Abstract:
A crank shaft thrust bearing half includes a semi-cylindrical bearing shell carrying a pair of radially outwardly projecting thrust flanges having outer thrust surfaces. Radial grooves are formed in the thrust face to divide the surface into thrust pads. Each thrust pad has hydrodynamic contours which provide a hydrodynamic wedging action during rotation of the shaft to support high thrust loads. The thrust pad at the trailing end of the flange has an elevated land area spaced about 30° away from the trailing end and includes a long relief area from the land to the end to provide a no contact zone of the flange. Locating the land away from the trailing end and providing the relief zone shifts the peak pressure location away from the end and toward the middle of the bearing to relieve the trailing end from stress. Having radially oriented oil grooves with suitable width enables adequate oil supply and improved hydrodynamic oil film generating capability.

Description:
BACKGROUND OF THE INVENTION 
     1. Technical Field 
     This invention relates to crank shaft bearings of the type having side flanges for accommodating high thrust loads. 
     2. Related Art 
     It is common practice to support internal combustion engine crank shafts with journal bearings located at spaced points along the crank shaft. At least one of the bearings is designed to support axial thrust forces applied through the crank shaft. Such thrust bearings are made in bearing halves, each half bearing including a semi-cylindrical bearing shell having a concave inner running surface and a pair of axially spaced flanges projecting radially outwardly of the shell and having axially outwardly facing thrust bearing surfaces. The thrust surfaces seat against associated side surfaces of the crank shaft arms, such that the shaft is prevented from shifting in the axial direction during operation. 
     In some applications, there is considerable axial loading and measures must be taken to protect the thrust surfaces of the bearing. U.S. Pat. No. 5,192,136 discloses such a bearing, wherein the thrust face is formed with a plurality of oil supply grooves that subdivide the thrust bearing surface into a plurality of thrust pads. Each thrust pad is contoured to generate a protective hydrodynamic wedge film thrust support action, in order to separate the two opposing surfaces and to prevent metal-to-metal contact between the thrust face and under axial loading. The thrust bearing surface profile shown generally in FIG. 5 is in use. Such a prior art bearing is also illustrated in FIG. 7 of the present application. It has been found that in some installations, thrust bearings of the above type tend to distort under heavy clamping load, as illustrated schematically in FIG. 8, and the degree of distortion is constantly changing with changes in operating temperature and axial load conditions on the bearing. Studies of the load profile across the prior thrust bearing surface are illustrated in FIG. 9 of the present application. The profile shows that high loading is carried at points along the thrust face including a location of an extremely high pressure peak at the trailing end of the half bearing in relation to the direction of rotation of the crank shaft across the bearing surface. Over time, the heavy loading and cyclic distortion changes that occur can fatigue the trailing end of the thrust face where the flange meets the bearing shell. In severe cases, such fatigue can lead to delamination of the bearing layer material and/or cracking of the thrust flange at the trailing end where it meets the bearing shell. 
     It is an object of the present invention to overcome the shortcomings of prior hydrodynamic thrust bearings. 
     SUMMARY OF THE INVENTION AND ADVANTAGES 
     A half bearing for a rotary shaft includes a bearing shell having a concave running surface. At least one thrust flange extends radially from the shell and has an axially facing thrust bearing surface extending circumferentially in a sliding direction of the thrust bearing surface corresponding to the direction of rotation of the rotary shaft. The thrust bearing surface has a leading end and a trailing end relative to the sliding direction of the bearing surface. The thrust bearing surface includes a plurality of circumferentially spaced oil supply grooves which extend radially of the thrust bearing surface and subdivide the surface into a plurality of thrust bearing pads. The thrust bearing pads includes a leading pad disposed between the leading end and leading oil supply groove. The thrust pads include a trailing pad disposed between the trailing end and groove of the thrust bearing surface. Each thrust bearing pad has a hydrodynamic contour which includes an inclined ramp surface extending in the sliding direction, and a raised land surface following the inclined ramp surface in the sliding direction. According to the invention, the spacing between the trailing groove and trailing end is greater than the distance between the leading groove and leading end. The land surface of the trailing thrust bearing pad is spaced from the trailing end of the thrust bearing surface by distance greater than a distance between the land surface of the leading thrust bearing pad and the leading groove. 
     This construction of the thrust face according to the invention has the advantage of controlling the distribution of the hydrodynamic load imparted by the rotating crankshaft on the bearing. More specifically, spacing the land surface of the trailing thrust pad further from the trailing end has the advantage of shifting the peak high pressure location, which occurs at the land, away from the trailing end of the thrust flange, thereby isolating the otherwise problematic corner region of the trailing end which is prone to fatigue failure as described in the previous section. In other words, adjusting the relative location of the oil supply grooves and high pressure land regions enables a corresponding shifting of the peak pressure location away from the trailing end of the thrust bearing, overcoming the fatigue and delamination failure problems associated with the prior bearings discusses in the previous section. The oil groove design also enables adequate oil supply and better hydrodynamic film generation. 
