Patent Publication Number: US-2016238073-A1

Title: Deep groove ball bearing

Description:
TECHNICAL FIELD 
     The present invention relates to a deep groove ball bearing including balls disposed between its outer race and inner race. 
     BACKGROUND ART 
     In a transmission of the type including a plurality of gear type speed reducing portions having different reduction ratios and disposed between an input shaft and an output shaft arranged coaxial with each other and a countershaft extending parallel to the input and output shafts, and capable of reducing the rotational speed of the input shaft in a stepwise manner and transmitting the thus reduced rotation to the output shaft, since helical gears are ordinarily used in the gear type speed reducing portions, thrust forces are applied to the input shaft, the output shaft and the countershaft while torque is being transmitted from the input shaft to the output shaft. 
     It is therefore necessary to use bearings capable of supporting both radial loads and thrust loads, to support the input shaft, the output shaft and the countershaft. 
     Since tapered roller bearings are large in load capacity and capable of supporting both radial loads and thrust loads, tapered roller bearings are suitable for use in transmissions. However, since tapered roller bearings tend to increase fuel consumption of the vehicle due to large torque loss, deep groove ball bearings, of which the torque loss is small, are now used in an increasingly large number of transmissions. 
     Deep groove ball bearings include balls disposed between the inner and outer races and retained by a retainer. Some of such retainers are made of metal, while others are made of synthetic resin. 
     The retainers used in conventional deep groove ball bearings are all capable of merely retaining the balls, so that deep groove ball bearings are inferior in oil circulating ability (lubricating ability) to tapered roller bearings. Deep groove ball bearings are therefore more likely to heat up under high-load and high-speed operating conditions such as when used in transmissions, so that deep groove ball bearings are less durable. 
     In order to solve this problem, the applicant of the present invention proposed, in the below-identified Patent document 1, a deep groove ball bearing including a retainer constituted by two retainer portions made of synthetic resin and having different outer diameters. The smaller-diameter retainer portion is fitted in the larger-diameter portion such that the retainer forms two cylindrical surfaces having different diameters on the outer periphery thereof. With this arrangement, due to the difference in peripheral speed between the two cylindrical surfaces resulting from the difference in diameter therebetween, the pump effect is generated which forcibly pulls lubricating oil into the bearing, thus more efficiently lubricating the interior of the bearing. 
     PRIOR ART DOCUMENT(S) 
     Patent Document(s) 
     
         
         Patent document 1: JP Patent Publication 2011-21661A 
       
    
     SUMMARY OF THE INVENTION 
     Object of the Invention 
     With the deep groove ball bearing disclosed in Patent document 1, since lubricating oil in the bearing is forcibly moved by the pump effect generated while the bearing is rotating, lubricating oil may excessively flow out of the bearing, causing insufficient lubrication. 
     An object of the present invention is to provide a deep groove ball bearing which can prevent an excessive amount of lubricating oil from flowing out of the bearing, thereby minimizing shortening of the life of the bearing due to insufficient lubrication. 
     In order to achieve this object, the present invention provides a deep groove ball bearing comprising an outer race formed with a raceway groove on a radially inner surface of the outer race, an inner race formed with a raceway groove on a radially outer surface of the inner race, balls disposed between the raceway groove of the outer race and the raceway groove of the inner race, and a retainer retaining the balls, wherein the retainer comprises a first cylindrical split retainer member, and a second cylindrical split retainer member inserted inside of the first split retainer member, wherein the first split retainer member includes a first axial portion having a first axial side surface formed with first cutouts circumferentially spaced apart from each other, and the second split retainer member includes a second axial portion having a second axial side surface axially opposed to the first axial side surface and formed with second cutouts circumferentially spaced apart from each other, wherein the second axial portion is fitted in the first axial portion such that each of the first cutouts and a corresponding one of the second cutouts define a circular pocket in which is retained each of the balls, wherein the first split retainer member includes a second axial portion opposite from the first axial side surface of the first split retainer member and protruding axially from the second axial side surface of the second split retainer member, wherein the second split retainer member includes a first axial portion opposite from the second axial side surface of the second split retainer member and protruding axially from the first axial side surface of the first split retainer member, wherein the deep groove ball bearing further comprises a coupling means axially inseparably coupling together the first and second split retainer members, and wherein a flange is provided at a second axial side of the first split retainer member opposite from the first axial side surface of the first split retainer member. 
