Patent Publication Number: US-8534389-B2

Title: Method and apparatus for reducing lubricant pressure pulsation within a rotary cone rock bit

Description:
TECHNICAL FIELD 
     The present invention relates generally to rock bit drilling tools, and more specifically concerns roller cone drilling tools and the lubrication and pressure compensation systems used within such roller cone drilling tools. 
     BACKGROUND 
     A roller cone rock bit is a commonly used cutting tool used in oil, gas, and mining fields for breaking through earth formations and shaping well bores. Reference is made to  FIG. 1  which illustrates a cross-sectional view of a portion of a typical roller cone rock bit.  FIG. 1  specifically illustrates the portion comprising one head and cone assembly of the bit. The general configuration and operation of such a bit is well known to those skilled in the art. 
     The head  10  of the bit includes a downwardly and inwardly extending bearing shaft  12 . A cutting cone  14  is rotatably mounted on the bearing shaft  12 . The bearing system for the head and cone assembly that is used in roller cone rock bits to rotatably support the cone  14  on the bearing shaft  12  typically employs either rollers as the load carrying element (a roller bearing system) or a journal as the load carrying element (a friction bearing system).  FIG. 1  specifically illustrates a friction journal bearing implementation including a bearing system defined by a first cylindrical friction bearing  16  (also referred to as the main journal bearing). The cone  14  is axially retained on the bearing shaft  12 , and further supported for rotation, by a set of ball bearings  18  provided within an annular raceway  20 . The bearing system for the head and cone assembly further includes second cylindrical friction bearing  22 , first radial friction (thrust) bearing  24  and second radial friction (thrust) bearing  26 . 
     The bearing system for the head and cone assembly of the bit is lubricated and sealed. The interstitial volume within the bearing system defined between the cone  14  and the bearing shaft  12  is filled with a lubricant (typically, grease). This lubricant is provided to the interstitial volume through a series of lubricant channels  28 . A pressure compensator  30 , usually including an elastomer diaphragm, is coupled in fluid communication with the series of lubricant channels  28 . The lubricant is retained within the bearing system by a sealing system  32  provided between the base of the cone  14  and the base of the bearing shaft  12 . The configuration and operation of the lubrication and sealing systems within roller cone drill bits are well known to those skilled in the art. 
     A body portion  34  of the bit, from which the head and cone assembly depends, includes an upper threaded portion forming a tool joint connection which facilitates connection of the bit to a drill string (not shown, but well understood by those skilled in the art). 
       FIG. 2  illustrates a cross-sectional view of the bit shown in  FIG. 1  focusing on a portion of the bearing system in greater detail. In particular,  FIG. 2  specifically focuses on the area of the first cylindrical friction bearing (main journal bearing)  16 . The first cylindrical friction bearing  16  is defined by an outer cylindrical surface  40  on the bearing shaft  12  and an inner cylindrical surface  42  of a bushing  44  which has been press fit into the cone  14 . This bushing  44  is a ring-shaped structure typically made of beryllium copper, although the use of other materials is known in the art. In a roller bearing system, the outer cylindrical surface  40  on the bearing shaft  12  would interact with roller bearings maintained, for example, in an annular roller raceway within the cone  14 . 
       FIG. 2  further shows that the ball bearings  18  ride in the annular raceway  20  defined at an interface between the bearing shaft  12  and cone  14 . The ball bearings  18  are delivered to the raceway  20  through a ball opening  46 , with that opening  46  being closed by a ball plug  48 . The ball plug  48  is shaped to define a portion of the lubricant channels  28  within the ball opening  46 . The ball bearing system as shown would typically also present in bearing system implementations which utilize roller bearings. 
     As discussed above, lubricant is retained within the bearing system by a sealing system  32 . The sealing system  32 , in a basic configuration, comprises an o-ring type seal member  50  positioned in a seal gland  52  between the cutter cone  14  and the bearing shaft  12  to retain lubricant and exclude external debris. A cylindrical surface seal boss  54  is provided at the base of the bearing shaft  12 . In the illustrated configuration, this surface of the seal boss  54  is outwardly radially offset (for example, by the thickness of the bushing  44 ) from the outer cylindrical surface  40  of the first friction bearing  16 . It will be understood that the seal boss  54  could exhibit no offset with respect to the main journal bearing  16  surface  40  if desired. The annular seal gland  52  is formed in the base of the cone  14 . The gland  52  and seal boss  54  align with each other when the cutting cone  14  is rotatably positioned on the bearing shaft  12 . The o-ring sealing member  50  is compressed between the surface(s) of the gland  52  and the seal boss  54 , and functions to retain lubricant within the bearing system. This sealing member  50  also prevents materials in the well bore (such as drilling mud and debris) from entering into the bearing system. 