     The invention has the further advantage of utilizing standard manufacturing techniques, coupled with controlled modification of the process, to yield a more reliable, high performance thrust bearing. Tests conducted on flange bearings constructed according to the present invention reveal that the subject thrust bearings are able to sustain a significantly higher average maximum load in use under severe testing conditions without experiencing high levels of localized torque, wear or fatigue failure associated with the prior hydrodynamic thrust bearing constructions operating under the same or comparable conditions. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     These and other features and advantages of the present invention will become more readily appreciated when considered in connection with the following detailed description and appended drawings, wherein: 
     FIG. 1 is an end view of an upper and lower bearing assembly constructed according to the invention; 
     FIG. 2 is a cross-sectional view taken generally along lines  2 — 2  of FIG. 1; 
     FIG. 3 is a enlarged end view of one of the half bearings of the assembly of FIG. 1 illustrating further features; 
     FIG. 4 is an enlarged cross-sectional taken along lines  4 — 4  of FIG. 3; 
     FIG. 5 a  is an end view of the bearing assembly of FIG. 1, but of the opposite side; 
     FIG. 5 b  is an enlarged cross-sectional view taken along lines  5   b — 5   b of FIG. 5 a;    
     FIG. 6 is another embodiment of a thrust bearing assembly according to the invention; 
     FIGS. 7 and 8 are perspective and top elevation views, respectively, of a prior art half bearing; 
     FIG. 9 is a schematic representation of the prior art surface profile and resultant hydrodynamic pressure profile across the surface in operation; and 
     FIG. 10 is a comparative surface and hydrodynamic pressure profile of a bearing constructed according to the invention. 
    
    
     DETAILED DESCRIPTION 
     FIGS. 1 and 2 show a bearing assembly  11  for rotatably supporting an engine crank shaft  13 . The crank shaft  13  is supported for rotation around axis  15  in a predetermined direction of rotation R. 
     The bearing assembly  11  includes an upper half bearing  17  and a lower half bearing  19 . The bearing halves  17 , 19  are collectively mounted as a unit on the lower end of an engine cylinder block by a cap structure  21 . A pair of bolts extend through ears on the cap structure  21  into threaded bores in a cylinder block wall  23  to rigidly attach the bearing assembly  11  to the cylinder block in a known fashion. 
     The lower face of wall  23  has a semi-circular cavity or recess which mates with a semi-cylindrical wall or shell  25  of the upper half bearing  17 , whereas the upper face of cap structure  21  has a semi-circular cavity or recess which mates with a semi-cylindrical wall or shell  27  of the lower half bearing  19 . When the cap structure is bolted to the cylinder block, the two half bearings  17 , 19  are rigidly clamped to the cylinder block. Such mounting arrangement for the bearing assembly  11  is conventional. 
     The upper and lower bearing halves  17 , 19  are preferably identical in structure. For purposes of simplicity, details concerning the construction and features of the bearing halves  17 , 19  will be directed to the lower bearing half  19 , it being understood that the same description is applicable to the upper half bearing  17 . 
     The half bearing  19  has at least one and preferably two arcuate flange walls  29 , 31  extending radially outwardly and perpendicularly from an outer convex surface  33  of the bearing shell  27 . The flange wall  29  has an external axially outwardly facing thrust bearing surface or face  35  and an axially inwardly facing surface  37 . Thrust flange  31  likewise has an axially outer facing thrust bearing surface or face  39  and an axially inner facing surface  41 . The inner surfaces  37 , 41  are substantially parallel. The outwardly facing surface  35 , 39  are also substantially parallel, apart from the hydrodynamic features to be described below. In the first illustrated embodiment, the side flanges  29 , 31  are formed as one integral piece with the bearing shell  27 . The bearing half  19  is formed into the generally conventional half thrust bearing configuration in part by conventional practice wherein the bearing half material is blanked from a strip which preferably is of a bi-metal construction wherein a relatively soft low friction sliding or bearing material, e.g., preferably but not limited to an aluminum-lead alloy is applied to a rigid steel backing. In the drawings, the bearing liner material is designated by numeral  43 , and the steel backing layer is designated by numeral  45 . The blanked strip is formed in one or more operations to provide the general shape of the bearing half  19 , including the semi-cylindrical shell  27  and arcuate flanges  29 , 31 . 