     With this deep groove ball bearing, when the outer race and the inner race rotate relative to each other, the balls revolve around the axis of the bearing while rotating about their respective centers, so that the retainer, which is constituted by the first split retainer member and the second split retainer member, rotates. 
     The outer periphery of the retainer is constituted by the radially outer surface of the first split retainer member and the radially outer surface of the second split retainer member. Since the two radially outer surfaces have different diameters from each other, the pump effect is generated in the bearing due to the difference in peripheral speed between the two radially outer surface resulting from the difference in diameter therebetween, whereby lubricating oil is pulled into the bearing, flows in the bearing, and flows out of the bearing from the second axial side of the first split retainer member. 
     At this time, the flange provided at the second axial side of the first split retainer member interfere with lubricating oil flowing out of the bearing, thus preventing an excessive amount of lubricating oil from flowing out of the bearing. 
     For reduced weight and cost of the deep groove ball bearing, the first split retainer member and the second split retainer member of the ball bearing are preferably made of synthetic resin. 
     Since the deep groove ball bearing is lubricated by lubricating oil, the first split retainer member and the second split retainer member are especially preferably made of oil-resistant synthetic resin. Such oil-resistant synthetic resins include polyamide 46 (PA46), polyamide 66 (PA66), polyamide 9T (PA9T), poly(ether ether ketone) resin (PEEK), and polyphenyl sulfide resin (PPS). 
     Advantages of the Invention 
     According to the present invention, the flange provided at the second axial side of the first split retainer member prevents an excessive amount of lubricating oil pulled into the bearing due to the pump effect generated while the retainer is rotating, from flowing out of the bearing. This prevents shortening of the life of the bearing due to shortage of lubricating oil in the bearing. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a vertical sectional view of a deep groove ball bearing according to the present invention. 
         FIG. 2  is a partially cutaway right-hand side view of a portion of a retainer shown in  FIG. 1 . 
         FIG. 3  is a left-hand side view of a portion of the retainer shown in  FIG. 1 . 
         FIG. 4  is an enlarged sectional view of a first split retainer member and a second split retainer member shown in  FIG. 1 , showing the connecting portion therebetween. 
         FIG. 5  is a plan view of portions of the first split retainer members and the second split retainer member. 
         FIG. 6(I)  is a plan view of the first split retainer member shown in  FIG. 5 , showing circumferential pocket gap between a pocket of the first split retainer member and a ball received in the pocket; and  FIG. 6 (II) is a plan view of the first split retainer member shown in  FIG. 5 , showing an axial pocket gap between a pocket of the first split retainer member and a ball received in the pocket. 
         FIG. 7  is a plan view of a portion of a different retainer. 
         FIG. 8  is a plan view of a first split retainer member shown in  FIG. 7 , showing a circumferential pocket gap between a pocket of the first split retainer member and a ball received in the pocket. 
     
    
    
     MODE FOR EMBODYING THE INVENTION 
     The embodiment of the present invention is now described with reference to the drawings. As shown in  FIG. 1 , the deep groove ball bearing A of the embodiment includes balls  30  disposed between a raceway groove  11  formed in the radially inner surface of an outer race  10  and a raceway groove  21  formed in the radially outer surface of the inner race  20 , and retained by a retainer  40 . 
     The outer race  10  has a pair of shoulders  12   a  and  12   b  which are formed, respectively, on one and the other sides of the raceway groove  11 , of which the shoulder  12   a  is higher than the corresponding shoulder of the outer race of a standard deep groove ball bearing, and the shoulder  12   b  is of the same height as the corresponding shoulder of the outer race of a standard deep groove ball bearing. The inner race  20  has a pair of shoulders  22   a  and  22   b  which are formed, respectively, on the one and the other sides of the raceway groove  11 , of which the shoulder  22   b  is higher than the corresponding shoulder of the inner race of a standard deep groove ball bearing, and the shoulder  22   a  is of the same height as the corresponding shoulder of the inner race of a standard deep groove ball bearing. However, the outer race  10  and the inner race  20  may be of the same structure as the outer race and the inner race of a standard deep groove ball bearing. Also, the shoulder  12   b , located on the other side of the outer race  10 , and the shoulder  22   a , located on the one side of the inner race  20 , may be lower in height than the corresponding shoulders of a standard deep groove ball bearing. A standard deep groove ball bearing refers to a bearing of which the pair of shoulders of the outer race and the pair of shoulders of the inner race are all of the same height. 
