Earth Boring Drill Bit With Snap Ring Cutter Retention

1. Field of the Invention:

This invention relates in general to earth boring drill bits and in particular to improved bearings, the means for retaining the rotatable cutters on such bits, and their lubrication.

2. Prior Art:

The success of rotary drilling enabled the discovery of deep oil and gas reservoirs. The rotary rock bit was an important invention which made that success possible. Only the soft formations could be commercially penetrated with the earlier drag bit, but the two cone rock bit invented by H. R. Hughes, U.S. Pat. No. 930,759, drilled the hard cap rock at the Spindletop Field near Beaumont, Texas with relative ease.

That distant invention, within the first decade of this century, could drill a scant fraction of the depth and speed of the modern rotary rock bit. If the original Hughes bit drilled for hours, the modern bit drills for days. Bits today sometimes drill for miles instead of feet. Many advances contributed to the impressive improvement in the performance of rock bits.

The original bit of Hughes had rotatable cutters, frictional bearing surfaces and ring-type retainers generally threaded for retention in the cutter. Because of the difficulty in providing seals with long life, relatively large lubricant reservoirs were required. By the decade of the thirties, drill bits with anti-friction bearing that were unsealed became commercially successful and replaced the sealed and lubricated drill bits. This type drill bit may be seen in the U.S. patent to Lewis E. Garfield et al., Pat. No. 2,030,442.

G. O. Atkinson et al obtained U.S. Pat. No. 3,075,781 on a seal that minimized leakage and required only a small reservoir of lubricant. The pressure compensator is an improvement that minimizes the pressure differential across the seal and increases its reliability. A recent version of the pressure compensator and pressure relief system is shown in the U.S. Pat. of Stuart C. Millsapps, Jr., U.S. Pat. No. 3,942,596. E. M. Galle patented an O-ring type seal, U.S. Pat. No. 3,397,928, that ultimately made friction bearings once again feasible in drill bits. An illustration of a recent bearing configuration is shown in the U.S. Pat. No. 3,922,038 issued to S. R. Scales. This bit has a cylindrical, friction bearing and a cutter retained by a ball bearing similar to that shown in the above patent to Lewis E. Garfield et al.

A ball bearing in a lubricated rock bit bearing has inherent disadvantages. A failure in any one of the numerous balls or the raceway in which they are positioned may permit metallic fragments to enter the friction bearing, with near certain damaging results. Metallic particles will often damage the seal ring, causing lubricant loss and rapid bearing failure. The necessity for a hole drilled into the ball race for introduction of the balls, retained by a welded plug, adds complexity to the bit that provides additional room for manufacturing error. The rock bit is the focal point of an expensive drilling process that has a low tolerance for failure.

There is shown in the prior art rock bit bearing and retainer means that are totally frictional; that is, without anti-friction ball or roller bearings. The original Hughes bit had friction bearings exclusively. A few of the many bits with only frictional bearings will be mentioned briefly.

F. L. Scott in U.S. Pat. No. 1,803,679 discloses a lubricated, tapered frictional rock bit bearing that retains its cutter by means of a resilient snap ring. J. C. Wright et al in U.S. Pat. No. 2,049,581 discloses a cutter rotatably secured with a snap sing to a bearing shaft releasably connected to the legs or head sections of a bit. The U.S. Pat. No. 2,814,465 to W. G. Green discloses a lubricated frictional bearing having a tapered and cylindrical portions joined by suitable means and secured to a cutter by a retainer ring. A lock ring is shown in Edward B. Williams, Jr.'s U.S. Pat. No. 3,844,363 for retaining a cutter and shaft with frictional bearings to a rock bit. A frictional bearing with tapered and cylindrical portions is disclosed by E. M. Galle in his U.S. Pat. No. 3,361,494, along with frictional plug lock and pin lock retaining systems. A recent revival of the earlier seen threaded ring retainer is seen in U.S. Pat. No. 3,971,600 of Murdoch et al. Notwithstanding this array of ideas, the only significantly successful rock bit in commerce today uses a ball bearing retainer.

