Patent Publication Number: US-2006002641-A1

Title: Fixed shaft type fluid dynamic bearing motor

Description:
This is a continuation-in-part application of Ser. No. 11/074,055 filed on Mar. 8, 2005. The entire content of the application is hereby incorporated by reference. 
    
    
     BACKGROUND OF THE INVENTION  
      1. Field of the Invention  
      The invention relates to a fluid dynamic bearing motor for a recording disk drive, and more particularly to a fixed shaft type fluid dynamic bearing motor which uses a novel lubricating fluid sealing structure as an alternative to conventional tapered seals.  
      2. Description of the Related Art  
      The dominant bearing structure in conventional fluid dynamic bearing motors for magnetic disk drives (HDDs) has been a rotating shaft structure in which a lubricating fluid and air form only a single interface to facilitate sealing in the lubricating fluid. However, such fluid dynamic bearing is suffering from a number of disadvantages, for example, it could be sensitive to external vibration, imbalances and shock.  
      A desirable solution to this problem would be to have the spindle motor attached to both the base and the top cover of the disk drive housing. This would increase overall drive performance. A motor attached at both ends is significantly stiffer than a rotational shaft bearing. And also, the existence of the motor shaft that supports the top cover of the housing should be big advantage for the extremely small disk drive.  
      All of the known fluid dynamic bearing designs for a motor attached at both ends has not been easy to realize. The reason for this is that in order to have top cover attachment, the motor and specifically the bearing would need to be open on both ends. Opening a motor at both ends greatly increases the risk of oil leakage out of the fluid dynamic bearing. This leakage is caused by, among other things, small differences in net flow rate created by differing pumping pressures in the bearing. If all of the flows within the bearing are not carefully balanced, a net pressure rise toward one or both ends may force fluid out through the capillary seal. Moreover, due to manufacturing imperfections of the bearing, the gap in the bearing may not be uniform along its length and this can create pressure imbalance in the bearing and hence, cause leakage when both ends of the fluid dynamic bearing are open. The net flow due to pressure gradients in a bearing has to be balanced by all the bearings individually for the fluid to stay inside the bearing. Any imbalances due to pumping by the grooves of the bearings will force the fluid out of the capillary until the meniscus at one end moves to a new equilibrium position.  
      Nevertheless, most of the fluid dynamic bearing motors fixed or attached at both ends achieved in the past are for large-sized structures which are adapted to carry a number of magnetic disks for high speed rotation. Thus, it is difficult to employ the structure of these motors for small-sized drives which carry and drive no more than two small magnetic disks or the like.  
      More specifically, the fluid dynamic bearing motors fixed or attached at both ends have many parts arranged in the axial direction, e.g., having one or more thrust plates. Thus, if such structure is simply miniaturized for use in a small sized motor, the same arrangement cannot secure the span between the upper and lower radial bearings, failing to maintain low non-repetitive runout during rotation. Above all, the greater number of parts makes cost reduction difficult.  
      For the fixed shaft type fluid dynamic bearing motors that are applicable to low-profile HDDs, single cone bearings have been proposed in Japanese Unexamined Patent Application Publication No. Hei 06-315242 and U.S. Pat. No. 6,686,674, and single thrust bearing structures have been proposed in U.S. Pat. No. 6,211,592 and Japanese Unexamined patent application Publication No. 2004-173377.  
      The single cone bearing proposed in Japanese Unexamined patent application Publication No. Hei 06-315242 and U.S. Pat. No. 6,686,674 are of a rotating shaft structure or single end-tied fixed shaft structure, and thus cannot be applied to fluid dynamic bearing motors with its shaft attached at both ends directly.  
      U.S. Pat. No. 6,211,592 proposes two types of structures in which the fixed shaft has a single radial bearing and a single thrust bearing. One of the structures employs herringbone grooves for single radial bearing and single thrust bearing. The other one employs an asymmetric herringbone groove and a spiral groove for single radial bearing and single thrust bearing respectively.  
      The former structure still has the possibility of leakage of the lubricating fluid in view of machining imperfections at the mass production stage. The latter structure is less likely to cause the leakage of the lubricating fluid, though it cannot produce enough rotational moment that is necessary to maintain low non-repetitive runout during rotation.  
      The structure proposed in Japanese Unexamined patent application Publication No. 2004-173377 looks good in sealing the lubricating fluid. Nevertheless, the upper and lower asymmetric herringbone grooves have their asymmetric portions at the top and bottom ends, respectively, in such directions as to press the lubricating fluid toward each other. This decreases the effective radial bearing space. Another concern lies in that the top end of the radial bearing theoretically has an unlubricated area and there is no means to prevent or to remove air bubbles entering into.  
      The tapered seal structure widely used in the lubricating fluid sealing structures of the fluid dynamic bearing motors also puts a strong constraint on low-profile HDDs.  
      The tapered seal is a method of sealing which utilizes the surface tension of the lubricating fluid. It is generally desirable that the tapered seal have an opening angle of 10 degrees or less, in view of sealing strength.  
      The tapered seal appropriately has a maximum gap of 0.3 millimeters or so. Even if the dimensional precision of the individual parts are increased to suppress the maximum gap to 0.2 millimeters, the tapered seal has a total length of 1.1 millimeters or more, given the opening angle of 10 degrees.  
      It can be said that, in order to achieve an HDD fluid dynamic bearing motor having a thickness of no greater than 3 millimeters or so, compromises must be made in various respects—including the sealing of the lubricating fluid—despite an awareness of inadequacies.  
     SUMMARY OF THE INVENTION  
      Thus, it is an object of the present invention to provide a fixed shaft type fluid dynamic bearing motor with its shaft attached or fixed at its both ends, with a reliable lubricating fluid sealing structure in which the bearing is open at both the upper and lower ends and ensuring highly precise rotational function.  
      Another object of the present invention is to provide a fluid dynamic bearing structure suitable for use in low profile motor for driving a few magnetic disk or the like at high precision.  
      A further object of the present invention is to provide a fluid dynamic bearing motor that has a single conical bearing surface, and suitable for low profile recording disk drive.  
      Yet further object of the invention is to provide a fluid dynamic bearing motor which has a cylindrical radial bearing and single thrust bearing, and suitable for low profile recording disk drive.  
