In these types of systems lubricating fluid, either gas or liquid, functions as the actual bearing surface between a shaft and a sleeve or hub. Liquid lubricants comprising oil, more complex fluids, or other lubricants have been utilized in such fluid dynamic bearings. The reason for the popularity of the use of such fluids is the elimination of the vibrations caused by mechanical contact in a ball bearing system, and the ability to scale the fluid dynamic bearing to smaller and smaller sizes. 
     In such designs, the changing viscosity of the fluid with operating temperature of the bearing and or motor is a significant restraint on available designs. Thus, as the temperature changes, the power required to spin the motor will vary, if the gap remains constant; further, the stiffness of the system will diminish as the system heats and fluid viscosity diminishes. 
     Past efforts to address this problem have included using different metals in the shaft and sleeve so that the gap would change with changes in temperature; however, such solutions are typically relatively expensive. Accordingly, it would be advantageous to design a fluid bearing which minimizes the power required at start-up and constant speed rotation even as the viscosity of the fluid undergoes substantial changes. 
     SUMMARY OF THE INVENTION 
     The present invention is intended to provide a fluid dynamic bearing assembly especially useful in a high speed spindle motor assembly. 
     More particularly, the present invention is intended to provide a fluid bearing assembly in which the temperature influence on power requirements, and stiffness, is diminished. 
     These and other advantages and objectives are achieved by providing a fluid bearing design wherein a fluid bearing supports the shaft for rotation, with its positioning being axially compensated by a magnetic preload. By this combination, as the motor speeds up and heats up, which would otherwise cause the fluid pressure in gap to change, the magnetic preload maintains the pressure in the fluid between relatively rotating rotor and stator. 
     The shaft may be supported for rotation by a conical bearing rotating within a sleeve. To prevent misalignment of the rotor and stator as the motor heats up and fluid viscosity changes, a magnetic preload is established; in a preferred embodiment, the magnetic preload is achieved using a magnetic back iron aligned with the stator magnet, the magnetic back iron being supported from the base. The shaft may further include a lower journal bearing for maintaining radial alignment and/or stiffness. 
     A shaft may be supported for rotation relative to a sleeve by a combination of journal bearing and thrust bearing whose gaps are connected and grooved to cooperate. The bearing system includes a magnetic preload at the end of the shaft distal from the journal bearing/thrust bearing combination, the magnetic force balancing the spiral groove thrust bearing to maintain the bearing support for the shaft and the load (including hub and disc) that it supports. Further, the journal bearing balances against the thrust bearing so that as fluid is drawn further into the thrust bearing, it is withdrawn from the journal bearing to reduce the working area of the journal bearing. A reservoir terminating in a capillary seal also provided on the far side of the thrust bearing from the journal bearing. This design allows the journal bearing to drain itself as the thrust bearing spins up and its pressure increases so that the pressure of the journal bearing matches the thrust bearing. 
     In sum, according to the present arrangement, as the temperature changes, the viscosity of the fluid in the bearing gap will change; due to the effect of the magnetic bias, the thrust gap will change and the pressure will remain constant maintaining system stiffness. Further, as the viscosity changes the journal length the effective length of the journal bearing will change so that the two bearings, the journal bearing and the thrust bearing remain in balance. 
