Patent Publication Number: US-6664687-B2

Title: Motor having single cone fluid dynamic bearing balanced with shaft end magnetic attraction

Description:
BACKGROUND OF THE INVENTION 
     1. Technical Field of the Invention 
     The present invention relates to a fluid dynamic bearing motor, and more particularly to a fluid dynamic bearing having a conical shape to enable the motor to be smaller in thickness and lower in cost. 
     2. Description of the Related Art 
     There has been a trend toward the fluid dynamic bearing motor as the power source for rotary memory devices, cooling fans, and the like, because of its quietness in operation and the necessity to reduce nonrepeatable runout (NRRO) of rotating parts. Portable applications of such electronic devices have been widespread, increasing the demands for further reduction in their thickness and required current. However, there are limitations on further reduction in thickness of the fluid dynamic bearings, because they need to have a certain span between the bearings for supporting the shaft in order to inhibit NRRO. Also, in order to maintain a constant clearance between the bearings, they must be machined with extreme precision in the order of submicrons, whereby it is difficult to produce them at low cost. 
     In order to make fluid dynamic bearings thinner, a novel structure is necessary which does not require two bearings for supporting the shaft at axially spaced positions. The bearings should have as little sliding area as possible so as to achieve a reduction in the required current. Further, cost reduction will be achieved through the development of a structure wherein the bearing clearance is maintained with necessary accuracy even with the components machined with a lower degree of precision. 
     Single cone fluid dynamic bearings, which can support loads of both radial and thrust directions, have attracted attention as having potentialities in many respects. However, while some single cone structures that help decrease the thickness of the bearing have been proposed, for example, in Japanese Laid-open Utility Model Publication No. Hei. 06-004731, these are for air dynamic bearings and anyway have not been very successful. The main reason is that the single cone bearing is structurally incapable of sufficiently inhibiting NRRO during rotation. Japanese Laid-open Patent Publications No. 2000-004557 and No. 2000-205248 propose combined use of a conical bearing and a cylindrical bearing to improve the overall performance. However, the cylindrical bearing requires high-degree machining precision for maintaining a constant bearing clearance, canceling out the advantages of the conical bearing. U.S. Pat. No. 5,854,524 discloses a single semi-spherical air dynamic bearing having a similar structure as that of the single cone bearing, but in this case also, the radius of two spherical surfaces must be strictly controlled to secure a sufficient radial load capacity, because of which cost reduction is hardly achievable. 
     Thus the problems yet to be resolved in single cone fluid dynamic bearing motors are how to improve the stability in its rotating attitude, and how to realize a structure which prevents leakage of the lubricant and yet is easy to assemble. 
     SUMMARY OF THE INVENTION 
     An object of the present invention is to resolve these problems and to provide a single cone fluid dynamic bearing motor which can be reduced in thickness and required current, and is simple and can be produced at lower cost. 
     A fluid dynamic bearing motor according to the invention includes a shaft having a diminishing conical taper surface, a sleeve having a conical concavity opposite the shaft, lubricant filled in a clearance between the shaft and the sleeve, and means for generating magnetic attraction between one end of the shaft and a cone apex of the sleeve, and an annular wall arranged around the shaft to face an outer circumferential wall of the sleeve, a clearance between the annular wall and the outer circumferential wall of the sleeve being increased in width toward an open end to form a taper seal of the lubricant. In this construction, a plurality of grooves are formed on a conical taper surface of one of the shaft and the sleeve, and the grooves are provided for creating load capacity when the motor rotates, whereby rotating parts of the motor are supported by axial components of the load capacity balanced with the magnetic attraction. The boundary of the lubricant is positioned around the sleeve, so as to enable a reliable seal to be formed even in high-speed operation. 
     A ring-shaped member is fixed to one end of the annular wall which is arranged around the shaft, and an annular recess is provided in the outer circumferential wall of the sleeve, the inner periphery of the ring-shaped member being positioned within the annular recess, so as to restrict an axial movable distance of the rotating parts. This structure serves as a stopper for the rotating parts in the case where the motor is subjected to a large shock. 
     The means for generating magnetic attraction includes a permanent magnet and a magnetic material, respectively provided inside the shaft and in the apex of the sleeve opposite to the shaft, or vice versa. Magnetic attraction developed at the end of the shaft acts on the shaft to adjust its position in cooperation with the load capacity created by the grooves, thereby ensuring the stable attitude of rotating parts. 
     Moreover, the following structures for a permanent magnet to protrude from one end of the shaft are proposed. The shaft includes a permanent magnet held inside. The permanent magnet is assembled with the shaft such that it is initially held movably but firmly enough to overcome the magnetic attraction as being substantially protruded from one end of the shaft, and is pressed into the shaft by a pressure larger than the magnetic attraction applied from both ends of the shaft and the sleeve to a predetermined position, where the cone apex of the sleeve or a plate spring interposed between the apex of the sleeve and the permanent magnet is resiliently deformed, whereby when the motor is stationary the permanent magnet and the apex of the sleeve or the plate spring make contact with each other, while they are brought out of contact when the motor is rotating, by a distance equal to or shorter than an axial flying height determined on conical surfaces of the shaft and sleeve. Thereby, the start-up failure caused by the conical surface of the shaft being fitted in the sleeve when the motor is not in operation can be avoided, improving the reliability of the motor. 
     Alternatively, the grooves may be formed on both opposite taper surfaces of the shaft and the sleeve at the almost same axial positions. In this constitution, the grooves have different angular length from each other in the circumferential direction. Thereby, each delay, from the time point when the bearing clearance becomes small until the time point when the pressure in the lubricant in the clearance becomes local maximum by the corresponding groove, is varied in proportion to the corresponding angular length of each of the grooves. Thereby, an improved constitution which can avoid half whirls and other unstable movements of the motor can be provided. 