      These and other objectives of the invention are achieved by a fixed shaft type fluid dynamic bearing motor according to the present invention. It comprises at least: a fixed shaft; a rotary portion including a sleeve which is rotatably fitted on the shaft with a small gap therebetween; an annular member fixedly provided to oppose a lower portion of the sleeve with a gap; a lubricating fluid lying in the gaps between the sleeve and the shaft, and between the sleeve and the annular member continuously, and having at least two interfaces with air near the top end of an inner periphery of the sleeve and around the lower part of the sleeve; and magnetic means for generating a magnetic attractive force in the axial direction between the shaft and the sleeve, a group of dynamic pressure generating grooves formed on either of the confronting surfaces of the sleeve and the shaft to support the rotary portion in a floated condition by the magnetic attractive force and an axial load due to pressure partially increased in the fluid by the grooves, the grooves being asymmetric herringbone grooves or spiral grooves to pump upward toward the upper end of the inner circumference of the sleeve, the fluid lying between the sleeve and shaft while the sleeve is rotating, and a channel formed in the sleeve and having an intake portion near the top end of the inner periphery of the sleeve and an outlet portion near the periphery of the bottom end of the sleeve, the intake portion being located radially inside the outlet portion, the channel continuously extending from the intake portion to the outlet portion, whereby the lubricating fluid is thrown out into the intake portion by centrifugal force near the top end of the inner periphery of the sleeve, and is conveyed from the intake portion to the outlet portion through the channel by centrifugal force and/or through a slanted channel in circumferential direction through the channel with the lubricating fluid being discontinuous.  
      According to an aspect of the present invention, the fluid dynamic bearing motor has one of the lubricating fluid interfaces with air at upper or lower side of the sleeve bottom level around the lower part of the sleeve. The fluid dynamic bearing motor which has the lubricating fluid interface at the lower part of the outer periphery of the sleeve enables thinner motor.  
      According to another aspect of the present invention, the fluid dynamic bearing motor realizes perfect sealing structure of the lubricating fluid by circulation of the lubricating fluid due to centrifugal force. During rotation of the motor, the lubricating fluid which is conveyed to the top of the sleeve inner surface by the pressure generating groove is thrown out into the channel in the sleeve. The channel desirably has a gap portion as small as the lubricating fluid can be retained therein by surface tension. At rest of the motor, the lubricating fluid is absorbed and retained in the channel. While the dimension of the gap of the channel may be as small as the lubricating fluid can be retained by surface tension, and the dimension varies depending on both the viscosity of the lubricating fluid and the surrounding materials. An appropriate value is no greater than 0.2 millimeters or so.  
      According to another aspect of the present invention, the fluid dynamic bearing motor has lubricating fluid pressure adjuster for adjusting the outward lubricating fluid pressure occurring in the channel around the channel outlet and/or in the channel. During rotation of the motor, when the lubricating fluid pressure at the channel outlet which is caused by the centrifugal force and/or by slanted channel in circumferential direction is too large, it may force the lubricating fluid interface move outward and then may cause the fluid leakage. The lubricating fluid pressure adjuster eases and adjusts the fluid pressure in the channel and stabilizes the fluid movement for perfect sealing.  
       FIG. 17  shows a sealing model of the fluid dynamic bearing motor according to the present invention. In the model, lubricating fluid is retained on outer periphery of the sleeve and in channel  171 ,  172  respectively.  173  represents pressure generating groove. Surface tension forces  176 ,  177  at the lubricating fluid accumulating portions  171 ,  172  are drawing respectively the lubricating fluid upwardly, and pressure generating groove  173  is drawing the lubricating fluid inversely, and drawn-in lubricating fluid by pressure generating groove  173  is thrown out into the accumulating portion of the channel  172  by the centrifugal force at the sleeve top  174 . And total quantity of the lubricating fluid in the lubricating fluid accumulating portions  171 ,  172  and in the bearing region that has groove  173  is constant. In this sealing model, the centrifugal force acting on the lubricating fluid in the channel should be a major cause to make the sealing unstable. The present invention employs lubricating fluid pressure adjusters to ease and to adjust the fluid pressure around the channel outlet, lubricating fluid pressure adjusters are, for example, gap diminishing region in the channel, pressure generating grooves between the channel outlet and the fluid boundary at the sleeve outside, means for pressing the lubricating fluid from the channel outlet toward the intake.  
      According to another aspect of the present invention, the fluid dynamic bearing motor has discontinuously filled lubricating fluid from the channel intake to the channel outlet. It makes easy that the fluid pressure diagram becomes continuous around the channel outlet so as to stabilize the fluid move.  
      According to yet another aspect of the present invention, the fluid dynamic bearing motor eliminates the need for a long tapered seal near the top end of the sleeve. At rest of the motor, most of the lubricating fluid is absorbed in the channel in the sleeve and during rotation, the lubricating fluid is thrown out into the channel near the top end of the sleeve by centrifugal force.  
      According to a further aspect of the invention, the fluid dynamic bearing motor effectively avoids leakage of the lubricating fluid. The lubricating fluid pumping capability of the dynamic-pressure generating groove, toward the top end of the sleeve is set sufficiently higher to compensate for such problems as imperfections in the dynamic-pressure generating groove, and the tilt of the gap in which the dynamic-pressure generating groove lies.  
      In a further aspect of the invention, the fluid dynamic bearing motor also has the function of removing air bubbles in the lubricating fluid. The lubricating fluid is influenced by the centrifugal force and is thrown out into the channel near the top of the sleeve inner surface. Meanwhile, the bubbles are released to the air since no centrifugal force acts thereon.  
      According to an aspect of an embodiment of the invention, the sleeve is composed of an outer barrel member and inner barrel member fixedly fitted in the outer barrel member with the channel being formed therebetween. Accordingly, it is easier to define the dimensions of the gap of the channel precisely and to control cross-sectional shape of the gap of the channel.  
      According to another aspect of the embodiment, the fluid dynamic bearing motor includes the fixed shaft of a conical or truncated conical shape with its diameter reducing toward the top end. The sleeve has a conical concave opening to fittingly receive the shaft. One or more sets or groups of dynamic-pressure generating grooves are formed on either of the shaft and the sleeve, with at least one of the dynamic-pressure generating grooves having capability of pumping the lubricating fluid toward the top end of the sleeve. This type of motor is suited for low profiles while securing the space for the dynamic-pressure generating grooves.  
      According to yet another aspect of the embodiment, the fluid dynamic bearing motor includes a fixed shaft of a cylindrical shape and a sleeve has a cylindrical opening to rotatably and fittingly receive the shaft. The sleeve opposes the annular member at its bottom end orthogonal to the shaft. Dynamic-pressure generating grooves are formed on either one of the outer periphery of the shaft and the inner periphery of the sleeve, and either one of the annular member and the bottom end of the sleeve, respectively. At least the dynamic-pressure generating groove formed on either the lower end of the sleeve or the surface opposing thereto is formed as an asymmetric herringbone groove or a spiral groove having capability of pumping the lubricating fluid radially inward.  