     It can further be seen that the design will be relatively easy to assemble requiring simply an injection of oil into the hub to fill the single gap between the hub/shaft combination and the sleeve. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       For a fuller understanding of the present invention, reference is made to the accompanying drawings in the following detailed description wherein  FIG. 1  is an elevation view and cross section of a computer hard disc drive spindle motor assembly taken along the axis of rotation of the spindle motor assembly; 
         FIG. 1  is a vertical sectional view of an exemplary embodiment of a disc drive motor incorporating a fluid dynamic bearing; 
         FIG. 1B  is a plan view of a typical grooving pattern used in a fluid dynamic bearing such as shown in  FIG. 1 ; 
         FIG. 2A  is a vertical sectional view of a constant pressure magnetically preload fluid dynamic bearing of a type incorporating the present invention; 
         FIG. 2B  is an expanded view of a section of the preferred embodiment of  FIG. 2A ; 
         FIG. 3  is a vertical sectional view of an alternative embodiment to the magnetically compensated constant pressure fluid dynamic bearing shown in FIG.  2 A; 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Reference will now be made in detail to exemplary embodiments of the invention, examples of which are illustrated in the accompanying drawings. While the invention will be described in conjunction with these embodiments, it is to be understood that the described embodiments are not intended to limit the invention solely and specifically to only those embodiments, or to use solely in the disc drive which is illustrated. On the contrary, the invention is intended to cover alternatives, modifications and equivalents which may be included within the spirit and scope of the invention as defined by the attached claims. Further, both hard disc drives, in which the present invention is especially useful, and spindle motors, where the invention is also especially useful are both well known to those of skill in this field. In order to avoid confusion while enabling those skilled in the art to practice the claimed invention, this specification omits such details with respect to known items. 
       FIG. 1  is a cross section through one embodiment of a spindle motor assembly which may readily be adapted to incorporate a fluid dynamic bearing arrangement according to the present invention. More specifically, the spindle motor shown in  FIG. 1  is of a fixed shaft design; whereas  FIGS. 2A ,  2 B, &amp;  3 A which are used to further illustrate the present invention are of a rotating shaft design. It will be apparent to a person of skill in the art that the present invention is useful with both either a fixed shaft or a rotating shaft design. 
       FIG. 1  illustrates a cross section through one embodiment of a spindle motor which may be adapted to incorporate a fluid dynamic arrangement according to the invention. The spindle motor assembly comprises a base  12  and a hub assembly  13 . A shaft  14  is mounted to the base  12  by a nut  16 . 
     The outer surface of the shaft  14  and the adjacent bore of the journal sleeve  26  together form hydrodynamic journal bearings  28 ,  30 . The dual reference numbers are used because the journal bearings are typically in two sections. The bearing gaps at the hydrodynamic journal bearings,  28 ,  30  are typically between 0.003 and 0.006. The journal bearings  28 ,  30  each include a grooved surface. The grooved surfaces may be provided either on the outer surface of the shaft  14 , or the inner bore surface of the journal sleeve  26 . 
     A thrust plate  32  is press fitted or formed or otherwise attached to an end of the shaft  14  and extends transversely to the shaft  14 . The thrust plate  32  is circular in form; the thrust plate  32  defines a first axial thrust surface  33  which, together with a facing sleeve thrust surface  35  extending transverse to the journal bearing defines a first fluid dynamic thrust bearing  34  in the gap between the two surfaces. As can be seen from  FIG. 1 , the disc thrust surface  35  at bearing  34  extends transversely to the journal at  30 , and the thrust bearing gap is connected to that journal bore. 
     A counterplate  36  is press fitted to or otherwise supported by the journal sleeve  26  adjacent the thrust plate surface  37  which is distal from the journal bearing  28 ,  30 . The counterplate  36  has a surface  39  which cooperates with the thrust plate surface  37  to define a gap in which fluid is maintained during rotational operation of the shaft and sleeve. Therefore, the counterplate  36  is sealed to the journal sleeve  26  by a O-ring  40  or similar means to prevent any loss of the fluid which appears in the gap between counterplate and thrust plate. 
     In use, the hub assembly  13  is rotated with respect to the base  12  by means of an electromagnetic motor. The electromagnet motor comprises a stator assembly  52  mounted to the base  2 , and a magnet  54  mounted to the journal sleeve  26 . 
     As can be appreciated from  FIG. 1 , the hub assembly  13 , which generally comprises the journal sleeve  26 , hub sleeve  50 , counterplate  36 , and first and second porous lubricant reservoirs  42  and  44 , is supported for rotation relative to the base  12  and shaft  14  on hydrodynamic bearings  28 ,  30 ,  34 , and  38 . 
