Abstract:
Fluid dynamic bearing systems produced by specified methods are provided. In one example, a method for designing a fluid dynamic bearing system includes determining a first stability ratio for a first journal bearing configuration. The method further includes determining a second stability ratio for a second journal bearing configuration. In a further example system, the first configuration has two sub-journal bearings and the second configuration has three sub-journal bearings. The method may comprise comparing the two stability ratios to determine whether adding a sub-journal increases the stability ratio of the bearing system.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
       [0001]    This divisional application claims priority from U.S. patent application Ser. No. 10/792,177, filed on Mar. 2, 2004, and which is herein fully incorporated by reference as if fully set forth and for all purposes. U.S. patent application Ser. No. 10/792,177 claims priority from U.S. Provisional Pat. App. No. 60/464,001, filed on Apr. 18, 2003, and which is herein fully incorporated by reference as if fully set forth and for all purposes. 
     
    
     BACKGROUND 
       [0002]    1. Field 
         [0003]    The present invention relates to fluid dynamic bearing motors and, more specifically, to a multi-journal fluid dynamic bearing motor assembly. 
         [0004]    2. Description of Related Art 
         [0005]      FIG. 1  provides a perspective view of a disc drive assembly  150 . In this arrangement, a plurality of discs  110 ′ are stacked vertically within the assembly  150 , permitting additional data to be stored, read and written. The drive spindle  151  receives the central openings  105  of the respective discs  110 . Separate suspension arms  156  and corresponding magnetic head assemblies  158  reside above each of the discs  110 . The assembly  150  includes a cover  130  and an intermediate seal  132  for providing an air-tight system. The seal  132  and cover  130  are shown exploded away from the disc stack  110 ′ for clarity. 
         [0006]    In operation, the discs  110  are rotated at high speeds about the spindle  151 . As the discs  110  rotate, an air bearing slider on the head  158  causes each magnetic head  158  to be suspended relative to the rotating disc  110 . The flying height of the magnetic head assembly  158  above the disc  110  is a function of the speed of rotation of the disc  110 , the aerodynamic lift properties of the slider along the magnetic head assembly  158  and, in some arrangements, a biasing spring tension in the suspension arm  156 . 
         [0007]    A servo spindle  152  pivots about pivot axis  140 . As the servo spindle  152  pivots, the magnetic head assembly  158  mounted at the tip of its suspension arm  156  swings through arc  142 . This pivoting motion allows the magnetic head  158  to change track positions on the disc  110 . The ability of the magnetic head  158  to move along the surface of the disc  110  allows it to read data residing in tracks along the magnetic layer of the disc. Each read/write head  158  generates or senses electromagnetic fields or magnetic encodings in the tracks of the magnetic disc as areas of magnetic flux. The presence or absence of flux reversals in the electromagnetic fields represents the data stored on the disc. 
         [0008]    Fluid dynamic bearings tend to generate less vibration and non-repetitive run-out in the rotating parts of motors than ball bearings and other types of bearings. For this reason, fluid dynamic bearing motors are oftentimes used in precision-oriented electronic devices to achieve better performance. For example, using a fluid dynamic bearing motor in a magnetic disc drive, such as magnetic disc drive  150  described above in conjunction with  FIG. 1 , results in more precise alignment between the tracks of the discs and the read/write heads. More precise alignment, in turn, allows discs to be designed with greater track densities, thereby allowing smaller discs and/or increasing the storage capacity of the discs. 
         [0009]    As persons skilled in the art are aware, an ongoing challenge in fluid dynamic journal bearings is balancing the tradeoff between motor performance and power consumption. For example, increasing the stiffness of the fluid dynamic journal bearings results in less vibration in a motor&#39;s rotating parts and, therefore, increased motor precision and performance. However, an increase in the stiffness of the bearings is usually accompanied by an increase in the power consumption of the motor. Therefore, there exists a need for a technique to increase the stability of a fluid dynamic bearing without increasing the amount of power consumed by the fluid dynamic bearing. 
       SUMMARY 
       [0010]    One embodiment of a method for designing and manufacturing a fluid dynamic bearing system includes determining a first stability ratio for a first journal bearing configuration. The method further includes determining a second stability ratio for a second journal bearing and installing the second journal bearing into the fluid dynamic bearing system. The two stability ratios may then be compared. Preferably, the first configuration has two sub-journal bearings and the second configuration has three sub-journal bearings. A third stability ratio for a third configuration may then be determined if the second stability ratio is greater than the first stability ratio. The third configuration may have four sub-journals. 