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
      In the accompanying drawings:  
       FIG. 1  is a vertical sectional view of a fixed shaft type fluid dynamic bearing motor which is a first embodiment of the present invention;  
       FIG. 2  is an enlarged perspective view of inner and outer cylindrical or barrel members which compose a sleeve shown in  FIG. 1 ;  
       FIG. 3  is an enlarged vertical sectional view of the bearing part of  FIG. 1 ;  
       FIG. 4  is an enlarged vertical sectional view of the bearing part of  FIG. 1 ;  
      FIGS.  5 ( a ) and  5 ( b ) illustrate in enlarged modeled forms the portion around the channel outlet and sleeve bottom of  FIG. 1  and the lubricating fluid pressure diagram;  
       FIG. 6  is a vertical sectional view of a fluid dynamic bearing motor which is a second embodiment of the present invention;  
       FIG. 7  is an enlarged vertical sectional view of the bearing part of  FIG. 6 ;  
       FIG. 8  is a vertical sectional view of a fluid dynamic bearing motor which is a third embodiment of the present invention;  
      FIGS.  9 ( a ) and  9 ( b ) are enlarged views of the bearing part of  FIG. 8 , showing the configuration of grooves which constitute a channel;  
       FIG. 10  is an enlarged view showing the configuration near the periphery of the bottom end of the sleeve of  FIG. 8 ; and  
       FIG. 11  is an enlarged perspective view of inner and outer cylindrical or barrel members which compose a sleeve shown in  FIG. 8 ;  
      FIGS.  12 ( a ),  12 ( b ) illustrate in enlarged modeled forms the portion around the channel outlet and sleeve bottom of  FIG. 8  and the lubricating fluid pressure diagram;  
      FIGS.  13 ( a ) and  13 ( b ) are sectional views of a low-profile recording disk drive which is a fourth embodiment of the present invention.  
       FIG. 14  is a vertical sectional view of a fluid dynamic bearing motor which is a fifth embodiment of the present invention;  
       FIG. 15  is an enlarged vertical sectional view of the bearing part of  FIG. 14 ;  
       FIG. 16  is an enlarged perspective view of inner and outer cylindrical or barrel members which compose a sleeve shown in  FIG. 14 ;  
       FIG. 17  illustrates a lubricating fluid sealing model of the present invention;  
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS  
      Hereinafter, embodiments, operating principles of a fixed shaft type fluid dynamic bearing motor according to the present invention will be described with reference to the drawings.  
       FIG. 1  is a vertical sectional view of a fixed shaft type fluid dynamic bearing motor which is a first embodiment of the present invention. A fixed shaft (hereinafter, referred to as a conical shaft  11  or a shaft  11 ) includes a truncated cone shape side wall diminishing its diameter toward an end of the shaft. A sleeve is composed of an outer member  12  and an inner member  13 . The inner member  13  has an inner wall forming a conical concavity accommodating the shaft  11  and surrounding the side wall, the inner wall opposing the wall of the shaft  11  with a clearance. The shaft  11  is positioned to a base plate  16  by using its radial side  1   e  at the bottom end, and is fixed to the base plate  16  with its axial side  1   d  being secured with a suitable adhesive strength.  
      To attract the rotating part, including the outer member  12  and the inner member  13 , magnetically along the axial direction, magnetic pieces  19  are embedded in the base plates  16  so as to face rotor magnets  17 . The numerals  15 ,  18 , and  1   a  represent a hub which supports one or more magnetic disks, a stator core, and a coil, respectively. An annular member  1   c , confronting to the lower periphery of the outer member  12 , is formed on a part of the base plate  16 . Numeral  14  represents a channel constituted as a gap between the outer member  12  and the inner member  13 .  
      A lubricating fluid is continuously filled into the gap between the shaft  11  and the inner member  13 , and the gap between the periphery of the outer member  12  and the annular member  1   c . As shown in  FIG. 3 , the interfaces  33 ,  34  of the lubricating fluid with the air lie near the top end of the inner member  13  and on the periphery of the outer member  12  respectively.  
      The numeral  1   b  represents stoppers for regulating the amount of axial movement of the rotating part. The stoppers are fixed to the top end of the annular member  1   c  and engaged with a stepped portion on the periphery of the outer member  12 .  
       FIG. 2  is a perspective view of the outer member  12  and the inner member  13  that constitute the sleeve of the fluid dynamic bearing motor shown in  FIG. 1 .  FIG. 2 ( a ) and  FIG. 2 ( b ) illustrate the outer member  12  and the inner member  13  respectively.  
      The outer member  12  is formed by press molding from Aluminium plate. And the inner member  13  is machined from SUS material. Two radial direction grooves for corresponding to the channel  14  and a tapered surface  23  for corresponding to a gap diminishing region in the channel are formed by machining on the surface of the inner member  13 . The numerals  21 ,  22  represent an opening of the outer member  12 , a through hole in which the shaft  11  is fitted loosely of the inner member  13  in respectively. And the numeral  24  represents outer surface of the inner member  13  besides the radial direction groove  25  and the tapered surface  23 .  
      The outer surface of the inner member  13  is fitted to the inner surface of the outer member  12  and fixed by bonding at the outer surface  24  of the inner member  13 . The opening  21  diameter of the outer member  12  is smaller than that of the hole  22  of the inner member  13 .  
      The tapered surface  23  constitute the gap diminishing region with the outer member  12 . The radial groove  25  has a retaining capability of the lubricating fluid as the depth around 20 micron meters at the top and around 50 micron meters at the side of the inner member  13 . It is easy to form the tapered surface  23  and radial grooves  25  on the surface of the inner member  13 .  
      The inner member  13  can be fabricated by molding of sintered material or resin also. In that case, the tapered surface  23  and radial direction grooves  25  are formed by molding die at the same time, production cost will be reduced. Also, when the outer member  12  is formed by press molding, pits and projections may be formed simultaneously in and on the inner periphery of the outer member  12  to constitute the channel  14 .  
       FIG. 2  shows the sleeve composition which has the conical bearing surface. Same sleeve composition is applicable for the bearing sleeve which has cylindrical bearing surface shown later.  
      Herringbone grooves  1   f  and  1   g  for generating dynamic pressures, are formed on the surface of the inner member  13  confronting to the surface of the shaft  11 . The outer herringbone groove  1   g  is formed asymmetrically so that it pumps the lubricating fluid toward the top end of the sleeve against centrifugal force. In order to provide the objective function of the present invention, the inner dynamic-pressure generating groove may also be given a lubricating fluid pumping capability. Nevertheless, this still leaves the possibility of negative pressure occurring between the dynamic-pressure generating grooves. In the presence of a plurality of dynamic-pressure generating grooves, it is desirable that the outermost dynamic-pressure generating groove have a lubricating fluid pumping capability so that it presses the lubricating fluid for circulation. Numeral  1   h  designates a pump-in spiral groove which contributes to the lubricating fluid sealing.  
       FIG. 3  is an enlarged view of the bearing part of the fluid dynamic bearing motor shown in  FIG. 1 . Description will now be given of the operating principle. For convenience of understanding,  FIG. 3  shows the channel  14  and the herringbone grooves  1   f  and  1   g  in the left half alone, while the directions of movement  31  and  32  of the lubricating fluid are shown by dotted lines in the right half. The numerals  33 ,  34  represent upper and lower lubricating fluid interfaces with air respectively. The numerals  35  represents a lubricating fluid interfaces with air in the channel  14 .  
