CA1292220C - Earth boring bit with two piece bearing and rigid face seal assembly - Google Patents
Earth boring bit with two piece bearing and rigid face seal assemblyInfo
- Publication number
- CA1292220C CA1292220C CA000559193A CA559193A CA1292220C CA 1292220 C CA1292220 C CA 1292220C CA 000559193 A CA000559193 A CA 000559193A CA 559193 A CA559193 A CA 559193A CA 1292220 C CA1292220 C CA 1292220C
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- Canada
- Prior art keywords
- cutter
- bearing
- seal
- shaft
- movement
- Prior art date
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Abstract
ABSTRACT OF THE DISCLOSURE
An earth boring bit having a cantilevered bearing shaft and a compensator system to equalize the pressure of the lubricant with the hydrostatic pressure of the drilling fluid surrounding the bit. A rigid face seal assembly, positioned between the cutter and bearing shaft of the bit, moves axially in response to, and to compensate for, dynamic pressure changes in the lubricant adjacent the seal. The positioning and sizing of resilient energizer rings in relationship to the geometries of the mating grooves between the cutter and shaft and the rigid sealing rings of the face seal assembly minimize axial seal movement relative to cutter and shaft during drilling to enhance seal life. A threaded bearing lug receives an internally threaded bearing sleeve. The journal bearing surface is approximately midway between the shaft seal groove, a portion of which is on its bearing sleeve and another portion on the cutter seal groove.
An earth boring bit having a cantilevered bearing shaft and a compensator system to equalize the pressure of the lubricant with the hydrostatic pressure of the drilling fluid surrounding the bit. A rigid face seal assembly, positioned between the cutter and bearing shaft of the bit, moves axially in response to, and to compensate for, dynamic pressure changes in the lubricant adjacent the seal. The positioning and sizing of resilient energizer rings in relationship to the geometries of the mating grooves between the cutter and shaft and the rigid sealing rings of the face seal assembly minimize axial seal movement relative to cutter and shaft during drilling to enhance seal life. A threaded bearing lug receives an internally threaded bearing sleeve. The journal bearing surface is approximately midway between the shaft seal groove, a portion of which is on its bearing sleeve and another portion on the cutter seal groove.
Description
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1 CROSS REFERE~CE TO RELATED PATENTS
1 CROSS REFERE~CE TO RELATED PATENTS
3 This application is related to "Earth Boring Bit 4 With Improved Rigid Face Seal Assembly", U.S. Patent No.
4,666,0Ul of May 19, 1987, which is related to "Earth 6 Boring Bit With Pressure Compensating Rigid Face Seal", 7 U.S. Patent No. 4,516,641 of May 14, 1985.
8 This application has disclosure in common with the 9 copending Canadian application of Joseph L. Kelly, Jr., entitled "VOL~ME ~ND PRESSURE BALANCED RIGID FACE SEAL
11 FOR ROCK BITS", Serial No. 559,906, filed February 26, 12 198~.
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BACKGROUND OF THE INVENTION
~ 1. Field of the Invention:
This invention relates to earth boring bits, lubricated ~vith a system ~ which includes a hydrostatic pressure compensator to balance the internal 6 pressure of the lubricant inside the bit with the hydrostatic pressurc of a 7 liquid drilling fluid that surrounds t~le bit during drilling. In this 8 cornbination, the specific improvement relates to the seal and bearing o assembly bctween cach clltter and bearing shaft.
2. Background In~ornt~tion:
2 The preferred embodiment in the above U.S. Patent No. 4,516,641 ~s utilizes a pair of rigid face seals positioned in a seal groove, including a 14 pair of resilient energizer rings, preferably of the O-ring type. The ~5 dimensional relationships o~ the sealiDg cornponents and the groove result in16 greater axial movement of the rigid face seals than the associated cutter.
17 As a bit rotates during drilling, its cutters move axially, or with a 18 rocking motion, on the bearing shafts because of the clearances and normal manufacturing tolerances. Some clearances are necessary to assemble the 20 cutters on the shafts. Axial and radial cutter movements which results from 21 the clearance between cutter and shaft c3uses rapid pressure variations in 22 the lubricant, or more accurately, voJume changcs in the lubricant in the 23 vici~ity of each seai. In the preferred embodiment of the above patent 2~ 4,516,641, the sisid face seals m~y move axialiy a distance greater than thc 25 sxial cutter ~ovemcnt by a ratio of about 1.88 to one.
26 Ihe disclcs~x irl the E~ of t:he a~ve U.S. Patents, No.
2Y 4,666,00l, teach~= ~he pOsitio~ ~ of the seal groc~ve an~l seal 28~a'S~51ibly in :relation to the ~ surfaoe such that rigid face s3 seal m~nt ~,~ decreased rela~ive ~ to axial cuttex mav~t ~o du~ing drilling to ~ anoe seal life. l~referably the r~tio of Sl ri~id r~r~ t to ax~al cutte~ mav~t is s~stan~ially one half ~o ar)e.
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1 A two piece bearing çonstruction is disclosed in U.S. Patent No.
2 4,600,064, "Earth Boring Bit With Bearing Sleeve", July 15, 1986. An 3 internally tapered and threaded bearing sleeve is made up on a mating, 4 externally threaded bea~ing lug. The mouth of the sleeve engages a shoulder 5 on the base region of the the bearing lug and has a selected radial thickness 6 such that the sleeve may be made up to a selected torque. Also, the bearing 7 sleeve has a length ~reater than that of the threaded portion to define a 8 thick walled inner end region to receive a resilient retainer ring in a groove D that provides a setected section over the threads on the ;nterior of the 10 sleeve.
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3 The general object of this invention is to provide an improved rigid 4 face seal assembly in a rock bit of the type susceptible to pressure or volume5 changes in the vicinity of the seal assembly when the associated cutter 6 rno~es durin~ drilling. The improvement decreases ~he pressure or volume 7 changes in the vicinity of the seal assembly by positioning the seal groove 8 and seal assembly in relation to the bearing surface such that rigid face sealmovement is decreased relative to axial cutter movement. In a preferred embodiment the axial movement of the rigid face seal with respect to the shaft is about one half that of the cutter, and is achieved by making the 12 bearing surface intermediate the radial thickness of the seal assembly 13 groove. Th~t is~ part of the groove is in the cutter and the other is in the 14 bearing shaft. Thc relationship of seal movement to cutter movement is 15 established by dimensioning consistent with the following formula, derived 16 for this invention l7 S- 2G-C
18 ~i+C+2A
Where: S = Seal Movement For Unit Cone Movement Relative to the Sh~ft 21 H = Effective Shaft O-Ring Annular Area 22 C - Effective Cutter O-Ring Annular Area 23 A = Effective Rigid Ring Annular Area 21 G = Effective Cutter Se;3l Groove Annular Area 26 The preferred embodiment uses a threaded bcaring lug upon which is 27 threaded a bearing sleeve containing about one half the radi~ hiclcness of 21~ the total sesl groove. This portion of the scal groove is called ~he shaft groove. The other radial portion Or the tot~l seal groove is formed in the 30 cutter.
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Addii;onal objects, features and advantages of the invention will 2 become ~pparent in the following description.
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DESCR~PTION OF THE DRAWINGS --3 Fig. 1 is a Yiew in longitudinal section of a portion of an earth 4 boring bit, showing the compensator system, bearing shaft, cutter and one ;, embocliment of a seal assembly.
s Fig. 2 is a fragmentary Yiew in longitudina1 scction of the lower 7 portion of a bit, enlarged with respect to Fig. 1, to better expose the seal 8 assembly 9 Fig. 3 is a fragmentary view in longitudinal section of yet another 10 portion of the cutter and bearing shaft~ showing the bearing seal assembly enlarged with respect to the illustration of Fig. 2.
Fig. 4 is an enlarged longitudinal section of one of the rigid rings of the seal assembly.
Fig. 5 is a fragmentary, longitudinal section of the seal seat and conical surface seal groove in the cutter.
Fig. 6 is a longitudinal section of the seal seat annular insert used on 17 the bearing shaft to form the conical, contoured surface that receives and 18 positions the scal assembly.
19 Fig. ~ is a fragmentary Yiew in longitudinal section of the lower 20 portion of an alternate embodiment of the invention shown in Fig. 1.
21 Fig. 8 is a vicw in longitudinal section of a portion of an earth 22 boring bit having an improved rigid face seal assembly to reduce during 2s drilling the amount of seal movemcnt relative to cutter movemcnt.
24 Fig. 9 is an cnlarged fragmcntary view in longitudinal section of the 25 preferred seai assernbly and groove embodiment shown in Fig. 8.
26 Fig~ lû is a fragmentary view of another embodimcnt of an earth 27 boring bit, partially in longitudinal section, that utilizcs a threaded bearing 2~ sleeve to form the journal bearing of the shaft and a portion of seal 20 groove, 3c Fig. I l is an enlarged fragmentary vicw in longitlldinal section of the ~I bearing, cutter and seal asscmbly ~ F~g. 10.
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DESCRIPTION OF THE PREFERRED ~MBODIMEN~S --2 Figs. 1-7 3 The improvement of this invention is best appreciated by initial 4 reference to the disclosure of one of the related paterlts, U.S. Pat~nt s 4,516,641, which is shown in Figs. 1-7 herein. The numeral 11 in Fig. 1 6 designates a lubricated, rotatable cone or cutter type earth boring bit having7 a body formed in three head sections or le&s 13, only one of which is shown.
8 Each leg 13 includes an oblique cantile~ered bearing sllaft 15 that depends g inwardly and downwardly tc support a rotatable cutter 17 haYing earth 10 disi~tegrating teeth 19. Lubricant passage 21 supplies lubricant to the bearing surfaces between the bearing shaft IS and cutter 17. A seal assembly 23 retains lubricant in the bearing and prevents borchole fluid 13 from entering the bearing. A hydrostatic pressure compensator is part of a lubrication system 25 connected with the lubricant passage 21 to equalize IS the pressure of the liquid lubricant inside the bearing with the hydrostatic pressure of the fluid in the borehole. A preferred compensator system is 17 shown in the patent to Stuart C. Millsapps, Jr., U S. Patent No. 4,276,946.
18 The geornetry of the bearings on the shaft 15 and within the cutter 17 are of a pr;or art configuration, including the use of a b~ll bearing 20 retainer 27, which with a plug 26 welded at 28 ret ins the cutter on the 21 bearing shaft, preferably as shown in the U~S. patent of Robert A.
2~ Cunningham, No. Re. 28,625.
23 Referring especially to Fig. 3, the cueter and shaft include an 24 annular seal groove or gland th~ t has axially spaced, generally radial end 25 walls ~9, 31 and inncr and outer circumfcrential walls 33, 35. End wall 31 20 and circumferential wall 33 are formed upon a seal seat inscrt 3~ secured to 27 the bearing shaft 15.
28 The seal assembly includcs a pair of annular rigid rings 37, 39 with 29 opposed radial faces 41, 43. The pair of rigid, prefcrably mctal, rings have 30 a radially measurcd thickncss lcss than ~hc annular spacc bctwccn the inner 5~ and outer side walls 33, 35 of thc groove and an axially measurcd width .