         [0011]    The disclosed method is especially useful for designing fluid dynamic bearing systems. One advantage of the disclosed method is that a journal arrangement designed according to the disclosed method has substantially greater stability than a journal arrangement not designed according to the disclosed method. Further, neither the radial stiffness nor the power consumption of the journal arrangement designed according to the disclosed method decreases appreciably. 
     
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         [0012]    So that the manner in which the above recited features of the present invention can be understood in detail, a more particular description of the invention, briefly summarized above, may be had by reference to the appended drawings. It is to be noted, however, that the appended drawings illustrate only typical embodiments of this invention and are therefore not to be considered limiting of its scope. 
           [0013]      FIG. 1  illustrates a perspective view of an exemplary disc drive assembly as might employ the improved spindle motor arrangement of the present invention, 
           [0014]      FIG. 2  illustrates a cross-sectional view of an improved spindle motor, according to one embodiment of the invention, in which three distinct journal bearings are provided, 
           [0015]      FIG. 3  illustrates the effects of the radial stiffness and the cross-coupled stiffness of a typical fluid dynamic journal bearing on the motion of a shaft, according to one embodiment of the invention, 
           [0016]      FIG. 4  illustrates the results of a simulation comparing the vibration responses of motors employing two, three, and four journal bearings, according to one embodiment of the invention, and 
           [0017]      FIG. 5  is a flow chart of method steps for using the stability ratio in the design of a fluid dynamic bearing, according to one embodiment of the present invention. 
       
    
    
     DETAILED DESCRIPTION 
       [0018]      FIG. 2  presents a partial cross-sectional view of an improved spindle motor arrangement  200  in one embodiment, in which three journal bearings  275 ,  280 , and  285  are provided. The motor  200  first comprises a hub  210 . The hub  210  includes an outer radial shoulder  205  for receiving a disc (not shown in  FIG. 2 ). Disposed within the hub  210  is a sleeve  250 . During operation, the sleeve  250  and hub  210  rotate together. In this arrangement, the sleeve  250  resides and rotates on a non-rotational thrust washer  265 . A shaft  255 , which is coupled to a base  230 , is provided along the inner diameter of the sleeve  250  to provide lateral support to the sleeve  250  while it is rotated. Joined to a bottom surface of the sleeve  250  is a shield  260 . The shield  260  encloses an outer portion of the thrust washer  265 . Between an inner side of the shield  260  and the enclosed portion of the thrust washer, a gap  295  is formed. The motor  200  also includes a stator  245 , which is mounted on a base  230 . The stator  245  typically defines an electric coil that, when energized, creates a magnetic field. The energized coil cooperates with magnets  215 , which are mounted from an inner surface of the hub  210  on a back iron  240 , to cause the hub  210  to rotate relative to the shaft. 
         [0019]    As persons skilled in the art will recognize, the motor  200  includes a hydrodynamic bearing system. More specifically, the thrust washer  265  is disposed proximally to a bottom surface of the sleeve  250 , and fluid is injected in gaps maintained between the sleeve  250  and surrounding parts, e.g., the shaft  255  and the thrust washer  265 , through a fill hole  270  disposed through the shield  260 . The fluid defines a thin fluid film that supports relative movement of the parts. The interface between the bottom of the sleeve  250  and the top of the thrust washer  265  thus defines a thrust bearing  290 . Liquid lubricant is provided along the thrust bearing gap  291  to provide a fluid bearing surface. Either a top face of the thrust washer  265  or a bottom surface of the sleeve  250  may include a grooved pattern (not shown) for receiving and holding liquid lubricant when the motor  200  is at rest. When the motor  200  is at rest, the sleeve  250  presses directly on the thrust washer  265 . Fluid is then extruded around the outer diameter of the shaft  155  and into a shaft-sleeve gap  292  and/or a shield-thrust washer gap  295 . 
