• 专利附录
  • 专利说明
  • 权利要求
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    • 专利摘要:
    A drilling tool includes a drill bit body, at least one bearing shaft, which extends from the drill bit body and a roller, which is mounted for rotating the bearing shaft. An outer bearing surface of the bearing shaft includes a non-load zone. A first track and a second track are formed in the outer bearing surface at the non-load zone. The first and second grooves are both circumferentially offset relative to each other and axially offset relative to each other. One or more of the grooves includes an opening for generating a fluid connection to an internal lubricant channel within this bearing shaft. The peripheral and axial displacement of the first and second grooves defines a plurality of damping zones, which act to limit the propagation of a pressure pulse due to roller pumping towards a sealing system for the drilling tool.
    公开号:SE539208C2
    申请号:SE1350951
    申请日:2011-11-03
    公开日:2017-05-09
    发明作者:Gallifet Thomas；Harrington David
    申请人:Varel Int Ind Lp；
    IPC主号:
    专利说明:

    DRILLING TOOL INCLUDING A DEVICE FOR REDUCING PULSATIONS IN THE LUBRICANT PRESSURE WITHIN A ROLLER DRILLING CROP TECHNICAL FIELD The present invention relates generally to drilling tools with a rock drill bit and a roller drill bit.
    BACKGROUND A roller drill bit is a commonly used cutting tool used in oil, gas and mining fields to break through soil formations and design drill heels for wells. Referring to Figure 1, which illustrates a cross-sectional view of a section of a typical roller drill bit. Figure 1 specifically illustrates the section comprising a unit with head and roller of the drill bit. The general configuration and operation of such a drill bit are well known to those skilled in the art.
    The drill bit head 10 includes a downwardly and inwardly extending bearing shaft 12. A cutting roller (cutting cone) 14 is rotatably mounted on the bearing shaft 12. The bearing system assembly with head and roller used in roller drill bits to rotatably support the roller (cone) 14 on the bearing shaft 12 typically uses either the load-bearing element (a roller bearing system) or a shaft journal as the load-bearing element (a plain bearing system). Figure 1 specifically illustrates a plain bearing implementation, including a bearing system defined by a first cylindrical plain bearing 16 (also referred to as the main plain bearing). The roller 14 is axially inclined to the bearing shaft 12 and further supported to rotate a set of ball bearings 18 provided within an annular bearing race 20. The bearing system of the unit head and roller further includes a second cylindrical plain bearing 22, a pre-radial plain (thrust) bearing 24 and a second radial sliding (pressure) bearing 26.
    The bearing system for the head and roller assembly of the drill bit is lubricated and sealed. The interstitial volume within the bearing system defined between the roller 14 and the bearing shaft 12 is filled with a lubricant (typically grease). This lubricant is supplied to the interstitial volume through a series of lubricant channels 28. A pressure compensator 30, usually including an elastomeric membrane, is connected in communication with the series of lubricant channels 28. The lubricant is retained within the bearing system between the bearing shaft 14 and the base shaft 32. The configuration and operation of the lubrication and sealing systems inside roller drill bits are well known to those skilled in the art.
    A body section 34 of the drill bit, from which the unit with head and roller hanger, includes an upper threaded section which forms a splice connection pre-tool, which facilitates connection of the drill bit to a drill string (not shown, but well understood by those skilled in the art).
    Figure 2 illustrates a cross-sectional view of the drill bit shown in Figure 1 and focuses on each section of the bearing system in more detail. In particular, Figure 2 specifically focuses on the area of the first cylindrical plain bearing (main plain bearing) 16. The first cylindrical plain 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-fitted into the roller 14. This bushing 44 is an annular structure, typically made of avberyllium copper, although the use of other materials is known in the art of rescue. In a roller bearing system, the outer cylindrical surface 40 of the bearing shaft 12 interacts with roller bearings, which are retained, for example, in an annular roller bearing path within the roller 14.