     Pressurized lubricating oil is supplied to the bearing assembly through a hole or port  47  in the cylinder block wall  23 . A mating slot or hole  49  in the shell  25  of the upper bearing half  17  delivers the oil to a groove  51  formed in the concave running surface  53  of the shell  25 . As the crank shaft  13  rotates around the shaft axis  15  in the direction of arrow R, the oil is carried by the rotating shaft surface onto the concave running surface  55  of the shell  27  of the lower bearing half  19 . In this fashion, a ring of oil encircles the shaft surface to provide hydrodynamic radial support for the shaft  13 . The concave surfaces  53 , 55  of the upper and lower shells  25 , 27  serve as radial bearing surfaces for the shaft  13 . 
     Referring again to the lower half bearing  19 , the shell  27  intersects the flanges  29 , 31  to provide a recessed corner  57  which may be chamfered, flat or rounded. Rotation of the shaft  15  drags oil along the recessed corner to provide added lubrication to the bearing surfaces. 
     The axially outer facing surface  35  of the flange  29  extends circumferentially in a sliding direction R which corresponds to the direction of rotation of the shaft  15 . With reference to the sliding direction R, the axially outer facing surface  35  has a leading end  59  at one end of the thrust flange  29 , and a trailing end  61  at the circumferentially opposite end of the thrust flange  29 . Thus, as the shaft  13  rotates and oil is directed onto the axially outer facing surface  35 , it passes first over the leading end  59  and is carried in the sliding direction R toward the trailing end  61 . 
     The axially outer facing surface  35  is formed with a plurality (i.e., two or more) of oil supply grooves. In the illustrated preferred embodiment, there are three such grooves and they are designated in the drawings by the reference numerals  63 ,  65  and  67 , respectively. The groove  63  is adjacent the leading end  59  and thus is designated the leading oil supply groove. Oil supply groove  67  is associated with the trailing end  61 , and thus is designated the trailing end groove. The remaining groove  65  is located between the leading and trailing grooves  63 , 67  and thus will be designated the intermediate groove. 
     As best shown in FIGS. 3 and 4, the grooves  63 ,  65 ,  67  each have an axis that passes through the center of the respective grooves. The axes, and thus the grooves  63 , 65  and  67 , extend radially of the shell  27  across the axially outer facing surface  35 , and thus are non-parallel to one another. It will be seen from FIG. 3 that the axes of the grooves intersect at the axial center line of the bearing assembly  11 , which corresponds to the axis  15  of rotation of the crank shaft  13 . The grooves  63 ,  65  and  67  are preferably narrow in relation to conventional grooves of the type disclosed in the aforementioned U.S. Pat. No. 5,192,136, which is incorporated herein by reference. The width of the grooves is preferably on the order of about 0.05-0.16 inches wide and have a depth of about 0.01-0.03 inches and are preferably formed entirely in the liner layer  43  as so as not to extend through the liner to the backing  45 . 
     The grooves  63 ,  65  and  67  are spaced circumferentially apart from one another and from the ends  59 , 61  of the axially outer facing surface  35 . The preferred relative spacing of the grooves with respect to one another and with respect to the ends is illustrated best in FIG.  3  and is expressed in degrees relative to the leading end  59 , with the overall circumferential length of the outer surface  35  being 180°. In FIG. 3, the locations of the axes of the grooves  63 ,  65  and  67  are designated by the letters C, E and G, respectively. It will be seen that the grooves  63 ,  65  and  67  are spaced equally from one another, preferably at a spacing of about 35° between their axes. It will also be seen that the leading groove  63  is closer to the leading end  59  than is the trailing groove  67  to the trailing end  61 . For reasons that will become more apparent below, the grooves are shifted toward the leading end  59  to provide a greater circumferential length of the outer surface  35  between the trailing groove  67  in the trailing end  61 . Preferably, the leading groove  63  is spaced about 45° from the leading end  59 , whereas the trailing groove  67  is spaced about 65° from the trailing end  61 . 
     Referring now more particularly to FIG. 4, a greatly exaggerated profile of the axially outer surface  35  is shown in which the hydrodynamic features are emphasized. It will be seen from FIGS. 3 and 4 that the grooves  63 ,  65  and  67  subdivide the axially outer facing thrust bearing surface  35  into a plurality of thrust bearing pads which are separated by the grooves. As shown in these two figures, there are four such thrust bearing pads and they include a leading thrust bearing pad  69  extending between the leading end  59  and leading groove  63 , a trailing thrust bearing pad  70  extending between the trailing groove  67  and the trailing end  61  and one or more (in this case two) intermediate thrust bearing pads  72  disposed between the leading and trailing grooves  63 , 67  and any intermediate grooves  65  therebetween. 