     Alternatively, conductive magnetic powder may be mixed in the lubricant so as to electrically connect the rotating parts to ground because of the powder gathered in the magnetic field between the end of the shaft and the apex of the sleeve. 
     According to the fluid dynamic bearing motor of the present invention, the load capacity created by the rotation of the motor acts vertically with respect to the conical surfaces, causing the shaft and the sleeve to rotate in non-contact relationship at a position where the axial components of the load capacity and the magnetic attraction are in equilibrium. The radial components of the load capacity counterbalance each other at respective circumferential points, thereby contributing to the centering of the rotating parts. The load capacity itself acts vertically on the tapered surface of the cone, and therefore it serves to adjust the attitude of rotating parts when they tilt with respect to the fulcrum conforming to the cone apex. 
     Magnetic attraction developed at the end of the shaft acts on the shaft to adjust its position in cooperation with the load capacity created by the grooves, thereby ensuring the stable attitude of rotating parts. 
     The main reason why the prior art single cone bearing has failed to maintain the attitude of rotating parts is that the bearing was provided only with a load equal to the weight of its own, or even less than that by using a magnetic bearing in order to avoid friction during the initial and final periods of operation as disclosed in Japanese Laid-open Utility Model Publication No. Hei. 06-004731. As has been explained above, a good balance is achieved between two forces of the axial component of load capacity of the bearing versus the load. Therefore, a small load can only create a small load capacity, which is insufficient to create forces for maintaining stable attitude of rotating parts. In the fluid dynamic bearing of the present invention, a large load is applied on the bearing by the magnetic attraction acted between the shaft and the sleeve. Therefore, the load capacity of the bearing, which counterbalances the load, can be set to a desired large value, whereby the stability of the attitude of rotating parts is improved. The magnetic attraction may be varied case by case depending on permissible level of NRRO, the size of the motor, and various other conditions. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     These and other objects and advantages of the present invention will become clear from the following description with reference to the accompanying drawings, wherein: 
     FIG. 1 is a cross sectional view showing a fluid dynamic bearing motor according to an embodiment of the present invention; 
     FIG. 2 illustrates the means for generating magnetic attraction and the lines of magnetic flux in the embodiment shown in FIG. 1; 
     FIG.  3 ( a ) and FIG.  3 ( b ) illustrate the bearing section in detail, FIG.  3 ( b ) being a plan view of a sleeve, and FIG.  3 ( a ) being a cross sectional view of a shaft and the sleeve; 
     FIG.  4 ( a ) illustrates a cross-section of the shaft and the sleeve, and component forces of load capacity, and FIG.  4 ( b ) illustrates the distribution of pressure developed during the rotation; 
     FIG.  5 ( a ) shows an embodiment in which magnetic attraction generating means is provided at one end of the shaft, and FIG.  5 ( b ) shows another embodiment using a rotor magnet as the magnetic attraction generating means, given in explanation of the difference in how a position adjusting force acts on the rotary section; 
     FIG. 6 is a detailed cross sectional view of the bearing section having a permanent magnet at one end of the shaft for limiting contact between the shaft and the sleeve when they are stationary; 
     FIG. 7 is an explanatory view illustrating how the permanent magnet of FIG. 6 is fitted in a predetermined position; 
     FIG.  8 ( a ) illustrates a cross-section of the bearing section having a crown, with a graph showing the pressure distribution, and FIG.  8 ( b ) illustrates how the load capacity acts on the rotary section when it is offset from the center; 
     FIG.  9 ( a ) illustrates the pressure distribution with a cross-section of the bearing section having a crown and spiral grooves, and FIG.  9 ( b ) illustrates how the load capacity acts on the rotary section when it is offset from the center; 
     FIG.  10 ( a ) and FIG.  10 ( b ) are detailed views of the bearing section having a modified construction wherein grooves are formed on both opposite surfaces of the shaft and the sleeve, FIG.  10 ( a ) being a plan view of the sleeve, and FIG.  10 ( b ) being a cross sectional view of the shaft and the sleeve. 
     FIG. 11 is a cross sectional view of a modified construction of the embodiment in which a channel is formed through the shaft; 
     FIG.  12 ( a ) and FIG.  12 ( b ) are explanatory views illustrating how a ring-shaped member can be axially adjusted, FIG.  12 ( a ) being an enlarged cross sectional view of the vicinity of the ring-shaped member, and FIG.  12 ( b ) being a cross sectional view of the bearing section; 
     FIG. 13 illustrates major parts of the bearing section having a modified construction in which is used conductive magnetic powder; and 
     FIG. 14 is a cross sectional view of the prior art fluid dynamic bearing motor. 
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     Preferred embodiments of a fluid dynamic bearing motor according to the present invention will be hereinafter described with reference to the accompanying drawings. 
     The prior art fluid dynamic bearing motor structure is reviewed by FIG. 14 before the description of present invention. The fluid dynamic bearing motor possesses two radial bearings which are provided on the surface of shaft  91  or cylindrical sleeve  92 , and two thrust bearings which are provided on the surface of both sides of a thrust plate  93 , and has herringbone grooves respectively in each bearing. The clearance between thrust plate  93 , sleeve  92 , and thrust bush  94  which compose thrust bearings are ten-micron meter level, and also the clearance between shaft  91  and sleeve  92  which compose radial bearings is two-micron meter level with the lubricant. 