    Figure 2 further shows that the ball bearings 18 are located in the annular bearing race 20 which is defined at an interface between the bearing shaft 12 and the roller 14. The ball bearings 18 supply the bearing race 20 through a ball opening 46, this opening 46 being closed by single ball plug 48. The ball plug 48 is formed so it defines a section of the lubricant channels 28 within the ball opening 46. The ball bearing system shown is typically also present in implementations of bearing systems using roller bearings.
    As discussed above, Lubricant is retained within the bearing system by a sealing system 32. In a basic configuration, the sealing system 32 includes an O-ring type sealing member 50 positioned in a sealing packing box 52 between the cutting roller 14 and the bearing shaft 12 to retain external lubricant and exclude external lubricant. A sealing hub with a cylindrical surface 54 is provided at the base of the bearing shaft 12. In the illustrated configuration, this surface of the sealing hub 54 is radially offset leaky (e.g. with the thickness of the bushing 44) from the outer cylindrical surface 40 of the first plain bearing 16. It will be appreciated that the sealing 54 if desired, need not exhibit any displacement with respect to the main slide bearing surface 16. The annular seal stuffing box 52 is formed in the base of the roller 14. The stuffing box 52 and the sealing hub 54 are aligned when the cutting roller 14 is rotatably positioned on the bearing shaft 12. The O-ring sealing member 50 the intermediate surface (s) of the stuffing box 52 and the sealing hub 54 and acts to retain lubricant within the bearing system. This sealing element 50 also prevents material in the wellbore (such as mud and dirt) from entering the bearing system.
    Over time, the rock drill bit industry has gone from a standard nitrile material densification element 50 to a highly saturated nitrile load for increased stability in terms of properties (heat resistance, chemical resistance). The use of a sealing system 32 in bearings for rock drill bits has dramatically increased the bearing life over the past fifty years. The longer the sealing system 32 acts to retain lubricant within the interstitial volume and the exclusion contaminant from the bearing system, the longer the life of the bearing and drill bit.
    The sealing system 32 thus constitutes a critical component in the rock drill bit. Referring again to Figure 1, the second cylindrical plain bearing 22 in the bearing system is defined by an outer cylindrical surface 60 on the bearing shaft 12 and an inner cylindrical surface 62 on the roller 14. The outer cylindrical surface 60 is radially offset from the outer cylindrical surface 40 (Figure 2). . The first radial plain bearing 24 in the bearing system is defined between the first and second cylindrical plain bearings 16 and 22 by a first radial surface 64 on the bearing shaft 12 and a second radial surface 66 on the roller 14. The second radial plain bearing 26 in the bearing system is located adjacent the second cylindrical plain bearing 22 at the axis of rotation of the roller and is defined by a third radial surface 68 on the bearing shaft 12 and a fourth radial surface 70 on the roller 14.
    The lubricant is provided in the interstitial volume defined between the surfaces 40 and 42 of the first cylindrical plain bearing 16, the surfaces 60 and 62 of the second cylindrical plain bearing 22, the surfaces 64 and 64 of the first radial plain bearing 24 and the surfaces 68 and 70 of the second radial plain bearing 26. The sealing system 32 with the O-ring type sealing element 50, which is positioned in the sealing packing box 52, acts to retain the lubricant within the lubrication system and specifically between the opposing radial and cylindrical surfaces of the bearing system.