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1 which is less than the wid~h or the distance between the cnd walls 29, 31 of 2 the groove, as will be explained more fully hereafter.
3 Each of a pair of resilicnt, energizer rings 45 or 47 extends betweeD
4 a seal seat 49 or 51 on one of the metal rings 37 or 39 and an opposed seal ~ seat 53 or 55 on the inner or outer circumferential walls 33, 35. Each seal 6 seat has an annular groove and configuration to position and retain the 7 associated energizer ring and the metal ring, which is suspended within the 8 groove intermediate the circumferential walls 33, 35 and the end walls 29, 9 31 to provide clearances Cl and C2 which exist when the thrust surfaces 32 o are in contact. The positions of seats 53 and 55 relative to each other are 1l selected such that, at assembly, the initial deflection of each seal half 12 relative to its adjacent end wall 29 or 31 will provide sufficient contact 13 pressure between radial faces 41 and 43 to maintain sealing contact between 1,~ all the elements of the seal assembly throughout the full range of seal 15 movements permitted by clearances Cl and C2 and the play between cutter 16 and shaft. See Patent No. 3,180,648 for a description of an earlier 17 construction of a conical, "Duo Cone" seal arrangement of the Caterpillar 18 Tractor Company, and patent Nos. 3~403,916, 3,524,654 and 4,077,634 for 19 improvements to such seals.
From Fig. 3 it may be seen that one of the metal rings 37, 39 is 21 inverted with respect to the other, a feature which permits the seal assembly22 to span the groove diagonally and engage opposite circumferential walls 33, 2~ 35. The clearances Cl and C2 are between each of the end walls 29, 31 of 2~ the groove al~d the engagcd rigid rings 37, 39. Drilling fluid fills the space 25 57 and acts upon the outcrmost side of the scal assembly 23, and lubricant 26 fills the space 59 and acts upon the ;nncrmost sidc of the assembly. The 27 rigid rings 37 or 3~ have a beveled, substantially conical portion on the 28 lubricant side of thc scal assembly to define a space 61 to feed lubr;cant to29 the en8aged radial faccs 41, 43, which regencrate inwardly as thcy wear in 30 serYice. Sce U.S. Patent No. 3,180,648 for a dcscription of one configur;ltion 31 of such seal faces.
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The dimensions provided below relate to the bit used in the first 2 field test of the invention, which was a "Hughes" 12-1/4 inch, J22 type bit.
3 With reference especially to Figs. 4 and 5, the radial thickness T of each of 4 the t}lree metal rings 37 was 0.200 inch, the axial width W was about 0.270 inch and the outside diametcr was about 3.449 inches. Angle d was ~bout s twenty degrees and the radii Rl and R2 averaged 0.048 and 0.080 inch 7 respectively, Rl being tangent with the conical surface 63. The axial 8 dimensions Y and Z averaged respectively 0.050 and 0.149 inch. The depth 9 D of the positioner groove 65 below the lip 67 averaged 0.009 inch to help o position and confine the associated energizer ring 4~ during assembly. The 11 radial thickness X o~ the radial sealing face 41 was about 0.050 incll with a12 surface finish of about one or two RMS and a tapered surface defined by a 13 spherical radius R3 of about 80 inches with a surface finish of about two or 14 three RMS.
The inverted and opposing rnetal rings 39 had a radial thickness T of .5 about 0.199 inch, a radial width W of about 0.247 inch and an outside 17 diameter of about 3.450 inches. Angle o~ was about 19 degrees, and the radii 18 Rl and R2 were both about 0.075 inch. The axial dimensions Y ~nd Z were respectively about 0.023 and 0.154 inch and the depth D of the positioner 20 groove being about 0.016 inch. There was a flat sealing face on the rings 39 21 that extended across the entire thickness T of the ring, that had a surface 22 finish similar to that of rings 37.
23 Ring 37 was purchased from Caterpillar lractor Company and is one 24 of thcir standard hard metal alloy rings. Ring 39 was manufac~ured by 25 Hughes Tool Company specifically for this invcntion from an air harderling 2f; tool Steel.
27 The configuration of thc scal scat in the cutter 17 of the bit is shown 2~ in Fig. ~. The seal seat was dcfined by a conical surface 70 having an 29 angle ~1 of about 19-112 degrces, a positioncr groove or seal seat 53 h~.~ving 30 a radius R4 of about 0.060 inch locatcd a distancc D2 of about 0.129 inch 31 from end wall 29, and a dcpth Dl of about 0.008 inch. Thc conical surface : : `
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1 70 intersected the groove 53 at a point 72 which was located radially from 2 the surface 35 a distance of about 0.02l inch.
3 A similar configuration was used for the seal seat on the bearin8 4 shaft 15, defincd in this instance by the seal seat insert 36 shown ill Fig. 65 having a thickness Tl of 0.105 inch. The positioner groove had a depth D3 6 of about 0.011 inch, l~ormed 'oy a radius R5 of about 0.060 inch located a 7 distance D4 about 0.140 inch from end wall 31. 92 was a conical allgle of 8 about 20 degrees located irl a manner similar to the conical surface 70 of g Fig. 5.
The O-rings or energizer rings 45, 47 after service had a cross-11 sectional thickness of about 0.168 inch, a hardness of about 59 durometer~
4,666,0Ul of May 19, 1987, which is related to "Earth 6 Boring Bit With Pressure Compensating Rigid Face Seal", 7 U.S. Patent No. 4,516,641 of May 14, 1985.
8 This application has disclosure in common with the 9 copending Canadian application of Joseph L. Kelly, Jr., entitled "VOL~ME ~ND PRESSURE BALANCED RIGID FACE SEAL
11 FOR ROCK BITS", Serial No. 559,906, filed February 26, 12 198~.
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BACKGROUND OF THE INVENTION
~ 1. Field of the Invention:
This invention relates to earth boring bits, lubricated ~vith a system ~ which includes a hydrostatic pressure compensator to balance the internal 6 pressure of the lubricant inside the bit with the hydrostatic pressurc of a 7 liquid drilling fluid that surrounds t~le bit during drilling. In this 8 cornbination, the specific improvement relates to the seal and bearing o assembly bctween cach clltter and bearing shaft.
2. Background In~ornt~tion:
2 The preferred embodiment in the above U.S. Patent No. 4,516,641 ~s utilizes a pair of rigid face seals positioned in a seal groove, including a 14 pair of resilient energizer rings, preferably of the O-ring type. The ~5 dimensional relationships o~ the sealiDg cornponents and the groove result in16 greater axial movement of the rigid face seals than the associated cutter.
17 As a bit rotates during drilling, its cutters move axially, or with a 18 rocking motion, on the bearing shafts because of the clearances and normal manufacturing tolerances. Some clearances are necessary to assemble the 20 cutters on the shafts. Axial and radial cutter movements which results from 21 the clearance between cutter and shaft c3uses rapid pressure variations in 22 the lubricant, or more accurately, voJume changcs in the lubricant in the 23 vici~ity of each seai. In the preferred embodiment of the above patent 2~ 4,516,641, the sisid face seals m~y move axialiy a distance greater than thc 25 sxial cutter ~ovemcnt by a ratio of about 1.88 to one.
26 Ihe disclcs~x irl the E~ of t:he a~ve U.S. Patents, No.
2Y 4,666,00l, teach~= ~he pOsitio~ ~ of the seal groc~ve an~l seal 28~a'S~51ibly in :relation to the ~ surfaoe such that rigid face s3 seal m~nt ~,~ decreased rela~ive ~ to axial cuttex mav~t ~o du~ing drilling to ~ anoe seal life. l~referably the r~tio of Sl ri~id r~r~ t to ax~al cutte~ mav~t is s~stan~ially one half ~o ar)e.
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1 A two piece bearing çonstruction is disclosed in U.S. Patent No.
2 4,600,064, "Earth Boring Bit With Bearing Sleeve", July 15, 1986. An 3 internally tapered and threaded bearing sleeve is made up on a mating, 4 externally threaded bea~ing lug. The mouth of the sleeve engages a shoulder 5 on the base region of the the bearing lug and has a selected radial thickness 6 such that the sleeve may be made up to a selected torque. Also, the bearing 7 sleeve has a length ~reater than that of the threaded portion to define a 8 thick walled inner end region to receive a resilient retainer ring in a groove D that provides a setected section over the threads on the ;nterior of the 10 sleeve.
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3 The general object of this invention is to provide an improved rigid 4 face seal assembly in a rock bit of the type susceptible to pressure or volume5 changes in the vicinity of the seal assembly when the associated cutter 6 rno~es durin~ drilling. The improvement decreases ~he pressure or volume 7 changes in the vicinity of the seal assembly by positioning the seal groove 8 and seal assembly in relation to the bearing surface such that rigid face sealmovement is decreased relative to axial cutter movement. In a preferred embodiment the axial movement of the rigid face seal with respect to the shaft is about one half that of the cutter, and is achieved by making the 12 bearing surface intermediate the radial thickness of the seal assembly 13 groove. Th~t is~ part of the groove is in the cutter and the other is in the 14 bearing shaft. Thc relationship of seal movement to cutter movement is 15 established by dimensioning consistent with the following formula, derived 16 for this invention l7 S- 2G-C
18 ~i+C+2A
Where: S = Seal Movement For Unit Cone Movement Relative to the Sh~ft 21 H = Effective Shaft O-Ring Annular Area 22 C - Effective Cutter O-Ring Annular Area 23 A = Effective Rigid Ring Annular Area 21 G = Effective Cutter Se;3l Groove Annular Area 26 The preferred embodiment uses a threaded bcaring lug upon which is 27 threaded a bearing sleeve containing about one half the radi~ hiclcness of 21~ the total sesl groove. This portion of the scal groove is called ~he shaft groove. The other radial portion Or the tot~l seal groove is formed in the 30 cutter.
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Addii;onal objects, features and advantages of the invention will 2 become ~pparent in the following description.
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DESCR~PTION OF THE DRAWINGS --3 Fig. 1 is a Yiew in longitudinal section of a portion of an earth 4 boring bit, showing the compensator system, bearing shaft, cutter and one ;, embocliment of a seal assembly.
s Fig. 2 is a fragmentary Yiew in longitudina1 scction of the lower 7 portion of a bit, enlarged with respect to Fig. 1, to better expose the seal 8 assembly 9 Fig. 3 is a fragmentary view in longitudinal section of yet another 10 portion of the cutter and bearing shaft~ showing the bearing seal assembly enlarged with respect to the illustration of Fig. 2.
Fig. 4 is an enlarged longitudinal section of one of the rigid rings of the seal assembly.
Fig. 5 is a fragmentary, longitudinal section of the seal seat and conical surface seal groove in the cutter.
Fig. 6 is a longitudinal section of the seal seat annular insert used on 17 the bearing shaft to form the conical, contoured surface that receives and 18 positions the scal assembly.
19 Fig. ~ is a fragmentary Yiew in longitudinal section of the lower 20 portion of an alternate embodiment of the invention shown in Fig. 1.
21 Fig. 8 is a vicw in longitudinal section of a portion of an earth 22 boring bit having an improved rigid face seal assembly to reduce during 2s drilling the amount of seal movemcnt relative to cutter movemcnt.
