Abstract:
A high-accuracy, long-life hydrodynamic bearing that does not cause oil film breakage in bearing clearances and a disc rotation apparatus using the bearing is disclosed. Oil film breakage is avoided as negative pressure is prevented from generating between the shaft and sleeve of the hydrodynamic bearing. Herringbone shaped dynamic pressure generating grooves, located on the thrust bearing section and the radial bearing section of the hydrodynamic bearing, are oil filled and have optimum shapes. The optimum shapes prevent the generation negative pressure and thus prevents the coagulation of air bubbles that can cause oil film breakage. The disc rotation apparatus, that holds a reproduction/recording disc, is concentrically secured to the hydrodynamic bearing and rotated. The disc is put into contact with magnetic or optical heads while rotating in the disc rotation apparatus. Both the hydrodynamic bearing and the disc rotation apparatus experience high reliability.

Description:
BACKGROUND ART 
   1. Technical Field 
   The present invention relates to a hydrodynamic bearing having fluid in its rotation section and a disc rotation apparatus having the same. 
   2. Prior Art 
   In recent years, in recording apparatuses using discs and the like, their memory capacities are increasing and their data transfer speeds are rising. Hence, a disc rotation apparatus for use in this kind of recording apparatus is required to rotate at high speed and with high accuracy, and a hydrodynamic bearing is used in its rotating main shaft section. 
   A conventional hydrodynamic bearing and an example of a disc rotation apparatus having the same will be described below referring to  FIG. 12  to  FIG. 16 .  FIG. 12  is a cross-sectional view showing the right portion of the center line C indicating the center of the rotation shaft of the conventional hydrodynamic bearing. In  FIG. 12 , a shaft  31  is rotatably inserted into a sleeve  32  having a bearing hole  32 A. At the lower end of the shaft  31 , a flange  33  is provided so as to be integrated therewith. The lower end of the flange  33  is accommodated in a recess portion formed by a hole in a base  35  and the sleeve  32  and rotatably held so as to be opposed to a thrust plate  34  mounted on the base  35 . A hub rotor  36 , a rotor magnet  38 , a plurality of discs  39 , spacers  40  and a clamper  41  are secured to the shaft  31 . A motor stator  37  opposed to the rotor magnet  38  is installed on the base  35 . Dynamic pressure generation grooves  32 B and  32 C indicated by broken lines are provided on the inner circumferential face of the bearing hole  32 A of the sleeve  32 . Dynamic pressure generation grooves  33 A are provided on the upper face of the flange  33 , a face opposed to the sleeve  32 . In addition, dynamic pressure generation grooves  33 B are provided on the lower face of the flange  33 , a face opposed to the thrust plate  34 . The clearances between the shaft  31  and the sleeve  32 , including the dynamic pressure generation grooves  32 B,  32 C,  33 A and  33 B, are filled with oil. 
   The operation of the conventional hydrodynamic bearing shown in  FIG. 12  will be described below. In  FIG. 12 , when electric power is applied to the coil of the stator  37 , a rotating magnet field is generated, and a rotation force is generated in the rotor magnet  38 , whereby the shaft  31  and the flange  33  rotate together with the hub rotor  36  and the discs  39 . During the rotation, dynamic pressures are generated in the oil by the dynamic pressure generation grooves  32 B,  32 C,  33 A and  33 B, and the shaft  31  is floated in the upward direction of the figure and rotates while holding space from the sleeve  32  and without making contact with the thrust plate  34  and the sleeve  32 . Magnet heads, not shown, make contact with the discs  39  and carry out the recording and reproduction of electrical signals. 
   The conventional hydrodynamic bearing configured as described above had problems described below.  FIG. 13  is a plan view of the flange  33  which is provided with a plurality of the dynamic pressure generation grooves  33 A indicated by black-colored regions.  FIG. 14  is a bottom view of the flange  33  which is similarly provided with a plurality of the dynamic pressure generation grooves  33 B indicated by black-colored regions. The outside diameters of the patterns of the dynamic pressure generation grooves  33 A and  33 B on the top and bottom faces are represented by D 1   o  and D 2   o , respectively, and their inside diameters are represented by D 1   i  and D 2   i , respectively. The diameters D 1   m  and D 2   m  of the respective turn-back parts of the dynamic pressure generation grooves  33 A and  33 B are set at sufficiently large values so that pumping pressures in the directions indicated by arrow E and arrow H, respectively, are raised. 