    During the operation of the drill bit, the rotating roller 14 oscillates along the head in at least one axial manner. This movement is generally referred to in the prior art as the 'cone pump'. Roll pumping is an inherent movement that moves hair from the external force applied to the roll by the rock during drilling. The oscillating frequency of this roller pump movement with respect to the head is related to the speed of the drill bit. The magnitude of the oscillating roller pump movement is related to the manufacturing clearance provided within the bearing system (more specifically the manufacturing clearance between the surfaces 40 and 42 of the first cylindrical plain bearing 16, the surfaces 60 and 62 of the second cylindrical plain bearing 22, the surfaces 64 and 64 of the first radial plain bearing 68 and 70 for the second radial bearing bearing 26). The size is further affected by the geometry and tolerances associated with the roller retaining system (e.g., the ball bearing path). As the roller pumping motion occurs, the interstitial volume defined between the preceding cylindrical and radial surfaces of the bearing system changes. This change in volume compresses the lubricant provided within the interstitial volume. The change in interstitial volume and the compression of the lubricant grease leads to the generation of a lubricant pressure pulse. For a very short period of time, in response to this pressure pulse, grease wastes 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 attenuate the pressure pulse by compensating for changes in volume through the diesel load diaphragm. However, it is known in the rescue art that the pressure pulse, despite the presence and activation of the pressure compensator 30, may also be noticeable in the sealing system 32 due to the presence of a separate second path for the fat degrading corresponding to this pressure pulse, between the opposing 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 the sealing function and can also reduce the life of the seal. For example, overpressure and negative pressure pulses may cause movement of the sealing member 50 within the sealing packing box due to roll pump movement. A gnawing pressure and abrasion of the sealing member 50 may be caused by this movement. In addition, an overpressure pulse due to roller pump movement may cause lubricant grease to leak past the sealing system 32. A negative pressure pulse due to the roller pump movement may or may not absorb material. 32 and into the storage system.
    We now refer to Figure 3, which shows a cross section of the bearing shaft 12, substantially at the site of the first plain bearing 16, taken along the dashed line 80 of Figure 2. As is known to those skilled in the art, the first plain bearing 16 of the bearing system includes a loading zone (having a bending angle of about 120 ° -180 °), which carries the load of the roller 14 and a non-load zone (having a bending angle of about 180 ° -240 °). The outer surface 40 of the bearing shaft 12 at the load zone is typically core welded (this is not shown explicitly, but is 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-load 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 latch 90 which is milled or machined in the outer surface 40. This latch 90 includes an opening 92 to provide an external connection into the lubricant channel 28.
    We now refer to Figure 4, which shows a side view of the bearing shaft 12 and focuses on the non-load zone. The circumferentially positioned groove 90 terminates the lubricant channel 28 at the outer surface 40 of the first plain bearing 16 bearing system using aperture 92. The axial width 94 of the groove 90 bridges most but not all of the axial width 96 of the surface 40 of the first plain bearing 16. in the storage system. For example, the axial width 94 is typically equal to the axial width 96 minus a constant (such as twice a fraction of 2.54 cm, for example 2 * 1 / 81.28 cm or 2 * 3 / 162.56 cm. the 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 and provides two equal damping zones 100. Due to the relative widths 94 and 96, the front damping zones 100 have a minimal dimension of the outer surface. 40 for the first sliding bearing 16, which is located axially adjacent to the groove 90 and is present along the path shown by arrow 98. This minimum dimension of the outer surface 40 is insufficient to limit the fat content and the passage of a pressure pulse between the bearing systems (at the surfaces 60). , 64 and 68) and the sealing system 32 (at surface 54) along path 98. More specifically, this minimal dimension of surface 40 along the path of arrow 98 provides only two relatively short (in an axial direction) attenuation zones 100 which can bis toe the evaporation of fl the fate of fat along the path of arrow 98 as hairpin from the axial passage of the pressure pulse. In this configuration, the pressure pulse may travel along the surface 40 and reach the sealing system 32 (at the surface 54) before being damped by the pressure compensator 30. As discussed above, this pressure pulse may have harmful effects on the sealing system 32 and in particular the sealing element 50. need to reduce or eliminate the pressure pulsation due to roller pumping from acting on the sealing system 32.