         [0020]    When the motor  200  is energized and the sleeve  250  and adjoining hub  210  are rotated, lubricating fluid is drawn into the thrust bearing region  290  to support relative rotation between the bottom end of the sleeve  250  and the facing surface of the thrust washer  265 . To limit the axial displacement of the sleeve  250  and adjoining hub  210  during operation, an axial bias force is typically introduced. In this embodiment, a bias ring  235  is utilized to prevent the thrust bearing gap  290  from becoming too large and reducing the stiffness of the thrust bearing. In such a configuration, an axially downward magnetic force results that pulls magnet  215  (and therefore hub  210 ) towards base  230 . The magnitude of this force is a function of, among other things, the size of a gap between the magnet  215  and the bias ring  235 . Alternatively, base  230  may comprise a magnetic metal such as a Series 400 steel or a low carbon steel. In other alternative embodiments, the biasing force may be created in any other feasible way such as, for example, by applying a spring force or a downward-acting pressure force on hub  210 . When the rotating sleeve  250  comes to rest, the sleeve end will rest on the thrust washer  265 . Although the volume of fluid is very small, it will tend to be forced back out into the gap(s)  295  and/or  292 . Therefore, space is preferably allowed in this gap(s)  295  and/or  292  for this fluid. To inhibit the loss of liquid lubricant from the gaps  290  and  292 , optional capillary seals are provided along gaps  295  and  294 , respectively. When the motor  200  is idle, the capillary seals  295 ,  294  aid in maintaining fluid within the bearing system. 
         [0021]    In addition to thrust bearing  290 ,  FIG. 2  shows that there are three sets of grooves disposed on an outer surface of the shaft  255  along a shaft-sleeve interface  292 . The grooved portions of the shaft along with the corresponding outer sections of the sleeve  250  comprise top  275 , intermediate  280 , and bottom  285  journal bearings, respectively. Upon rotation of the sleeve  250 , the grooved patterns of each of the journal bearings  275 ,  280 ,  285  create a high pressure region at the center of each pattern to retain fluid and provide stiffness to the motor  200 . Although, only one intermediate journal  285  is depicted in  FIG. 2 , any number of intermediate journals  285  may be provided, depending upon design considerations of a particular motor. 
         [0022]    The substantially chevron-shaped groove patterns of the journals  275 ,  280 , and  285  shown in  FIG. 2  are preferred but other patterns, such as spiral patterns, would suffice. Preferably, the groove patterns are disposed on the shaft  255 , however, they may also be disposed on an inner surface of the sleeve  250 . It can seen that the groove patterns are each identical to one another and spaced equidistantly along the shaft  255 . This is only a preferred embodiment, however. In alternative embodiments, the journal bearings  275 ,  280 , and  285  may each comprise a different number and configuration of grooves and may be unequally spaced along the gap  292 . Further, the groove patterns may overlap one another as long as they are substantially separate, i.e., as long as the apex(es) of each groove pattern is/are longitudinally spaced apart. 
         [0023]    It is understood that the motor seen in  FIG. 2  is exemplary only. The present invention may be employed in various motor configurations. For example, the motor may be configured so that the shaft is coupled to the hub and rotates around a stationary sleeve. 
         [0024]      FIG. 3  illustrates the effects of the radial stiffness  320  and the cross-couple stiffness  325  of a typical fluid dynamic journal bearing  330  on the motion of a shaft  305 , according to one embodiment of the invention. As shown, the shaft  305  is subject to an excitation force that causes the shaft  305  to move in the horizontal direction towards a sleeve  310 . As a result of this horizontal motion  315 , the fluid dynamic journal bearing  330  exerts a force on the shaft  305  in a direction parallel and opposite to the horizontal motion  315 . The radial stiffness  320  (k. xx ) of the fluid dynamic journal bearing  330  causes this parallel and opposite reaction force. This type of stiffness is a desirable quality in fluid dynamic journal bearing  330  because the radial stiffness  320  tends to reduce the amplitude of the horizontal motion  315 . In addition to the radial stiffness  320 , the fluid dynamic journal bearing  330  is configured to have a cross-coupled stiffness  325  (k xy ), which acts orthogonally to the radial stiffness  320 . As persons skilled in the art will understand, the cross-coupled stiffness  325  causes the fluid dynamic journal bearing  330  to exert a force on the shaft  305  in a direction orthogonal to the horizontal motion  315 . This orthogonal force, in turn, introduces an orthogonal component  316  to the motion of the shaft  305 , thereby decreasing the stability and performance of fluid dynamic journal bearing  330 . For this reason, cross-coupled stiffness  325  is not a desirable quality in fluid dynamic journal bearing  330 . 
         [0025]    The stability of a bearing may be gauged by the radial stiffness divided by the cross coupled stiffness (stability ratio). A high ratio indicates a relatively greater radial stiffness, meaning that the radial stiffness governs the behavior of the shaft more than the cross-coupled stiffness. Thus, with a greater ratio, the propensity of a fluid dynamic journal bearing, through the radial stiffness, to limit the shaft motion outweighs the propensity of the journal, through the cross-coupled stiffness, to add unwanted motion to the shaft. 