     FIG. 15  and  FIG. 16  are views showing the cross sections of relevant parts in the vicinity of the lower end of the shaft  31  and showing pressures on the surfaces of the flange  33  and the shaft  31  of the above-mentioned conventional hydrodynamic bearing. If the pumping pressures in the directions indicated by arrows E and H shown in  FIG. 13  and  FIG. 14 , respectively, are raised too high, a negative pressure with respect to atmospheric pressure is generated at the central portion of the lower face of the flange  33  as indicated by curve P 1  in  FIG. 15 , whereby air bubbles mixed in the oil are coagulated and air is accumulated in a region  43 B having a constant size. 
   In  FIG. 16 , the dynamic pressure generation grooves  32 B and  32 C of the sleeve  32  are made so that dimension L 1  in the figure is larger than dimension L 2 , (L 1 &gt;L 2 ), and so that dimension L 4  is larger than dimension L 3 , (L 4 &gt;L 3 ). In addition, the dimensional difference (L 1 −L 2 ) is selected so as to be nearly equal to the dimensional difference (L 4 −L 3 ), that is, (L 1 −L 2 )≈(L 4 −L 3 ). As shown by ΔL in  FIG. 16 , in the case that the amount of the oil becomes slightly insufficient and the upper face of the oil is at the position lower than the upper ends of the dynamic pressure generation grooves  33 B by dimension  4 L, no oil is present in the portion corresponding to the dimension ΔL of the upper ends of the dynamic pressure generation grooves  33 B, whereby the pressure distribution of oil is represented by curve P 2  shown in  FIG. 16 . In addition, a negative pressure is generated at the lower portion of the range of the dimension L 4  in the figure. Hence, air bubbles are accumulated in a region  43 A, whereby there is a fear of breaking the oil film in this region  43 A and of causing friction between the shaft  31  and the sleeve  32 . 
   SUMMARY OF THE INVENTION 
   The present invention purports to provide a hydrodynamic bearing in which a negative pressure is prevented from generating between the shaft and the sleeve, whereby oil film breakage due to locally accumulated air in oil does not occur. 
   A hydrodynamic bearing in accordance with the present invention comprises a sleeve having a bearing hole at the nearly central portion thereof, a shaft rotatably inserted into the bearing hole of the above-mentioned sleeve, and a nearly disc-shaped flange, secured to one end of the above-mentioned shaft, one face of which is opposed to the end face of the sleeve  1  and the other face of which is opposed to a thrust plate provided to hermetically seal a region including the above-mentioned end face of the above-mentioned sleeve, wherein herringbone-shaped first and second dynamic pressure generation grooves are provided on at least one of the inner circumferential face of the above-mentioned sleeve and the outer circumferential face of the above-mentioned shaft so as to be arranged in the direction along the shaft, herringbone-shaped third dynamic pressure generation grooves are provided on at least one of the opposed faces of the flange and the thrust plate, the above-mentioned first, second and third dynamic pressure generation grooves are filled with oil having a kinematic viscosity of 4 cSt (centi-stokes) or more at 40° C. of temperature, one of the above-mentioned sleeve and the above-mentioned shaft is secured to a base and the other is secured to a rotatable hub rotor, and when the outside diameter of the herringbone pattern of the above-mentioned third dynamic pressure generation groove is designated as d 1   o , the inside diameter thereof is designated as d 1   i , the diameter of the turn-back part is designated as d 1   m , the value of the diameter d 1   m  being in the range of 1 mm or more and 10 mm or less, and the diameter of the turn-back part of the herring pattern, wherein the oil pressure generated by the above-mentioned third dynamic pressure generation grooves in the direction from the outer circumference to the inner circumference of the flange becomes equal to the oil pressure generated in the direction from the inner circumference to the outer circumference thereof, is designated as dsy and is represented by:
 
 dsy={ ( d 1 i   2   +d 1 o   2 )/2} 1/2 ,
 
where the diameter d 1   m  of the turn-back part is determined so that the value obtained by subtracting the diameter d 1   m  from the diameter dsy, (dsy−d 1   m ), is in the range of 0.05 mm or more and 0.8 mm or less, that is, d 1   m =dsy−(0.05 to 0.8 mm).