     As shown in  FIGS. 1 to 3 , the retainer  40  comprises a first split retainer member  41 , and a second split retainer member  42  inserted in the first split retainer member  41 . 
     As shown in  FIG. 5 , the first split retainer member  41  comprises an annular member  43  made of a synthetic resin and formed on a first axial side surface thereof with opposed pairs of pocket claws  44  that are circumferentially equidistantly spaced apart from each other. The annular member  43  is further formed with first cutouts  45  each between one opposed pair of pocket claws  44  and having, in plan view, the shape of a partial circle larger than a semicircle. The annular member  43  further includes a flange  60  provided on the second axial side thereof opposite from the first axial side surface thereof to extend radially inwardly and radially outwardly. 
     As shown in  FIG. 1 , the annular member  43  has an inner diameter substantially equal to the pitch circle diameter (PCD) of the balls  30 , and has an outer diameter within the range between the inner diameter of the shoulder  12   a , i.e. the higher of the two shoulders  12   a  and  12   b , of the outer race  10 , and the inner diameter of the lower shoulder  12   b  of the outer race  10 , whereby the first split retainer member  41  can be inserted into the bearing from its side where there is the lower shoulder  12   b . The cutouts  45  have spherical inner surfaces extending along the outer peripheries of the respective balls  30 . 
     The second split retainer member  42  comprises an annular member  48  made of a synthetic resin and formed, on a second axial side surface thereof which is opposed to the first axial side surface of the annular member  43 , with opposed pairs of pocket claws  49  that are circumferentially equidistantly spaced apart from each other. The annular member  48  is further formed with second cutouts  50  each between one opposed pair of pocket claws  49  and having, in plan view, the shape of a partial circle larger than a semicircle. 
     As shown in  FIG. 1 , the annular member  48  has an outer diameter substantially equal to the pitch circle diameter (PCD) of the balls  30 , and has an inner diameter within the range between the inner diameter of the shoulder  22   b , i.e. the higher of the two shoulders  22   a  and  22   b , of the inner race  20 , and the inner diameter of the lower shoulder  22   a  of the inner race  20 , whereby the second split retainer member  42  can be inserted into the bearing from its side where there is the lower shoulder  22   a , and fitted into the first split retainer member  41 . The cutouts  50  have spherical inner surfaces extending along the outer peripheries of the respective balls  30 . 
     As shown in  FIGS. 4 and 5 , between the first split retainer member  41  and the second split retainer member  42 , there are provided coupling means X which axially inseparably hold together the first split retainer member  41  and the second split retainer member  42  when a second axial portion of the second split retainer member  42  which includes the above-mentioned second axial side surface is inserted into a first axial portion of the first split retainer member  41  which includes the above-mentioned first axial side surface until a circular pocket is defined by each opposed pair of cutouts  45  and  50 . 
     Each of the coupling means X includes an inwardly extending engaging claw  46  formed on the distal end of a crossbar  43   a  defined between an adjacent pair of the cutouts  45  of the first split retainer member  41 , a groove-shaped engaging recess  47  formed in the radially inner surface of the annular member  43  so as to be axially aligned with the engaging claw  46 , an outwardly extending engaging claw  51  formed on the distal end of a crossbar  48   a  defined between the adjacent pair of the cutouts  50  corresponding to the adjacent pair of the cutouts  45 , and an engaging recess  52  formed in the radially outer surface of the annular member  48  so as to be axially aligned with the engaging claw  51 . By engaging the engaging claws  46  of the first split retainer member  41  into the respective engaging recesses  52  of the second split retainer member  42 , and engaging the engaging claws  51  of the second split retainer member  42  into the respective engaging recesses  47  of the first split retainer member  41 , the first split retainer member  41  and the second split retainer member  42  are axially inseparably held together. 