The threaded retainer ring has some disadvantage that may have limited its success. The threads act as stress raisers and the ring must be secured against rotation while the cone is mated with it. This generally requires a drilled hole in the head section.

Snap rings may have stress raising groove configurations and exhibit a tendency to retract during drilling into the assembly groove and permit accidental cutter loss. When a cutter is thrust inward, as when remaing for example, the leading of the ring causes stresses that tend to urge the ring into the assembly groove, with the possibility of cutter loss.

The invention may be summarized as an improved rock bit having exclusively friction bearing means between the rotatable cutter and bearing shaft. The cutter and shaft are secured by the snap ring between registering grooves, the ring having an opening for the passage of lubricant. Lubricant is supplied by a system that includes a compensator that controls the pressure differential across a seal between the cutter and shaft. This produces a high strength bearing and retainer means that in combination with the lubrication system can be used to simplify construction and increase reliability.

Additional objects, features and advantages of the invention will become apparent from the following description.

FIG. 1 is a fragmentary view, partially in section, of a rock bit which embodies the features of the invention.

FIG. 2 is a full view of a snap ring used to rotatably retain the cutter on the bearing shaft or pin of the rock bit of FIG. 1.

FIG. 3 is an enlarged, fragmentary view in longitudinal section of a portion of the cutter and the bearing pin of FIG. 1.

FIG. 4 is a fragmentary view, in longitudinal section, of a region of the cutter and bearing pin around the cross-sectioned snap ring, of FIG. 1, the cutter being thrust outwardly on the bearing shaft.

FIG. 5 is, like FIG. 4, a fragmentary view, in longitudinal section but with the cutter thrust inwardly on the bearing shaft.

FIG. 6 is a fragmentary view, in longitudinal section, of a second construction with the cutter thrust inward.

FIG. 7 is a fragmentary view, in longitudinal section, of a third construction, with the cutter thrust inward.

Portions of an earth boring drill bit 11 are shown in FIG. 1, including a body 13 formed of three head sections 15 there are typically joined by a welding process. Threads 17 are formed on the top of body 13 for connection to a conventional drill string (not shown). Each head section 15 has a cantilevered shaft or bearing pin 19 having its unsupported end oriented inward and downward. A generally conically shaped cutter 21 is rotatably mounted on each bearing pin 19. Cutter 21 has earth disintegrating teeth 23 on its exterior and a central opening or bearing recess 25 in its interior for mounting on the bearing pin 19. Friction bearing means formed on the bearing pin 19 and cutter bearing recess 25 are connected with lubricant passages 27. (See U.S. Pat. Nos. 3,397,928 and 4,012,238 for bearing surface treatments.) A pressure compensator 29 and associated passages constitute a lubricant reservoir that limits the pressure differential between the lubricant and the ambient fluid which surrounds the bit after flowing through the nozzle means 32. (See the copending application of Stuart C. Millsapps, Jr., Lubricant Pressure Compensator For An Earth Boring Drill Bit, Ser. No. 687,131, filed May 17, 1976 now U.S. Pat. No. 4,055,225.) An O-ring seal 31 located between each bearing pin 19 and cutter 21 at the base of the bearing pin prevents egress of lubricant and ingress of borehole fluid. (See U.S. Pat. No. 3,397,928 for O-ring seal discloure and U.S. Pat. No. 3,935,114 for lubricant.)

Referring to FIGS. 3, 4 and 5, an annular assembly groove 33 is formed on the cylindrical surface 34 of the bearing pin 19. Groove 33 has a circular bottom portion 35 formed about a radius located on centerline 41. An outer and an inner sidewall designated respectively sidewall 37 or 39 extend from the semi-circular bottom portion 35 to the surface of the bearing pin 19, outer sidewall 37 being the one farther from the inner end or nose 43 of the bearing pin 19. Outer sidewall 37 is planar and perpendicular to the coaxis (not shown) of the bearing pin 19 and cutter 21.