      The herringbone grooves are each made of a pair of spiral grooves for pumping the lubricating fluid toward each other. When the pumping capabilities of the lubricating fluid are configured unevenly, these spiral grooves exert the lubricating fluid pumping capability in one direction as an asymmetric herringbone groove. The herringbone groove  1   g  is set to have a lubricating fluid pumping capability directed radially inward, i.e., toward the top end of the sleeve.  
      When the inner member  13  and the outer member  12  are rotated, the herringbone grooves  1   f  and  1   g  increase the pressure of the lubricating fluid locally near their respective centers, thereby supporting the inner member  13  and the outer member  12  without contact.  
      Meanwhile, the herringbone groove  1   g , having the asymmetric configuration, pumps the lubricating fluid toward the top end of the sleeve. The lubricating fluid thus flows in the direction shown by the dotted line  31 , and is thrown out into the channel  14  by centrifugal force acting on the lubricating fluid near the top end of the inner member  13 —the intake portion. The thrown out lubricating fluid joins with the lubricating fluid at the boundary  35  and then is lead to the channel outlet. The dotted line  32  shows the direction of flow of the lubricating fluid within the channel  14 .  
      Near the top end of the sleeve where the leakage of the lubricating fluid is the most probable, the lubricating fluid is thrown out into the channel  14  by the centrifugal force acting directly on the lubricating fluid, and thus is prevented from leakage completely.  
      The foregoing structure for sealing the lubricating fluid also has the function of removing air bubbles. More specifically, if bubbles exist between the shaft  11  and the inner member  13 , they are conveyed to near the top end of the sleeve by the flow of the lubricating fluid shown by the dotted line  31 . At the top end of the sleeve, the lubricating fluid experiences the centrifugal force and is thrown out as shown by the dotted line  32 . Meanwhile, the bubbles are released to the air since no centrifugal force acts thereon. The behavior of the lubricating fluid at rest, and during rotation, will be described further with reference to  FIG. 4 . The left half of the diagram shows the state at rest, in which part of the inner member  13  is in contact with the shaft  11 . The right half shows the state of during rotation, in which the inner member  13  floats without contact with the shaft  11 . What is worth noting in the left and right halves of  FIG. 4  is the positions of the lubricating fluid.  
      The gap inside the channel  14  is as small as less than 50 micrometers. When at rest, the lubricating fluid lying between the outer member  12  and the annular member  1   c , i.e., in a gap greater than the gap, is absorbed through the outlet portion. During rotation, the lubricating fluid is supplied from the channel  14  to between the outer member  12  and the annular member  1   c , and to between the shaft  11  and the inner member  13 , by centrifugal force. Consequently, near the top end of the sleeve, if the gap of the space  43  formed by the three members (the outer member  12 , the inner member  13 , and the shaft  11 ) is set greater than the gap that constitutes the channel  14 , the lubricating fluid at rest is drawn into the channel  14  by surface tension and is no longer present in the foregoing space  43 .  
      During rotation, the lubricating fluid is thrown out into the channel  14  by centrifugal force, and is no longer present in the foregoing space  43  again. This allows effective sealing of the lubricating fluid, with an axial space shorter than in conventional tapered seal structures.  
      The opening diameter in the top end of the outer member  12  is smaller than the diameter of the inner periphery of the inner member  13 . The top end of the outer member  12  with small opening diameter not only promises the operation of damming up lubricating fluid that flows along the inner periphery of the inner member  13  during rotation, but also ensures the provision of perfect leakage prevention since the lubricating fluid is thrown out to the channel  14  before reaching the top end of the outer member  12 .  
      The lubricating fluid at rest is retained by the channel  14 , and the interface between the lubricating fluid and the air lies in the intake portion. The top end of the outer member  12  with small opening diameter keeps the interface of the lubricating fluid being positioned away from the exterior to the interior of the motor, thus playing a significant role in reducing the possibility of leakage.  
      Furthermore, the reduced gap between the opening in the top end of the outer member  12  and the shaft  11  provides the effect that the vapor pressure of the lubricating fluid within the space  43  is increased to suppress the evaporation of the lubricating fluid.  
      The lubricating fluid in the channel  14  is pressed outwardly by the centrifugal force and/or by the slanted channel in circumferential direction. The gap diminishing region in the channel  14  and the spiral groove  1   h  are employed as the lubricating fluid pressure adjuster for adjusting the fluid pressure occurring in the channel. While the spiral groove  1   h  lies between the channel outlet and the interface  51 (numeral  34 ) with air on the lower portion of outer periphery of the sleeve as shown in  FIG. 5 ( a ). Along the dotted line  52 , lubricating fluid pressure diagram is shown in  FIG. 5 ( b ). The horizontal axis indicates the location of points on the dotted line  52 , and the vertical axis indicates the lubricating fluid pressure referring the atmospheric pressure P 0 .  
      The fluid pressure at the point  53  inside of the interface  51  is lower than the atmospheric pressure P 0 , and the fluid pressure at the point  54  is slightly higher than that by the centrifugal force. Then the fluid pressure at the point  55  is increased from the pressure at the point  54  by the spiral groove  1   h . The fluid pressure at the point  56  inside of the interface  35  is lower than the atmospheric pressure P 0 , and the fluid pressure at the point  55  is increased from the point  56  by the centrifugal force. There is some possibility that the interface  35  is not clear enough, because the lubricating fluid is continiously flowing in. In this embodiment, the interface  35  is wide in circumferential direction enough to reduce the effect of flowing of the fluid into the interface  35 .  
      The fluid pressure should be continuous as shown in  FIG. 5 ( b ) during rotation. When the quantity of the lubricating fluid at outer periphery of the sleeve increases, the interface  51  moves outward, and then the fluid pressure at the point  53  becomes higher towards P 0  because that a radius of the interface  51  curve becomes larger. While the lubricating fluid in the channel increases, the pressure difference between the points  56  and  55  also becomes larger. Accordingly, the quantity of the lubricating fluid around the channel outlet is properly divided in the channel and at outer periphery of the sleeve as the fluid pressure is continuous as shown in  FIG. 5 ( b ).  
      When the lubricating fluid pressure adjuster is not employed, the stabilization condition of the fluid around the channel outlet is that the location of the point  54  is radially outward from the point  53  as the pressure at the point  54  becomes larger by the centrifugal force. Then there exist strict constraints about the outer member  12  shape and dimensions. The present embodiment applying the gap diminishing region in the channel and the spiral groove  1   h  between the channel outlet and the fluid interface  51  makes the design flexible.  
      The fluid dynamic bearing motor of the present invention, has discontinuously filled lubricating fluid from the channel intake to the channel outlet. It makes the fluid pressure balance around the channel outlet easy and contributes to the stable fluid sealing. In case that there is continuously filled lubricating fluid in the channel, it is hard to balance the fluid pressure generated by the grooves and the centrifugal force with the pressure near the fluid interface during rotation.  