     Since the first split retainer member  41  and the second split retainer member  42  are brought into contact with lubricating oil for lubricating the deep groove ball bearing, they are made of an oil-resistant synthetic resin. Such synthetic resins include polyamide 46 (PA46), polyamide 66 (PA66), polyamide 9T (PA9T), poly(ether ether ketone) (PEEK), and polyphenylene sulfide (PPS). A suitable one of them should be selected according to the type of the lubricating oil used. 
     In assembling the deep groove ball bearing of the embodiment, the inner race  20  is inserted into the outer race  10  and a required number of the balls  30  are mounted in position between the raceway groove  21  of the inner race  20  and the raceway groove  11  of the outer race  10 . 
     In order to mount the balls  30  in position, the inner race  20  is radially moved relative to the outer race  10  until the radially outer surface of the inner race  20  is partially brought into abutment with the radially inner surface of the outer race  10 , and a crescent moon-shaped gap is defined between the inner and outer races  20  and  10  at a position diametrically opposite from the portion where the inner and outer races  20  and  10  are in abutment with each other, and then the balls  30  are fitted into the bearing through one side of the crescent moon-shaped gap. 
     When fitting the balls  30 , if the shoulder  12   a  of the outer race  10 , to which a thrust force is applied, or the higher of the shoulders of the inner race  20 , i.e. the shoulder  22   b  is higher in height H 1  than necessary, the shoulder  12   a  or  22   b  may interfere with the balls  30  when fitting the balls  30 . Thus, in the embodiment, the ratio of either of the heights H 1  to the diameter of each ball  30 , i.e. the ratio H 1 /d is determined at less than 0.50 so that the balls  30  can be reliably fitted into the bearing. In order to prevent a ball  30  from landing onto the shoulder  12   a  or  22   b , the ratio of either of the heights H 1  to the diameter of each ball  30 , namely, the ratio H 1 /d is determined at 0.25 or over. 
     After fitting the balls  30 , with the inner and outer races  20  and  10  positioned such that their center axes coincide with each other, and with the balls  30  arranged so as to be circumferentially equidistantly spaced from each other, the first split retainer member  41  is inserted between the inner and outer races  20  and  10  from the side where there is the lower shoulder  12   b  of the outer race  10 , until the balls  30  are fitted into the respective cutouts  45  of the first split retainer member  41 . 
     Next, the second split retainer member  42  is inserted between the inner and outer races  20  and  10  from the side where there is the lower shoulder  22   a  of the inner race  20 , until the above-mentioned second axial portion of the second split retainer member  42  is fitted into the first axial portion of the first split retainer member  41 , with the balls  30  fitted into the respective cutouts  50  of the second split retainer member  42 . 
     By fitting the second split retainer member  42  into the first split retainer member  41 , the deep groove ball bearing is assembled, with the engaging claws  46  and  51  of the split retainer members  41  and  42  engaged in the engaging recesses  52  and  47  of the split retainer members  42  and  41 , respectively, as shown in  FIGS. 1 and 4 . 
     After fitting the balls  30  between the raceway groove  11  of the outer race  10  and the raceway groove  21  of the inner race  20 , by, as described above, simply inserting the first and second split retainer members  41  and  42  between the outer race  10  and the inner race  20  from the respective sides until the second split retainer  42  is fitted into the first split retainer member  41 , the deep groove ball bearing A is assembled. 
     If, for example, this deep groove ball bearing A is used to support a torque transmission shaft, such as the input shaft of a transmission, when the torque transmission shaft rotates, and the inner and outer races  20  and  10  rotate relative to each other, the balls  30  are rotated about their own axes, while revolving around the axis of the bearing. Due to the revolution of the balls  30 , the retainer  40  is rotated, so that if lubricating oil is present between the inner and outer races  20  and  10 , the lubricating oil is rotated due to contact with the retainer  40 . 
     Since the first split retainer member  41  and the second split retainer member  42  have different outer diameters, the lubricating oil portion that is being rotated due to contact with the first split retainer member  41  flows faster than the lubricating oil portion being rotated due to contact with the second split retainer member  42 . This causes the lubricating oil portion flowing at a lower speed to be pulled toward the lubricating oil portion flowing at a higher speed, thus generating a pump effect. Due to the pump effect, lubricating oil is sucked into the bearing, flows through the bearing, and is discharged from the bearing through the other axial side of the first split retainer member  41  as shown by arrows in  FIG. 1 . 