    SUMMARY A drilling tool includes a drill bit body, at least one bearing shaft, which extends from the drill bit body and a roller, which is mounted for rotation on the bearing shaft. An outer bearing surface of the bearing shaft includes a non-load zone. A first track and a second track are formed in the outer bearing surface at the non-load zone. The first track is circumferentially offset from the second track. The outer bearing surface of the bearing shaft forms a cylindrical surface which is axially located between a source of entry pressure pulse due to roller pumping and a sealing system for the roller and bearing shaft. The first and second grooves are positioned in the non-load zone of the outer bearing surface, so that each defines a first damping zone and a second damping zone, the first and second damping zones axially limiting the propagation of the pressure pulse due to the roller pumping towards the sealing system. The first and second grooves are axially offset relative to each other. In one embodiment, openings are provided in the first and second grooves for an external connection to an internal lubrication channel for the tool.
    The peripheral displacement of the first and second grooves provides a peripheral attenuation zone for limiting the propagation of a pressure pulse due to roller pumping from a pressure source towards a sealing system for the drilling tool.
    The axial displacement of the first and second grooves provides a plurality of axial damping zones to limit the propagation of a pressure pulse track due to roller pumping from a pressure source towards a sealing system for the drilling tool.
    BRIEF DESCRIPTION OF THE DRAWINGS Figure 1 illustrates a cross-sectional view of a section of a typical roller drill bit; Figure 2 illustrates a cross-sectional view of the typical roller drill bit shown in Figure 1 and focuses on the bearing system in more detail; Figure 3 illustrates a cross section of the bearing shaft, taken at the location of the dashed line in Figure 2; Figure 4 illustrates a side view of the bearing shaft of Figure 2; Figure 5 illustrates a cross-sectional view of a roller drill bit and focuses on one embodiment of a bearing system in more detail; Figure 6 illustrates a cross section of the bearing shaft, taken at the location of the dashed line in Figure 5; and Figure 7 illustrates a side view of the bearing shaft of Figure 5.
    DETAILED DESCRIPTION OF THE DRAWINGS Figure 5 illustrates a cross-sectional view of a roller drill bit and focuses on one embodiment of the present invention for combating pulsations the ice lubricant pressure originating with the bearing system. Figure 5 is specifically directed to the area of the cylindrical plain bearing (main plain bearing) 116. The cylindrical plain 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-fitted into a roller 114 mounted to rotate about the bearing shaft 112. The bushing 144 is an annular structure, typically made of beryllium copper, although the use of other materials is known in the art of rescue. In a roller bearing system, the outer cylindrical surface 140 of the bearing shaft 112 interacts with roller bearings, which are retained, for example, in an annular roller bearing path within the roller 114.
    The bearing system further includes ball bearings 118, which are located in an annular bearing race 120 defined at the interface between the bearing shaft 112 and the roller 114. The ball bearings 118 are fed to the bearing race 120 through a ball opening 146, this opening 146 being closed by a ball plug 148. The ball plug 148 is configured to define a section of a lubricant channel 128. The ball bearing system shown is typically also present in implementations of bearing systems using roller bearings.
    Lubricant is provided in the interstitial volume between the surfaces 140 and 142 for the cylindrical plain bearing 116 as well as in the annular bearing race 120 and other opposing cylindrical and radial bearing surfaces (as discussed above) the intermediate roller 114 and the shaft 112. The lubricant is retained within the bearing sealing system 132. in a basic configuration, an O-ring type sealing member 150, which is positioned in a seal packing box 152 between the cutting roller 114 and the bearing shaft 112 to retain lubricant and exclude external dirt. A sealing hub with a cylindrical surface 154 is provided at the base of the bearing shaft 112. In the illustrated configuration, this surface of the sealing hub 154 is radially offset outwardly (for example with the thickness of the bushing 144) from the outer cylindrical surface 140 of the first plain bearing 116. , if desired, need not be displaced with respect to the surface 40 of the main slide bearing. The annular sealing stuffing box 152 is formed in the base of the roller 114. The stuffing box 152 and the sealing hub 154 are aligned when the notch roller 114 is rotatably positioned on the bearing shaft 112. the intermediate surface (s) of the stuffing box 152 and the sealing hub 154 and acts to retain lubricant within the bearing system. This sealing element 150 also prevents material (drilling mud and dirt) in the well bore from entering the bearing system.
