Title:
Valve for hydraulic rock drill
Kind Code:
A1


Abstract:
A concentric valve for a hydraulic rock drill featuring simple and direct fluid connections, especially suited for use with fluid storage volumes.



Inventors:
Paul Jr., Campbell B. (Troutville, VA, US)
Application Number:
10/178980
Publication Date:
01/09/2003
Filing Date:
06/25/2002
Assignee:
CAMPBELL PAUL B.
Primary Class:
International Classes:
B25D9/12; B25D9/14; B25D9/20; E21B1/26; (IPC1-7): E21B1/00
View Patent Images:
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Primary Examiner:
LOPEZ, MICHELLE
Attorney, Agent or Firm:
Paul B. Campbell, Jr. (Troutville, VA, US)
Claims:

I claim:



1. A hydraulic rock drill, comprising a. a housing having a piston bore and valve bore formed therein b. a piston axially disposed in said piston bore and a valve axially disposed in said valve bore c. said valve bore being concentric with said piston bore d. said housing containing a high pressure fluid reservoir e. said housing containing a low pressure fluid reservoir f. said piston bore having short and direct connections to said high pressure reservoir g. said piston bore having short and direct connections to said low pressure reservoir h. said valve bore having short and direct connections to said high pressure reservoir i. said valve bore having short and direct connections to said low pressure reservoir

2. A hydraulic rock drill as in claim 1 wherein said valve moves under the influence of pressure alone without physical contact between said valve and said piston.

Description:

CROSS-REFERENCE TO RELATED APPLICATIONS

[0001] This invention is used on the hydraulic rock drill of co-pending applications entitled “Steel Retainer for Rock Drill” and “Operating System for Hydraulic Rock Drill”, both filed Jun. 25, 2002. This application is entitled to the benefit of Provisional Patent Application Ser. No. 60/300,891, filed Jun. 25, 2001.

BACKGROUND

[0002] 1. Field of Invention

[0003] This invention relates specifically to a hydraulic rock drill, particularly in the way hydraulic fluid is directed to generate reciprocating motion of an impact piston.

[0004] 2. Description of Prior Art

[0005] A percussive rock drill is a device that, in conjunction with a drill bit, uses rotation and percussive energy to drill a hole in rock for purposes of blasting, etc. Every fluid operated percussive rock drill includes certain basic features. A striking piston imparts impact energy to a drill steel and bit, and a valve mechanism directs the working fluid so as to cause reciprocating motion of the piston. A drill bit is attached to the end of the drill steel. A rotation mechanism rotates the drill steel to give the bit a fresh rock surface to strike with each blow, and the combination of impact and rotation causes the drill bit to penetrate the rock. A drill steel retention mechanism allows removal of the drill steel and bit when the hole is completed. Flushing fluid (typically air or water) travels through holes in the drill steel and bit to blow rock cuttings out of the drilled hole.

[0006] In a typical operation of a hydraulic rock drill, the striking piston is caused to reciprocate by variable hydraulic forces. The hydraulic forces are generated by hydraulic fluid that is directed by a valve mechanism. The drill steel is constrained and located by a chuck mechanism and a steel retainer, and is caused to rotate by a mechanism such as a hydraulic motor driving through a gear reduction. Finally, some type of fluid energy storage mechanism is used to provide relatively constant pressure sources of working fluid for the piston and rotation.

SUMMARY

[0007] An object of the present invention is to eliminate the need for troublesome nitrogen-charged accumulators by providing fluid reservoirs that minimize pressure spikes in hoses and provide relatively constant pressure sources of working fluid. A second object is to provide a simplified fluid valving mechanism that obtains the maximum benefit from the fluid reservoirs.

DRAWING FIGURES

[0008] FIGS. 1 through 6 show the operation of the improved valve mechanism in conjunction with the fluid storage volumes.