 
   A hydrodynamic bearing in accordance with another aspect of the present invention comprises a sleeve having a bearing hole at the nearly central portion thereof, a shaft rotatably inserted into the bearing hole of the above-mentioned sleeve, and a nearly disc-shaped flange, secured to one end of the above-mentioned shaft, one face of which is opposed to the end face of the sleeve  1  and the other face of which is opposed to a thrust plate provided to hermetically seal a region including the above-mentioned end face of the above-mentioned sleeve, wherein herringbone-shaped first and second dynamic pressure generation grooves are provided on at least one of the inner circumferential face of the above-mentioned sleeve and the outer circumferential face of the above-mentioned shaft so as to be arranged in the direction along the shaft, among the above-mentioned first and second dynamic pressure generation grooves, when the grooves away from the above-mentioned thrust plate are designated as the first dynamic pressure generation grooves and the grooves close thereto are designated as the second dynamic pressure generation grooves, the first length L 1  of the groove portion, away from the above-mentioned thrust plate, of the above-mentioned herringbone-shaped first dynamic pressure generation groove in the direction of the shaft is larger than the second length L 2  of the groove portion close to the above-mentioned thrust plate in the direction of the shaft, when the diameter of the above-mentioned shaft is in the range of 1 mm or more and 10 mm or less, the value obtained by subtracting the length L 2  from the length L 1  is set in the range of 0.05 or more and 1.5 mm or less, herringbone-shaped third dynamic pressure generation grooves are provided on at least one of the opposed faces of the flange and the thrust plate, the above-mentioned first, second and third dynamic pressure generation grooves are filled with oil having a kinematic viscosity of 4 cSt or more at 40° C. of temperature, and one of the above-mentioned sleeve and the above-mentioned shaft is secured to a base and the other is secured to a rotatable hub rotor. 
   A hydrodynamic bearing in accordance with another aspect of the present invention comprises a sleeve having a bearing hole at the nearly central portion thereof, a shaft rotatably inserted into the bearing hole of the above-mentioned sleeve, and a nearly disc-shaped flange, secured to one end of the above-mentioned shaft, one face of which is opposed to the end face of the sleeve  1  and the other face of which is opposed to a thrust plate provided to hermetically seal a region including the above-mentioned end face of the above-mentioned sleeve, wherein herringbone-shaped first and second dynamic pressure generation grooves are provided on at least one of the inner circumferential face of the above-mentioned sleeve and the outer circumferential face of the above-mentioned shaft, among the above-mentioned first and second dynamic pressure generation grooves, when the grooves away from the above-mentioned thrust plate are designated as the first dynamic pressure generation grooves and the grooves close thereto are designated as the second dynamic pressure generation grooves, the first length L 1  of the groove portion, away from the above-mentioned thrust plate, of the above-mentioned herringbone-shaped first dynamic pressure generation groove in the direction of the shaft is larger than the second length L 2  of the groove portion close to the above-mentioned thrust plate in the direction of the shaft, the above-mentioned herringbone-shaped second dynamic pressure generation groove is made symmetric with respect to a line passing through the herringbone-shaped turn-back parts, the value of a calculation expression, (L 1 +L 2 )/(2×L 2 ) represented by using the above-mentioned first length L 1  and the above-mentioned second length L 2 , is in the range of 1.02 to 1.60, herringbone-shaped third dynamic pressure generation grooves are provided on at least one of the opposed faces of the flange and the thrust plate, the above-mentioned first, second and third dynamic pressure generation grooves are supplied with oil having a kinematic viscosity of 4 cSt or more at 40° C. of temperature, one of the above-mentioned sleeve and the above-mentioned shaft is secured to a base and the other is secured to a rotatable hub rotor, and when the outside diameter of the herringbone pattern of the above-mentioned third dynamic pressure generation groove is designated as d 1   o , the inside diameter thereof is designated as d 1   i , the diameter of the turn-back part thereof is designated as d 1   m , and the diameter of the turn-back part of the herring pattern, wherein the oil pressure generated by the above-mentioned third dynamic pressure generation grooves in the direction from the outer circumference to the inner circumference of the flange becomes equal to the oil pressure generated in the direction from the inner circumference to the outer circumference thereof, is designated as dsy and is represented by:
 
 dsy={ ( d 1 i   2   +d 1o 2 )/2} 1/2 ,
 
   the diameter d 1   m  of the turn-back part is determined so that when the diameter of the above-mentioned shaft is in the range of 1 mm or more and 10 mm or less, the value obtained by subtracting the above-mentioned length L 2  from the above-mentioned length L 1  is set in the range of 0.05 mm or more and 1.5 mm or less, the diameter d 1   m  is in the range of 1 mm or more and 10 mm or less, and the value obtained by subtracting the diameter d 1   m  from the diameter dsy is in the range of 0.05 mm or more and 0.8 mm or less, that is, d 1   m =dsy: (0.05 to 0.8 mm). 