    We now refer to Figure 6, which shows a cross section of the bearing shaft 112, substantially at the location of the bearing bearing 116 and taken along the dashed line 180 of Figure 5. The bearing bearing bearing bearing 116 includes a load zone (having an arc angle of about 120 ° -180 °). , which carries the load 114 of the roller and a non-load zone (having an arc angle of about 180 ° -240 °). The outer surface of the bearing shaft 112 at the load zone is typically hard welded (not explicitly known, but will be appreciated 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-load zone (in this embodiment, two such terminations are shown, but it will be appreciated that three or more terminations are provided). Each end of the lubricant channel 128 on the outer surface 140 of the bearing shaft 112 is provided at a circumferentially positioned groove 190 which is milled or machined in the outer surface 140 of the bearing shaft 112. This groove 190 includes an opening 192 into the lubricant channel 128.
    Figure 6 specifically shows the presence of two grooves 190 formed in the outer surface 140 of the bearing shaft 112. It will be appreciated that three or more grooves 190 are provided on the edge. The included grooves are 190 years circumferentially offset from each other (with 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 appreciated that this is not required. A groove 190 without any opening 192 into the lubricant channel 128 could instead be provided. In fact, neither of the two grooves 190 of Figure 6 need have an opening 192 against the lubricant channel 128 as long as any other mechanism is provided for securing the delivery of lubricant to the plain bearing 116.
    When comparing the grooves 190 with the openings 192 in Figure 6 and the groove 90 with the opening 92 in Figure 3, it should be appreciated that the openings 192 in Figure 6 into the ice lubricant channel 128 have a smaller diameter than the opening 92 in Figure 3. The smaller openings 192 serve to limit the fate of lubricant grease through the openings 192. Although two grooves 190 are shown in Figure 6, it will be appreciated that more than two circumferentially offset grooves 190 may be provided.
    For example, the peripheral length 208 of each groove 190 may extend over an arc angle of between about 10-300, and more preferably extend over an arc angle of between about 15-20 °.
    We now refer to Figure 7, which shows a side view of the bearing shaft 112 and focuses on the non-load zone. Each circumferentially positioned groove 190 terminates the lubricant channel 128 at the sliding 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 Figure 4. In a preferred embodiment, the axial width 194 of each groove 190 is equal to 70% of the axial width 196 of the bearing bearing 116 of the bearing system. In a preferred implementation, a ratio between the peripheral length 208 and the 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 Figure 6 into the lubricant channel 128 have a smaller diameter than the opening 92 in Figure 3. A reduction in the dimensions of the opening 192 (compared to the opening 92) limits the fat flow through the opening 192 and thus assists in attenuating the pressure pulse and grease. as years associated with occurrences of roll pumping. In a preferred embodiment, the cross-sectional area of the aperture 192 constitutes less than 150% of the annular area of fate of the bearing in the vicinity of the groove 190 between the surfaces 140 and 142. Mathematically, this can be expressed as follows: D zk * ((4 / Tt) * (C * L)) 0.5 wherein: D = the diameter of the opening 192; k is a constant, for example greater than 1, such as 1.5; C = the diametrical play of the bearing; and L = the cup length of the rafter 190 (see reference 208 in Figures 6 and 7).
    Alternatively, this can be mathematically expressed as follows: D2 s k * ((D1 + C) ^ 2-D1 ^ 2) ^ 0,5 where: D2 = the diameter of the opening 192; k is a constant, for example a fracture fraction less than 1 such as 0.9; D1 = the diameter of the shaft at the surface 140 and C = the diametrical play of the bearing.