Inner sidewall 39 is a conical thrust surface or inclined wall nonperpendicular to the axis of the bearing pin, oriented at an angle .alpha. of preferably 15.degree. with respect to a plane perpendicular to axis of the bearing pin. This produces an entrance to groove 33 larger than the diameter of its circular bottom portion 35 and is a beneficial thrust surface as will be explained subsequently.

A registering retainer groove 45 is formed in the bearing recess 25 of cutter 21. Groove 45 has sidewalls 46 and 50, with outer sidewall 46 located farthest from nose 43 of bearing pin 19. As seen in section, sidewall 46 has the form of a circular arc equal to one quarter of a circle, with a radius substantially equal to the radius of the bottom portion 35 of the previously described groove 33. The depth of groove 45 is greater than the radius of the circular arc. As seen in section in FIG. 4, inner sidewall 50 consists of a straight line connected at one end by means of a small radius with the quarter-circular arc of sidewall 46 and its other end extends to form one surface of a tapered or conical frictional nose bearing 52.

Grooves 33 and 45 are located so that they register to define an irregularly shaped annular cavity with curved bottoms. A snap ring 47 with a curved boundary and preferably circular cross-section is located in this cavity. It is formed of resilinet metal, preferably piano wire, and contains a gap 49 (see FIG. 2) so that its annular diameter may be compressed or expanded and also so that lubricant may flow past the ring. The annular diameter of ring 47 is selected so that it will expand tightly into groove 45 when relaxed. The cross-sectional diameter of 47 is substantially equal to the diameter of the circular portion 35 of groove 33.

The depth of assembly groove 33 is at least as large as the cross-sectional diameter of ring 47. This enables ring 47 to be stressed and retract fully into grooves 33 for assembling cutter 21 over bearing pin 19. The depth of retainer groove 45 is greater than one-half the crosssectional diameter of ring 47 but less than its cross-sectional full diameter. This depth is greater than the thickness of one-half the diameter of the cross section of ring 47 by an aount, conveniently called the "offset", which in the preferred form is in the range from 0.005 inch to 0.020 inch. This places the cross-sectional centerline 51 of ring 47 outside the cylindrical surface 34 of the bearing pin and also outside the edge of the retainer groove by the same quantity.

To assembly the cutter 21 over bearing pin 19, a tool (not shown) is used to compress ring 47 into groove 33. The cutter 21 is then inserted partially over the bearing pin 19 and the tool removed. Cutter 21 is then forced outwardly and against the bearing pin until engagement of the opposing surfaces of the tapered nose bearing 52, as shown in FIG. 4. Groove 45 will be sufficiently aligned with groove 33 for ring 47 to snap into it. Ring 47 is shown in FIG. 4 against sidewall 46 of groove 45, and is not touching sidewall 50. Sidewall 39 of groove 33 also is separated from the ring by a clearance space C.sub.1 to assure movement into retainer groove 45. The centerline 51 of ring 47 is offset to a position outside the surface 34 of the bearing pin 19 by the preferred distances of 0.005 to 0.020 inch.

Centerlines 41 and 51 do not lie on a common plane perpendicular to the bearing pin axis. Rather, centerline 51 is spaced closer to the nose 43 than centerline 41. Because of clearance C.sub.1 the cutter 21 may move inward on bearing pin 19 to the position shown in FIG. 5. Here, ring 47 is confined between sidewall 46 of groove 45 and the corner of sidewall 39 of groove 33. A clearance C.sub.2 appears between the opposing surfaces of conical or tapered thrust bearing 52.