   In accordance with the above-mentioned configurations of the present invention, the patterns of the dynamic pressure generation grooves in the thrust bearing section and the radial bearing section have optimum shapes, whereby no negative pressure is generated inside the bearing. Hence, since air accumulation due to the coagulation of air bubbles can be prevented, it is possible to provide a hydrodynamic bearing not causing oil film breakage. 
   A disc rotation apparatus using the hydrodynamic bearing in accordance with the present invention records or reproduces signals, wherein a recording/reproduction disc is concentrically secured to the hub rotor of the hydrodynamic bearing in accordance with claims  1  to  5  and rotated, magnetic heads or optical heads are provided so as to be opposed to the faces of the above-mentioned rotating disc, and the magnetic heads or optical heads are configured so as to be movable in parallel with the faces of the above-mentioned disc. By using the hydrodynamic bearing in accordance with the present invention, it is possible to obtain a disc rotation apparatus being high in reliability like that of the bearing. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
       FIG. 1  is a cross-sectional view of a hydrodynamic bearing in accordance with a preferred embodiment of the present invention; 
       FIG. 2  is a bottom view of the flange  3  of the hydrodynamic bearing in accordance with this embodiment; 
       FIG. 3  is a plan view of the flange  3  of the hydrodynamic bearing in accordance with this embodiment; 
       FIG. 4  is a graph showing the relationship between the pump pressure in the dynamic pressure generation grooves  3 A of the flange  3  and the dimensional distribution of the diameter dsy of the turn-back part and the inside diameter d 1   i  of the dynamic pressure generation groove  3 A in the hydrodynamic bearing in accordance with this embodiment; 
       FIG. 5  is a cross-sectional view of a relevant part showing the distribution of oil pressure generated by the dynamic pressure generation grooves  3 A and  3 B in the case when the floating distance S 1  between the flange  3  and the thrust plate  4  is sufficiently small in the hydrodynamic bearing in accordance with this embodiment; 
       FIG. 6  is a cross-sectional view of a relevant part showing the distribution of oil pressure generated by the dynamic pressure generation grooves  3 A and  3 B in the case when the floating distance S 2  between the flange  3  and the thrust plate  4  is sufficiently large in the hydrodynamic bearing in accordance with this embodiment; 
       FIG. 7  is a cross-sectional view of a relevant part showing the distribution of oil pressure in the radial bearing section and the distribution of oil pressure generated by the dynamic pressure generation grooves  3 A and  3 B of the flange  3  of the hydrodynamic bearing in accordance with this embodiment; 
       FIG. 8  is a cross-sectional view of a relevant part showing the distribution of oil pressure in the radial bearing section and the distribution of oil pressure generated by the dynamic pressure generation grooves  3 A and  3 B of the flange  3  in the case when the amount of oil is smaller than a specified amount in the hydrodynamic bearing in accordance with this embodiment; 
       FIG. 9  is a graph showing the relationship between the oil pressure generated by the dynamic pressure generation grooves  1 A and  1 B and the dimensional distribution of the dynamic pressure generation grooves  1 A and  1 B of the hydrodynamic bearing in accordance with this embodiment; 
       FIG. 10  is a graph showing the relationship between the bubble entry amount in oil and the kinematic viscosity of oil at 40° C. of temperature in the hydrodynamic bearing in accordance with this embodiment; 
       FIG. 11  is a cross-sectional view of a disc rotation apparatus using the hydrodynamic bearing in accordance with this embodiment of the present invention; 
       FIG. 12  is the cross-sectional view showing the right half of the conventional hydrodynamic bearing; 
       FIG. 13  is the plan view of the flange  33  of the conventional hydrodynamic bearing; 
       FIG. 14  is the bottom view of the flange  33  of the conventional hydrodynamic bearing; 
       FIG. 15  is the cross-sectional view of a relevant part showing the distribution of oil pressure generated by the dynamic pressure generation grooves  33 A and  33 B of the flange  33  in the conventional hydrodynamic bearing; and 
       FIG. 16  is the cross-sectional view of a relevant part showing the distribution of oil pressure generated in the radial direction by the dynamic pressure generation grooves  32 B and  3 C of the sleeve  32  of the conventional hydrodynamic bearing. 