     The operation of a hydrodynamic bearing can best be understood by reference to  FIG. 1B , which illustrates a plan view of one of the surfaces of a hydrodynamic thrust bearing. The illustrated hydrodynamic bearing surface, generally indicated by the numeral  60 , comprises a series of alternating grooves  62  and lands  64 . Each groove  62  comprises a leg which extends outward from the inner radius  66  of the hydrodynamic bearing surface  60  and a leg which extends inward from the outer radius  68  of the hydrodynamic bearing surface  60 . The two legs end in a point at an intermediate radius  70 . The plurality of grooves  62  and lands  64  together form a curved herringbone pattern as illustrated in the figure. 
     A hydrodynamic thrust bearing is formed when the bearing surface  60  is placed adjacent to an opposed bearing surface with a film of lubricant between the two surfaces. When the bearing surface  60  is rotated in the direction  72 , that is against the herringbone pattern, the grooves  62  and lands  64  tend to draw lubricant from the inner and outer radii  66  and  68  towards the points of the herringbone pattern at  70 . This creates a radial pressure distribution within the lubricant which serves to keep the bearing surfaces apart under external loading. 
     By varying the pattern of grooves  62  and the lands  64  in a known fashion, the pressure distribution across the hydrodynamic bearing can be varied. In particular, if the pressure in the bearing lubricant is greater at the inner radius  66  than at the outer radius  68  during operation, a net flow of lubricant from the inner radius  66  to the outer radius  68  will result, and vice versa. This can be done, for example, by having the intermediate radius  70 , at which the points of the herringbone pattern are located, closer to the outer radius  68 . Other ways in which the pressure distribution across the hydrodynamic bearing can be varied include altering the depth or width of the grooves, the number of grooves, or the angle the grooves make with a radial axis. The significance of having a net flow of lubricant across the bearing surface is discussed below. 
     The grooves  62  and  64  may be formed in the hydrodynamic bearing surface by any number of means including, for example, electroforming or stamping. 
     Although the operation of a hydrodynamic bearing has been discussed with reference to a hydrodynamic thrust bearing, it will be appreciated that the above principles can be applied to a hydrodynamic journal bearing such as the hydrodynamic journal bearings  28  and  30  illustrated in FIG.  1 . In particular, the pattern of the grooves and lands of the hydrodynamic journal bearings  28 ,  30  can be arranged to create a net flow of lubricant in a direction along the longitudinal axis of the shaft  14 , i.e. towards or away from the base  12 ; in this case it is toward the thrust bearing. 
     It will also be appreciated that a hydrodynamic bearing is not limited to the use of a herringbone pattern of grooves  62  and lands  64 . For example, a spiral or sinusoidal pattern may be used as an alternative to the herringbone pattern. The herringbone pattern is however preferred for thrust bearing arrangements as it generates a pressure distribution across the bearing surface which provides improved bearing rocking stiffness. Bearing rocking stiffness is a measure of the ability of a thrust bearing to resist rotation of the bearing surfaces relative to one another about an axis trasverse to the axis of rotation of the thrust bearing. 
     Referring again to  FIG. 1 , in use the hub assembly  13  (generally comprising the journal sleeve  26 , counterplate  36 , and the hub sleeve  50 , although it may include the hub and the shaft in embodiments such as shown in  FIGS. 2A ,  2 B and  3 A) is rotated relative to the base  12  by means of an electromagnetic motor comprising stator assembly  52  and magnet  54 . The hub assembly is supported for smooth rotation on the shaft  14  and thrust plate  32  by the pressures generated in the lubricant at the hydrodynamic bearings  28 ,  30 ,  34  and  38 . 