     Over time, the rock bit industry has moved from a standard nitrile material for the seal member  50 , to a highly saturated nitrile elastomer for added stability of properties (thermal resistance, chemical resistance). The use of a sealing system  32  in rock bit bearings has dramatically increased bearing life in the past fifty years. The longer the sealing system  32  functions to retain lubricant within the interstitial volume, and exclude contamination of the bearing system, the longer the life of the bearing and drill bit. The sealing system  32  is, thus, a critical component of the rock bit. 
     With reference once again to  FIG. 1 , the second cylindrical friction bearing  22  of the bearing system is defined by an outer cylindrical surface  60  on the bearing shaft  12  and an inner cylindrical surface  62  on the cone  14 . The outer cylindrical surface  60  is inwardly radially offset from the outer cylindrical surface  40  ( FIG. 2 ). The first radial friction bearing  24  of the bearing system is defined between the first and second cylindrical friction bearings  16  and  22  by a first radial surface  64  on the bearing shaft  12  and a second radial surface  66  on the cone  14 . The second radial friction bearing  26  of the bearing system is adjacent the second cylindrical friction bearing  22  at the axis of rotation for the cone and is defined by a third radial surface  68  on the bearing shaft  12  and a fourth radial surface  70  on the cone  14 . 
     The lubricant is provided in the interstitial volume that is defined between the surfaces  40  and  42  of the first cylindrical friction bearing  16 , the surfaces  60  and  62  of the second cylindrical friction bearing  22 , the surfaces  64  and  64  of the first radial friction bearing  24  and the surfaces  68  and  70  of the second radial friction bearing  26 . The sealing system  32  with the o-ring type seal member  50  positioned in the seal gland  52  functions to retain the lubricant within the lubrication system and specifically between the opposed radial and cylindrical surfaces of the bearing system. 
     During operation of the bit, the rotating cone  14  oscillates along the head in at least an axial manner. This motion is commonly referred to in the art as a “cone pump.” Cone pumping is an inherent motion resulting from the external force that is imposed on the cone by the rocks during the drilling process. The oscillating frequency of this cone pump motion with respect to the head is related to the rotating speed of the bit. The magnitude of the oscillating cone pump motion is related to the manufacturing clearances provided within the bearing system (more specifically, the manufacturing clearances between the surfaces  40  and  42  of the first cylindrical friction bearing  16 , the surfaces  60  and  62  of the second cylindrical friction bearing  22 , the surfaces  64  and  64  of the first radial friction bearing  24  and the surfaces  68  and  70  of the second radial friction bearing  26 ). The magnitude is further influenced by the geometry and tolerances associated with the retaining system for the cone (for example, the ball race). When cone pump motion occurs, the interstitial volume defined between the foregoing cylindrical and radial surfaces of the bearing system changes. This change in volume squeezes the lubricant provided within the interstitial volume. The change in interstitial volume and squeezing of the lubricant grease results in the generation of a lubricant pressure pulse. Over a very short period of time, responsive to this pressure pulse, grease flows along a first path between the bearing system and the pressure compensator  30  through the series of lubricant channels  28 . The pressure compensator  30  is designed to relieve or dampen the pressure pulse by compensating for volume changes through its elastomer diaphragm. However, it is known in the art that the pressure pulse, notwithstanding the presence and actuation of the pressure compensator  30 , can also be felt at the sealing system  32  due to the presence of a separate second path for the flow of grease, responsive to this pressure pulse, between the opposed radial and cylindrical surfaces of the bearing system and the sealing system  32 . 
     The flow of grease along this second path in response to the pressure pulse is known to be detrimental to seal operation and can also reduce seal life. For example, positive and negative pressure pulses due to cone pump motion may cause movement of the sealing member  50  within the seal gland. A nibbling and wearing of the seal member  50  may result from this movement. Additionally, a positive pressure pulse due to cone pump motion may cause lubricant grease to leak out past the sealing system  32 . A negative pressure pulse due to cone pump motion may pull materials from the well bore (such as drilling mud and debris) past the sealing system  32  and into the bearing system. 