   

   DESCRIPTION OF PREFERRED EMBODIMENT 
   A preferred embodiment of a hydrodynamic bearing in accordance with the present invention will be described below referring to  FIGS. 1 to 10 .  FIG. 1  is a cross-sectional view of a hydrodynamic bearing in accordance with an embodiment of the present invention. In  FIG. 1 , a sleeve  1  has a bearing hole  20  at its nearly central portion, and herringbone-shaped dynamic pressure generation grooves  1 A and  1 B are formed on the inner circumferential face of the bearing hole  20 . A recess portion  1 C is formed at the lower end of the sleeve  1 . A shaft  2  is rotatably inserted into the bearing hole  20 . A flange  3  is secured to the lower end of the shaft  2  so as to be accommodated in the recess portion  1 C at the lower end of the sleeve  1 . A thrust plate  4  is secured to the recess portion  1 C of the sleeve  1  by a securing method, such as laser welding, precision crimping or bonding, and the recess portion  1 C including the flange  3  is hermetically sealed. The sleeve  1  is secured to a base  6 . The shaft  2  is secured to a hub rotor  7 . Dynamic pressure generation grooves are provided on one of the opposed faces of the flange  3  and the thrust plate  4 . In  FIG. 1 , dynamic pressure generation grooves  3 A are provided on the lower face of the flange  3 . Dynamic pressure generation grooves  3 B are also provided on the upper face of the flange  3  opposed to the recess portion  1 C of the sleeve  1 . The insides of the dynamic pressure generation grooves  1 A,  1 B,  3 A and  3 B are filled with oil or grease. A rotor magnet  9  is installed in the hub rotor  7 . In addition, a stator  8  is installed on the base  6  so as to be opposed to the above-mentioned rotor magnet  9 . Two discs  10 , for example, are installed on the hub rotor  7  via a spacer  12 . The discs  10  are secured by a damper  11  installed on the shaft  2  by a screw  13 . 
   The operation of the hydrodynamic bearing in accordance with this embodiment configured as mentioned above will be described with reference to  FIGS. 1 to 10 . In  FIG. 1 , first, when electric power is applied to the coil of the stator  8 , a rotating magnet field is generated, and the rotor magnet  9  receives a rotation force, and the hub rotor  7 , the shaft  2  and the discs  10  rotate together with the damper  11  and the spacer  12 . By the rotation, the dynamic pressure generation grooves  1 A,  1 B,  3 A and  3 B rake up oil, and pressures are generated between the dynamic pressure generation grooves  1 A and  1 B and the shaft  2  and between the dynamic pressure generation grooves  3 A and the thrust plate  4 . Hence, the shaft  2  is floated in the upward direction of the figure and rotates without making contact with the thrust plate  4  and the sleeve  1 . 
     FIG. 2  is a view of the lower face of the flange  3 , that is, the bottom face thereof opposing to the thrust plate  4 , and the black-colored portions indicate the dynamic pressure generation grooves  3 A. The outside diameter of the pattern of the dynamic pressure generation groove  3 A is designated as d 1   o , the inside diameter thereof is designated as d 1   i  and the diameter of the turn-back part is designated as d 1   m . When the flange  3  rotates inside the recess portion  1 C of the sleeve  1 , an oil pressure G is generated on the face of the flange  3  in the direction from the outer circumference to the inner circumference thereof. Furthermore, an oil pressure H is also generated in the direction from the inner circumference to the outer circumference thereof. The diameter of the turn-back part wherein the pressure G becomes equal to the pressure H is represented by dsy. Usually, the dynamic pressure generation grooves  3 A are designed so that the pressure G becomes equal to the pressure H. For this purpose, the diameter d 1   m  is determined by equation (1), a well-known equation in hydrodynamics.
   d 1 m={ ( d 1 i   2   +d 1 o   2 )/2} 1/2   (1) 
   However, the hydrodynamic bearing in accordance with the present invention is designed so that the pressure G becomes larger than the pressure H. In other words, when the diameter d 1   m  has a value in the range of 1 mm or more and 10 mm or less and the relationship represented by equation (2) is established, the value of the diameter d 1   m  is set so that the value obtained by subtracting the diameter d 1   m  from the diameter dsy is in the range of 0.05 or more and 0.8 mm or less as represented by equation (3).