     The embodiments of the present invention are intended to minimize power consumption and maintain stability of the rotating hub. The problem is complicated by the fact that the relative rotation of the hub sleeve shaft combinations is typically supported by fluid whose viscosity changes with temperature. Moreover, the power consumption also changes with the change in viscosity of the fluid. At low temperature the viscosity is high and the power consumption is also relatively high. The larger the grooved areas, the greater the power consumption. The power consumption and also stiffness change with the width of the gap in which the bearing is established. In the typical designs as exemplified in  FIG. 1 , the gap is constant, and therefore the power consumption and stiffness vary as the viscosity of the fluid changes. 
       FIG. 2A  shows a fluid bearing comprising a sleeve  200  and a shaft  202  supporting a hub  204  for rotation in which the design is modified to maintain stiffness with changes in viscosity. The hub supports one or more discs  206 ,  208 . The design includes a conical fluid dynamic bearing  210  comprising a gap between the outer surface  212  of shaft  202  and the inner surface  214  of sleeve  200 . One of those two surfaces has grooves to maintain the pressure of a fluid  216  maintained in this gap to support the relative rotation of the shaft and sleeve. 
     The design shown includes a stator  222  supported on the outer surface of the base  224 , and cooperating with magnet  226  so that appropriate energization of the stator causes high speed rotation of the hub  204  and therefore the discs  206 ,  208 . A biasing magnet or magnet preload  230  is mounted on an axially facing surface of the disc base  224  and across a gap  232  from the magnet  226  which is used to drive the motor rotation. This is the simplest approach to establishing a constant magnetic axial bias against the shaft, to axially position the shaft  202  relative to the sleeve; of course a separate magnet which is not the magnet incorporated in the motor could also be used. 
     The magnet  226  is of course primarily in plane to interact with the stator  222  to drive the motor in constant speed rotation. However, there is also magnetic flux which leaks out of the base region  240  of the magnet  222  and interacts with the magnetic keeper or back iron  230  across the gap  232 . This force can be calibrated against the axial force established across the gap  210  of the fluid bearing between shaft  202  and sleeve  200 . Once this force is established, as the temperature changes, and the viscosity of the fluid changes, the fluid bearing gap will adjust so that the axial force across the gap remains substantially stable with changes in temperature. Further, with the use of the conical design, which provides both axial and radial support for the relatively rotating parts, good misalignment stiffness is established. It should be noted as to the magnet  226  that no major modifications to the motor magnet are required to implement this invention. It will be necessary to calibrate the gap  232  to establish the force across the gap relative to the desired force which is needed to establish and maintain the pressure in the gap  210  with changes in temperature of the fluid so that the fluid bearing is properly temperature compensated. 
     To reproduce the motor in high volume production, the gap should be set accurately so that by utilizing magnet  226  cooperating with back iron  230  of constant size separated by a constant gap, a constant force can be established; this force will establish the parameters for the rest of the motor so that a constant force is established across the bearing gap. 
     It should be noted that in this particular embodiment, a further fluid bearing  250  is defined between the outer surface of the shaft  202  and the inner surface of the sleeve  200 . This bearing is defined using well established technology, imposing grooves on either the outer surface of the shaft or the  202  or the inner surface of sleeve  200  with fluid in the gap supporting the relative rotation of the shaft and sleeve. 
     As shown in greater detail in  FIG. 2B , two other features need to be considered. To maintain the fluid in the gap  210  of the conical bearing, a radial capillary seal  270  is incorporated into the design. According to this feature, the sleeve  200  either incorporates or has integrated therewith or attached thereto a vertical arm  272  which has a shoulder  274  extending radially toward the shaft  202 . The sleeve itself further incorporates a slanted surface  280  which diverges slightly from the shoulder  274 . The angular surface  280  and the facing lower surface  282  of shoulder  274  together define an opening which is connected to the upper end of the gap  210  between sleeve  200  and shaft  202 . As the sleeve  202  rotates relative to the sleeve  200  and the shoulder  274 , any excessive fluid is thrown into this diverging gap where it is trapped so that the fluid cannot escape to the region outside the fluid bearing which incorporates gap  210 . This radial capillary seal  270  also functions as a reservoir so that fluid is always available for the conical bearing. A similar capillary seal may also be incorporated into the design adjacent the end of the journal bearing. 