      The magnetic pieces  19  are made of a magnetic substance such as silicon steel plates, ferrite, and permalloy. The magnetic piece  19  generates a magnetic attractive force between the rotating part and the stationary part, in cooperation with the rotor magnets  17 .  
      The rotor magnets  17  are magnetized so as to alternate in magnetization, and thus cause eddy currents in the magnetic piece  19  during rotation. Permalloy or ferrite is less susceptible to eddy currents than silicon steel plates, and thus is suitable for high speed rotation.  
      If the magnetic attractive force resulting from the magnetic piece  19  alone is insufficient, the stator core  18  and the rotor magnets  17  may be displaced axially relative to each other to generate additional magnetic attractive forces.  
      The rotating part is floated and supported at the position where the axial components of the load capacities created by the herringbone grooves  1   f  and  1   g , and the magnetic attractive force are balanced. The load from the weight of the rotating part on the bearing part varies, while the amount of float of the rotating part varies depending on the orientation of the motor in use, such as being erect, inverted, or sideways. The magnetic attractive force is set at around three to five times the total weight of the rotating part including the outer member  12 , the inner member  13 , the hub  15 , the rotor magnets  17 , and the magnetic disk or the like to be mounted thereon. This applies an axial downward load above a certain level irrespective of the orientation of the HDD, whereby the present fluid dynamic bearing can achieve low non-repetitive runout during rotation.  
      The foregoing has dealt with the case where the dynamic-pressure generating grooves are composed of the two herringbone grooves formed in the conical surface of the shaft  11 . It is possible, however, for only a single series of asymmetric herringbone groove formed in the conical surface of the shaft  11  to float and support the rotating part, and to achieve low non-repetitive runout during rotation. In this case, a fluid dynamic bearing motor of lower profile can be constructed. The structure of the bearing part and the principle of operation in case of a single herringbone groove formed in the conical surface are disclosed in detail in a U.S. Pat. No. 6,686,674 that is owned by the same applicant of the present application, and disclosure of the patent is incorporated herein by reference.  
      In the present embodiment, the channel  14  is formed as the gap between the inner member  13  and the outer member  12 . Nevertheless, the inner member  13  of the sleeve may be made of a porous material having a number of small gaps so that the small gaps form the channel  14 . A sintered alloy material may be filled into the outer member  12  to form the inner member  13 , and to form the herringbone grooves  1   f  and  1   g  simultaneously.  
      Since small gaps also exist in the surface of the area where the herringbone grooves  1   f  and  1   g  are formed, the lubricating fluid might permeate into the inner member  13  through those gaps in the surface, possibly causing shortage of the lubricating fluid in the herringbone groove  1   f . In this case, the small gaps in the surface of the inner member  13 , excluding near the interface with the outer member  12 , are filled with a resin having a high lubricity for caulking.  
      The novel lubricating fluid sealing structure, of which the structure and principle of operation have been described in the present embodiment, is characterized in that the axial space necessary near the top end of the sleeve can be made smaller. Referring to  FIG. 4 , the necessary axial space is the sum of the thickness of the outer member  12  and the axial length of the space  43  formed by the three members: the outer member  12 , the inner member  13 , and the shaft  11 .  
      If the outer member  12  is formed by pressing or drawing a thin plate of 0.2 millimeters or so, and the latter dimension is set at 0.1 millimeters (which is greater than the opening gap of the channel  14 , or 50 micrometers) then the entire lubricating fluid sealing structure can be formed in 0.3 millimeters. These values can also be reduced further, and it is possible to achieve a reliable lubricating fluid sealing structure with considerably smaller axial dimensions as compared to conventional tapered seals.  
      While the first embodiment has dealt with an example of a cone bearing, a second embodiment shown in  FIG. 6  will deal with an example where the lubricating fluid sealing structure of the present invention is applied to a cylindrical shaft.  
      The sleeve, which rotatably fits to a T-shaped cylindrical shaft  61 , is composed of an inner cylinder  63  and an outer cylinder  62  corresponding to an inner member  13  and an outer member  12  respectively in  FIG. 1 . A channel  64  is formed in the gap between the outer cylinder  62  and the inner cylinder  63 . A lubricating fluid continuously lies between the shaft  61  and the inner cylinder  63  and between the outer cylinder  62  and an annular member  1   c.    
      The inner periphery of the inner cylinder  63  is provided with a single herringbone groove  66 , which constitutes a radial bearing. A flange  65  of the shaft  61  confronting the bottom end of the inner cylinder  63  is provided with an asymmetric herringbone groove  67  which has a radially inward lubricating fluid pumping capability against centrifugal force. This constitutes a thrust bearing. The annular member  1   c  which is a part of the base plate  16 , and the flange  65  which is a part of the shaft  61  are corresponding to the annular member defined in the claim  1  and  2 . The gap diminishing region in the channel  64  alone is employed as the lubricating fluid pressure adjuster in this embodiment. The other parts are the same as in the first embodiment shown in  FIG. 1 . The same members will be designated by identical numerals.  
      As far as the combination of the dynamic-pressure generating grooves alone is concerned, the radial and thrust bearings are close to those of U.S. Pat. No. 6,211,592. There is a difference, however, in that the asymmetric herringbone groove  67  is arranged near the bottom end of the inner cylinder  63  to make the lubricating fluid flow toward the top end of the inner periphery of the inner cylinder  63 .  
      In the case of U.S. Pat. No. 6,211,592, the individual herringbone grooves can cause flows of the lubricating fluid due to imperfections in mass production, possibly causing leakage of the lubricating fluid with a considerable probability. In the case of  FIG. 6 , on the other hand, the lubricating fluid sealing structure of the present invention is adopted to prevent the lubricating fluid from leaking.  
       FIG. 7  is an enlarged view of the shaft  61  and in the vicinity of the sleeve of the fluid dynamic bearing motor shown in  FIG. 6 . The asymmetric herringbone groove  67  is provided with a lubricating fluid pumping capability sufficient to pump the lubricating fluid toward the top end of the inner cylinder  63  against centrifugal force.  
      The lubricating fluid is pumped from the periphery of the herringbone groove  67  to near the top end of the inner periphery of the inner cylinder  63 , as shown by the dotted line  71 . Near the top end of the inner periphery of the inner cylinder  63 , the lubricating fluid is thrown out into the channel  64  by centrifugal force, and returns to near the periphery of the herringbone groove  67 . The gap diminishing region in the channel  14  alone is employed as the lubricating fluid pressure adjuster. So this embodiment is suitable for small profile and low rotational speed application.  
      As in the first embodiment described in conjunction with  FIGS. 3 and 4 , the lubricating fluid is sealed effectively, and bubbles are separated by the same principle. The dotted lines  71  and  72  correspond to the dotted lines  31  and  32  of  FIG. 3 .  
