Patent Grant
7042125
The present invention relates to the field of fluid dynamic bearings, and more particularly to a fluid bearing incorporating an asymmetric journal bearing, to reduce cost and/or power requirements.
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 of 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.
Many current fluid dynamic bearing designs are a combination of journal and thrust bearings. Frequently, these designs include a shaft journal bearing design having a thrust plate at an end thereof, or a dual conical bearing design, including a conical bearing at or close to either end of the shaft. The conical bearings typically include a grooved surface on each cone; the thrust plate bearings typically include two grooved surfaces, one facing each of the gaps defined by the thrust plate and sleeve, and by the thrust plate and counterplate. Because of these multiple grooved surfaces, it is difficult to make motors which are low in cost, and have low power requirements, because of the fact that each of the grooved surfaces, both at start-up and at steady state, require a power budget to start and maintain constant speed rotation.
The present invention intended to produce a fluid dynamic bearing assembly which is especially useful in a high speed spindle motor assembly.
More particularly, the present invention is intended to provide a fluid bearing assembly in which one aspect power consumption by the bearing assembly is reduced.
In another aspect of the invention, a fluid dynamic bearing assembly is proposed in which the cost of assembly of the motor is reduced.
These and other advantages and objectives are achieved by providing a design including a shaft having a journal bearing for providing radial support, and a conical or thrust plate or similar bearing at or near an end of the shaft.
A fluid bearing design is provided which according to one aspect includes a shaft defining together with a surrounding sleeve an asymmetric journal bearing, and a thrust bearing at or near an end of the shaft towards which the asymmetric journal bearing is pumping, with that end of the shaft being closed off. The journal bearing asymmetry establishes a hydraulic pressure toward the closed end of the shaft. This pressure provides an axial thrust to set the bearing gap for the conical bearing. The conical bearing itself is a relatively balanced bearing, although it may have a bias pumping toward the shaft and the journal bearing.
A pressure closed equalization path from the journal bearing through the conical bearing to the end of the shaft may be established to maintain a constant hydraulic force across the conical bearing, and which may also prevent any asymmetry in the conical bearing from affecting the net thrust force acting upon the end of the shaft where the conical bearing is located.
Alternatively, in a fluid dynamic bearing design comprising a shaft and a thrust plate at or near an end of the shaft, asymmetry is again established along the journal bearing to establish a pressure gradient directed toward the thrust bearing. In a preferred aspect of the invention, the gap between the thrust plate and the counterplate is made somewhat wider than normal. Journal bearing asymmetry establishes a hydraulic force to the bottom end of the shaft, distal from the location of the thrust plate. The pressure is intended to provide axial thrust against the end of the shaft where the thrust plate is mounted to establish and maintain the bearing gap for the thrust bearing. In a preferred aspect of the invention, a pressure equalization path from the thrust bearing toward the distal end of the shaft is established to help prevent asymmetry in the thrust bearing from affecting the net thrust force acting upon the end of the shaft adjacent the thrust bearing.
Other features and advantages of the invention will be apparent to a person of skill in the art who studies the disclosure. Of some preferred embodiments and aspects of the invention given with respect to the following figures.
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.
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 0.001 and 0.006 although other gap widths may be useful. 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
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.
The hub assembly 13 is rotated with respect to the base 12 in use by means of an electromagnetic motor. The electromagnet motor comprises a stator assembly 52 mounted to the base 12, and a magnet 54 mounted to the journal sleeve 26.
As can be appreciated from
The operation of a hydrodynamic bearing can best be understood by reference to
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 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 new 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 lands 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
It will also be appreciated that a hydrodynamic bearing is not limited to the use of a particular pattern of grooves 62 and lands 64. For example, a spiral or sinusoidal pattern may be used as an alternative to the herringbone pattern.