    While a reduction in the diameter of the aperture 192 is a preferred option, another alternative is to insert a choke structure (such as a choke plate or constrictor) into a larger dimension aperture such as the aperture 92 shown in Figure 3, this choke structure effectively providing a contracted aperture. in the manner described above. Although Figure 7 shows that each latch 190 includes an opening 192 for the lubricant channel 128, it will be appreciated that only one of the latches 190 may have an opening 192, the other latch 190 including a blind area formed on the bearing surface 140. Further, it should It will be appreciated that neither of the circumferentially offset latches 190 need to have an opening 92 toward the lubricant channel 128, provided that some other mechanism exists to ensure delivery of lubricant to the plain bearing 116.
    In a preferred embodiment, each opening 192 is axially offset to a position closer to an edge of the surface 140 of the sliding bearing 116. In other words, the openings 192 are not axially centered on the surface 140 of the sliding bearing 116. For example, the left opening 192 in Figure 7 is shown to have an axial displacement to a position closer to an upper edge 210 of the surface 140 of the sliding bearing 116, while the right opening 192 in Figure 7 is shown to have an axial displacement to a position closer to a lower edge 212 the surface 140 of the sliding bearing 116. In a preferred implementation the openings are 192 axially offset in opposite directions, as shown in Figure 7. However, it should be appreciated that both apertures 192 may be axially offset against one and the same edge (210 or 212) of the surface 140.
    Axially displacing the openings 192 in the manner described and providing the relative widths 194 and 196, increases (in comparison with Figure 4) the dimension of the outer surface 140 of the first plain bearing 116 which is axially adjacent to the groove 190 and is along the paths shown by arrows 198. The increased dimension of the outer surface 140 limits the boat surface fett of grease and the passage of a pressure pulse intermediate storage system (at surfaces 160, 164 and 168) and the sealing system 132 (at surface 154). As a result of the axial displacement, the increased dimension of the surface 140 at each arrow 198 (in an axial direction) provides a relatively shorter damping zone 200 on one side of the groove 190 and a relatively longer damping zone 202 on the other side of the groove 190. This configuration with longer attenuation zones 202 provide improved performance compared to the configuration of Figure 4 iterms of an attenuation of the fate of fat due to the axial passage of the pulse pulse. The additional attenuation as a hair tube from the presence of the relatively longer attenuation zones 202 further assists in the protection of the sealing system 132 (at the surface 154) from the pressure pulse and supports the attenuation function of the pre-pressure compensator 30 (see Figure 1). In a preferred implementation, the ratio between the axial width of the relatively longer damping zone 202 and the axial width of the relatively shorter damping zone 200 is between about 3 to 1 and about 6 to 1. It is preferred that the axial displacement of the grooves 190 should retain at least a small dimension. of peripheral axial overlap 216 intermediate grooves, especially in cases where one of the grooves is a blind groove without any opening 192 (but it should also be appreciated that no axial overlap 206 may need to be necessary in some implementations).
    The peripheral displacement of the two grooves 190 along the relative widths 13 194 and 196 and the axial displacement of the grooves 190 further provide an additional attenuation zone 204, which is circumferentially located between the two grooves 190. The degree of peripheral displacement is chosen in such a way that the peripheral pressure damping between the grooves is approximately equal to the axial pressure damping between a groove and an additional breath of the bearing. In other words, the peripheral displacement of the grooves 190 is chosen so that it is about as difficult for the pressure pulse for grease to travel between the spirit of the bearing system and the track longitudinal for arrow 198 as it is for the pressure pulse for grease to travel between tracks longitudinal for arrow 206. sowed year both possible trajectories for the fat due to imprints essentially equally damped.
    When the roller pump movement occurs, the lubricant supplied to the bearing system is compressed by the interstitial volume (with the shaft 116 surfaces 140, 160, 164 and 168). This leads to the generation of a pressure pulse. In response to the pressure pulse fl lubricant grease destroys through the series of lubricant channels 28 the intermediate bearing system and the pressure compensator 30 (see Figure 1). The pressure compensator 30 is designed to attenuate or relieve the pressure pulse by compensating for changes in volume through its elastomeric membrane. However, the paths provided by arrows 198 and 206 are also available for fat loss in response to the pressure pulse. The attenuation zones 200, 202 and 204 are provided to limit the flow of fat along these paths and thus reduce or eliminate the pressure pulsation due to roller pumping from acting on the sealing system 132. Although Figures 5-7 specifically illustrate the use of a plain bearing system, it will be appreciated that the grooves 190 (with or without openings 192) may alternatively be used in conjunction with a roller bearing system.