     Two radial bearings and the existence of thrust plate  93  make the entire motor thin difficult. The bearing clearance and also the right angle degree of shaft  91  and hub  95 , shaft  91  and thrust plate  93  should be well controlled at the mass production stage because the load capacity of the bearing depends on the clearance. These are factors of increasing the cost. Moreover the joint part of the thrust bush  94  and the sleeve  92  in a portion which the lubricant contacts, is joined to provide a seal by swaging, bonding, or laser welding. The lubricant leakage may be caused from the joint part space and a serious trouble is often invited. Reference numerals  96 ,  97 ,  98 , and  99  respectively represent a rotor magnet, a stator core, coils, and a base. 
     FIG. 1 is a cross sectional view of a fluid dynamic bearing motor according to a embodiment of the present invention. A shaft  11  has a diminishing conical taper, and a sleeve  12  arranged opposite the shaft  11  has a conical concavity. The clearance between the shaft  11  and the sleeve  12  is filled with magnetic oil as the lubricant. The shaft  11  is surrounded by an annular wall  23 , and the clearance between the annular wall  23  and the outer circumference of the sleeve  12  becomes wider in an axial direction, thereby forming a taper seal, where there is the boundary  17  of the lubricant. A permanent magnet  35  is provided within a through hole  39  in the shaft  11 , so as to generate magnetic attraction between itself and the top of the sleeve  12  which is made of a magnetic material. 
     The permanent magnet  35  is fitted within the through hole  39  such that it is initially held with clearance so as to be movable as being largely protruded from the shaft  11 . It is then brought into contact with the inside top limit of the sleeve  12 , and is adjusted via the through hole  39  and fixed in position at the time of assembling such that the permanent magnet makes contact with the sleeve when the motor is stationary, while they are brought out of contact when the motor is rotating, by a distance equal to or shorter than an axial flying height determined on conical surfaces of the shaft and sleeve. 
     Rotary section is composed of the shaft  11 , the annular wall  23 , a hub  41 , a rotor magnet  44 , and others, and fixed section is composed of the sleeve  12 , a base  43 , a stator core  47 , coils  50 , and others. 
     The bearing section is constituted by the shaft  11 , the sleeve  12 , and a series of herringbone grooves, to be described later, provided in one of the conical taper surfaces  13  of the shaft  11  and the sleeve  12 . The grooves serve to pump the lubricant toward their center to increase the pressure of the lubricant. The load capacity thereby created is in inverse proportion to the size of the clearance between the shaft and sleeve. Therefore, the clearance size is determined such that the axial components of the load capacity and the above-mentioned magnetic attraction are in equilibrium, while radial components of the load capacity are used for the centering of the shaft  11 . Accordingly, the magnetic attraction, which determines the load capacity, is set so that the load capacity is large enough to support the rotary section during rotation. The clearance, accordingly, is approximately several micrometers wide. When the apical conical angle of the bearing section is large, the axial components of the load capacity may be given more consideration, while the radial components play a more important role when the apical conical angle is small. In this embodiment, the angle of the cone apex is slightly smaller than 60° so as to give more weight to the radial components to ensure precise centering of the shaft. 
     The stator core  47  and the coils  50  cooperate with the rotor magnet  44  to drive the rotary section. The rotary section further includes a magnetic or optical disk or the like carried thereon as a load. The force applied to the interface between the shaft  11  and the sleeve  12  varies depending on the manner in which the memory device is installed in a normal state or inverted state. That is, if the device is set in a normal state, the bearing receives the weight of the movable parts in addition to the magnetic attraction. If the device is set in an inverted state, the bearing receives a load less than the magnetic attraction because the weight of the movable parts is subtracted therefrom. In light of this, the magnetic attraction should be approximately three times larger than the weight of the movable parts, which has empirically been confirmed to ensure stable rotating attitude of the rotary section. If the magnetic attraction is increased so as to create accordingly larger load capacity, precession of the shaft can further be restricted and its attitude can be made more stable. On the other hand, it has been ascertained that such increase in the magnetic attraction causes the sliding friction to become larger at the time of starting up or stopping the motor, resulting in shorter operable life of the bearing. Therefore, in the case of the fluid dynamic bearing motor for a small magnetic disk device, magnetic attraction should be approximately five times larger than the weight of the movable parts, which is the sum of the weight of the rotary section and the load weight. Such settings may be determined case by case depending on the required precision for the rotating attitude of rotary section. 
     FIG. 2 illustrates a cross-section of the shaft  11 , permanent magnet  35 , sleeve  12  and others, together with lines of magnetic flux. The shaft  11  is made of a non-magnetic material while having the permanent magnet  35  inside, which is made of highly magnetic rare earth. Provided that the bearing section has about 5 mm diameter, the design value allotted to the diameter of the permanent magnet  35  is 1 to 2 mm. Since the top of the shaft  11  and the sleeve  12  are arranged with a very small clearance of about 10 micrometers therebetween, magnetic attraction remains constant irrespective of the variations in this clearance. Thus the tolerance in machining and assembling can be set larger. Reference numeral  55  indicates the direction of magnetization of the permanent magnet  35 . The distal end of the permanent magnet  35  is formed spherical so as to concentrate the magnetic flux. The magnetic flux  56  thus intensified enters the truncated cone top end of the sleeve  12 , and as shown by the reference numerals  57  and  58 , it passes through the sleeve  12  and returns to the other end of the permanent magnet  35  via the conical tapered surface. The magnetic flux  58  from the conical tapered surface to the end of the permanent magnet  35  flies a long distance and over a large area and thus is distributed and low in intensity. Accordingly, the magnetic attraction between the conical tapered surface of the sleeve  12  and the permanent magnet  35  is weak. 
     In this specific example, magnetic oil is used as the lubricant. Therefore, a centripetal force acts on the magnetic oil around the truncated conical top of the sleeve  12  where the magnetic intensity is high, whereby air bubbles are positively eliminated from the bearing section. Normal oil can also be used for achieving the intended purpose of the bearing section. 