24 Fig. 9 is an cnlarged fragmcntary view in longitudinal section of the 25 preferred seai assernbly and groove embodiment shown in Fig. 8.
26 Fig~ lû is a fragmentary view of another embodimcnt of an earth 27 boring bit, partially in longitudinal section, that utilizcs a threaded bearing 2~ sleeve to form the journal bearing of the shaft and a portion of seal 20 groove, 3c Fig. I l is an enlarged fragmentary vicw in longitlldinal section of the ~I bearing, cutter and seal asscmbly ~ F~g. 10.
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DESCRIPTION OF THE PREFERRED ~MBODIMEN~S --2 Figs. 1-7 3 The improvement of this invention is best appreciated by initial 4 reference to the disclosure of one of the related paterlts, U.S. Pat~nt s 4,516,641, which is shown in Figs. 1-7 herein. The numeral 11 in Fig. 1 6 designates a lubricated, rotatable cone or cutter type earth boring bit having7 a body formed in three head sections or le&s 13, only one of which is shown.
8 Each leg 13 includes an oblique cantile~ered bearing sllaft 15 that depends g inwardly and downwardly tc support a rotatable cutter 17 haYing earth 10 disi~tegrating teeth 19. Lubricant passage 21 supplies lubricant to the bearing surfaces between the bearing shaft IS and cutter 17. A seal assembly 23 retains lubricant in the bearing and prevents borchole fluid 13 from entering the bearing. A hydrostatic pressure compensator is part of a lubrication system 25 connected with the lubricant passage 21 to equalize IS the pressure of the liquid lubricant inside the bearing with the hydrostatic pressure of the fluid in the borehole. A preferred compensator system is 17 shown in the patent to Stuart C. Millsapps, Jr., U S. Patent No. 4,276,946.
18 The geornetry of the bearings on the shaft 15 and within the cutter 17 are of a pr;or art configuration, including the use of a b~ll bearing 20 retainer 27, which with a plug 26 welded at 28 ret ins the cutter on the 21 bearing shaft, preferably as shown in the U~S. patent of Robert A.
2~ Cunningham, No. Re. 28,625.
23 Referring especially to Fig. 3, the cueter and shaft include an 24 annular seal groove or gland th~ t has axially spaced, generally radial end 25 walls ~9, 31 and inncr and outer circumfcrential walls 33, 35. End wall 31 20 and circumferential wall 33 are formed upon a seal seat inscrt 3~ secured to 27 the bearing shaft 15.
28 The seal assembly includcs a pair of annular rigid rings 37, 39 with 29 opposed radial faces 41, 43. The pair of rigid, prefcrably mctal, rings have 30 a radially measurcd thickncss lcss than ~hc annular spacc bctwccn the inner 5~ and outer side walls 33, 35 of thc groove and an axially measurcd width .
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1 which is less than the wid~h or the distance between the cnd walls 29, 31 of 2 the groove, as will be explained more fully hereafter.
3 Each of a pair of resilicnt, energizer rings 45 or 47 extends betweeD
4 a seal seat 49 or 51 on one of the metal rings 37 or 39 and an opposed seal ~ seat 53 or 55 on the inner or outer circumferential walls 33, 35. Each seal 6 seat has an annular groove and configuration to position and retain the 7 associated energizer ring and the metal ring, which is suspended within the 8 groove intermediate the circumferential walls 33, 35 and the end walls 29, 9 31 to provide clearances Cl and C2 which exist when the thrust surfaces 32 o are in contact. The positions of seats 53 and 55 relative to each other are 1l selected such that, at assembly, the initial deflection of each seal half 12 relative to its adjacent end wall 29 or 31 will provide sufficient contact 13 pressure between radial faces 41 and 43 to maintain sealing contact between 1,~ all the elements of the seal assembly throughout the full range of seal 15 movements permitted by clearances Cl and C2 and the play between cutter 16 and shaft. See Patent No. 3,180,648 for a description of an earlier 17 construction of a conical, "Duo Cone" seal arrangement of the Caterpillar 18 Tractor Company, and patent Nos. 3~403,916, 3,524,654 and 4,077,634 for 19 improvements to such seals.
From Fig. 3 it may be seen that one of the metal rings 37, 39 is 21 inverted with respect to the other, a feature which permits the seal assembly22 to span the groove diagonally and engage opposite circumferential walls 33, 2~ 35. The clearances Cl and C2 are between each of the end walls 29, 31 of 2~ the groove al~d the engagcd rigid rings 37, 39. Drilling fluid fills the space 25 57 and acts upon the outcrmost side of the scal assembly 23, and lubricant 26 fills the space 59 and acts upon the ;nncrmost sidc of the assembly. The 27 rigid rings 37 or 3~ have a beveled, substantially conical portion on the 28 lubricant side of thc scal assembly to define a space 61 to feed lubr;cant to29 the en8aged radial faccs 41, 43, which regencrate inwardly as thcy wear in 30 serYice. Sce U.S. Patent No. 3,180,648 for a dcscription of one configur;ltion 31 of such seal faces.
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The dimensions provided below relate to the bit used in the first 2 field test of the invention, which was a "Hughes" 12-1/4 inch, J22 type bit.
3 With reference especially to Figs. 4 and 5, the radial thickness T of each of 4 the t}lree metal rings 37 was 0.200 inch, the axial width W was about 0.270 inch and the outside diametcr was about 3.449 inches. Angle d was ~bout s twenty degrees and the radii Rl and R2 averaged 0.048 and 0.080 inch 7 respectively, Rl being tangent with the conical surface 63. The axial 8 dimensions Y and Z averaged respectively 0.050 and 0.149 inch. The depth 9 D of the positioner groove 65 below the lip 67 averaged 0.009 inch to help o position and confine the associated energizer ring 4~ during assembly. The 11 radial thickness X o~ the radial sealing face 41 was about 0.050 incll with a12 surface finish of about one or two RMS and a tapered surface defined by a 13 spherical radius R3 of about 80 inches with a surface finish of about two or 14 three RMS.
The inverted and opposing rnetal rings 39 had a radial thickness T of .5 about 0.199 inch, a radial width W of about 0.247 inch and an outside 17 diameter of about 3.450 inches. Angle o~ was about 19 degrees, and the radii 18 Rl and R2 were both about 0.075 inch. The axial dimensions Y ~nd Z were respectively about 0.023 and 0.154 inch and the depth D of the positioner 20 groove being about 0.016 inch. There was a flat sealing face on the rings 39 21 that extended across the entire thickness T of the ring, that had a surface 22 finish similar to that of rings 37.
23 Ring 37 was purchased from Caterpillar lractor Company and is one 24 of thcir standard hard metal alloy rings. Ring 39 was manufac~ured by 25 Hughes Tool Company specifically for this invcntion from an air harderling 2f; tool Steel.
27 The configuration of thc scal scat in the cutter 17 of the bit is shown 2~ in Fig. ~. The seal seat was dcfined by a conical surface 70 having an 29 angle ~1 of about 19-112 degrces, a positioncr groove or seal seat 53 h~.~ving 30 a radius R4 of about 0.060 inch locatcd a distancc D2 of about 0.129 inch 31 from end wall 29, and a dcpth Dl of about 0.008 inch. Thc conical surface : : `
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1 70 intersected the groove 53 at a point 72 which was located radially from 2 the surface 35 a distance of about 0.02l inch.
3 A similar configuration was used for the seal seat on the bearin8 4 shaft 15, defincd in this instance by the seal seat insert 36 shown ill Fig. 65 having a thickness Tl of 0.105 inch. The positioner groove had a depth D3 6 of about 0.011 inch, l~ormed 'oy a radius R5 of about 0.060 inch located a 7 distance D4 about 0.140 inch from end wall 31. 92 was a conical allgle of 8 about 20 degrees located irl a manner similar to the conical surface 70 of g Fig. 5.
The O-rings or energizer rings 45, 47 after service had a cross-11 sectional thickness of about 0.168 inch, a hardness of about 59 durometer~
12 Shore A, inside diameters of about 3.057 and 2.760 inches respectively, and a3 high resilience, measured to be about 43 percent rebound using the above 14 described O-rings and a Shore Resiliometer, Model SR- I . The radial end 5 walls 29, 31 of the seal groove were located a width of about 0.580 inch 16 apart with bearing thrust surfaces 32 in mutual contact. Using the abovc 17 components, the assembly loading on the faces of the rigid rings was about 18 40 to 60 pounds, as determined from load deflection curves. The clearances lg Cl and C2 were respectively about 0.035 and 0.029 inch at assembly with 20 thrust surfaccs 32 in contact to define the minimum groove width and the 21 diameters of the circumferential walls 33, 35 were respectively about 2.969 22 and 3.529 inches 23 For thc filst bit tested, the axial bearing play of each of the cntters, 24 after testing, was:
Axial Pl~ly (inch) No. I Cutter 0.012 No. 2 Cutter 0.015 ~8 No. 3 Cutter 0.012 31 In operation, and during drilling in a well bore fillcd with liquid, the compcnsator 25 acts to baiancc the hydrostatic pressurc of the liquid in thc .
22~
1 well bore with the pressure of the liquid lubricant inside the bearing.
2 HoweYer, cutter movemcnts during drilling, caused by the complex forces 3 applied to a cutter, and the clearances which are of necessity used to enable 4 assembly of the parts, produce rapid changes in the volume defined by the 5 space S9. The viscosity of the lubricant and flow restrictions between the ff spac~ S9 and hydrostatic pressure compensator 59 do not allow compensation 7 for the volume changes in space 59 as rapidly as they occur. Nevertheless, 8 seal assembly 23 will move sufficiently to provide the required volume 9 change and thereby minimize the pressure changes experienced by th seal o which would otherwise cause rapid depletian of the lubricant supply or entry of borehole fluids into the bearing, with resulting bearing and seal 12 damage.
Axial Pl~ly (inch) No. I Cutter 0.012 No. 2 Cutter 0.015 ~8 No. 3 Cutter 0.012 31 In operation, and during drilling in a well bore fillcd with liquid, the compcnsator 25 acts to baiancc the hydrostatic pressurc of the liquid in thc .
22~
1 well bore with the pressure of the liquid lubricant inside the bearing.
2 HoweYer, cutter movemcnts during drilling, caused by the complex forces 3 applied to a cutter, and the clearances which are of necessity used to enable 4 assembly of the parts, produce rapid changes in the volume defined by the 5 space S9. The viscosity of the lubricant and flow restrictions between the ff spac~ S9 and hydrostatic pressure compensator 59 do not allow compensation 7 for the volume changes in space 59 as rapidly as they occur. Nevertheless, 8 seal assembly 23 will move sufficiently to provide the required volume 9 change and thereby minimize the pressure changes experienced by th seal o which would otherwise cause rapid depletian of the lubricant supply or entry of borehole fluids into the bearing, with resulting bearing and seal 12 damage.