The distance from the offset annular edge of wall 39 of the assembly groove 33 to the opposite corner or outer wall 46 in retainer groove 45 is less than the thickness or diameter of snap ring 47 when the cutter is thrust inward to the position shown in FIG. 5. That is, the diameter of the snap ring in FIG. 5 is represented by the letter x+y. The distance from the edge of wall 39 to the opposite corner of groove 46 is a lesser distance x. Hence, when the cone is thrust inward, the ring cannot be retracted into the assemby groove 33 but must remain in the retainer groove 45 to prevent accidental cutter loss.

In operation, the most normal drilling condition produces outard thrust of cutter 21 on bearing pin 19 as seen in FIG. 4. No external forces are exerted on ring 47 due to engagement of the surfaces of tapered bearing 52 and the resulting clearance C.sub.3 between the ring 47 and wall 37 of groove 33. If thrust on the cutter is inward, as sometimes occurs when reaming an undersized hole, the tendency will be for the cutter 21 to be pushed inwardly and from the bearing pin 19. This tendency will be resisted by ring 47, which transmits the thrust between sidewall 46 and sidewall 39 as shown in FIG. 5. Inner sidewall 39 serves as a thrust surface, providing a compressive reacting force.

The forming of inner sidewall 39 at an angle .alpha. generates a force component F (see FIG. 6) at the same angle .alpha. and in a direction oblique to the axis of rotation of the cutter 21. Thus, the configuration of the groove 33 produces a three dimensional conical force distribution F that expands the ring 47 outward into groove 45 in cutter 21. This prevents compression of ring 47 into the groove 33 of bearing pin 19 and the consequent loss of cutter 21.

FIG. 6 is a second construction from that shown in 33' FIGS. 4 and 5 in that the walls 37', 39' of assembly groove 33' are parallel and formed at the angle .alpha.. The "offset" is smaller in FIG. 6 and is not necessarily utilized since here the tapered wall 39' exerts a force component F normal to the surface of ring 47' to urge the ring into retainer groove 45'.

In FIG. 7 is shown a third construction in which the walls 37", 39" of assembly groove 33" are parallel and perpendicular the centerline of the bearing pin. Here the edge or outer wall 39" of groove 33" engages the ring 47" at a distance from the centerline 51" of the ring 47" equal to the "offset" from cylindrical bearing surface 34". A force component F is exerted obliquely against the ring, at an angle normal to the surface of ring 47", to urge the ring into the retainer groove 45' when the cone is thrust inward.

The preferred construction of FIGS. 4 and 5 is a combination of the features of FIGS. 6 and 7 in that wall 39 (but not wall 37) of groove 33 is tapered and the "offset" feature is also used. The quantity of offset in the specified preferred range is such that angle .alpha. of preferably 15.degree. results in surface 39 contact with the ring rather than sharp edge contact. This provides a superior wear condition since a sharp edge will wear the ring more rapidly.

Each of the three constructions of FIG. 4 through 7 have a feature that further minimizes loss of a cutter, as previously explained in connection with FIG. 5. That is, the cross-sectional thickness x+y of th ring 47 is greater than the distance x from the edge of wall 39 to the opposite corner of groove 46 when the cutters are thrust inwardly on their bearing shafts.

It should be apaprent from the foregoing that improvements having significant advantages have been made. The utilization of a bearing structure which consists of lubricated frictionnal bearing surfaces, including a frictional retainer means, simplifies construction and increases reliability. The utilization of the disclosed snap ring and groove configuration minimizes the danger of loss of the cutter. The previously described offset position between the centerline of the ring and the surface of the bearing near the assembly groove results in an oblique force component that urges the ring into its retainer groove. The smooth angular thrust surface in the assembly groove to engage the ring when the cone is thrust inward provides an advantageous wear environment. Further, the inclined orientation of this surface, which acts as a thrust surface, also urges the retainer ring into its retaining groove when the cone is thrust inward. The ring configuration and thickness being lesser than the distance across the groove assures cutter retention.

While the invention has been shown in only one of its forms, it should be apparent to those skilled in the art that is is not thus limited but is susceptible to various changes and modifications without departing from the spirit thereof.