     Reference is now made to  FIG. 3  which shows a cross-section of the bearing shaft  12  generally at the location of the first friction bearing  16  taken along dotted line  80  of  FIG. 2 . As is known by those skilled in the art, the first friction bearing  16  for the bearing system includes a loading zone (having an arc angle of about 120°-180°) which bears the load of the cone  14  and a non-loading zone (having an arc angle of about 180°-240°). The outer surface  40  of the bearing shaft  12  at the loading zone is typically hardfaced (not explicitly shown, but known to those skilled in the art). One of the lubricant channels  28  for the lubrication system terminates at the outer cylindrical surface  40  of the bearing shaft  12  in the area of the non-loading zone. The termination of the lubricant channel  28  on the outer surface  40  of the bearing shaft  12  is typically provided by a circumferentially positioned groove  90  that is milled or machined into the outer surface  40 . This groove  90  includes an opening  92  for providing fluid communication into the lubricant channel  28 . 
     Reference is now made to  FIG. 4  which shows a side view of the bearing shaft  12  focusing on the non-loading zone. The circumferentially positioned groove  90  terminates the lubricant channel  28  at the outer surface  40  of the first friction bearing  16  for the bearing system using opening  92 . The axial width  94  of the groove  90  spans most, but not all, of the axial width  96  of the surface  40  for the first friction bearing  16  of the bearing system. For example, the axial width  94  is typically equal to the axial width  96  minus a constant (such as twice a fraction of an inch, for example, 2* 1/32″ or 2* 3/64″. In this way, the axial width  94  is typically greater than 80-90% of the axial width  96 . The groove  90  is typically axially centered with respect to the surface  40  providing two equally sized attenuation zones  100 . Because of the relative widths  94  and  96 , the attenuation zones  100  present a minimal amount of outer surface  40  for the first friction bearing  16  that is located axially adjacent the groove  90  and present along the path shown by arrow  98 . This minimal amount of outer surface  40  is insufficient to restrict the flow of grease and the passage of a pressure pulse between the bearing system (at surfaces  60 ,  64  and  68 ) and the sealing system  32  (at surface  54 ) along path  98 . More specifically, this minimal amount of surface  40  along the path of arrow  98  provides only two relatively short (in an axial direction) attenuation zones  100  which might assist in attenuating the flow of grease along the path of arrow  98  resulting from the axial passage of the pressure pulse. In this configuration, the pressure pulse may travel along surface  40  and reach the sealing system  32  (at surface  54 ) before being dampened by the pressure compensator  30 . As discussed above, this pressure pulse may have detrimental effects on the sealing system  32  and particularly the sealing member  50 . There is accordingly a need in the art to reduce, or eliminate, the pressure pulsation due to cone pumping from acting on the sealing system  32 . 
     SUMMARY 
     A drill tool includes a bit body, at least one bearing shaft extending from the bit body and a cone mounted for rotation on the bearing shaft. An outer bearing surface of the bearing shaft includes a non-loading zone. In an embodiment, a first groove and a second groove are formed in the outer bearing surface at the non-loading zone. The first and second grooves are both circumferentially offset from each other and axially offset from each other. The circumferential and axial offsetting of the first and second grooves define a plurality of attenuation zones that function to restrict propagation of a cone pumping pressure pulse towards a sealing system of the drill tool. 
     In an embodiment, a drill tool comprises: a bit body; at least one bearing shaft extending from the bit body; a cone mounted for rotation on the bearing shaft; a first groove formed in a non-loading zone of an outer bearing surface of the bearing shaft; and a second groove formed in the non-loading zone of the same outer bearing surface of the bearing shaft; wherein the first groove is circumferentially offset from the second groove. 
     In a further embodiment, the first and second grooves are axially offset from each other on the outer bearing surface of the bearing shaft. 
     In an embodiment, openings are provided in the first and second grooves for fluid communication to an internal lubrication channel of the tool. 
     The circumferential offset of the first and second grooves provides a circumferential attenuation zone to restrict propagation of a cone pumping pressure pulse from a pressure source towards a sealing system of the drill tool. 