13 Use of the seal assembly 23 described above in a bit which includes a 14 hydrostatic pressure cornpensator minimizcs the pressure differentials to which the seal assembly is exposed through the cooperative relationship of 16 the hydrostatic pressure compensator and the dynamic pressure 17 compensating abilities of the seal assembly. The seal assembly is one which 18 spans diagonally the seal groove such that one of the resilient energizer 19 rings engages a wall of thc cutter, while the other energizer ring engages a 20 wall of the shaft. Thus, the outermost portion of each of the energizer rings21 is exposed to the fluid in the borehole, while an innermost portion of each 2~ of the energizcr rings is exposed to the lubricant inside the bearing. Every 23 pressure differential is therefore sensed by the seal assembly, which is 24 moved by cach such press~re differential. A seal assembly which cannot be 25 moved by the differential pressure cannot effcctivcly con~pcnsate for 26 dynamic changes in the volume of space 59. Preferably thc seal half 27 consisting of energizcr ring 47 and rigid ring 39 should have the same axial ~8 load deflection charac$eristics as the mating half consistir~g of energizer 29 ring 45 and rigid ring 37 to bal;incc and m;nimize the incrense in the 30 loadin~ of engaged radial faces 41 and 43 causcd by prcssurc differentials.
31 Another rGq~lircment for a satisfactory seal asscmbly, and thc groove in which it is placcd, is that thc assembly bc positioncd bct~vccn thc end :
.C~2 ~ z~
walls of the groovc to permit unrestrictGd axial movement of the rigid rings 2 between the walls of the groove in respODSe to sensed pressure differentials.
3 If the bearing lubricant could freely enter and lea~e space 59 as the voiume " of space 59 is changed by cutter movement, the p~essure differentials acting on the seal would be negligible and the movement of the rigid rin~s w~uld t; be less than the cutter movement. Furthermore, if the load de~lection 7 characteristics of each halî were equal, as p~eferred, the rigid ring 8 movement would be one-half the cutter movement in the above described g embodiment. However, because lubricant movement is restricted, grcater rigid ring movement must be provided for. The required clearances Cl and C2 were determined by building a model of the seal~ cutter and shaft 12 assembly and measuring the movement of the rigid rings in response to 13 cutter movement with the exit from space ~9 blocked; for example, by a conventional O-ring seal betwecn the bearing surfaces of the simulated 15 cutter and shaft. To be sure that accurate rigid ring moYement takes place 16 in the model, it is important to have space ~9 completely filled with an 17 incompressible fluid that is free of any air or vapor pockets. Furthermore, 18 in some cases, it may be necessary to pressurize space 57 with air to insure 19 complete rigid ring movement in response to movement of end wall 29 away 20 from end wall 31.
31 Another rGq~lircment for a satisfactory seal asscmbly, and thc groove in which it is placcd, is that thc assembly bc positioncd bct~vccn thc end :
.C~2 ~ z~
walls of the groovc to permit unrestrictGd axial movement of the rigid rings 2 between the walls of the groove in respODSe to sensed pressure differentials.
3 If the bearing lubricant could freely enter and lea~e space 59 as the voiume " of space 59 is changed by cutter movement, the p~essure differentials acting on the seal would be negligible and the movement of the rigid rin~s w~uld t; be less than the cutter movement. Furthermore, if the load de~lection 7 characteristics of each halî were equal, as p~eferred, the rigid ring 8 movement would be one-half the cutter movement in the above described g embodiment. However, because lubricant movement is restricted, grcater rigid ring movement must be provided for. The required clearances Cl and C2 were determined by building a model of the seal~ cutter and shaft 12 assembly and measuring the movement of the rigid rings in response to 13 cutter movement with the exit from space ~9 blocked; for example, by a conventional O-ring seal betwecn the bearing surfaces of the simulated 15 cutter and shaft. To be sure that accurate rigid ring moYement takes place 16 in the model, it is important to have space ~9 completely filled with an 17 incompressible fluid that is free of any air or vapor pockets. Furthermore, 18 in some cases, it may be necessary to pressurize space 57 with air to insure 19 complete rigid ring movement in response to movement of end wall 29 away 20 from end wall 31.
21 A model as hereinabove described was used to mcasure the movement 22 of thc rigid rings in responsc to cutter movcment for the shaft, cutter and 23 seal used in the first test bit. Air pressure in spacc 57 was not rcquired for 24 this test bccause the pressure in space 59 did not drop bclow the vapor 25 pressure of the fluid used to fill the space. Thc ratio of the rig;d ring 26 movement to cutter movement was dctcrmined from thc measuremcnts to be 27 1.88:1. This ratio is in1ucnced by the gcometry of space 59, the size, shape28 and elastic properties of the energizer rings and the manner in which the 29 energszer rings are deformcd by the rigid rings and wall of thc seal grooYc.
30 Thus, a changc in any of these paramctcrs is lilcely to cause a change in 31 rcquired clcarances Cl and C2.
z~z~
After the ratio of the rigid ring movemcnt to cutter movement has 2 been established, as described above, the minimum values for Cl and C2 3 may be calculatcd. The maximum seal or rigid ring movement with respect to the bearing shaft is calculated by multiplying the axial play between the 5 cutter and shaft by the ratio of rigid rin~ movement to cutter movement.
6 When bearing thrust suIfaces 32 are in contact, the first axial clearance C
7 between rigid ring 37 and the inner wall 29 of the groove should be greater 8 $han the maximum rigid ring movement less ~he axial play between the g cutter and the shaft. The second axial clearance C2 between rigid ring 39 io and the oute} end wall 31 of the groove measllred with the thrust surfaces 32 in contact should be greater than a value equal to the maximum rigid 12 ring movement as calculated above less the displacement which rigid rings 3 37 and 39 undergo when the space between end walls 29 and 31 is increased 14 by axial play from its minimum length to its maximum length in the 15 absence of any pressure differential across the seal. This displacement of 16 rigid rings 37 and 39 in the absence of a pressure differential can be 17 determined with the model if space 59 is vented or it can be calctllated with18 the aid of the load deflect;on curves for the seal halves.
19 While the embodiment of the invention disclosed above was that of 20 the initial tes~, the commercial embodiment is expectcd to be closer to that 21 shown in Fig. 7. The leg 101 includes an oblique cantilevered bearing shaft 22 103 that depcnds inwardly and downwardly to support a rotatablc cutter 105 23 having earth disintegrating tecth 107. A lubricant passage 109 supplics 24 lubricant to the bearing surfaccs betwcen the bearing shaft 103 and the ~5 CUttCF 105. A snap ring retainer 106? similar to that shown in U.S. Patent 28 No. 4,344,658 is used in placc of the ball retainer shown in Fig. 1.
27 A seal assembly 111 retains lubric~nt and excludcs borehole fluids.
28 This seal assembly has thc salne configuration as assembly 23 of Fiæ. 3, ,?,9 however, thc inncrmost energizer ring 113 engagcs dircctly thc journal 30 bearing cylindrical surface 115, rathcr than a scal seat insert 36. A seal seat 31 configuration is providcd similar to thc scal seat 55 and inner circumfercntial wall 33 iD thc Fig. 3 embodim nt. This reduccs thc diamete~
:
~Z~2;21~
of the seal seat of Fig. 7, as comparcd to the diameter of the seal seat in Fig.2 3. This reduction i~ diameter of the scal seat in relation to the diameter of 3 journal bearing cylindrical surface 115 reduces the ratio of sigid ring " rnovcment to rutter movement. This ratio, determined by mak;n8 a model 5 similar to the a~e described in eonnection with the Fig. 3 embodiment, 6 except using the Fig. 7 bear;ng configuration, is 1.28:1. The materials for 7 the various components of the seal assembly are identical with the matcrials 8 used in the embodiment of Figs. 1-4 except both r;gid rings are pseferably 5~ constructed of the same hard metal alloy as ring 37.
1l Figs. 8-9 1~
An alternate to the disclo6~re above aFpears ~n U.S. :Patent 4,666,001, arxl is E~ in Fir~;. 8 and 9.
In Fig. 8 the Dumeral 121 designates ~nc section of a rock bit having 6 a threaded upper end or shank 123, a lubrication system 12~ of the type 17 previossly described, which feeds lubricant a the passage 127, and to a 18 passage 129 formed in a ball plug 131 secured by weld 133 to the section 19 121. Additional passage means 134 enables lubricant communication with 20 the bearing shaft 135, which has on its interior and cantilelrered cnd a pilot 21 pin 137 and ball bearing raceway 139.
22 The ball plug 131 retains plural ball bearings 143 in the ball bearing 23 raceway 139 and in a mating raCcwQy 145 in the cutter 147. A bearing 24 sleeve 149 has an intcrior surfacc 151 which engages with interference fit a 2s mating portion on thc bearing sha~t l 3~. An annular ~xterior bcarin~
26 surface 175 eng~ges a rn~ting cylindrical bearing surface 173 in the cutter 147. These be~ring elcmcnts cn~ble rotation Or the cutter 147 such that the 2n earth disintcgrating teeth 153 engage and disintegrate geological form3tions 2~ during drilling To seal lubric~nt wi~hin the cuttcr 147 for lubric~..tion of the aboYe 31 describcd bearing surfaccs, a seal asscmbly I~S is providcd, which may be bettcr seen with rcfcrencc to Fig. 9.
~29;~
The seal assernbly 155 is disposed partially within a generally L
2 shaped groove in the shaft 135, as best seen in Fig. 9, having a radial end 3 wall 157 and a circumferential wall 159. Another groove is located within 4 the cutter 147, having a radial end wall 161 and circumferential wall 163.
Confined within the grooves are a pair of energizer rings 165j 167 6 and rigid annular rings 169, 171. The geometries of the groove surfaces, the 7 energi~er rings and the ~nnular rings shown here are identical with those 8 described in the preceding embodiment. However, the bearing surface 1?5 is 9 located substantially midway betweerl the circumferelltial walls 159, 163 and o is defirled by a bearing sleeve 149 having an annular surface 151 retained by interference fit in a mating annular slot with a small shoulder 177 located in the bearing shaft.
13 The improvernent of Figs. 8 and 9 decreases the lubricant pressure and volume changes in the vicinity of the seal assembly by positioning thc 15 seal groove and seal assembly in relation to the bearing surface 175 such 16 that the ratio of rigid face seal movement to cutter movement is about one 17 half to one. This is achieved by making the bearing surface 175 18 intermediate the radial thickness of the seal assembly grooves or thc lg diameters of the circumfcrential surfaces 159j 163. That is, part of the 20 groove is in the cuttcr and the other is in thc bcaring shaf t. Thc 21 relationship of scal assembly movement to cutter movement is established by 22 dimensioning consistent with the following formula, derived for this 23 invention 24 ~ 2C~-C~
2s H~2A
: ~6 :~: 7 Where: S - Seal Movement For Unit Cone Movemcnt ~ Relati~e to Shaft 28 H = Effective Sh3ft O-Ring 165 Annular Arca ~9 C = Effective Cuttcr O-Ring 167 Annular Area A - Effectlve Rigid Ring Annular Are3 31 ` G - Effective Cuttcr Seal Groove Annular Arca :
: :::
~ ~ - 1 6 -:~
:: :
: :: ~ :
~2~Z~
The ratio of seal movement to cutter movement of about one half to -2 one is achieved using the abo~e formula, a circumferential wall 159 3 diameter of 2.970 inch, and the following annular areas:
H = 0.94~ sq. inch 0 C = 1.041 sq. inch 7 A = 0.587 sq. inch 8 G = 1.311 sq. inch ,.,; i o The formula assurnes no radial bearing clearance, no lubricant flow to or from the seal assembly, no or off- setting changes in annular areas H, C, A and G, and rolling, nonsliding movement of the energizer rings 165, 1 67.