         [0026]    As persons skilled in the art are aware, shorter journal bearings tend to have greater stability ratios than longer ones. However, simply reducing the length of the journal bearings in a conventional two-journal motor to improve stability would have negative consequences, such as a loss of radial stiffness. In order to obtain the stability benefits of a shorter bearing while not sacrificing the radial stiffness of a longer bearing, the overall journal length of the two-journal motor may be broken into a greater number of sub-journals. Doing this, effectively takes a larger journal with a lower stability ratio and divides it into smaller journals with greater stability ratios. Because these smaller sub-journals act in parallel, the overall effective journal length remains approximately constant, so the radial stiffness remains relatively constant. Further, because the radial stiffness doesn&#39;t change, power consumption remains constant (or, more importantly, does not increase). Because smaller-journals with greater stability ratios are used, the effective stability ratio for the overall journal increases, thereby decreasing or masking the influence that the cross-coupled stiffness has on shaft motion. 
         [0027]      FIG. 4  shows the results of a simulation measuring the response of a fluid dynamic bearing motor with two  400 , three  410 , and four  420  journal configurations at different operating frequencies, according to one embodiment of the invention. More specifically, the performance of each motor configuration is shown as the operational vibration (op-vibe) of the motor as a function of frequency. The journal gap of each motor configuration is the same. The total journal bearing length of each motor is the same, therefore, power consumption is approximately the same. In each of the three simulations, the total bearing length was divided into two  400 , three  410 , and four  420  “sub-journals,” respectively. As  FIG. 4  shows, for frequencies between about 200 and about 600 Hertz, the four-journal configuration  420  exhibits stability superior to either the two  400  or three  410  journal configurations. Further, for frequencies between 400 and about 800 Hertz, the three journal configuration  410  exhibits superior stability to the two journal  400  configuration. The foregoing shows that in certain circumstances, dividing up the total journal length into smaller sub-journals may increase performance without increasing the power consumption of the motor. 
         [0028]      FIG. 5  is a flow chart of method steps  500  for using the stability ratio in the design of a fluid dynamic bearing, according to one embodiment of the present invention. Although the method steps  500  are described in the context of the systems illustrated in  FIGS. 1-4 , any system configured to perform the method steps in any order is within the scope of the invention. 
         [0029]    As shown in  FIG. 5 , the method of using the stability ratio starts in step  510  where fluid dynamic journal bearing parameters are determined according to total system requirements, such as, for example, power consumption, total radial stiffness, and loading. These requirements are usually supplied by a customer. Typically, these requirements fix bearing parameters, such as the shaft diameter, the journal gap between the shaft and the sleeve, and the total length of the journal bearings. 
         [0030]    At step  520 , the motor is initially designed with a conventional two journal bearing configuration. At step  530 , the stability ratio of the two journal motor is determined. The stability ratio may be determined theoretically or empirically. At step  540 , an additional journal bearing is added to the initial two-journal design. Preferably, this is done by taking the total journal bearing length of the two journal design and providing three sub-journals, each one third the length of the total journal bearing length. 
         [0031]    At step  550 , the stability ratio of the motor with the three journal configuration is determined. As previously discussed herein, decreasing the individual sub-journal bearing lengths may improve the stability ratio of a given motor. There are situations, however, in which this will not be the case. For example, referring back to  FIG. 4 , it can be observed that for frequencies less than about 200 Hertz, the stability ratio of the two journal configuration is optimal and for frequencies greater than about 600 Hertz and less than about 800 Hertz, the stability ratio of the three journal configuration is optimal. Thus, at step  560 , the stability ratio of the three journal design is compared to that of the conventional two journal design to see if the stability ratio has increased by adding another sub-journal. If so, then the method returns to step  540 , and the process is repeated until adding an additional sub-journal no longer increases the stability ratio of the motor. At that point, the stability ratio is optimized and the design is concluded. If the stability ratio of the three journal design is not greater than that of the two journal design, then the two journal configuration may be deemed optimal and the design is concluded. 
         [0032]    The invention has been described above with reference to specific embodiments. Persons skilled in the art, however, will understand that various modifications and changes may be made thereto without departing from the broader spirit and scope of the invention as set forth in the amended claims.