    Furthermore, although Figures 5-7 specifically illustrate the provision of grooves 190 (with or without openings 192) in connection with the main bearing in the bearing system (either sliding or rolling), it will be appreciated that the grooves 190 (with or without openings 192) may alternatively be provided. in conjunction with any suitable bearing surface of shaft 116 (including, but not limited to, surfaces 140, 160, 14, 164 and 168) in either a plain bearing or roller bearing implementation. Although explained in the context of a drilling tool designed primarily for use in a drilling application at an oil field, it should be understood that the description is not limited thereto and that the bearing system described can be used in any roller drilling tool (rotary cone drilling tool), including tools such as used in applications which do not concern oil fields. Specifically, the drilling tool may be configured for use with any suitable drilling fluid, including air, mist, foam or liquid (water, sludge or oil based), or any combination of the foregoing. Furthermore, although described in this context, it is a solution to the problems associated with co-roll pumping and lubricant pressure pulsations in sealed and pressure compensated systems. The solutions described herein are equally applicable to pearl drill bits (rotary cone drill bits) which are lubricated but not pressurized. and membrane systems. Although preferred embodiments of the method and apparatus of the present invention have been illustrated in the following drawings and described in the foregoing detailed description, it should be understood that the invention is not limited to the disclosed embodiments, but is capable of numerous rearrangements, modifications, and substitutions. as defined and defined by the following claims.
    权利要求:
    Claims (21)
    [1]
    A drilling tool, comprising: a drill bit body (34): at least one bearing shaft (112) extending from the drill bit body (34): a roller (114) mounted for rotation on the bearing shaft (112); a first groove (190) formed in a non-load zone for an outer bearing surface (140) the bearing shaft (112); and a second groove (190) formed in the non-load zone for the same outer bearing surface (140) of the bearing shaft (112); the first groove (190) being circumferentially offset from the second groove (190), the outer bearing surface (140) of the bearing shaft (112) being a cylindrical surface which is axially located between a cold to a pressure pulse due to unrolling pumping and a sealing system (132) for the roller (114) and the bearing shaft (112), the first and second grooves (190) being positioned in the non-load zone of the outer bearing surface (140), so that each defines a first damping zone (200) and second damping zone (202), the first and second damping zones (200,202) axially limiting the propagation of the pressure pulse due to the roller pumping towards the sealing system (132), characterized in that the first and second grooves (190) are axially displaced relative to each other.
    [2]
    The drilling tool of claim 1, wherein the bearing shaft (112) further includes an internal lubrication channel (128) and further includes a first opening (192) within the first groove (190), the first opening (192) providing an outer connection between the first groove. (190) and the internal lubrication channel (128).
    [3]
    Drilling tool according to claim 2, wherein the first opening (192) has an end diameter D which satisfies the following equation: D ß k * ((4 / Tt) * (C * L)) ^ 0.5; wherein: coagulates a constant larger than one; C = the diametrical play of the bearing; and L = the arc length of the first track (190). 16
    [4]
    Drilling tool according to claim 2, wherein the first opening (192) has an end diameter D2 which satisfies the following equation: D2 sk * ((D1 + C) ^ 2-D1 ^ 2) ^ 0,5; wherein: k is a constant smaller than one; D1 = the diameter of the outer surface (140) of the shaft (112); and C = the diametrical play of the bearing.
    [5]
    The drilling tool of claim 2, wherein the first opening (192) has a cross-sectional area, which constitutes less than 150% of an annular. Circumferential area along the outer bearing surface (140) of the bearing shaft (112) in an environment of the first rafter (190).