     The axial offset of the first and second grooves provides a plurality of axial attenuation zones to restrict propagation of a cone pumping pressure pulse from a pressure source towards a sealing system of the drill tool. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  illustrates a cross-sectional view of a portion of a typical roller cone rock bit; 
         FIG. 2  illustrates a cross-sectional view of the typical roller cone rock bit shown in  FIG. 1  focusing on the bearing system in greater detail; 
         FIG. 3  illustrates a cross-section of the bearing shaft taken at the location of the dotted line in  FIG. 2 ; 
         FIG. 4  illustrates a side view of the bearing shaft of  FIG. 2 ; 
         FIG. 5  illustrates a cross-sectional view of roller cone rock bit focusing on an embodiment of a bearing system in greater detail; 
         FIG. 6  illustrates a cross-section of the bearing shaft taken at the location of the dotted line in  FIG. 5 ; and 
         FIG. 7  illustrates a side view of the bearing shaft of  FIG. 5 . 
     
    
    
     DETAILED DESCRIPTION OF THE DRAWINGS 
       FIG. 5  illustrates a cross-sectional view of a roller cone rock bit focusing on an embodiment of the present invention for addressing lubricant pressure pulsation originating at the bearing system.  FIG. 5  is specifically directed to the area of the cylindrical friction bearing (main journal bearing)  116 . The cylindrical friction bearing  116  is defined by an outer cylindrical surface  140  on a bearing shaft  112  and an inner cylindrical surface  142  of a bushing  144  which has been press fit into a cone  114  mounted to rotate about the bearing shaft  112 . The bushing  144  is a ring-shaped structure typically made of beryllium copper, although the use of other materials is known in the art. In a roller bearing system, the outer cylindrical surface  140  on the bearing shaft  112  would interact with roller bearings maintained, for example, in an annular roller raceway within the cone  114 . 
     The bearing system further includes ball bearings  118  which ride in an annular raceway  120  defined at the interface between the bearing shaft  112  and cone  114 . The ball bearings  118  are delivered to the raceway  120  through a ball opening  146 , with that opening  146  being closed by a ball plug  148 . The ball plug  148  is shaped to define a portion of a lubricant channel  128 . The ball bearing system as shown would typically also present in bearing system implementations which utilize roller bearings. 
     Lubricant is provided in the interstitial volume between the surfaces  140  and  142  of the cylindrical friction bearing  116  as well as in the annular raceway  120  and other opposed cylindrical and radial bearing surfaces (as discussed above) between the cone  114  and the shaft  112 . The lubricant is retained within the bearing system by a sealing system  132 . The sealing system  132 , in a basic configuration, comprises an o-ring type seal member  150  positioned in a seal gland  152  between the cutter cone  114  and the bearing shaft  112  to retain lubricant and exclude external debris. A cylindrical surface seal boss  154  is provided at the base of the bearing shaft  112 . In the illustrated configuration, this surface of the seal boss  154  is outwardly radially offset (for example, by the thickness of the bushing  144 ) from the outer cylindrical surface  140  of the first friction bearing  116 . It will be understood that the seal boss could exhibit no offset with respect to the main journal bearing surface  40  if desired. The annular seal gland  152  is formed in base of the cone  114 . The gland  152  and seal boss  154  align with each other when the cutting cone  114  is rotatably positioned on the bearing shaft  112 . The o-ring sealing member  150  is compressed between the surface(s) of the gland  152  and the seal boss  154 , and functions to retain lubricant within the bearing system. This sealing member  150  also prevents materials (drilling mud and debris) in the well bore from entering into the bearing system. 
     Reference is now made to  FIG. 6  which shows a cross-section of the bearing shaft  112  generally at the location of the friction bearing  116  and taken along dotted line  180  of  FIG. 5 . The friction bearing  116  for the bearing system includes a loading zone (having an arc angle of about 120°-180°) which bears the load of the cone  114  and a non-loading zone (having an arc angle of about 180°-240°). The outer surface of the bearing shaft  112  at the loading zone is typically hardfaced (not explicitly known, but understood by those skilled in the art). At least one of the lubricant channels  128  for the lubrication system terminates at the outer surface  140  of the bearing shaft  112  in the area of the non-loading zone (in this embodiment, two such terminations are shown, but it will be understood that three or more terminations could be provided). Each termination of the lubricant channel  128  on the outer surface  140  of the bearing shaft  112  is provided at a circumferentially positioned groove  190  that is milled or machined into the outer surface  140  of the bearing shaft  112 . This groove  190  includes an opening  192  into the lubricant channel  128 . 
       FIG. 6  specifically shows the presence of two grooves  190  formed in the outer surface  140  of the bearing shaft  112 . It will be understood that three or more grooves  190  could be provided. The included grooves  190  are circumferentially offset from each other (by an arc angle of between about 45-120°). Although both grooves  190  are shown to include openings  192  into the lubricant channel  128 , it will be understood that this is not required. A groove  190 , without an opening  192  into the lubricant channel  128 , could instead be provided. Indeed, neither of the two grooves  190  of  FIG. 6  is required to have an opening  192  to the lubricant channel  128  as long as some other mechanism is provided for ensuring the delivery of lubricant to the friction bearing  116 . 