     However, the flange  60  at the other axial side of the first split retainer member  41  interferes with lubricating oil leaving the bearing, thus preventing leakage of excessive amount of lubricating oil, which in turn prevents shortening of the life of the bearing due to shortage of lubricating oil in the bearing. 
     With the deep groove ball bearing of the embodiment, a pair of the pocket claws  44  and a pair of the pocket claws  49  are provided, respectively, at the open end of each cutout  45  of the first split retainer member  41  and at the open end of the corresponding cutout  50  of the second split retainer member  42 , so as to embrace the ball  30  received in the two cutouts  45  and  50  such that the pair of the pocket claws  44  and the pair of the pocket claws  49  extend in opposite directions to each other when they are combined. With the claws  44  combined with the claws  49 , the engaging claws  46  and the  51  are engaged in the respective engaging recesses  47  and  52 , thereby axially inseparably coupling together the first split retainer member  41  and the second split retainer member  42 . This positively prevents separation of the retainer  40  even if the some balls  30  are moved faster or slower than the others under a large moment load. 
     As shown in  FIGS. 5 and 6 (I), the sizes δ 1  of the circumferential gaps  70  defined between the engaging claws  46 ,  51  and the corresponding engaging recesses  47 ,  52  are preferably larger than the sizes δ 2  of the circumferential pocket gaps  71  defined between the balls  30  and the corresponding cutouts  45 ,  50 . With this arrangement, even if some balls  30  are moved faster or slower than the others under a large moment load, and as a result, the first split retainer member  41  and the second split retainer member  42  are rotated relative to each other, none of the engaging claws  46  and  51  will abut one of the circumferentially opposed side walls of the corresponding engaging recess  47 ,  52 . This effectively prevents damage to the engaging claws  46  and  51 . 
     As shown in  FIGS. 4 and 6 (II), the sizes δ 3  of the axial gaps  72  defined between the engaging claws  46  and  51  are preferably larger than the sizes δ 4  of the axial pocket gaps defined between the balls  30  and the corresponding cutouts  45 ,  50 . With this arrangement, while the inner surfaces of the opposed pairs of pocket claws  44  and  49  may abut the outer peripheries of the balls  30  under an axial force that tends to move the first split retainer member  41  and the second split retainer member  42  away from each other, the engaging claws  46  and  51  are positively prevented from coming into abutment with the axial end surfaces of the engaging recesses  47 ,  52 . This effectively prevents damage to the engaging claws  46  and  51 . 
     In  FIGS. 1 and 5 , the cutouts  45  and  50  are formed by cutting out the annular members  43  and  48 , and each, in plan view, in the shape of a partial circle larger than a semicircle. The cutouts  45  and  50  are however not limited thereto, and may be each, in plan view, in the shape of the letter U so that each opposed pair of the cutouts  45  and  50  form a cylindrical pocket with the first split retainer member  41  and the second split retainer member  42  fitted together. 
     If the cutouts  45  and  50  shown in  FIG. 7  are used, as shown in  FIGS. 7 and 8 , the sizes δ 5  of the circumferential gaps  74  defined between the engaging claws  46 ,  51  and the corresponding engaging recesses  47 ,  52  are preferably larger than the sizes δ 6  of the circumferential pocket gaps  75  defined between the balls  30  and the corresponding cutouts  45 ,  50 . With this arrangement, even if some balls  30  are moved faster or slower than the others under a large moment load, and as a result, the first split retainer member  41  and the second split retainer member  42  are rotated relative to each other, none of the engaging claws  46  and  51  will abut one of the circumferentially opposed side walls of the corresponding engaging recess  47 ,  52 . This effectively prevents damage to the engaging claws  46  and  51 . 
     DESCRIPTION OF THE NUMERALS 
     
         
           10 . Outer race 
           11 . Raceway groove 
           20 . Inner race 
           21 . Raceway groove 
           30 . Ball 
           40 . Retainer 
           41 . First split retainer member 
           42 . Second split retainer member 
           45 . Cutout 
           50 . Cutout 
           60 . Flange 
         X. Coupling means