 
 dsy={ ( d 1 i   2   +d 1 o   2 )/2} 1/2   (2)
 
 dsy−d 1 m= 0.05 to 0.8 mm  (3)
 
     FIG. 3  is a plan view of the flange  3 , and the black-colored portions indicate the dynamic pressure generation grooves  3 B. The dynamic pressure generation grooves  3 B are designed so that the pressure in the direction indicated by arrow E from the inner circumference to the outer circumference is nearly balanced with the pressure in the direction indicated by arrow F from the outer circumference to the inner circumference. In other words, when the outside diameter of the pattern of the dynamic pressure generation groove  3 B is designated as d 2   o , the inside diameter thereof is designated as d 2   i  and the diameter of the turn-back part thereof is designated as d 2   m , a relationship represented by equation (4) is established.
   d 2 m={ ( d 2 o   2   +d 2 i   2 )/2} 1/2   (4) 
   The vertical axis of the graph in  FIG. 4  represents the oil pressure (pascal) in the dynamic pressure generation groove  3 A, which is variable depending on the value of the diameter difference (dsy−d 1   m ). If asymmetry is insufficient in the pressures inside the bearing, a partially negative pressure portion is generated somewhere inside the bearing, and air may be accumulated there. On the other hand, if asymmetry is excessive, the internal pressure becomes too high, and there arises a danger of causing cavitation or microbubbles. Relating to the hydrodynamic bearing in accordance with this embodiment, a hydrodynamic bearing is made by using transparent materials for the sake of observation, and experiments are carried out. As a result, it was found that when the value of the above-mentioned dsy−d 1   m  was in the range of 0.05 or more to 0.99 or less, the amount of air bubbles entered and the amount of air coagulated during rotation were minimal, whereby this range was an appropriate range and air is least likely to be accumulated in oil. 
     FIG. 5  is a cross-sectional view showing the cross-section of a relevant part and the pressure distribution of oil by the dynamic pressure generation grooves  3 A and  3 B with reference to the atmospheric pressure in the case that the floating amount (S 1 ) of the flange  3  from the thrust plate  4  is sufficiently small. In the hydrodynamic bearing in accordance with the present invention, only the positive pressure indicated by curve P 10  representing the pressure distribution of oil is generated and no negative pressure is generated. For this reason, a phenomenon of air accumulation between the flange  3  and the thrust plate  4  hardly occurs. 
     FIG. 6  is a cross-sectional view showing the cross-section of a relevant part and the pressure distribution of oil by the dynamic pressure generation grooves  3 A and  3 B as indicated by pressure curves P 11  and P 12  in the case that the floating amount (S 2 ) is sufficiently large. Even in this case, no negative pressure is generated inside the bearing as indicated by the pressure curve P 11 . In  FIG. 6 , the positive pressure indicated by the curve P 12  of the pressure generated by the dynamic pressure generation grooves  3 B on the upper face of the flange  3  prevents collision between the flange  3  and the sleeve  1 . 
     FIG. 7  and  FIG. 8 , views showing the cross-sections of a relevant part and the pressure distributions, show detailed pressure distributions regarding the pressures generated in the radial direction (the left-to-right direction in the figure) of the dynamic pressure generation grooves  1 A and  1 B.  FIG. 7  shows a case wherein the clearance portions of the hydrodynamic bearing are wholly filled with oil  5  and the liquid face is above the upper ends of the dynamic pressure generation grooves  1 A. The dynamic pressure generation grooves  1 A are provided in the upper portion of the sleeve  1  and made asymmetric such that the groove portion  28 A in the range of the upper half dimension L 1  is longer than the groove portion  29 A in the range of the lower half dimension L 2 . Hence, the oil is pressed downward by the effect of dynamic pressure, thereby being prevented from leaking outside. The acute connection part of the groove portion  28 A and the groove portion  29 A is referred to as a turn-back part. The groove portion  28 A and the groove portion  29 A of the dynamic pressure generation groove  1 A have the same inclination angle. In the configuration shown in  FIG. 7 , if the difference between the dimension L 1  and the dimension L 2  of the dynamic pressure generation groove  1 A is too small, there is a danger of causing oil leakage. On the other hand, if the difference is too large, the internal pressure becomes too high, and there is a danger of generating cavitation or microbubbles. 