     Apart from the structure shown in FIG. 1, magnetic attraction could be developed using the rotor magnet  44  and the stator  47  axially offset from each other, or the rotor magnet  44  and the magnetic piece arranged below the rotor magnet. However, the former has a disadvantage that it produces vibration, and the latter causes an increase in consumed current because of the Foucault current developed in the magnetic piece. The magnetic attraction generating means in this embodiment can resolve all these problems encountered by the above-mentioned other mechanisms. 
     FIG.  3 ( a ) and FIG.  3 ( b ) illustrate the structure of the bearing section of the embodiment shown in FIG. 1 in more detail. FIG.  3 ( b ) is a plan view of the sleeve  12 , and FIG.  3 ( a ) is a cross sectional view of the shaft  11  and the sleeve  12 . As shown in FIG.  3 ( b ), a series of herringbone grooves  18  is provided on the taper surface  13  of the sleeve  12 . The grooves  18  are V-shaped shallow recesses of about several micrometers depth. When the motor rotates, the grooves  18  pump the lubricant from the outer and inner peripheral sides toward their central pointed ends to increase the pressure of the lubricant, so as to lift the shaft  11  from the sleeve  12  and support it in a flying state. In this embodiment, the grooves are formed so that the pumping capacity from the outer peripheral side toward the inner peripheral side is larger than that from the inner peripheral side toward the outer peripheral side, whereby the pumping capacity towards the inner peripheral side remains and the pressure of the lubricant on the inner peripheral side can be increased swiftly when starting up the motor, so as to decrease the sliding friction between the shaft  11  and the sleeve  12 . The grooves  18  illustrated in FIG.  3 ( b ) have larger groove length on the inner peripheral side, but this does not contradict the description in the foregoing, since the pumping capacity is determined by the diminishing degree of the circumferential length of the grooves and the radial length of the grooves. 
     The clearance between the annular wall  23  and the outer circumference of the sleeve  12  becomes wider in an axial direction, where a taper seal is formed, which provides a seal by the surface tension of the lubricant. To one end of the annular wall  23  is fixed a ring-shaped member  24 , of which inner periphery fits in an annular recess  26  formed on the outer circumferential wall of the sleeve  12 , thereby restricting displacement of the rotary section in axial directions. The ring-shaped member  24  is either resilient or partially cut out so as to be rotatably fitted into the annular recess  26  in advance during the assembly of the bearing components. Thereafter, the ring-shaped member  24  is fixed to the end face of the annular wall  23  by spot-welding or bonding through access holes  25 . Three such access holes  25  are provided at circumferentially spaced points so as to evenly secure the ring-shaped member  24 . 
     Since the taper seal of the lubricant is formed not on the outer periphery of the conical bearing surface but on the outer circumference of the sleeve  12 , the overall thickness of the motor can be made smaller. Meanwhile, the taper seal can have a sufficient space in the axial direction, whereby the taper angle can be made as small as 10° or lower to form a strong seal of the lubricant. The boundary  17  of the lubricant is therefore formed not between conical surfaces, but between substantially vertical outer walls of the sleeve  12  and the annular wall  23 . Therefore there is no risk that the lubricant may leak under centrifugal force even in high-speed operation. 
     FIG.  4 ( a ) and FIG.  4 ( b ) illustrate the distribution of pressure developed in the lubricant when the motor rotates and the component forces of the load capacity applied to the interface between the shaft  11  and the sleeve  12  in accordance with the pressure distribution. These drawings are given in explanation of how the rotating attitude of the shaft is self-adjusted. 
     FIG.  4 ( b ) shows various features  62 ,  63 ,  64 , and  65  of the pressure distribution of the lubricant caused by the grooves  18  in operation. The y-axis  60  represents pressure, while the x-axis  61  indicates radial coordinates corresponding to FIG.  4 ( a ). The pressure reaches a highest point  63 ,  65  at positions substantially corresponding to the pointed ends of the V-shaped grooves  18 . The drawing shows the pressure distribution without the influence of the atmospheric pressure, and therefore the pressure  62  at an outer peripheral point is almost zero. On the other hand, the pressure  64  at an inner peripheral point is higher than the atmospheric pressure, because the grooves  18  are formed to have larger pumping capacity towards the inner peripheral side. 
     FIG.  4 ( a ) shows a cross-section of the shaft  11  and the sleeve  12 . Reference numerals  67 ,  68  in FIG.  4 ( a ) represent the load capacity created as the pressure in the lubricant increases. It should be noted that such a load capacity is created at each one of the several circumferentially located points, but only two of these are shown in a cross-section for the ease of explanation. 
     Reference numerals  69 ,  71  represent the axial components of the load capacity  67 ,  68 , respectively. Reference numerals  70 ,  72  represent respective radial components thereof. Since the load capacity  67 ,  68  is substantially in inverse proportion to the size of the clearance between the shaft  11  and the sleeve  12 , the clearance is determined such that the axial components  69 ,  71  and the magnetic attraction between the rotary section and the fixed section are in equilibrium. The radial components  70 ,  72  act in opposite directions so that they counterbalance each other, whereby the shaft  11  is centered. 
     The load capacity  67 ,  68  acts vertically to the conical surfaces. Thus, it acts on the shaft  11  as moment, i.e., the distance L multiplied by the load capacity  67 ,  68 , where L is the distance from an imaginary fulcrum  66  corresponding to the cone apex and the point from which the load capacity  67 ,  68  acts. The moment resulting from the load capacity  67 ,  68  acts in reverse directions, and because the load capacity  67 ,  68  is substantially in inverse proportion to the nearby clearance between the shaft  11  and the sleeve  12 , the moment caused by the load capacity  67 ,  68  acts around the fulcrum  66  as a position adjusting force for the shaft  11 , counterbalancing each other to equalize the clearance between the shaft  11  and the sleeve  12 . Thereby, the attitude of the shaft  11  is maintained upright, and its precession is restricted. 