      The structure near the top end of the sleeve is shown enlarged further in the circle shown by the dotted line  73 . The numeral  74  designates the shoulder line of the shaft  61 . The area below the numeral  74  is the area of the radial bearing. The top end of the herringbone groove  66  lies near the level of this numeral  74 .  
      The numeral  75  represents an annular projection which is formed around the inner periphery of the inner cylinder  63 . In  FIG. 7 , the annular projection is given a height (in the radial direction) of approximately one half of 2 micrometers (which is the gap width between the inner cylinder  63  and the shaft  61 ).  
      The lubricating fluid  76  pumped by the asymmetric herringbone groove  67  is thrown out to the channel  64  beyond this annular projection  75 , whereby an accumulation of the lubricating fluid  76  is constantly formed near the top end of the herringbone groove  66 . The annular projection  75  desirably has a height close to the gap between the inner cylinder  63  and the shaft  61 . Greater heights have little further effect. Instead of forming the annular projection  75 , the top end of the herringbone groove  66  may be extended to near the intake portion of the channel  64 , with the effect of putting the lubricating fluid near the top end of the herringbone groove  66 .  
      In  FIGS. 6 and 7 , the radial bearing is made with only a herringbone groove  66  in the inner periphery of the inner cylinder  63 . It is possible to center the rotating part to the shaft  61 , but not to generate a moment for restoring orientation when the rotating part tilts. In the present embodiment, the moment for restoring the orientation of the rotating part is generated by the asymmetric herringbone groove  67 —the thrust bearing.  
      More specifically, when the rotating part tilts, the bottom end of the inner cylinder  63  also tilts to change the gap with the flange  65 . In the vicinities of the areas where the gap varies in size, the asymmetric herringbone groove  67  increases the local pressure at its radial center by a degree inversely proportional to the gap. A moment for restoring the orientation of the rotating part occurs thus, and the orientation of the rotating part is restored. Having a single radial bearing alone, the present embodiment is suited to low-profile HDDs.  
      In the case of the fixed shaft structure as shown in  FIG. 6 , the shaft is usually pressed into the base plate, followed by adhesive bonding. In view of the fastening strength and the precision of the rectangularity between the two, an axial thickness of 1 millimeter or a little less is desirably secured for the fastening portion.  
      Suppose that the portion opposed to the bottom end of the inner cylinder  63 , serving as the flange  65 , is given a minimum necessary dimension of around 0.5 millimeters and is integrated with a shaft to form the T-shaped shaft  61  for use. Then, a radial bearing space of 0.5 millimeters or so can be managed.  
      This makes it difficult, however, to secure the fastening strength and the rectangularity between the T-shaped shaft  61  and the base plate  16 . Thus, in  FIG. 6 , the position of the T-shaped shaft  61  with respect to the base plate  16  is regulated by the radial side  1   e  of the flange  65  of the T-shaped shaft  61 . The adhesive bonding strength and the rectangularity are secured by the outer periphery  1   d  of the surface in which the thrust bearing is formed.  
      According to the present embodiment, the magnetic attractive force is balanced with the axial load capacity, which only a single thrust bearing generates by dynamic pressure. Nevertheless, this consequently causes the bottom end of the inner cylinder  63  to make contact and slide over the flange  65  under the magnetic attractive force when at the start of rotation, and at halt.  
      In the present embodiment, in order to avoid damage ascribable to the contact and slide, a solid lubricant comprising mainly molybdenum disulfide is applied to approximately 10 micrometers on the bottom end of the inner cylinder  63 . Alternatively, a DLC film of 1 micrometer or so may be formed effectively as a solid lubricating film.  
      In another possible method, projections having a height of several micrometers may be formed in a circumferential configuration, or in a spot configuration on the bottom end of the inner cylinder  63  or part of the flange  65  so that the frictional force at the time of rotation is reduced for easier startup. This is already public knowledge in flying head technology, and description thereof will thus be omitted.  
       FIG. 8  shows a third embodiment. Like the second embodiment, this third embodiment will deal with an example of cylindrical shaft. Description will thus be concentrated on differences from the second embodiment shown in  FIG. 6 .  
      The sleeve, which rotatably fits to a T-shaped cylindrical shaft  61 , is composed of an inner cylinder  85  and an outer cylinder  84 . A channel  86  is formed in the gap between the outer cylinder  84  and the inner cylinder  85 . That is, two herringbone grooves  81  and  82  are formed as radial bearings between the T-shaped shaft  61  and the inner periphery of the inner cylinder  85 .  
      The lower herringbone groove  82  is formed asymmetric so as to have a downward lubricating fluid pumping capability. In addition, a pump-in spiral groove  83  is formed on the flange  65  of the shaft  61  opposed to the bottom end of the inner cylinder  85 .  
      During rotation, the asymmetric herringbone groove  82  and the spiral groove  83  press the lubricating fluid toward each other to increase the pressure of the lubricating fluid at the bottom end of the inner cylinder  85 . The rotating part is supported without contact at the point where the resulting axial load capacity and the magnetic attractive force are balanced.  
      When an outer cylinder  84  and the inner cylinder  85  constituting the sleeve are rotated, the pressure of the lubricating fluid is increased locally by the herringbone grooves  81 ,  82  and the spiral groove  83 , whereby the outer cylinder  84  and the inner cylinder  85  are supported without contact. Here, the herringbone groove  82  is configured to have the downward lubricating fluid pumping capability, and the lubricating fluid pumping capability of the herringbone groove  82  is set smaller than that of the spiral groove  83  at a predetermined rotational speed. The lubricating fluid thus keeps flowing across the herringbone grooves  82  and  81  toward the top end of the inner cylinder  85 .  
      The lubricating fluid is thrown out into a channel  86  formed in the gap between the outer cylinder  84 , and the inner cylinder  85 , by centrifugal force. The lubricating fluid is further conveyed to near the inner periphery of outer cylinder  84 , and finally reaches the outlet portion near the lower periphery of the inner cylinder  85 .  
      The channel  86  is different in its shape from the channel of the second embodiment, details are illustrated in  FIGS. 9, 10 ,  11 .  FIG. 9 ( a ) is showing the plain drawing of the sleeve top, an air section  91 , a lubricating fluid region  93 , and their interface  92  in the gap dimishing region of the channel  86  are illustrated. Numeral  94  shows rotational direction of the inner cylinder  85  and the outer cylinder  84 .  
      For the sake of regulation of the amount of axial movement of the rotating part, an engaging portion  88  is formed in an area where the lubricating fluid is in contact. As shown enlarged in  FIG. 10 , the lower periphery of the outer cylinder  84  reduces in diameter with an increasing distance from the bottom end to above, and the gap from the annular member  87  (corresponding to the annular member  1   c  shown in  FIG. 1 ) is increased gradually to form a tapered seal portion.  