14 The preferred assembly method uses the steps of forming about one lS half the groove in the cutter 147 and the other half in the shaft 135, 16 assembling one rigid ring 169 and energizer ring 165 in the seal region of 17 the shaft 135, assembling thc sleeve 149 and then the other rigid ring 171 18 and the mating energizer ring 167 in the cutter 147, and then assembling the ~9 cutter 147 and sleeve 149 on the shaît 135 against shoulder 177, the slee~e 20 149 engaging the shaft 135 with an interference fit.
21 During drilling with the embodiment shown in Figs. 8 and 9, cutter 22 147 will move on the shaft 135 in a complex, wobbling fashion. rhc above 23 formula assumes simpler, axial movemcnts of thc cutter but is thought to be 2~ sufficicntly accurate to be useful in cstablislling design parameters fos 25 groove and se~l assembly geomctries.
26 Tho adYantages of the embodiment of Figs. 8 ~nd 9 over the 27 embodiment of Figs. 2 through 7 are: (1) Pressure changcs in thc lubricant 2~ adjaeent the seal assembly during drilling and conscquent cutter movcmcnt 29 are greatly reduced, ideally eliminated; (2) variations in loading on the faces 30 of the rigid `rings 169, 171 are minimized; (3) rolling motions of the 31 energizcr rings 165,167 are more nearly equali2ed; ~4) sincc any axial ::
' ;,: ' :
:~
O
movement of the seal assembly is less than axial movement of the eutter, the 2 clearanccs with end walls 157, 161 may be minimized.
4 Figs lO-II
6 The improvement of this application is disclosed in Figs. 10-6 I l. Here, an earth boring bit 201 has three head sections 203 that are 7 welded to form a body. Extending inwardly and downwardly from each of 8 the sections 203 is a cantilevered bearing lug 205, threaded at 207 to rcceivean interna11y threaded bearin~ sleevc 209. This type of bearing arrangement lO is shown in U.S. Paten~ No. 4~600,064, "Earth Boring Bit With Bearing 11 Sleeven, July IS~ 1986.
12 In an upper portion of each head section 203 is a compensa~or 211 lS that forms part of a compensator system that includes pass~ges 213, a 14 pseferred form of compensator being showII in U.S. Patent No. 4,276,946, 15 "Biased Lubricant Compensator For An Earth Boring Drill Bit", July 7, 1981.
16 Convcntional noz~le means 214 direct drilling fluid toward 2 borehole 17 bottom. A part of the passage 213 extends through the leg 215 of the section 18 203 into intersection with another passage 217, formed in ~his instance 19 coaxially with the bearing lug 20~. Lubricant is introduced from passage 20 217 through a passage 218 in a pilot pin 220 formed on sleeve 209.
21 The thread 207 formed OD the oxterior of the bcaring lug 205 22 diverges outwardly from an inner end region 2l9 to: an outer base region 2:!~ 221, having a shoulder 223 which is transverse or perpendicular wilh respcçt 4 to the rotation~l a~cis of thc lug 2D5. ~he thread form is an A.P.I. V-.040 25 with a 2.708 inch pitch diameter at 0.625 inch from shoulder 223.
2~ ~ The bcaring sleeve 209 attached ~o the bearing lug 205 has a 27 ~tantially~ ri~ ~rior j~rnal bearing ~rfaoe 225, a tapered an~
28 threaded interior portion 227 tha~ mates with thc tlire~ds 207 :oî the lu~ 20~, 9 and a: ~raverse mouth 229 that m~tes with the shoulder 223 of the base 30 region 221 of the lug. The radi~l thickness of the mouth 2~9 is slJbstantinlly Sl 0-345 inch, for a 14 3/4 inch bit providcd by way of examplc, 1he slcevc ~:
~: ~ - 1 8 -being made f~om a metal alloy with a minimum yield strcngth so as to -2 provide a torsional yicld strength of substantially 11,000 foot pounds.
3 From the drawing it îs apparent that the bearing slee~e 209 and its 4 cylindrical journal bearing portion 225 has a length greater than that of its threaded portion 227 to define a 6 thick walled inner end portion 231 in which is formed an 7 assembly grooYe 233 that opposes or r~gist~rs with a retainer groove 235 8 formed in the rylindrical bearing portio~ 237 of the cutter 239. The 9 minimum thickness of the metal be~ween the assembly groove 233 and the threaded portion 227 of the bearing sleeve 209 for the 14 3/4 diameter bit 11 pro~ided by way of example should be substantially 0.422 inch. A drive pin 12 hole 241 provides a meQns to apply a selected torque of about 2500 foo~-13 pounds to the sleeve on assembly with the lug 205. The cueter is of a 14 conventional configuration, with earth disintegrating teeth 243, and is retained on the bearing sleeve 209 with a resilient snap ring 245 having a 16 curved cross section and grooYe configur~tion with cur~ed bottom portion of 17 the type disclosed in U.S. Patent No. 4,236,764, "Earth Boring Drill Bit With 18 Snap Ring Cutter Retentionn, Deccmber 2, 1980.
19 The sleeve 209 hns a bsronizing treatment of the type describe in U.S.
Patent 3,922,038, "Wear Resistant Boronized Surfaces And Boronizing 21 Methods", Novembcr 25, 1975, on ~he ex2erior cylindrical surface 225 to 22 improve wear resistance. This treatment providcs the requisite improvement 23 to wear resistance without causing a substantial weakening of the sleeve.
24 A bearing seal assembly 247 is esscntially that as shown in connection with the above description of Figs. 1-7 but the seal groove arrangement is 26 similar to that of Fig 8-9. Referring to Fig. 11, the seal assembly 247 is 27~ disposed par~ially within a generally L shapcd, shaft se31 groo~e 249, a 28 portion of which is in the outcr end of the sleeve 20g. The shaft se~l groove 29 249 ir~l~ a ck~fer~nti~, ger~rally c~ylindrical a~er wall 251 and a radial 3 0 inner end wall 253. A second portion 255 of the shart seal groovç 249 31 includes a circumfcrenti~l wall 257 and a radial outer end wall 259 formed 32 on thc shoulder 223 on the base region Or the bearing lug 205. Thc mouth '21~
1 229 of the slee~re 209 is transYerse and sealingly engages an opposed surface 2 263 of thc shoulder 2~3 of the bearing lug 20~.
3 An opposed, generally L-shapcd cutte~ seal groov~ 265 is formed în 4 the outcr, mouth region 2S7 of thc cutter 239, and includes a ~; circumferenSial, generally cylindrical wall 269 and a radial end wall 271.
Confined within the abo~e dcscribed cuttcr seal groove 26~ and shaft ~ seal groove 249 are a pair of encrgizer rings 273, 275 and a pair of ~igid 8 annular rings 277, 279. The geometries of the groove surfaces, the energizer ~, rings and the annular rlngs shown here are identical with those deseribed in connection with Figs. 8-9. The journal bearing surface 225 of the bearing sleeve is 10cated substantially midway between the circumferential walls 2 251, 269 in the preferred embodiment.
3 The improvement of Figs. 10-11 decseases the lubricant pressure and 4 volume changes in the vicinity oï the seal assembly by positioning the seal 5 groove and seal assembly in relation to the journal bearing surface such that 6 the ratio of rigid ring seal movement to cutter movement is abollt one half.
7 Seal movement for unit cone movement is established by using dimensions A, C, G and H consistent with the formula provided in describing Figs. 8-9.
19 The advantages of the embodiment of Figs. 10-11 are similar to those 20 of ~igs. 1-7 and 8-9, but in addition, the assembly problems associated with 21 Fi8s B-9 are avoided.
22 While the invention h~s been shown in only the preferr~d forms, it 2~ should be apparent to those skilled in the art that it is not thus limi~ed, but 24 is susceptible to various changes ancl modifications without departing from 25 the spirit thereof.
:
- ~0 -
30 Thus, a changc in any of these paramctcrs is lilcely to cause a change in 31 rcquired clcarances Cl and C2.
z~z~
After the ratio of the rigid ring movemcnt to cutter movement has 2 been established, as described above, the minimum values for Cl and C2 3 may be calculatcd. The maximum seal or rigid ring movement with respect to the bearing shaft is calculated by multiplying the axial play between the 5 cutter and shaft by the ratio of rigid rin~ movement to cutter movement.
6 When bearing thrust suIfaces 32 are in contact, the first axial clearance C
7 between rigid ring 37 and the inner wall 29 of the groove should be greater 8 $han the maximum rigid ring movement less ~he axial play between the g cutter and the shaft. The second axial clearance C2 between rigid ring 39 io and the oute} end wall 31 of the groove measllred with the thrust surfaces 32 in contact should be greater than a value equal to the maximum rigid 12 ring movement as calculated above less the displacement which rigid rings 3 37 and 39 undergo when the space between end walls 29 and 31 is increased 14 by axial play from its minimum length to its maximum length in the 15 absence of any pressure differential across the seal. This displacement of 16 rigid rings 37 and 39 in the absence of a pressure differential can be 17 determined with the model if space 59 is vented or it can be calctllated with18 the aid of the load deflect;on curves for the seal halves.
19 While the embodiment of the invention disclosed above was that of 20 the initial tes~, the commercial embodiment is expectcd to be closer to that 21 shown in Fig. 7. The leg 101 includes an oblique cantilevered bearing shaft 22 103 that depcnds inwardly and downwardly to support a rotatablc cutter 105 23 having earth disintegrating tecth 107. A lubricant passage 109 supplics 24 lubricant to the bearing surfaccs betwcen the bearing shaft 103 and the ~5 CUttCF 105. A snap ring retainer 106? similar to that shown in U.S. Patent 28 No. 4,344,658 is used in placc of the ball retainer shown in Fig. 1.
27 A seal assembly 111 retains lubric~nt and excludcs borehole fluids.
28 This seal assembly has thc salne configuration as assembly 23 of Fiæ. 3, ,?,9 however, thc inncrmost energizer ring 113 engagcs dircctly thc journal 30 bearing cylindrical surface 115, rathcr than a scal seat insert 36. A seal seat 31 configuration is providcd similar to thc scal seat 55 and inner circumfercntial wall 33 iD thc Fig. 3 embodim nt. This reduccs thc diamete~
:
~Z~2;21~
of the seal seat of Fig. 7, as comparcd to the diameter of the seal seat in Fig.2 3. This reduction i~ diameter of the scal seat in relation to the diameter of 3 journal bearing cylindrical surface 115 reduces the ratio of sigid ring " rnovcment to rutter movement. This ratio, determined by mak;n8 a model 5 similar to the a~e described in eonnection with the Fig. 3 embodiment, 6 except using the Fig. 7 bear;ng configuration, is 1.28:1. The materials for 7 the various components of the seal assembly are identical with the matcrials 8 used in the embodiment of Figs. 1-4 except both r;gid rings are pseferably 5~ constructed of the same hard metal alloy as ring 37.