     Viscosity of the oil used as the lubricant generally decreases at a high temperature, leading to a decrease in the load capacity. It is the practice in the prior art to set the load capacity high to allow for the decrease in pressure over a maximum limit of the temperature range for use, as a result of which there are the problems of excessive load capacity and large current at lower temperatures. According to the invention, the clearance between the shaft  11  and the sleeve  12  is changed corresponding to the equilibrium between the axial components  69 ,  71  of the load capacity  67 ,  68  and the magnetic attraction, and therefore the load capacity is kept substantially constant irrespective of the temperature. That is, a temperature compensation is automatically provided. This allows the load capacity to be set constant over the entire range of temperatures, eliminating the problems of excessive load capacity or current at low temperatures, and enabling a design with low current to be made. 
     Furthermore, the motor according to the invention is low in respect of bearing loss. Bearing loss of the fluid dynamic bearing is mainly caused by friction between the surfaces of the shaft  11  and sleeve  12  and the lubricant in small clearances where the grooves exist. The bearing according to the invention has only a series of grooves, which is a practical minimum, and thereby can achieve a reduction in required current. 
     The moment which acts on the shaft  11  to maintain its attitude is defined by the product which is obtained by multiplying the distance L by the load capacity  67 ,  68  as noted above. Therefore, there is no need to provide two series of grooves with a large span therebetween in an axial direction as in the prior art. The motor according to the invention needs only one series of grooves  18 , therefore the structure is more simple and thinner than the prior art. 
     FIG.  5 ( a ) and FIG.  5 ( b ) are given in explanation of how the magnetic attraction at one end of the shaft effectively acts to restore its rotating attitude. FIG.  5 ( a ) illustrates one example of the embodiment in which magnetic attraction is developed at one end of the shaft, and FIG.  5 ( b ) illustrates another example having a rotor magnet as the magnetic attraction generating means. Both of these drawings show a state wherein the upper part of the shaft  11  is tilted leftwards, and the load capacity created by the dynamic pressure is denoted at reference numerals  67 ,  68  on the left and right sides similarly to FIG.  4 ( a ). The load capacity  67  is larger than the load capacity  68  because of the difference in the bearing clearance, thereby acting as a moment force on the shaft  11  to restore its rotating attitude as has been described with reference to FIG.  4 ( a ). It will be understood from FIG.  5 ( a ) that the magnetic attraction  83  acting from the top end of the shaft  11  serves as the moment to adjust the rotating attitude jointly with the load capacity  67 ,  68 . In the example shown in FIG.  5 ( b ) using the rotor magnet  44  as the magnetic attraction generating means, the magnetic forces  84 ,  85  on both sides are substantially balanced with each other, but these magnetic forces  84 ,  85  developed between the rotor magnet  44  and the magnetic piece  53  are in inverse proportion to the clearance therebetween. Therefore, the magnetic force  84  on the side on which the clearance is smaller becomes larger than the magnetic attraction  85 , disturbing the balance between the load capacity  67 ,  68 . Thus the structure in which one end of the shaft has magnetic attraction can more advantageously help maintain the stable attitude of rotary section. 
     FIG.  6  and FIG. 7 are detailed views of the bearing section illustrating a modified construction of the embodiment wherein the permanent magnet prevents the shaft and the sleeve from making surface contact with each other when they are stationary. As shown in FIG. 6, the permanent magnet  35  is provided at the top end of the shaft  11 , such as to contact a plate spring  33  placed at the inside top limit of the conical sleeve  12  when stationary. The dotted lines  11   a  illustrate the position of the shaft when stationary, while the solid lines indicate the position of the shaft  11  when rotating. The permanent magnet  35  protrudes by a predetermined amount such that f≧d, where d is the distance between the top of the permanent magnet  35  and the inside top limit of the sleeve  12 , and f is the axial flying height of the shaft  11  from the sleeve  12  measured at conical surfaces. To be specific, the permanent magnet  35  is protruded so that f—d is about 5 micrometers if the flying height is within the range of 10 to 20 micrometers, taking into account that the flying height f of the shaft  11  varies depending on temperatures. Thus the top of the shaft  11  flies up from the plate spring  33  at least about 5 micrometers during rotation, while its conical surface flies up to an axial height of about 10 to 20 micrometers, maintaining a stable rotating attitude. 
     Conical bearings have a potential risk that the shaft fits into the sleeve, increasing the friction therebetween, resulting in start-up failure. This is caused by various factors such as the intensity of magnetic attraction, the apical conical angles, and the hardness of the material making up the shaft and sleeve, correlating with each other. Small motors to which the present invention is applied are relatively free of such troubles, but the structure shown in FIG. 6 further ensures that no such troubles will occur. 
     FIG. 7 is given in explanation of how the permanent magnet shown in FIG. 6 is adjusted in position. The permanent magnet  35  is initially fitted in the cylinder  32  inside the shaft  11  with clearance so as to be movable, but firmly enough to overcome the magnetic attraction. For assembling the permanent magnet  35 , it is placed upon the shaft  11  as being protruded substantially therefrom, and the sleeve  12  is coupled thereon. Pressure that is larger than the magnetic attraction is then applied to the sleeve  12  and the shaft  11  so that the permanent magnet  35  contacts the plate spring  33  placed at the inside top limit of the sleeve  12 , until the shaft  11  and the sleeve  12  make surface contact with each other on their conical surfaces and the plate spring  33  is resiliently deformed. The dotted lines  11   b  show the position of shaft under the pressure and the dotted lines of plate spring  33  indicate the deformed one under pressure, while the solid lines indicate the plate spring  33  having restored initial shape, after the pressure has been removed. As the plate spring  33  resiliently returns into its initial shape, a clearance is created between the conical surfaces of the shaft  11  and the sleeve  12 . The resilient deformation of the top of the sleeve may be arranged on the inside top limit of the sleeve  12  instead of utilizing a plate spring. 