     In comparing the grooves  190  with openings  192  in  FIG. 6  to the groove  90  with opening  92  in  FIG. 3 , it will be noted that the openings  192  in  FIG. 6  into the lubricant channel  128  have a smaller diameter than the opening  92  in  FIG. 3 . The smaller openings  192  serve to restrict the flow of lubricant grease through the openings  192 . 
     Although two grooves  190  are shown in  FIG. 6 , it will be understood that more than two circumferentially offset grooves  190  could be provided. 
     The circumferential length  208  of each groove  190  may, for example, extend over an arc angle of between about 10-30°, and more preferably extend over an arc angle of between about 15-20°. 
     Reference is now made to  FIG. 7  which shows a side view of the bearing shaft  112  focusing on the non-loading zone. Each circumferentially positioned groove  190  terminates the lubricant channel  128  at the friction bearing  116  for the bearing system using an opening  192 . The two grooves  190  are circumferentially offset from each other. The axial width  194  of each groove  190  is shorter than the axial width  94  of the groove  90  in  FIG. 4 . In a preferred embodiment, the axial width  194  of each groove  190  is no more than 70% of the axial width  196  of the friction bearing  116  for the bearing system. In a preferred implementation, a ratio of circumferential length  208  to axial width  194  of each groove  190  is between about 2-to-1 and about 4-to-1. 
     As discussed above, the openings  192  in  FIG. 6  into the lubricant channel  128  have a smaller diameter than the opening  92  in  FIG. 3 . Reducing the size of the opening  192  (in comparison to the opening  92 ) restricts the flow of grease through the opening  192  and thus assists in attenuating the pressure pulse and grease flow associated with instances of cone pumping. In a preferred embodiment, the cross sectional area of the opening  192  is less than 150% of the annular flow area of the bearing in the vicinity of the groove  190  between the surfaces  140  and  142 . Mathematically, this may be expressed as follows:
 
 D≈k *((4/π)*( C*L ))^0.5
 
     wherein: D=diameter of the opening  192 ; k is a constant, for example, greater than 1 such as 1.5; C=diametrical clearance of the bearing; and L=arc length of the groove  190  (see, reference  208  in  FIGS. 6 and 7 ). 
     Alternatively, this may be mathematically expressed as follows:
 
 D 2≦ k *(( D 1+ C )^2− D 1^2)^0.5
 
     wherein: D2=diameter of the opening  192 ; k is a constant, for example, a fraction less than 1 such as 0.9; D1=diameter of the shaft at the surface  140  and C=diametrical clearance of the bearing. 
     While reducing the diameter of the opening  192  is one preferred option, another option is to insert a choke structure (such as a choke plate or constrictor) in a larger sized opening such as the opening  92  shown in  FIG. 3 , this choke structure effectively providing a constricted opening in the manner described above. 
     Although  FIG. 7  shows that each groove  190  includes an opening  192  to the lubricant channel  128 , it will be understood that only one of the grooves  190  could have an opening  192 , with the other groove  190  comprising a blind area formed on the bearing surface  140 . Still further, it will be understood that neither of the circumferentially offset grooves  190  need have an opening  92  to the lubricant channel  128  provided some other mechanism exists for ensuring the delivery of lubricant to the friction bearing  116 . 
     In a preferred embodiment, each opening  192  is axially offset to a position closer to one edge of the surface  140  for the friction bearing  116 . In other words, the openings  192  are not axially centered on the surface  140  for the friction bearing  116 . For example, the left opening  192  in  FIG. 7  is shown to have an axial offset to a position closer to an upper edge  210  of the surface  140  for the friction bearing  116 , while the right opening  192  in  FIG. 7  is shown to have an axial offset to a position closer to an lower edge  212  of the surface  140  for the friction bearing  116 . In a preferred implementation, the openings  192  are axially offset in opposite directions, as shown in  FIG. 7 . It will be understood, however, that both openings  192  can be axially offset towards a same edge ( 210  or  212 ) of surface  140 . 