1l Figs. 8-9 1~
An alternate to the disclo6~re above aFpears ~n U.S. :Patent 4,666,001, arxl is E~ in Fir~;. 8 and 9.
In Fig. 8 the Dumeral 121 designates ~nc section of a rock bit having 6 a threaded upper end or shank 123, a lubrication system 12~ of the type 17 previossly described, which feeds lubricant a the passage 127, and to a 18 passage 129 formed in a ball plug 131 secured by weld 133 to the section 19 121. Additional passage means 134 enables lubricant communication with 20 the bearing shaft 135, which has on its interior and cantilelrered cnd a pilot 21 pin 137 and ball bearing raceway 139.
22 The ball plug 131 retains plural ball bearings 143 in the ball bearing 23 raceway 139 and in a mating raCcwQy 145 in the cutter 147. A bearing 24 sleeve 149 has an intcrior surfacc 151 which engages with interference fit a 2s mating portion on thc bearing sha~t l 3~. An annular ~xterior bcarin~
26 surface 175 eng~ges a rn~ting cylindrical bearing surface 173 in the cutter 147. These be~ring elcmcnts cn~ble rotation Or the cutter 147 such that the 2n earth disintcgrating teeth 153 engage and disintegrate geological form3tions 2~ during drilling To seal lubric~nt wi~hin the cuttcr 147 for lubric~..tion of the aboYe 31 describcd bearing surfaccs, a seal asscmbly I~S is providcd, which may be bettcr seen with rcfcrencc to Fig. 9.
~29;~
The seal assernbly 155 is disposed partially within a generally L
2 shaped groove in the shaft 135, as best seen in Fig. 9, having a radial end 3 wall 157 and a circumferential wall 159. Another groove is located within 4 the cutter 147, having a radial end wall 161 and circumferential wall 163.
Confined within the grooves are a pair of energizer rings 165j 167 6 and rigid annular rings 169, 171. The geometries of the groove surfaces, the 7 energi~er rings and the ~nnular rings shown here are identical with those 8 described in the preceding embodiment. However, the bearing surface 1?5 is 9 located substantially midway betweerl the circumferelltial walls 159, 163 and o is defirled by a bearing sleeve 149 having an annular surface 151 retained by interference fit in a mating annular slot with a small shoulder 177 located in the bearing shaft.
13 The improvernent of Figs. 8 and 9 decreases the lubricant pressure and volume changes in the vicinity of the seal assembly by positioning thc 15 seal groove and seal assembly in relation to the bearing surface 175 such 16 that the ratio of rigid face seal movement to cutter movement is about one 17 half to one. This is achieved by making the bearing surface 175 18 intermediate the radial thickness of the seal assembly grooves or thc lg diameters of the circumfcrential surfaces 159j 163. That is, part of the 20 groove is in the cuttcr and the other is in thc bcaring shaf t. Thc 21 relationship of scal assembly movement to cutter movement is established by 22 dimensioning consistent with the following formula, derived for this 23 invention 24 ~ 2C~-C~
2s H~2A
: ~6 :~: 7 Where: S - Seal Movement For Unit Cone Movemcnt ~ Relati~e to Shaft 28 H = Effective Sh3ft O-Ring 165 Annular Arca ~9 C = Effective Cuttcr O-Ring 167 Annular Area A - Effectlve Rigid Ring Annular Are3 31 ` G - Effective Cuttcr Seal Groove Annular Arca :
: :::
~ ~ - 1 6 -:~
:: :
: :: ~ :
~2~Z~
The ratio of seal movement to cutter movement of about one half to -2 one is achieved using the abo~e formula, a circumferential wall 159 3 diameter of 2.970 inch, and the following annular areas:
H = 0.94~ sq. inch 0 C = 1.041 sq. inch 7 A = 0.587 sq. inch 8 G = 1.311 sq. inch ,.,; i o The formula assurnes no radial bearing clearance, no lubricant flow to or from the seal assembly, no or off- setting changes in annular areas H, C, A and G, and rolling, nonsliding movement of the energizer rings 165, 1 67.
14 The preferred assembly method uses the steps of forming about one lS half the groove in the cutter 147 and the other half in the shaft 135, 16 assembling one rigid ring 169 and energizer ring 165 in the seal region of 17 the shaft 135, assembling thc sleeve 149 and then the other rigid ring 171 18 and the mating energizer ring 167 in the cutter 147, and then assembling the ~9 cutter 147 and sleeve 149 on the shaît 135 against shoulder 177, the slee~e 20 149 engaging the shaft 135 with an interference fit.
21 During drilling with the embodiment shown in Figs. 8 and 9, cutter 22 147 will move on the shaft 135 in a complex, wobbling fashion. rhc above 23 formula assumes simpler, axial movemcnts of thc cutter but is thought to be 2~ sufficicntly accurate to be useful in cstablislling design parameters fos 25 groove and se~l assembly geomctries.
26 Tho adYantages of the embodiment of Figs. 8 ~nd 9 over the 27 embodiment of Figs. 2 through 7 are: (1) Pressure changcs in thc lubricant 2~ adjaeent the seal assembly during drilling and conscquent cutter movcmcnt 29 are greatly reduced, ideally eliminated; (2) variations in loading on the faces 30 of the rigid `rings 169, 171 are minimized; (3) rolling motions of the 31 energizcr rings 165,167 are more nearly equali2ed; ~4) sincc any axial ::
' ;,: ' :
:~
O
movement of the seal assembly is less than axial movement of the eutter, the 2 clearanccs with end walls 157, 161 may be minimized.
4 Figs lO-II
6 The improvement of this application is disclosed in Figs. 10-6 I l. Here, an earth boring bit 201 has three head sections 203 that are 7 welded to form a body. Extending inwardly and downwardly from each of 8 the sections 203 is a cantilevered bearing lug 205, threaded at 207 to rcceivean interna11y threaded bearin~ sleevc 209. This type of bearing arrangement lO is shown in U.S. Paten~ No. 4~600,064, "Earth Boring Bit With Bearing 11 Sleeven, July IS~ 1986.
12 In an upper portion of each head section 203 is a compensa~or 211 lS that forms part of a compensator system that includes pass~ges 213, a 14 pseferred form of compensator being showII in U.S. Patent No. 4,276,946, 15 "Biased Lubricant Compensator For An Earth Boring Drill Bit", July 7, 1981.
16 Convcntional noz~le means 214 direct drilling fluid toward 2 borehole 17 bottom. A part of the passage 213 extends through the leg 215 of the section 18 203 into intersection with another passage 217, formed in ~his instance 19 coaxially with the bearing lug 20~. Lubricant is introduced from passage 20 217 through a passage 218 in a pilot pin 220 formed on sleeve 209.
21 The thread 207 formed OD the oxterior of the bcaring lug 205 22 diverges outwardly from an inner end region 2l9 to: an outer base region 2:!~ 221, having a shoulder 223 which is transverse or perpendicular wilh respcçt 4 to the rotation~l a~cis of thc lug 2D5. ~he thread form is an A.P.I. V-.040 25 with a 2.708 inch pitch diameter at 0.625 inch from shoulder 223.
2~ ~ The bcaring sleeve 209 attached ~o the bearing lug 205 has a 27 ~tantially~ ri~ ~rior j~rnal bearing ~rfaoe 225, a tapered an~
28 threaded interior portion 227 tha~ mates with thc tlire~ds 207 :oî the lu~ 20~, 9 and a: ~raverse mouth 229 that m~tes with the shoulder 223 of the base 30 region 221 of the lug. The radi~l thickness of the mouth 2~9 is slJbstantinlly Sl 0-345 inch, for a 14 3/4 inch bit providcd by way of examplc, 1he slcevc ~:
~: ~ - 1 8 -being made f~om a metal alloy with a minimum yield strcngth so as to -2 provide a torsional yicld strength of substantially 11,000 foot pounds.
3 From the drawing it îs apparent that the bearing slee~e 209 and its 4 cylindrical journal bearing portion 225 has a length greater than that of its threaded portion 227 to define a 6 thick walled inner end portion 231 in which is formed an 7 assembly grooYe 233 that opposes or r~gist~rs with a retainer groove 235 8 formed in the rylindrical bearing portio~ 237 of the cutter 239. The 9 minimum thickness of the metal be~ween the assembly groove 233 and the threaded portion 227 of the bearing sleeve 209 for the 14 3/4 diameter bit 11 pro~ided by way of example should be substantially 0.422 inch. A drive pin 12 hole 241 provides a meQns to apply a selected torque of about 2500 foo~-13 pounds to the sleeve on assembly with the lug 205. The cueter is of a 14 conventional configuration, with earth disintegrating teeth 243, and is retained on the bearing sleeve 209 with a resilient snap ring 245 having a 16 curved cross section and grooYe configur~tion with cur~ed bottom portion of 17 the type disclosed in U.S. Patent No. 4,236,764, "Earth Boring Drill Bit With 18 Snap Ring Cutter Retentionn, Deccmber 2, 1980.
19 The sleeve 209 hns a bsronizing treatment of the type describe in U.S.
Patent 3,922,038, "Wear Resistant Boronized Surfaces And Boronizing 21 Methods", Novembcr 25, 1975, on ~he ex2erior cylindrical surface 225 to 22 improve wear resistance. This treatment providcs the requisite improvement 23 to wear resistance without causing a substantial weakening of the sleeve.
24 A bearing seal assembly 247 is esscntially that as shown in connection with the above description of Figs. 1-7 but the seal groove arrangement is 26 similar to that of Fig 8-9. Referring to Fig. 11, the seal assembly 247 is 27~ disposed par~ially within a generally L shapcd, shaft se31 groo~e 249, a 28 portion of which is in the outcr end of the sleeve 20g. The shaft se~l groove 29 249 ir~l~ a ck~fer~nti~, ger~rally c~ylindrical a~er wall 251 and a radial 3 0 inner end wall 253. A second portion 255 of the shart seal groovç 249 31 includes a circumfcrenti~l wall 257 and a radial outer end wall 259 formed 32 on thc shoulder 223 on the base region Or the bearing lug 205. Thc mouth '21~
1 229 of the slee~re 209 is transYerse and sealingly engages an opposed surface 2 263 of thc shoulder 2~3 of the bearing lug 20~.
3 An opposed, generally L-shapcd cutte~ seal groov~ 265 is formed în 4 the outcr, mouth region 2S7 of thc cutter 239, and includes a ~; circumferenSial, generally cylindrical wall 269 and a radial end wall 271.
Confined within the abo~e dcscribed cuttcr seal groove 26~ and shaft ~ seal groove 249 are a pair of encrgizer rings 273, 275 and a pair of ~igid 8 annular rings 277, 279. The geometries of the groove surfaces, the energizer ~, rings and the annular rlngs shown here are identical with those deseribed in connection with Figs. 8-9. The journal bearing surface 225 of the bearing sleeve is 10cated substantially midway between the circumferential walls 2 251, 269 in the preferred embodiment.
3 The improvement of Figs. 10-11 decseases the lubricant pressure and 4 volume changes in the vicinity oï the seal assembly by positioning the seal 5 groove and seal assembly in relation to the journal bearing surface such that 6 the ratio of rigid ring seal movement to cutter movement is abollt one half.