      In addition, a side  101  of the annular member  87 , and a side  102  protruded from the periphery of the inner cylinder  85 , are engaged to form an engaging portion. The tilted inner periphery of the annular member  87  toward the periphery of the outer cylinder  84  may be sharpened to engage with the tilt of the lower periphery of the outer cylinder  84  for the sake of a structure providing positional regulation. Even when the sleeve is moved upward by excessive impact and the engaging portion makes contact or slides, the presence of the lubricating fluid can avoid serious problems such as damage or the production of abrasive dust.  
      The peripheral portion of the spiral groove  83  is where negative pressure can easily occur during high speed rotation. Countermeasures will now be described with reference to  FIG. 10 .  
      While the spiral groove  83  pumps the lubricating fluid radially inward, the radially-outward centrifugal force acting on the peripheral portion can lower the pressure of the lubricating fluid to a negative pressure. This makes it easier for bubbles to reside. The numerals  105  represents an intersection of the outer cylinder  84  with the interface  104  between the lubricating fluid with the air, while the numeral  106  represents an intersection of the annular member  87  with the interface. The portion of the lubricating fluid interface  104  around the intersection  105  is moving rapidly with the outer cylinder  84 , and the portion of the lubricating fluid interface  104  around the intersection  106  is at rest with the annular member  87 . In the present embodiment, the spiral groove  83  is given an outer diameter greater than the outer diameter of the outer cylinder  84 , i.e., it is arranged radially outside the high-speed flow side ( 105 ) of the interface  104  of the lubricating fluid  103 .  
      Consequently, the centrifugal force acting on the lubricating fluid that is rotating and flowing at high speed is integrated along the surface of the outer cylinder  84 . The pressure of the lubricating fluid reaches its maximum near the periphery of the bottom end of the inner cylinder  85 . In this structure, the centrifugal force is then utilized to apply pressure to near the periphery of the spiral groove  83 , thereby avoiding the occurrence of negative pressure.  
       FIG. 11 ( b ) shows a perspective view of the inner cylinder  85 . A groove  25 ′ formed on the surface of the inner cylinder  85  is different from the groove  25  in its shape. The groove  25  is linear and the groove  25 ′ is spiral shape. The direction of the spiral shape groove  25 ′ is to press the lubricating fluid from the channel outlet toward the inlet during rotation.  
       FIG. 11 ( a ) shows a perspective view of the combination of the inner cylinder  85  and the outer cylinder  84 . Lower parts of the outer cylinder  84  have differnt length as shown in numerals  112 ,  113 . Numeral  112  indicates a rear part, and numeral  113  indicates a front part regarding the channel outlet  111 . The rear part  112  is an extended part of the outer cylinder  84 , and it makes gap smaller between the outer cylinder  84  and the annular member  87 . Therefore, the lubricating fluid is pressed into the channel outlet  111  during rotation.  
      FIGS.  12 ( a ) and  12 ( b ) show the enlarged view of an accumulation of the lubricating fluid of outer periphery of the sleeve and the channel close to its outlet, and the lubricating fluid pressure diagram. Numeral  103  indicates the lubricating fluid at the outer periphery of the sleeve, numeral  121  (also numeral  111 ) indicates the outlet of the channel  86 , and numeral  122  indicates the lubricating fluid in the channel  86 , respectively. Along the dotted line  125 , the point  126  inside of the interface  104 , the point  127  around the outlet  121 , the point  128  close to the bottom of the channel  86 , the point  129  inside of the interface  92  of the fluid  122  are shown in  FIG. 12 ( a ). Fluid pressures at these points are indicated in  FIG. 12 ( b ). The horizontal axis means the location of points on the dotted line  125 , and the vertical axis means the lubricating fluid pressure referring the atmospheric pressure P 0 .  
      The fluid pressure at the point  126  inside of the interface  104  is lower than the atmospheric pressure P 0 , and the fluid pressure at the point  127  is slightly higher than that by the centrifugal force. The fluid pressure at the point  129  inside of the interface  92  is lower than P 0 . The fluid pressure at the point  128  is increased by the centrifugal force acting on the lubricating fluid in the channel  86  from the point  129 . Pressure difference between the points  128  and  127  is the effect of that the lubricating fluid is pushed from the channel outlet  121  ( 111 ) during rotation.  
      The fluid pressure should be continuous as shown in  FIG. 12 ( b ) during rotation. While the quantity of the lubricating fluid at outer periphery of the sleeve increases, the interface  104  moves outward, and then the fluid pressure at the point  126  becomes higher towards P 0  because that a radius of the interface  104  curve becomes larger. Accordingly, the quantity of the lubricating fluid around the channel outlet  121  ( 111 ) is properly divided in the channel and at outer periphery of the sleeve as the fluid pressure is continuous as shown in  FIG. 12 ( b ).  
      In the embodiment shown in  FIGS. 8, 9 ,  10 ,  11  and  12 , the gap diminishing in the channel, slanted channel, and the structure to press the lubricating fluid from the channel outlet  121  ( 111 ) are employed as the lubricating fluid pressure adjuster for adjusting the outward/downward lubricating fluid pressure occurring in the channel around the channel outlet. And the embodiment can seal the lubricating fluid steadily even in case of higher rotational speed.  
      For a fourth embodiment of the present invention, description will be given of an example where a low-profile HDD is formed. FIGS.  13 ( a ) and  13 ( b ) show an example of configuration of the low-profile HDD which is formed by incorporating the third embodiment of the present invention, or the fluid dynamic bearing motor of the fixed shaft structure of  FIG. 8 .  
      The low-profile HDD shown in  FIG. 13 ( a ) has a fluid dynamic bearing motor  136  of fixed shaft structure which is formed on a case  131 , or on the base plate  16 . A magnetic disk  133  is loaded on the motor  136 . An actuator  135  for positioning a magnetic head  134  at a predetermined position on the magnetic disk  133  is provided. A cover  132  is fixed to the case  131 . The shaft  61  makes contact with the cover  132  from below, thereby supporting the cover  132 . None of electronic circuits and filter mechanisms for controlling the environment inside the HDD is shown.  
      In FIGS.  13 ( a ) and  13 ( b ), the fluid dynamic bearing motor  136  is shown with the internal bearing alone.  FIG. 13 ( b ) shows an enlarged view. In the present embodiment, it is assumed that the magnetic disk has a diameter of 25 millimeters or so, and the low-profile HDD has a thickness of 2.5 millimeters or so.  
      Due to the limitation on the thickness of the HDD, bolts for fixing the shaft  61  to the cover  132  are omitted. The shaft  61  is used as a supporting column which makes contact with the cover  132  from inside, and avoids inward deformation of the cover  132 . The numeral  138  designates the thickness of the cover  132 , the numeral  139  the distance from the inside of the cover  132  to the surface of the sleeve, the numeral  13   a  the axial thickness of the outer cylinder  84 , the numeral  13   b  the distance from the bottom of the outer cylinder  84  to the shoulder of the shaft  61  (the level of top end of the herringbone groove  81 ), and the numeral  137  the distance from the bottom of the case  131  to the bottom end of the herringbone groove  82 . The numeral  13   c  designates the distance from the top end of the herringbone groove  81  (the shoulder of the shaft  61 ) to the bottom end of the herringbone grove  82 , showing the length secured for the radial bearing part.  