Referring again to
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 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
The present design provides a fluid dynamic bearing in which a net hydraulic pressure is generated by a asymmetric journal bearing located on the shaft of the motor, establishing a thrust force toward a second bearing surface which is located off axis from the shaft. An advantage of this approach is that the need for accurate setting of axial play in the off-axis surface is diminished. The pressure created by asymmetric journal bearing exerts a positive thrust force on the end of the shaft which displaces the shaft axially, such that a second bearing surface located off-axis from the shaft moves against a counter bearing surface, thus creating a counter thrust force. Thus, the hydraulic force and the counter thrust force so established combine to form a thrust bearing. The present invention, is intended to simplify the design, reduce cost and/or power consumption which is typically associated with the use of thrust bearing and conical bearing designs; it also may reduce the overall height of the design. The fewer grooved surfaces that are created provide these beneficial results.
A first aspect of the present invention appears in
This conical bearing comprises an outer surface 312 of cone 314, the inner surface 316 of sleeve 302 and the fluid in the gap 318 which is thereby to support relative rotation of the sleeve and the counterplate 320 that is supported over and around the shaft 300 and the cone 314. To establish the fluid pressure which supports the relative rotation of the sleeve 306 over the cone 314, one of the two surfaces of the non-axial (in this instance conical) bearing defined by surfaces 312 and 316 is grooved in accordance with known principles. Relative rotation of these two surfaces will build up pressure in the gap to keep the surfaces out of contact and allow for relative rotation. In operation, the asymmetric journal bearing 302 defined on the shaft 300 provides a net pressure P2 in the direction of the arrow 307 above ambient pressure P0 which exists at the capillary seal 322 that closes off the bottom of the gap 315 along which the journal bearing is defined. This pressure generally indicated by the arrow 307 provides a lifting force against the distal end 326 of the shaft 300, that is the end which is most distant from the capillary seal 322 and which is at or near the conical bearing. This lifting force raises the shaft 300 off the counterplate 320 and closes the conical gap 318 between the cone 314 and the sleeve 306. The net pressure P4 bearing against the bottom of the shaft will then substantially match the net pressure from the P2 from the journal bearing, and come to bear against the end 326 of the shaft, raising the shaft and supporting it above the facing surface 330 of counterplate 320.
In this particular approach, the journal bearing 404 creates an asymmetric pressure toward the conical bearing 408 which is at the distal end of the shaft from the base 400. The conical bearing 408 comprises an outer surface of a generally conical element on a shaft and a facing surface of the sleeve, with fluid in the gap supporting relative rotation. The design includes a cone 408 at or near the end of the shaft 402 distal from the base. This cone provides both axial and radial support for the sleeve rotating around the shaft by virtue of grooves provided on either the outer surface 422 of the cone or the inner surface 424 of the sleeve. An equalization path 426 is provided from the gap 428 which is defined between the axial surface 430 of the cone and the end surface 432 of the shaft 402 and the facing axial surface 434 of counter plate 436. According to the embodiment, grooves in this region are not necessary. Rather, the asymmetric pressure represented by the arrow 406 is created against the end of the shaft 432. At the end 432 of the shaft 402 so that as the sleeve rotates over the shaft, a lifting force is created which lifts the counter plate 436 away from the shaft end 432.
As with the previous embodiment, the objective is to establish by virtue of the asymmetric journal bearing 406 a net pressure P2 on the base 432 of the shaft 402 above ambient pressure P0 which would be measurable at or near the end of the gap which defines the journal bearing and ends in the capillary seal 440. This pressure provides a lifting force that raises that, as the sleeve rotates relative to the shaft under the impetus of the motor not shown in which this is typically incorporated which closes the conical bearing gap 423 between the sleeve surface 424 and the cone surface 422. The design objective which is intended to be met for optimum performance in that the net pressure P4 generated in the conical gap 423 will match the net pressure from the P2 journal bearing 404.
This application claims priority to a provisional application Ser. No. 60/377,329, filed May 1, 2002, entitled Hydraulic FDB Motor for HDD Applications invented by Jeffry Arnold LeBlanc, Alan Lyndon Grantz, Troy Michael Herndon, Michael David Kennedy, Anthony Joseph Aiello and Robert Alan Nottingham and incorporated herein by reference.
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| Number | Date | Country | |
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| Number | Date | Country | |
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| 60377329 | May 2002 | US |