7 Seal movement for unit cone movement is established by using dimensions A, C, G and H consistent with the formula provided in describing Figs. 8-9.
19 The advantages of the embodiment of Figs. 10-11 are similar to those 20 of ~igs. 1-7 and 8-9, but in addition, the assembly problems associated with 21 Fi8s B-9 are avoided.
22 While the invention h~s been shown in only the preferr~d forms, it 2~ should be apparent to those skilled in the art that it is not thus limi~ed, but 24 is susceptible to various changes ancl modifications without departing from 25 the spirit thereof.
:
- ~0 -
Claims (15)
1. An earth boring bit with an improved pressure compensating, rigid face seal assembly having one rigid ring and sealing face carried by a bearing shaft and an opposed sealing face carried by the cutter, said bit comprising:
a body;
a threaded and cantilevered bearing lug positioned to extend obliquely inwardly and downwardly from the body;
an internally threaded bearing sleeve screwed onto the bearing lug to form a journal bearing on the bearing shaft;
a cutter having a generally cylindrical bearing formed internally wherein, secured for rotation about the journal bearing of the shaft;
a lubrication system in the body, including a hydrostatic pressure compensator to lubricate said bearings;
a cutter seal groove formed near the outermost region of the cylindrical bearing in the cutter to have a circumferential, generally cylindrical wall;
a shaft seal groove formed at least partially in the bearing sleeve to oppose the cutter seal groove and having a circumferential, generally cylindrical wall;
at least one rigid ring positioned between the shaft and cutter grooves with its sealing face engaging the opposed sealing face carried by a selected one of the shaft and the cutter;
at least one resilient energizer ring sealingly engaging a selected circumferential wall of the cutter and shaft seal groove and sealingly engaging the rigid ring;
the diameter of the journal bearing being intermediate the circumferential walls of the cutter and shaft seal grooves to decrease axial movement of the rigid rings with respect to axial cutter movement to a selected ratio during drilling.
a body;
a threaded and cantilevered bearing lug positioned to extend obliquely inwardly and downwardly from the body;
an internally threaded bearing sleeve screwed onto the bearing lug to form a journal bearing on the bearing shaft;
a cutter having a generally cylindrical bearing formed internally wherein, secured for rotation about the journal bearing of the shaft;
a lubrication system in the body, including a hydrostatic pressure compensator to lubricate said bearings;
a cutter seal groove formed near the outermost region of the cylindrical bearing in the cutter to have a circumferential, generally cylindrical wall;
a shaft seal groove formed at least partially in the bearing sleeve to oppose the cutter seal groove and having a circumferential, generally cylindrical wall;
at least one rigid ring positioned between the shaft and cutter grooves with its sealing face engaging the opposed sealing face carried by a selected one of the shaft and the cutter;
at least one resilient energizer ring sealingly engaging a selected circumferential wall of the cutter and shaft seal groove and sealingly engaging the rigid ring;
the diameter of the journal bearing being intermediate the circumferential walls of the cutter and shaft seal grooves to decrease axial movement of the rigid rings with respect to axial cutter movement to a selected ratio during drilling.
2. The invention defined by Claim 1 wherein the ratio of the seal movement to axial cutter movement is selected to decrease seal movement, consistent with the following formula:
Where: S = Seal Movement For Unit Cutter Movement Relative to the Shaft H = Effective Shaft O-Ring Annular Area C = Effective Cutter O-Ring Annular Area A = Effective Rigid Ring Annular Area G = Effective Cutter Seal Groove Annular Area
Where: S = Seal Movement For Unit Cutter Movement Relative to the Shaft H = Effective Shaft O-Ring Annular Area C = Effective Cutter O-Ring Annular Area A = Effective Rigid Ring Annular Area G = Effective Cutter Seal Groove Annular Area
3. The invention defined by Claim 2 wherein the ratio of rigid ring movement to axial cutter movement is substantially one half to one.
4. An earth boring bit with an improved pressure compensating, rigid face seal assembly having one rigid ring and sealing face carried by a bearing shaft and an opposed sealing face carried by the cutter, said bit comprising:
a body;
a threaded and cantilevered bearing lug positioned to extend obliquely inwardly and downwardly from the body;
an internally threaded bearing sleeve screwed onto the bearing lug to form a generally cylindrical journal bearing on the bearing shaft;
a cutter having a generally cylindrical bearing formed internally therein, secured for rotation about the journal bearing of the shaft;
a lubrication system in the body, including a hydrostatic pressure compensator to lubricate said bearings;
a cutter seal groove formed near the outermost region of the cylindrical bearing in the cutter to have a circumferential, generally cylindrical wall;
a shaft seal groove formed at least partially in the bearing sleeve to oppose the cutter seal groove and having a circumferential, generally cylindrical wall;
a pair of rigid rings positioned in the seal groove to have opposed sealing faces;
a pair of resilient energizer rings, each of which sealingly engages a respective one of the rigid rings and one of the oppositely facing circumferential walls of the cutter and shaft seal grooves to define a seal assembly;
the diameter of the journal bearing being intermediate the diameters of the circumferential walls of the cutter and shaft seal grooves to decrease axial movement of the rigid rings with respect to axial cutter movement to a selected ratio when the cutter moves axially on the bearing shaft during drilling.
a body;
a threaded and cantilevered bearing lug positioned to extend obliquely inwardly and downwardly from the body;
an internally threaded bearing sleeve screwed onto the bearing lug to form a generally cylindrical journal bearing on the bearing shaft;
a cutter having a generally cylindrical bearing formed internally therein, secured for rotation about the journal bearing of the shaft;
a lubrication system in the body, including a hydrostatic pressure compensator to lubricate said bearings;
a cutter seal groove formed near the outermost region of the cylindrical bearing in the cutter to have a circumferential, generally cylindrical wall;
a shaft seal groove formed at least partially in the bearing sleeve to oppose the cutter seal groove and having a circumferential, generally cylindrical wall;
a pair of rigid rings positioned in the seal groove to have opposed sealing faces;
a pair of resilient energizer rings, each of which sealingly engages a respective one of the rigid rings and one of the oppositely facing circumferential walls of the cutter and shaft seal grooves to define a seal assembly;
the diameter of the journal bearing being intermediate the diameters of the circumferential walls of the cutter and shaft seal grooves to decrease axial movement of the rigid rings with respect to axial cutter movement to a selected ratio when the cutter moves axially on the bearing shaft during drilling.
5. The invention defined by Claim 4 wherein the ratio of the seal movement to axial cutter movement is selected to decrease seal movement, consistent with the following formula:
Where: S = Seal Movement For Unit Cutter Movement Relative to the Shaft H = Effective Shaft O-Ring Annular Area C = Effective Cutter O-Ring Annular Area A = Effective Rigid Ring Annular Area G = Effective Cutter Seal Groove Annular Area
Where: S = Seal Movement For Unit Cutter Movement Relative to the Shaft H = Effective Shaft O-Ring Annular Area C = Effective Cutter O-Ring Annular Area A = Effective Rigid Ring Annular Area G = Effective Cutter Seal Groove Annular Area
6. The invention defined by Claim 5 wherein the ratio of rigid ring movement to axial cutter movement is substantially one half to one.
7. An earth boring bit with an improved pressure compensating, rigid face seal assembly having one rigid sealing face carried by a bearing shaft and an opposed sealing face carried by the cutter, said bit comprising:
a body which includes at least one leg and a cantilevered and threaded bearing lug that extends downwardly and inwardly;
a lubrication system formed at least partially in the body, including a hydrostatic pressure compensator to lubricate said bearings;
the bearing lug including a threaded inner end and an outer base region with a transverse shoulder secured to the body;
a bearing sleeve with a substantially cylindrical journal bearing surface containing an assembly groove for a cutter retainer means, a threaded interior secured to the the threaded end of the bearing lug, and a mouth engaging the transverse shoulder of the lug;
a rotatable cutter having an interior cylindrical bearing surface, a retainer groove that registers with the assembly groove in the bearing sleeve, and an annular, cutter seal groove in the outer end of the bearing surface;
retainer means between the assembly groove and the retainer groove;
a cutter seal groove formed near the outermost region of the cylindrical bearing in the cutter to have a circumferential, generally cylindrical wall and a generally radial end wall;
a shaft seal groove formed at least partially in the outer end of the bearing sleeve and the transverse shoulder of the bearing lug to oppose the cutter seal groove, and having a circumferential, generally cylindrical wall and a generally radial end wall;
at least one rigid ring positioned between the seal and cutter grooves with its sealing face engaging the opposed sealing face carried by a selected one of the shaft and the cutter;
at least one resilient energizer ring sealingly engaging a selected circumferential wall of the cutter and shaft seal groove and sealingly engaging the rigid ring;
the axial width of the rigid ring being less than the axial width of the shaft and cutter grooves when the cutter is thrust outward on the bearing shaft to define at least one axial clearance to permit unrestricted axial movement of the seal assembly between the radial end walls;
the diameter of the journal bearing being intermediate the circumferential walls of the cutter and shaft seal grooves to decrease axial movement of the rigid ring with respect to axial cutter movement to a selected ratio during drilling.
a body which includes at least one leg and a cantilevered and threaded bearing lug that extends downwardly and inwardly;
a lubrication system formed at least partially in the body, including a hydrostatic pressure compensator to lubricate said bearings;
the bearing lug including a threaded inner end and an outer base region with a transverse shoulder secured to the body;
a bearing sleeve with a substantially cylindrical journal bearing surface containing an assembly groove for a cutter retainer means, a threaded interior secured to the the threaded end of the bearing lug, and a mouth engaging the transverse shoulder of the lug;
a rotatable cutter having an interior cylindrical bearing surface, a retainer groove that registers with the assembly groove in the bearing sleeve, and an annular, cutter seal groove in the outer end of the bearing surface;
retainer means between the assembly groove and the retainer groove;
a cutter seal groove formed near the outermost region of the cylindrical bearing in the cutter to have a circumferential, generally cylindrical wall and a generally radial end wall;
a shaft seal groove formed at least partially in the outer end of the bearing sleeve and the transverse shoulder of the bearing lug to oppose the cutter seal groove, and having a circumferential, generally cylindrical wall and a generally radial end wall;
at least one rigid ring positioned between the seal and cutter grooves with its sealing face engaging the opposed sealing face carried by a selected one of the shaft and the cutter;
at least one resilient energizer ring sealingly engaging a selected circumferential wall of the cutter and shaft seal groove and sealingly engaging the rigid ring;
the axial width of the rigid ring being less than the axial width of the shaft and cutter grooves when the cutter is thrust outward on the bearing shaft to define at least one axial clearance to permit unrestricted axial movement of the seal assembly between the radial end walls;
the diameter of the journal bearing being intermediate the circumferential walls of the cutter and shaft seal grooves to decrease axial movement of the rigid ring with respect to axial cutter movement to a selected ratio during drilling.