    [6]
    The drilling tool of claim 2, further comprising a second opening (192) within the second latch (190), the second opening (192) providing a clear connection between the second latch (190) and the internal lubrication channel (128).
    [7]
    The drilling tool of claim 1, wherein the peripheral displacement of the first notch (190) from the second notch (190) defines an additional damping zone (204) extending circumferentially along the outer bearing surface (140) of the bearing shaft (112) between the first the rafter (190) and the other rafter (190).
    [8]
    The drilling tool of claim 7, wherein the additional damping zone (204) between the first rafter (190) and the second rafter (190) provides a peripheral damping length which is approximately equal to an axial damping width provided between either the first rafter (190) or the second latch (190) and a further end of the outer bearing surface (140) of the bearing shaft (112).
    [9]
    The drilling tool of claim 1, wherein the outer cylindrical surface (140) is a main slide bearing surface.
    [10]
    The drilling tool of claim 1, wherein the bearing shaft (112) supports a plain bearing (116).
    [11]
    The drilling tool of claim 1, wherein the outer bearing surface (140) of the bearing shaft (112) is axially defined between a first edge (210) and a second edge (212), and 17 wherein the first groove (190) is axially offset and located closer to the first edge (210) of the outer bearing surface (140) than the second rafter (190) and the second rafter (190) is located closer to the second edge (212) than the first rafter (190).
    [12]
    The drilling tool of claim 11, wherein the first damping zone (200) extends along the outer bearing surface (140) of the bearing shaft (112) axially between that first edge (210) and the first rafter (190) and between the second edge (212) and the second rafter (190) and the second damping zone (202) along the outer bearing surface (140) of the bearing shaft (112) extend axially between the first rafter (190) and the second edge (212) and between the first edge (210) and the second rafter (190).
    [13]
    The drilling tool of claim 1, wherein each of the first and second rafters (190) has an axial width (194), and wherein the axial width (194) of each of the first and second rafters (190) does not constitute more than 70% of an axial width (196) of the outer bearing surface (140) of the bearing shaft (112) between the first and second edges (210, 212).
    [14]
    The drilling tool of claim 1, each of the first and second rafters (190) having an axial width (194) and a peripheral length (208), and wherein a ratio between the peripheral length (208) and the axial width (194) ) for each of the first and second latches (190) being between about 2 to 1 and about 4 to 1.
    [15]
    The drilling tool of claim 12, wherein a first axial width of the first damping zone (200) differs from a second axial width of the second damping zone (202).
    [16]
    The drilling tool of claim 15, wherein a ratio between the first axial width and the second axial width is between about 3 to 1 and about 6 to 1.
    [17]
    The drilling tool of claim 7, wherein the first, second and additional damping zones (200, 202, 204) axially limit the propagation of the pressure pulse due to the roller pumping towards the sealing system (132). 18
    [18]
    The drilling tool of claim 2, wherein the first opening (192) provides propagation of the pressure pulse due to the roller pumping.
    [19]
    The drilling tool of claim 1, wherein the sealing system (132) comprises a annular sealing gasket box (152) and a sealing member (150) received in the annular sealing gasket box (152).
    [20]
    The drilling tool of claim 19, wherein the sealing member (150) is of the O-ring type.
    [21]
    The drilling tool of claim 1, wherein the first and second grooves (190) have a peripheral axial overlap (216).
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    公开号 | 公开日 | 专利标题
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    CN103328758A|2013-09-25|
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    法律状态:
    优先权:
    申请号 | 申请日 | 专利标题
    US13/016,399|US8534389B2|2011-01-28|2011-01-28|Method and apparatus for reducing lubricant pressure pulsation within a rotary cone rock bit|
    PCT/US2011/059189|WO2012102772A1|2011-01-28|2011-11-03|Method and apparatus for reducing lubricant pressure pulsation within a rotary cone rock bit|
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