     After the position alignment, the permanent magnet  35  should preferably be fixed in position by bonding or welding, so as to withstand large shocks. Further, it is preferable to provide antifriction measures on the top of the permanent magnet  35  and the opposite plate spring  33  placed at inside top limit of the sleeve  12  such as application of a ceramic material or plating treatment, so as to ensure stable performance over a long time. 
     Single cone bearings have the characteristics that even when the shaft and the sleeve have slightly different diameters, they still can face each other at given axial positions, whereby the tolerance of their dimensions can be made large, offering the advantage of lower cost. The permanent magnet  35  shown in FIG. 6 could initially be fixed to the shaft  11 , but in that case the diameters of the shaft  11  and the sleeve  12  and the protruding amount of the permanent magnet  35  must precisely be controlled. If the demands for the performance of the fluid dynamic bearing motor in regard to inhibition of NRRO are relatively low, then such control of dimensions could easily be achieved, while it is not if the demands are high. Thus the total cost would be lower with the structure wherein the permanent magnet allows itself to be positionally adjusted as in this embodiment. 
     FIG.  8 ( a ) shows a shaft fixed type embodiment in which has herringbone grooves and crown in the conical surface. The herringbone grooves in the conical surface are formed to have flat region in central parts. While the grooves  20 ,  21  on both sides are shown in the cross sectional view so that their positions are more clearly understood, they are actually formed on the surface of the conical shaft  11 , having a several micrometers depth. The shaft  11  has a slightly bulging crown  19  on its conical surface so as to have a flat band region where the bearing clearance is minimum. Correspondingly, a circumferential groove  40  of about 10 micrometers depth is provided in the sleeve  12  opposite the flat band region formed by the crown  19 . Specific dimensions of the crown  19  may differ case by case depending on various conditions, but basically they are set such that the bearing clearance at the outermost periphery of the conical shaft  11  and the sleeve  12  is several micrometers larger than that in the flat band region. With this construction, even if the apical conical angles of the shaft  11  and the sleeve  12  are not precisely in conformity with each other, edge contacts at the inner and outer peripheries can be prevented. Therefore, the machining tolerance of the components can be made larger. 
     The herringbone grooves are made up of two types of spiral grooves for pumping in and pumping out purposes. In other words, pumping-out spiral grooves  20  are positioned on the inner peripheral side, while pumping-in spiral grooves  21  are arranged on the outer peripheral side, with the crown  19  for making the bearing clearance minimum positioned therebetween. The number of grooves per one round, the inclination angle of the grooves, and other features of the grooves can suitably be set according to their purposes. 
     FIG.  8 ( a ) shows the pressure distribution observed during the operation of the bearing having the above-described grooves. The y-axis  73  indicates axial coordinates, while the x-axis  74  represents pressure. Reference numerals  75 ,  76 ,  77 ,  78 , and  79  represent mean pressure values in a circumferential direction at respective axial positions. The drawing shows the pressure distribution without the influence of the atmospheric pressure, and therefore the pressure  75  at an outer peripheral point is zero. The pressure increases as denoted by the reference numeral  76  because of the grooves  21 , and becomes constant in the central region as indicated by the reference numeral  77 . The pressure decreases at a position where the grooves  20  are formed as indicated by the reference numeral  78 . At the top  14  of the cone, the pressure is slightly higher than the atmospheric pressure as indicated by the reference numeral  79 . 
     The attitude of the rotary section is basically maintained by the high pressure  77  in the central region. A more specific account of the position adjusting mechanism will be given below with reference to FIG.  8 ( b ). The pressure values  75 ,  76 ,  77 ,  78 , and  79  in the pressure distribution of FIG.  8 ( a ) are mean values in circumferential directions and they may locally vary if the sleeve  12  comes off-center or tilts with respect to the shaft  11 . FIG.  8 ( b ) illustrates a state wherein the sleeve  12  is rotating as being inclined leftward at the upper part thereof and rightward at the lower part thereof with respect to the shaft  11 . The load capacity, created by the grooves  20  in the central region where the clearance is made small by the crown  19 , becomes uneven in the circumferential direction, i.e., the load capacity F 11  on the right side becomes larger than the load capacity F 12  on the left side because the bearing clearance is smaller on the right side. Similarly, the pressure developed by the grooves  21  becomes uneven, the load capacity F 21  on the right side being smaller than the load capacity F 22  on the left side where the bearing clearance is smaller. Here, the load capacity acts on the upper part of the sleeve  12  as moment of L 1 *(F 11 -F 12 ), while it acts on the lower part of the sleeve  12  as moment of L 2 *(F 21 -F 22 ), where L 1 , L 2  are the distances from an imaginary fulcrum  66  corresponding to the cone apex and the respective points from which the load capacity F 11 , F 12 , F 21 , F 22  acts. The moment acts around the fulcrum  66  as a force to make the bearing clearance at respective points equal. It should be noted that the description given above is simplified and the moment actually counterbalances each other at all circumferential and axial points, not only on the left and right sides. 