8. The invention defined by Claim 7 wherein the ratio of the seal movement to axial cutter movement is selected to decrease seal movement, consistent with the following formula:
Where: S = Seal Movement For Unit Cutter Movement Relative to the Shaft H = Effective Shaft O-Ring Annular Area C = Effective Cutter O-Ring Annular Area A = Effective Rigid Ring Annular Area G = Effective Cutter Seal Groove Annular Area
Where: S = Seal Movement For Unit Cutter Movement Relative to the Shaft H = Effective Shaft O-Ring Annular Area C = Effective Cutter O-Ring Annular Area A = Effective Rigid Ring Annular Area G = Effective Cutter Seal Groove Annular Area
9. The invention defined by Claim 8 wherein each of the resilient energizer rings is of the O-ring type and the ratio of rigid ring movement to axial cutter movement is substantially one half to one.
10. An earth boring bit with an improved pressure compensating, rigid face seal assembly having one rigid ring and scaling face carried by a bearing shaft and an opposed sealing face carried by the cutter, said bit comprising:
a body which includes at least one leg and a cantilevered and threaded bearing lug that extends downwardly and inwardly;
a lubrication system formed at least partially in the body, including a hydrostatic pressure compensator to lubricate said bearings;
the bearing lug including a threaded inner end and an outer base region with a transverse shoulder secured to the body;
a bearing sleeve with a substantially cylindrical journal bearing surface containing an assembly groove for a cutter retainer means, a threaded interior secured to the the threaded end of the bearing lug, and a mouth engaging the transverse shoulder of the lug;
a rotatable cutter having an interior cylindrical bearing surface, a retainer groove that registers with the assembly groove in the bearing sleeve, and an annular, cutter seal groove in the outer end of the bearing surface;
retainer means between the assembly groove and the retainer groove;
a cutter seal groove formed near the outermost region of the cylindrical bearing in the cutter to have a circumferential, generally cylindrical wall and a generally radial end wall;
a shaft seal groove formed at least partially in the outer end of the bearing sleeve and the transverse shoulder of the bearing lug to oppose the cutter seal groove, and having a circurmferential, generally cylindrical wall and a generally radial end wall;
a pair of rigid rings positioned in the seal groove to have opposed, sealing faces;
a pair of resilient energizer rings, each of which sealingly engages a respective one of the rigid rings and one of the oppositely facing circumferential walls of the cutter and shaft seal grooves to define a seal assembly;
the axial width of the rigid rings being less than the axial width of the shaft and cutter grooves when the cutter is thrust outward on the bearing shaft to define at least one axial clearance to permit unrestricted axial movement of the seal assembly between the radial end walls;
the diameter of the journal bearing being intermediate the circumferential walls of the cutter and shaft seal grooves to decrease axial movement of the seal assembly to a selected ratio when the cutter moves axially on the bearing shaft during drilling.
a body which includes at least one leg and a cantilevered and threaded bearing lug that extends downwardly and inwardly;
a lubrication system formed at least partially in the body, including a hydrostatic pressure compensator to lubricate said bearings;
the bearing lug including a threaded inner end and an outer base region with a transverse shoulder secured to the body;
a bearing sleeve with a substantially cylindrical journal bearing surface containing an assembly groove for a cutter retainer means, a threaded interior secured to the the threaded end of the bearing lug, and a mouth engaging the transverse shoulder of the lug;
a rotatable cutter having an interior cylindrical bearing surface, a retainer groove that registers with the assembly groove in the bearing sleeve, and an annular, cutter seal groove in the outer end of the bearing surface;
retainer means between the assembly groove and the retainer groove;
a cutter seal groove formed near the outermost region of the cylindrical bearing in the cutter to have a circumferential, generally cylindrical wall and a generally radial end wall;
a shaft seal groove formed at least partially in the outer end of the bearing sleeve and the transverse shoulder of the bearing lug to oppose the cutter seal groove, and having a circurmferential, generally cylindrical wall and a generally radial end wall;
a pair of rigid rings positioned in the seal groove to have opposed, sealing faces;
a pair of resilient energizer rings, each of which sealingly engages a respective one of the rigid rings and one of the oppositely facing circumferential walls of the cutter and shaft seal grooves to define a seal assembly;
the axial width of the rigid rings being less than the axial width of the shaft and cutter grooves when the cutter is thrust outward on the bearing shaft to define at least one axial clearance to permit unrestricted axial movement of the seal assembly between the radial end walls;
the diameter of the journal bearing being intermediate the circumferential walls of the cutter and shaft seal grooves to decrease axial movement of the seal assembly to a selected ratio when the cutter moves axially on the bearing shaft during drilling.
11. The invention defined by Claim 10 wherein the ratio of the seal movement to axial cutter movement is selected to decrease seal movement, consistent with the following formula:
Where: S = Seal Movement For Unit Cutter Movement relative to the Shaft H = Effective Shaft O-Ring Annular Area C = Effective Cutter O-Ring Annular Area A = Effective Rigid Ring Annular Area G = Effective Cutter Seal Groove Annular Area
Where: S = Seal Movement For Unit Cutter Movement relative to the Shaft H = Effective Shaft O-Ring Annular Area C = Effective Cutter O-Ring Annular Area A = Effective Rigid Ring Annular Area G = Effective Cutter Seal Groove Annular Area
12. The invention defined by Claim 11 wherein each of the resilient energizer rings is of the O-ring type and the ratio of rigid ring movement to axial cutter movement is substantially one half to one.
13. An earth boring bit with an improved seal means and pressure compensating system, said bit comprising:
a body which includes at least one leg and a cantilevered and threaded bearing lug that extends downwardly and inwardly;
a lubrication system formed at least partially in the body, including a hydrostatic pressure compensator to lubricate said bearings;
the bearing lug including a threaded inner end and an outer base region with a transverse shoulder secured to the body;
a bearing sleeve with a substantially cylindrical journal bearing surface containing an assembly groove for a cutter retainer means, a threaded interior secured to the the threaded end of the bearing lug, and a mouth engaging the transverse shoulder of the lug;
a rotatable cutter having an interior cylindrical bearing surface, a retainer groove that registers with the assembly groove in the bearing sleeve, and an annular, cutter seal groove in the outer end of the bearing surface;
retainer means between the assembly groove and the retainer groove;
a cutter seal groove formed near the outermost region of the cylindrical bearing in the cutter to have a circumferential, generally cylindrical wall and a generally radial end wall;
a shaft seal groove formed at least partially in the end of the bearing sleeve and the transverse shoulder of the bearing lug to oppose the cutter seal groove, and having a circumferential, generally cylindrical wall and a generally radial end wall;
a pair of rigid rings positioned in the seal groove to have opposed and engaging sealing faces;
a pair of resilient energizer rings, each of which sealingly engages a respective one of the rigid rings and one of the oppositely facing circumferential walls of the seal groove to define a seal assembly positioned between between the end walls of the seal groove;
the seal assembly being exposed to and moved axially by dynamic pressure differentials between the lubricant and the ambient drilling fluids;
the axial width of the engaged rigid rings and seal assembly being less than the axial, minimum width of the seal groove when the cutter is thrust outwardly on the bearing shaft to define at least one axial clearance to permit axial movement of the rigid rings between the end walls of the groove when the cutter moves relative to the bearing shaft;
the diameter of the journal bearing being intermediate the circumferential walls of the seal groove to decrease axial movement of the seal assembly to a selected ratio when the cutter moves axially on the bearing shaft.
a body which includes at least one leg and a cantilevered and threaded bearing lug that extends downwardly and inwardly;
a lubrication system formed at least partially in the body, including a hydrostatic pressure compensator to lubricate said bearings;
the bearing lug including a threaded inner end and an outer base region with a transverse shoulder secured to the body;
a bearing sleeve with a substantially cylindrical journal bearing surface containing an assembly groove for a cutter retainer means, a threaded interior secured to the the threaded end of the bearing lug, and a mouth engaging the transverse shoulder of the lug;
a rotatable cutter having an interior cylindrical bearing surface, a retainer groove that registers with the assembly groove in the bearing sleeve, and an annular, cutter seal groove in the outer end of the bearing surface;
retainer means between the assembly groove and the retainer groove;
a cutter seal groove formed near the outermost region of the cylindrical bearing in the cutter to have a circumferential, generally cylindrical wall and a generally radial end wall;
a shaft seal groove formed at least partially in the end of the bearing sleeve and the transverse shoulder of the bearing lug to oppose the cutter seal groove, and having a circumferential, generally cylindrical wall and a generally radial end wall;
a pair of rigid rings positioned in the seal groove to have opposed and engaging sealing faces;
a pair of resilient energizer rings, each of which sealingly engages a respective one of the rigid rings and one of the oppositely facing circumferential walls of the seal groove to define a seal assembly positioned between between the end walls of the seal groove;
the seal assembly being exposed to and moved axially by dynamic pressure differentials between the lubricant and the ambient drilling fluids;
the axial width of the engaged rigid rings and seal assembly being less than the axial, minimum width of the seal groove when the cutter is thrust outwardly on the bearing shaft to define at least one axial clearance to permit axial movement of the rigid rings between the end walls of the groove when the cutter moves relative to the bearing shaft;
the diameter of the journal bearing being intermediate the circumferential walls of the seal groove to decrease axial movement of the seal assembly to a selected ratio when the cutter moves axially on the bearing shaft.
14. The invention defined by Claim 13 wherein the ratio of the seal movement to axial cutter movement is selected to decrease seal movement, consistent with the following formula:
S = Where: S = Seal Movement For Unit Cutter Movement Relative to the Shaft H = Effective Shaft O-Ring Annular Area C = Effective Cutter O-Ring Annular Area A = Effective Rigid Ring Annular Area G = Effective Cutter Seal Groove Annular Area
S = Where: S = Seal Movement For Unit Cutter Movement Relative to the Shaft H = Effective Shaft O-Ring Annular Area C = Effective Cutter O-Ring Annular Area A = Effective Rigid Ring Annular Area G = Effective Cutter Seal Groove Annular Area
15. The invention defined by Claim 14 wherein each of the resilient energizer rings is of the O-ring type and the ratio of rigid ring movement to axial cutter movement is substantially one half to one.
Applications Claiming Priority (2)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| US07/023,170 US4753303A (en) | 1983-10-17 | 1987-03-09 | Earth boring bit with two piece bearing and rigid face seal assembly |
| US023,170 | 1987-03-09 |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| CA1292220C true CA1292220C (en) | 1991-11-19 |
Family
ID=21813499
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA000559193A Expired - Fee Related CA1292220C (en) | 1987-03-09 | 1988-02-18 | Earth boring bit with two piece bearing and rigid face seal assembly |
Country Status (3)
| Country | Link |
|---|---|
| CA (1) | CA1292220C (en) |
| IT (1) | IT1216038B (en) |
| MX (1) | MX166269B (en) |
-
1988
- 1988-02-18 CA CA000559193A patent/CA1292220C/en not_active Expired - Fee Related
- 1988-03-09 IT IT8819706A patent/IT1216038B/en active
- 1988-03-09 MX MX1070188A patent/MX166269B/en unknown
Also Published As
| Publication number | Publication date |
|---|---|
| IT8819706A0 (en) | 1988-03-09 |
| IT1216038B (en) | 1990-02-22 |
| MX166269B (en) | 1992-12-28 |
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Legal Events
| Date | Code | Title | Description |
|---|---|---|---|
| MKLA | Lapsed |