      Suppose here that the dimensions designated by the numerals  138 ,  139 ,  13   a , and  13   b  are set at 0.1 millimeters each, and the dimension designated by the numeral  137  is set at 0.5 millimeters. The total thickness of the HDD of 2.5 millimeters then allows 1.6 millimeters for the effective length  13   c  of the radial bearing part. Since it is enough to assign 0.7 millimeters or so to each of the herringbone grooves  81  and  82 , the low-profile HDD having a thickness of 2.5 millimeters can be formed even if the asymmetric portion of the herringbone groove  82  is arranged in the middle.  
       FIG. 14  shows a fifth embodiment. Like the third embodiment, this fifth embodiment will deal with an example of cylindrical shaft. Description will thus be concentrated on differences from the third embodiment shown in  FIG. 8 . Major different points are an inner cylinder  142 , an outer cylinder  141 , and a cover  143  to constitute a sleeve and also the structure for the lubricating pressure adjuster. The centrifugal force acting on the lubricating fluid in the channel is one of major cause to destabilize the lubricating fluid sealing, this fifth embodiment makes radial thickness of the lubricating fluid in the channel minimum to suppress the effect of the centrifugal force. The gap diminishing region in the channel is arranged in parallel with axis, therefore the gap diminishing region has enough capability for the lubricating fluid sealing.  
       FIG. 15  is an enlarged view of the bearing right half of the fluid dynamic bearing motor shown in  FIG. 14 . Description will now be given of the detail structure. A sleeve is composed of an outer cylinder  141  and an inner cylinder  142  and a cover  143 . A channel  144  is formed as gap between the inner cylinder  142 , the outer cylinder  141 , and the cover  143 . The gap diminishing region in the channel  144  is arranged in parallel with the shaft  61 , its radial gap width being smaller toward the channel outlet  154  at the sleeve bottom end. The inner cylinder  142  has a gap around 50 micrometer with the cover  143 , and has concave groove  151  in radial direction. The concave groove  151  lead the thrown out lubricating fluid by the centrifugal force to the lubricating fluid in the channel  144  with minimum exposing area size against the air. Numerals  152 ,  153  indicate the interfaces with the air in the channel and at the sleeve outside, respectively.  
      Radial thickness of the interface  152  region is possible to set less than around 50 micrometer, fluid pressure increase by the centrifugal force should be small. And also, the inner cylinder  142  and the outer cylinder  141  are fixed each other, therefore minimum gap width of the gap diminishing region can be around zero. Therefore surface tension force of the lubricating fluid in the gap diminishing region is enough to retain the lubricating fluid against the centrifugal force.  
       FIG. 16  shows the inner cylinder  142 , the outer cylinder  141 , and the cover  143  to constitute the channel  144  of the fifth embodiment.  FIG. 16 ( a ) shows a perspective view of the inner cylinder  142 .  FIG. 16 ( b ) shows a perspective view of the combination of the inner cylinder  142 , the outer cylinder  141 , and the cover  143 .  
      The inner cylinder  142  shown in  FIG. 16 ( a ) has the concave groove  151  at its top. The numeral  161  represents a through hole in which the shaft  61  is fitted loosely, and the numeral  163  represents a slanted flat surface of the outer surface of the inner cylinder  142 , the numeral  162  represents the outer surface of the inner cylinder  142  besides the slanted flat surface  163 . As shown in  FIG. 16 ( b ), the outer surface  162  of the inner cylinder  142  is fittingly fixed with inner surface of the outer cylinder  141 , the slanted flat surface  163  and the outer cylinder  141  constitute the gap diminishing region that the gap width is being smaller toward the bottom. A hatched area shows the lubricating fluid staying zone  165 , numeral  164  shows air zone, numeral  152  shows the lubricating fluid interface with the air.  
      The fifth embodiment shown in  FIG. 16  constitutes the gap diminishing region by the slanted flat surface of the inner cylinder  142  and the outer cylinder  141 . Employing the inner cylinder  142  as a truncated cone shape side wall diminishing its diameter toward its upper end, a circular gap diminishing region can be made.  
      The lubricating fluid pressure adjuster employed in this embodiment is the diminishing gap in parallel with the axis  61 . Radial thickness of the interface  152  region is so small that fluid pressure increase by the centrifugal force should be small. And also, the minimum gap width of the gap diminishing region can be around zero. Therefore surface tension force of the lubricating fluid in the gap diminishing region is enough to retain the lubricating fluid against the centrifugal force even in the case of high speed rotation.  
      It is important to control the radial thickness of the interface  152  region in this embodiment, because lubricant pressure increase by the centrifugal force is proportional to radial depth of the lubricating fluid. In mass production stage, parts dimensions and also the lubricating fluid quantity may vary, confirmation of the lubricating fluid position or its span should be highly desirable. This embodiment enables to fix the cover  143  after confirming those dimensions, or to monitor those dimensions after finishing bearing assembly by applying transparent plate as the cover  143 .  
      The foregoing has shown that the fixed shaft type fluid dynamic bearing motor of the present invention is suited to achieving a low-profile HDD. This indicates the high potential of the present invention. When applied to an HDD having a sufficient thickness, the present invention realizes a fluid dynamic bearing motor having shaft vibrations significantly smaller than conventional structures.  
      In principle, the present invention is suitable for high speed rotations, and is suited to server-class HDDs of small sizes which require rotations as high as around 20,000 RPM.  
      In the present invention, a new lubricating fluid sealing method alternative to conventional tapered seals has been proposed, and the characteristics thereof have been described along with the principle of operation.  
      The embodiments have dealt with application examples such as a cone bearing and a cylindrical bearing which have a straight bearing surface. In addition thereto, structures having a curved bearing surface are also applicable.  
      Up to this point, the principle of operation and structure of the present invention have been described in conjunction with the embodiments. The foregoing embodiments are no more than a few examples given for the sake of describing the principle of operation of the present invention, and it is understood that modifications may be made to the materials, structures, and the like without departing from the spirit of the present invention, and the foregoing description by no means limits the scope of the present invention.  
      From the studies on the behavior of the lubricating fluid in fluid dynamic bearings, a fixed shaft type fluid dynamic bearing motor which has a low-profile and is free from the leakage of the lubricating fluid has been achieved. This motor is particularly suitable for a recording disk drive motor for high speed rotation in which low non-repetitive runout can be, and a low-profile recording disk drive whose case cover requires a supporting column.  
      The present application claims Convention priority based on a Japanese patent application 2004-196174, 2004-199022, 2004-227399 of which disclosure is incorporated herein by reference.