     It should further be noted that a pin or the like  290  is impressed though an opening in the sleeve  200  so that it fits into a recess  292  which is defined in the outer surface of the shaft  202 . This recess  292  is a groove which extends all the way around the outer surface of the sleeve so that the shaft  202  may rotate past the sleeve and past the pin  290 ; however, the pin  290  will prevent axial separation or undue axial movement of the shaft  202  and sleeve  200 . 
     It should be noted that the spring pin is essentially a rolled piece of metal which can be inserted in the opening and then allowed to spring back against the sides of the opening thereby capturing the shaft relative to the sleeve. 
     Other features and advantages of the invention will become apparent to a person of skill in the art who studies the following disclosure of preferred embodiments. 
     Referring next to  FIG. 3A , we see a vertical sectional partial view of a alternative embodiment of a compensated constant pressure fluid bearing design. In this design, shaft  312  is supported for rotation relative to the sleeve  314 , and supports in turn a hub and one or more discs. A magnetic circuit comprising a pair of magnetic keepers  300 ,  302  and an intervening magnet  304  are supported from the base  310  adjacent the shaft  312 . A magnetic gap is established between the magnetic keeper  300  and the base of the shaft  312  which will create a magnetic preload axially along the shaft  312  tending to pull it toward the base. In a potential alternative, a ferro fluid  314  to enhance the magnetic circuit may be incorporated in the gap  316  between the magnetic keeper  300  and the base of the shaft  312  which is of a magnetic material. This magnetic preload, established across the gap  316 , with or without a ferro fluid incorporated therein, is balanced against the rotation supporting pressure of a combined thrust bearing  320  and journal bearing  322 . This magnetic bias is provided, to provide an offset and bias against the temperature dependent pressure variations of a combined journal bearing  322  and thrust bearing  320  which are in fluid communication to support rotation of the shaft  312  and hub  313 . The thrust bearing is defined by surfaces  326  and  328  which define a gap  330  in which fluid is maintained and pressurized by grooves on one of the two surfaces  326 ,  328 . The fluid is maintained in this gap in part by a reservoir and capillary seal which is defined adjacent to the thrust bearing. This reservoir  337  and capillary seal  339  are defined by surfaces  335 ,  336  which establish a gap  334  with diverging walls, the gap being adjacent to and parallel to the journal bearing  322 . The journal bearing  322  is defined by wall surfaces  340 ,  342 ,  344  which establish a gap  340 . Fluid  346  is maintained in this gap and pressurized therein by rotation of the sleeve which is supported from the base and the shaft  312  which rotates with hub  313 . 
     As described with the previous embodiment, the magnetic force across the gap  316  is primarily balanced against the pressure established in the thrust bearing at gap  330 ; as the viscosity changes, with temperature, the pressure across the gap is maintained by the fact that the shaft can adjust its position axially under the influence of the magnetic field so that the gap  330  pressure is reliably maintained to support rotation of the shaft and hub. The journal bearing established in the gap  340  is biased (by appropriate design of the groove pattern or the like) to pump toward the thrust bearing  330 . As the viscosity changes, the journal bearing will pump fluid toward the thrust bearing to maintain fluid therein; further, the groove pattern on the journal bearing is set to have a length such that the effectiveness of this journal bearing is diminished as fluid is pumped up toward and into the thrust bearing; this can occur because as fluid is pumped toward the thrust bearing, some portion of the journal bearing gap will have little or no fluid. The provision of a reservoir in the gap  335  which ends in the capillary seal  334  allows the journal bearing to balance against the thrust bearing; the reservoir which is on the far side of the thrust bearing but fluidly connected thereto from the journal bearing allows the journal bearing to drain itself until the pressure of the journal bearing matches that of the thrust bearing, with both of these bearings being balanced against to the magnetic bearing and the force which it establishes to maintain shaft stiffness.