     In this way, by arranging a series of herringbone grooves on the conical surface with a small clearance region therebetween, a moment force is generated that acts on the rotary section to equalize the upper and lower clearances between the shaft  11  and the sleeve  12 , thereby adjusting the rotating attitude of the rotary section. Thus the precession is further restricted in the fluid dynamic bearing motor of this embodiment. When the sleeve  12  comes off center with respect to the shaft  11 , the pressure in the lubricant locally increases because of the wedge effect in the intermediate small-clearance band region formed by the crown  19 . A delay from the time when the bearing clearance is reduced until the time when a large pressure is developed may induce half whirls or other unstable movements of the rotary section. This is why the circumferential groove  40  is provided, as it distributes the locally collected lubricant in circumferential directions, thereby enhancing the position adjusting effect by the grooves and preventing half whirls. 
     FIG.  9 ( a ) and FIG.  9 ( b ) illustrate the bearing section having spiral grooves formed on the conical taper surface of the shaft  11 . The conical shaft  11  has a crown  19  so that the clearance between its intermediate band region and the sleeve  12  becomes minimum. The spiral grooves  22  for the pumping-in purpose are provided on the surface on the outer peripheral side of the shaft  11 . Reference numerals  80 ,  81 , and  82  denote mean values of pressure at respective axial positions. As shown, the pressure becomes constant on the inner peripheral side from the spiral grooves  22  as indicated by the numeral  82 . As can be seen from FIG.  9 ( b ), the pressure may vary in circumferential directions in accordance with the change in the clearance between the shaft  11  and the sleeve  12  over the area from the grooves to the small-clearance band region. FIG.  9 ( b ) illustrates the load capacity F 21 , F 22  in a state wherein the sleeve  12  is tilted leftwards and the bearing clearance is small on the lower left side. Since the load capacity is in inverse proportion to the bearing clearance, F 22  is larger than F 21 . Thus, it acts on the sleeve  12  as moment of L 2 *(F 21 -F 22 ), where L 2  is the distance from the imaginary fulcrum  66  conforming to the cone apex to the point from which the load capacity F 22  acts. The moment acts to equalize the bearing clearance, as a result of which the attitude of the sleeve  12  is adjusted. It should go without saying that the moment force acts circumferentially on the sleeve  12 , although the drawing illustrates moment forces acting from only both sides for the ease of explanation. 
     In this embodiment, even without the crown  19 , whenever the shaft comes off-center, the pressure distribution becomes uneven in the circumferential direction, whereby the moment acts on the sleeve  12  to adjust its rotating attitude. However, the crown  19  causes the pressure distribution to become uneven at a more peripherally outer position, whereby the moment force L 2 *(F 21 -F 22 ) can be made larger. 
     FIG.  10 ( a ) and FIG.  10 ( b ) show the vicinity of the bearing section according to a further modified construction of the embodiment in which grooves are formed on both opposite surfaces of the bearing section. FIG.  10 ( b ) is a cross-section of the shaft and the sleeve. The shaft  11  has a permanent magnet  35  inside for generating magnetic attraction. On its outer surface, a series of spiral grooves  20  is formed on its upper part for the pumping-out purpose, and another series of spiral grooves  21  is formed on its lower part for the pumping-in purpose. FIG.  10 ( a ) shows a bearing surface of the sleeve  12  in a plan view. As shown, the sleeve  12  has on its bearing surface a plurality of herringbone grooves  27  on its bearing surface. The grooves  20 ,  21 , and  27  have a depth of about several micrometers, and grooves  20 ,  21  on the surface of the shaft  11  and those  27  on the sleeve  12  have different angular lengths in the circumferential direction. In the specific example given in these drawings, the grooves  27  on the surface of the sleeve  12  have angular lengths of less than half as large as that of the grooves  20 ,  21  on the shaft  11  in the circumferential direction. The arrows  29 ,  30  indicate the direction in which the sleeve  12  rotates. 
     Grooves pump the lubricant when the bearing rotates to increase the pressure in the lubricant. The increased pressure, which is substantially in inverse proportion to the bearing clearance, causes a force to act on the rotary section to adjust its rotating attitude. Since the grooves are arranged at circumferentially spaced positions, even if the sleeve comes off-center with respect to the shaft and the bearing clearance becomes locally small, there is a delay until the balance in the circumferential pressure distribution is disturbed. This delay or time lag is in proportion to the angular length of the grooves in the circumferential direction. It is known that control systems with the time lag between the change in the controlled variable and the control over the change are susceptible to a resonant phenomenon, which, in the case of the fluid dynamic bearing, takes the form of precession, oil whip or other unstable movements. 
     In order to avoid such unstable movements, for example, the circumferential length of the grooves  21  may be varied so that the time lag is varied. However, if the angular lengths of only several grooves in one round are changed, the possibility of the position adjusting force not acting evenly increases, or other problem may arise. Therefore, in this embodiment, the grooves on the shaft  11  and those on the sleeve  12  are varied in their angular lengths in the circumferential direction so as to both achieve the circumferential evenness in the position adjusting force which is created by the increased pressure in lubricant, and the variety in the angular length of the grooves in the circumferential direction. Machining of the grooves is generally not easy and forming them on both bearing surfaces may lead to an increase in cost. However, the conical shaft  11  and the sleeve  12  in this embodiment can both be produced by molding, and therefore such grooves can be provided without increasing cost. Thus a fluid dynamic bearing motor with limited precession is realized. 
     FIG. 11 shows another modified construction of the embodiment having a channel  34  that runs through the shaft  11  from its truncated cone top  14  to the outer periphery thereof. The channel  34  is provided for circulating the lubricant compressed towards the top  14  of the shaft  11  to the outside of the cone. The channel  34  is filled with fibrous or porous material to adjust the flow resistance such that pressure remains at the top  14  of the cone, whereby the sleeve  12  can fly up swiftly at the time of start-up, and whereby shock-absorbing effects are achieved because of the compressed lubricant that escapes and adjusts the damping level. Moreover, galls produced on the sliding parts can be removed with the structure of this example. 