     Axially offsetting the openings  192  in the manner described, and providing the relative widths  194  and  196 , increases (in comparison to  FIG. 4 ) the amount of outer surface  140  for the first friction bearing  116  that is axially adjacent the groove  190  and present along the paths shown by arrows  198 . The increased amount of outer surface  140  better restricts the flow of grease and the passage of a pressure pulse between the bearing system (at surfaces  160 ,  164  and  168 ) and the sealing system  132  (at surface  154 ). As a result of the axial offset, the increased amount of surface  140  at each arrow  198  provides (in an axial direction) a relatively shorter attenuation zone  200  on one side of the groove  190  and a relatively longer attenuation zone  202  on the other side of the groove  190 . This configuration with longer attenuation zones  202  provides improved performance over the configuration of  FIG. 4  in terms of attenuating the flow of grease due to the axial passage of the pressure pulse. The additional attenuation resulting from the presence of the relatively longer attenuation zones  202  further assists in protecting the sealing system  132  (at surface  154 ) from the pressure pulse and supports the damping operation of the pressure compensator  30  (see,  FIG. 1 ). In a preferred implementation, the ratio of axial width of the relatively longer attenuation zone  202  to the axial width of the relatively shorter attenuation zone  200  is between about 3-to-1 and about 6-to-1. It is preferred that the axial offsetting of the grooves  190  should preserve at least a small amount of circumferential axial overlap  216  between the grooves, especially in instances where one of the grooves is a blind groove without an opening  192  (but, it should also be understood that no axial overlap  206  may be necessary in some implementations). 
     The circumferential offset of the two grooves  190 , along with the relative widths  194  and  196  and axial offset of the grooves  190 , further provides an additional attenuation zone  204  circumferentially located between the two grooves  190 . The degree of circumferential offset is selected such that circumferential pressure attenuation between the grooves is approximately equal to the axial pressure attenuation between a groove and a further end of the bearing. In other words, the circumferential offset of the grooves  190  is selected so that it is approximately equally difficult for the grease pressure pulse to travel between the end of the bearing system and the groove along the path of arrow  198  as it is for the grease pressure pulse to travel between grooves along the path of arrow  206 . In this way, both possible paths of grease pressure travel are substantially equally attenuated. 
     When cone pump motion occurs, the lubricant provided in the interstitial volume bearing system (with shaft  116  surfaces  140 ,  160 ,  164  and  168 ) is squeezed. This results in the generation of a pressure pulse. In response to the pressure pulse, lubricant grease flows through the series of lubricant channels  28  between the bearing system and the pressure compensator  30  (see,  FIG. 1 ). The pressure compensator  30  is designed to dampen or relieve the pressure pulse by compensating for volume changes through its elastomer diaphragm. The paths provided by arrows  198  and  206 , however, are also available for grease flow in response to the pressure pulse. The attenuation zones  200 ,  202  and  204  are provided to restrict the flow of grease along these paths and thus reduce, or eliminate, the pressure pulsation due to cone pumping from acting on the sealing system  132 . 
     Although  FIGS. 5-7  specifically illustrate the use of a friction journal bearing system, it will be understood that the grooves  190  (with or without openings  192 ) could alternatively be used in connection with a roller bearing system. 
     Furthermore, although  FIG. 5-7  specifically illustrate the provision of grooves  190  (with or without openings  192 ) in connection with the main bearing of the bearing system (whether journal or roller), it will be understood that the grooves  190  (with or without openings  192 ) could alternatively be provided in connection with any suitable bearing surface of shaft  116  (including, but not limited to, surfaces  140 ,  160 ,  164  and  168 ) in either a friction journal bearing or roller bearing implementation. 
     Although explained in the context of a drilling tool designed primarily for use in an oilfield drilling application, it will be understood that the disclosure is not so restricted and that the bearing system as described could be used in any rotary cone drilling tool including tools used in non-oil field applications. Specifically, the drilling tool can be configured for use with any suitable drilling fluid including air, mist, foam or liquid (water, mud or oil-based), or any combination of the foregoing. Furthermore, although described in the context of a solution to the problems associated with cone pumping and lubricant pressure pulsation in sealed and pressure compensated systems, the solutions described herein are equally applicable to rotary cone bits which are lubricated but do not include a pressure compensator and diaphragm system. 
     Although preferred embodiments of the method and apparatus of the present invention have been illustrated in the accompanying Drawings and described in the foregoing Detailed Description, it will be understood that the invention is not limited to the embodiments disclosed, but is capable of numerous rearrangements, modifications and substitutions without departing from the spirit of the invention as set forth and defined by the following claims.