   In the dynamic pressure generation groove  1 B, the groove portion  28 B of the upper half is made symmetric with the groove portion  29 B of the lower half. Since the dynamic pressure generation groove  1 A is made asymmetric, the pressure inside the bearing becomes positive as indicated by pressure curve P 13 . Since no negative pressure is generated inside the bearing even in this case, air accumulation hardly occurs. The pressures in the thrust direction become positive as indicated by pressure curves P 14  and P 15 , whereby no negative pressure is generated. 
     FIG. 8  shows a case wherein the oil inside the bearing decreases and becomes insufficient by the amount corresponding to the dimension ΔL. Even in this case, only the positive pressure is generated as indicated by pressure curve P 17 , whereby no negative pressure is generated inside the bearing. 
     FIG. 9  shows the appropriate range of the asymmetry of the dynamic pressure generation groove  1 A. It is desirable that the dimension L 2  of the groove portion  29 A is smaller than the dimension L 1  of the groove portion  28 A, that is, the portion on the opposite side, and that the value of the relational expression shown on the left side of equation (5), wherein the difference between the dimensions L 1  and L 2 , (L 1 −L 2 ), is set in the range represented by equation (5), is in the value range shown on the right side when the diameter of the shaft  2  is in the range of 1 mm or more and 10 mm or less.
 ( L 1 −L 2)=0.05 mm to 1.5 mm  (5) 
   As the results of various experiments, in the range shown in equation 5, the entry of air and the entry of microbubbles hardly occurred. 
     FIG. 10  shows the relationship between the kinematic viscosity of oil or the kinematic viscosity of the base oil of grease and the bubble mixing rate into the clearances of the bearing, obtained from the observation results of the experimental bearing made of the transparent materials. The bubble mixing rate is represented by the percentage of the volume of bubbles with respect to the volume of oil. According to the observation results, it was found that the bubble mixing rate was very low in the case when oil or the base oil of grease had a kinematic viscosity of 4 cSt or more at 40° C. of temperature. 
   The configuration and operation of a disc rotation apparatus using the hydrodynamic bearing in accordance with the present invention will be described by using  FIG. 11 . In  FIG. 11 , on a hydrodynamic bearing provided inside a box-shaped base  6  and comprising a sleeve  1 , a shaft  2 , a flange  3 , a thrust plate  4 , a hub rotor  7 , a stator  8  and a rotor magnet  9 , two discs  10  are installed while space is provided therebetween by using a spacer  12 . Heads  25  respectively supported by arms  15  are opposed to both faces of the disc  10 . The arms  15  rotate while being supported by a head support shaft  16 . The upper face of the base  6  is hermetically sealed by an upper lid  14  so as to prevent the entry of dust and the like. When electric power is applied to the motor stator  8 , a rotating magnet field is generated, and the rotor magnet  9  starts rotating together with the hub rotor  7 , the shaft  2  and the discs  10 . The dynamic pressure generation grooves  1 A,  1 B,  3 A and  3 B rake up oil by pumping forces and generate pressures, whereby the bearing portion floats and rotates with high accuracy in a noncontact state. The heads  25  make contact with the rotating discs  10 , thereby recording or reproducing electrical signals. 
   Although the thrust plate is secured to the sleeve  1  in  FIG. 1 , it may be secured to the base  6  if the interior of the bearing can be hermetically sealed. 
   Even if helical dynamic pressure generation grooves, in which d 1   m =d 1   o , are used as a modification application example of the dynamic pressure generation grooves  3 A shown in  FIG. 2 , instead of the herringbone-shaped grooves, nearly equivalent performance can be obtained. 
   As mentioned above, with the hydrodynamic bearing in accordance with this embodiment, the entry of air into the hydrodynamic bearing section is prevented, and the breakage of oil film, having been apt to occur in bearings, is prevented. As a result, a long-life disc rotation apparatus capable of rotating discs with high accuracy is obtained by using the hydrodynamic bearing in accordance with the present invention. 
   In addition, the design conditions of the dynamic pressure generation grooves are combined with the selection conditions of the kinematic viscosity of oil so that the accumulation of air inside the bearing due to the pumping forces in the dynamic pressure generation grooves is prevented during rotation, therefore the breaking of oil film in the clearances of the bearing does not occur, whereby the hydrodynamic bearing in accordance with the present invention has high accuracy and long life.