     FIG.  12 ( a ) and FIG.  12 ( b ) illustrate a modified construction of the embodiment wherein the ring-shaped member can be adjusted in axial directions. FIG.  12 ( b ) is a cross sectional view of the bearing section, and FIG.  12 ( a ) is an enlarged cross sectional view of part  89  of the ring-shaped member and other components. In this example, the annular wall  23  has a protrusion  86  on its upper end, while the ring-shaped member  24  has a corresponding through hole to match this protrusion. The ring-shaped member  24  is preliminarily coupled into the annular recess  26  around the sleeve  12  and assembled to the shaft  11 . Access holes  25  are provided, through which the protrusion  86  and the through hole of the ring-shaped member  24  are engaged with each other. Then, using a jig  88 , the inner periphery of the ring-shaped member  24  is abutted onto the end face  87  of the annular recess  26 . The ring-shaped member is thus coupled to the protrusion  86  as being resiliently deformed. 
     In this assembling process, the ring-shaped member  24  is resiliently deformed in an axial direction by about 20 micrometers, while being coupled to the protrusion  86  firmly. Thereby, axial displacement of the rotary section including the hub  41  is restricted to be about 20 micrometers even if it is subjected to large shocks. In the case of hard disk drives, there is a strong demand for restricting axial displacement of the magnetic disk to a minimum By utilizing resilient deformation of the ring-shaped member  24  as in this embodiment, such requirements can be met without higher demands for the tolerance of various components. Alternatively, the ring-shaped member  24  and the protrusion  86  may be joined after the assembly by bonding or welding to have a higher strength to withstand large impacts. 
     FIG. 13 is a detailed cross sectional view of the top end of the shaft  11  and the sleeve  12  according to another modified construction of the embodiment, wherein conductive magnetic powder is mixed in the lubricant. The drawing is given in explanation of how the magnetic powder gathers in the magnetic field between the top end of the shaft  11  and the sleeve  12  so as to achieve electrical conduction between the rotary section and the fixed section. In the center at the inside top limit of the sleeve  12  is provided a boss  36 . Thus, the conductive magnetic powder  37  mixed in the lubricant builds up between the central boss  36  in the sleeve  12  and the magnet  35  where the magnetic flux is intense, thereby bridging the two and forming a conductive path therebetween. The conductive magnetic powder  37  may be, for example, fine particles of ferrite magnet having a size of about 0.2 to 0.3 micrometers so as to remain magnetic, with a coating of about 100 angstrom thick gold. 
     The embodiments shown in FIG. 6, FIG. 11 employ a construction wherein no weld joints are formed between the members in a portion which the lubricant contacts. In the prior art, separate components were joined to provide a seal by swaging, bonding, or laser welding, but this was a major cause of later leakage of lubricant because of frequent bond failure, leading to a fatal fault. The present invention provides a fluid dynamic bearing motor free of the risk of oil leakage, as it eliminates joints in an area where the lubricant flows as shown in this embodiment. 
     For the material of the bearing components such as shaft and sleeve, any of the metal materials such as stainless steel or copper alloy which have commonly been used for the fluid dynamic bearing can be used. Preferably, a thin film of nickel, titanium, diamond-like-carbon, or molybdenum disulfide should be formed on one of the conical taper surfaces, so as to decrease the friction at the time of starting up and stopping the motor. 
     Regarding the manufacturing method of the bearing components, not to mention the shaft having a convex shape, the sleeve, although having a concave shape, it can be easily released, because its tapered top is opened. Therefore they both can be formed at one time including the grooves, by any known techniques such as press molding or injection molding. Accordingly, the bearing components can also be made of a ceramics or sintered alloy by molding, or of a resin material having superior antifriction properties such as polyphenyl sulfide resin (PPS) containing carbon fiber by molding, whereby a reduction in production cost is achieved. 
     Although the embodiments shown in above have been described as having the sleeve  12  or the shaft  11  and the hub  41  formed in one piece, they may be separate components and assembled together. Whether they should be produced in one piece or separately may be determined case by case so that the cost is lower, taking into consideration the characteristics and specifications required for each component. In the application of the invention to a hard disk drive as has been shown in these embodiments, however, there are stringent specifications with regard to the height and tilt of the install surface of the magnetic disk. Since these are strongly affected by their positional relationship with the bearing surface, it is more preferable to form the sleeve  12  or the shaft  11  and the hub  41  in one piece to achieve higher precision. The fluid dynamic bearing motor according to the present invention enables the integral structure of the sleeve or the shaft and the hub and realizes a high-precision, low-cost motor. 
     According to a fluid dynamic bearing motor of the present invention, the bearing section has a simple structure wherein grooves are formed on a conical taper surface for increasing the pressure in lubricant and creating a load capacity, which is balanced with magnetic attraction. With this structure, the attitude of the rotary section in the bearing is made stable, and a reliable seal of the lubricant is achieved even in high-speed operation. The bearings can be mass-produced at low cost by molding, and the total thickness of the motor can be reduced. Further, a temperature compensation of the load capacity for supporting the rotary section is achieved, and the current required for operating the motor is reduced. Therefore, the motor according to the invention is particularly suitable for small, rotary memory devices such as magnetic or optical disk devices, or cooling fans for CPUs. 
     While there has been described what are at present considered to be preferred embodiments of the present invention, it will be understood that various modifications may be made thereto, and it is intended that the appended claims cover all such modifications as fall within the true spirit and scope of the invention.