     Each thrust bearing pad  69 ,  70  and  72  is contoured to induce hydrodynamic support action, sometimes referred to as film wedge hydrodynamic action, to the thrust surface  35  suitable for handling relatively high thrust loads to prevent or minimize metal-to-metal contact between the thrust surface  35  and the shaft  13 . As the shaft rotates, oil is drawn across the thrust face  35  and is forced into narrowing spaces formed between the thrust face  35  and an opposing thrust shoulder of the shaft. To achieve the hydrodynamic effect, each thrust bearing pad is provided with an inclined ramp surface  74  that extends in the direction of inclination with the sliding direction R of the shaft  13  across the thrust face  35 . The ramp  74  of the intermediate and trailing thrust pads  72 , 70  preferably has a height of about 0.002-0.004 inches. The ramp surface  74  at the leading thrust bearing pad  69  has a height of about 0.002-0.006 inches. The length of the ramp surface  74  on the leading thrust bearing pad  69 , given in degrees, is about 37°. The length of the ramp  74  in the intermediate and trailing thrust bearing pad  72 , 70  is about 27°. 
     In each thrust bearing pad  69 ,  70  and  72 , the inclined ramp  74  transitions into an elevated land area  76  which preferably is flat and planar. The width of the land area for each pad is preferably about 8° or less. The land  76 , in turn, transitions into the grooves  63 ,  65  and  67 , except for the land  76  of the trailing thrust bearing pad  70 . Preferably, the land  76  of the leading thrust bearing pad  69  and the intermediate thrust bearing pads  72  transition abruptly into their leading and intermediate grooves  63 , 65  in the sliding direction R. In such case, there is no spacing between the point where the land  76  terminate and the grooves  63 , 65  begin. The invention does contemplate, however, a transition between the lands  76  and grooves  63 , 65 . However, the trailing thrust bearing pad  70  is provided with a declining relief ramp area  78  that extends from the land  76  to the trailing end  61 . The relief area  78  spans preferably about 30° from the end of the land  76  to the trailing end  61 , such that the trailing end  61  resides at a level below the land  76  of the trailing thrust bearing pad  70 . Referring to FIG. 10, the resultant hydrodynamic pressure profile is illustrated by way of comparison to the prior art profile of FIG.  9 . In FIG. 10, the spacing of the land  76  from the trailing end  61  along with the declining relief ramp  78  provides a “no contact” zone of the flange adjacent the trailing end  61 . Consequently, the pressure peak location at the land  76  is spaced substantially from the trailing end  61  (by about 30° or more) which acts to locate the peak hydrodynamic load inwardly of the trailing end  61  to relieve localized stress at the trailing end  61  of the flange  29 . It will be seen from FIG. 10 that the pressure profile peaks are about equal in magnitude across the lands  76 , avoiding sharp concentrations at the ends, and particularly the trailing end  61  according to the prior profile of FIG.  3 . 
     FIG. 5 a  shows the opposite side of the bearing assembly  11  as that seen in FIG. 1, and thus the flanges  31  of the upper and lower bearing halves  17 , 19  are shown. The grooves  80  may be identical in location, size and orientation to the grooves  63 ,  65  and  67 . However, as shown in FIG. 5 b , the surface regions between adjacent grooves  80  and between the endmost grooves  80  on the ends of the thrust pad are substantially planar and do not incorporate the hydrodynamic ramp and land features of the opposite flange  29 . Moreover, while a particular groove configuration  80  is shown in FIGS. 5 a  and  5   b , the invention contemplates more or less grooves (e.g., five such grooves  80 ) or alternative groove configurations, such as parallel thumbnail-style grooves shown in U.S. Pat. No. 5,192,136, the disclosure of which is incorporated herein by reference. The grooves  80  serve to channel lubricating oil across the flanges  31 . 
     Referring now to FIG. 6, an alternative embodiment 11′ of the bearing assembly is shown, wherein everything is the same as described above with respect to the first embodiment except that the flanges  29 ′ and  31 ′ are formed separately from the shell  25 , 27 . The same reference numerals are used to designate like features, but are primed. The flanges  29 ′, 31 ′ are connected by any suitable means to achieve a unitized structure such as in the manner disclosed, for example, in U.S. Pat. No. 4,714,356, the disclosure of which is incorporated herein by reference. While this patent discloses welding the flanges  29 ′, 31 ′ to the shell  27 ′, other conventional joining processes may be used, for example providing radial lugs on the flanges which engage in slots at the edges of the shell, such as shown in U.S. Pat. No. 3,713,714 also incorporated herein by reference. 
     Obviously, many modifications and variation of the present invention are possible in light of the above teachings. It is, therefore, to be understood that within the scope of the appended claims, the invention may be practiced otherwise than as specifically described. The invention is defined by the claims.