[0009] The first improvement relates to the method of providing for suppression of pressure spikes in fluid hoses without the use of nitrogen-charged accumulators. U.S. Pat. No. 3,918,532 (Lester A. Amtsberg, Nov. 11, 1975) describes a “bulk oil accumulator” that is part of the housing of the tool and provides a reservoir of high pressure fluid. The claimed benefits are short passages between the fluid reservoir and the impact piston and valve. The present invention offers these benefits to a greater degree, and adds others. The subject of the previous patent did not take maximum advantage of the possibility of short passages because the valve axis was oriented at right angles to the piston axis and hence the valve was removed from the piston by a finite distance. In the present invention, the valve is oriented coaxially with the piston, with the shortest possible fluid connections between the fluid reservoir and the working shoulders on the piston. In addition, the previous invention did not provide a reservoir to receive fluid that was exhausted from the valve. Most commercial hydraulic drills in use today have both high pressure and exhaust pressure accumulators because the pressure spikes in the exhaust pressure line can be just as detrimental to hose life as those in the high pressure line. Large fluid reservoirs offer two primary advantages over nitrogen-charged accumulators. The first advantage is in the simplification of internal porting, for improved performance and ease of manufacture. Connections between the valve and piston and the fluid reservoirs can be made with short straight holes, rather than a series of longer intersecting holes drilled at angles to one another. The second advantage is in improved maintenance and serviceability. Unlike nitrogen-charged accumulators, fluid reservoirs require no nitrogen gas bottles or charging valves and have no dynamic seals or diaphragms to fail. The function of the high pressure and exhaust pressure fluid reservoirs will be more fully described in the working cycle description as shown in FIGS. 1 through 6.

[0010] The second improvement relates to a simplified fluid valve mechanism. Fluid valve mechanisms that cause reciprocating motion of an impact piston in a hydraulic impact device may be roughly categorized into two types. The first type is concentric or coaxial with the piston. Various concentric valve configurations are the subjects of U.S. Pat. Nos. 3,766,830 (Roger Montabert, Oct. 23, 1973) and 4,005,637 (John V. Bouyoucos et al, Feb. 1, 1977). These and other concentric valve configurations commonly have one of two characteristics. They are either switched by physical contact with the impact piston (with the possibility of unpredictable valve motion because of rebound effects), or they are switched by contact with varying fluid pressure that requires long passages in the drill housing. The second type of valve mechanism might best be described as a spool valve configuration wherein the valve is installed adjacent to the piston but not concentric with it. The valve may be switched by pilot pressure that is controlled by the piston position or by some other means such as a motor. Various spool valve configurations are the subjects of U.S. Pat. Nos. 3,918,532 (Lester A. Amtsberg, Nov. 11, 1975), 4,192,447 (Eugene L. Krasnoff et al, Mar. 6, 1979), and 5,031,505 (Paul B. Campbell, Jul. 16, 1991). The single common characteristic of spool valves is the requirement for multiple flow passages in the drill housing. These flow passages may be used to conduct working fluid to and from the impact piston, or to conduct pilot fluid to the valve, or both, but they always add complication to the drill assembly. The present invention accomplishes reliable and predictable valve switching while requiring neither physical contact with the piston nor long passages in the drill housing. The operation of the cycle is described and shown in FIGS. 1 through 6.

[0011] Impact piston 12 is installed in bore of body 10 and is free to slide axially therein, subject to pressures acting on shoulder areas 72 and 48. Shoulder 48 is substantially larger than shoulder 72. Shoulder 72 is subjected to constant supply pressure by connection through passages 22 and 24 to supply pressure volume 20. Supply pressure volume 20 is connected to a source of high pressure fluid through passage 30. Shoulder 48 is subjected to varying drive pressure. Drive pressure is either supply pressure or exhaust pressure, depending on the position of valve 14.

[0012] Valve 14 is installed in bore of body 10 and bore of valve seat 16 and is free to slide axially therein, subject to pressures acting on shoulder areas 50, 60, 62, 66, and 68. The sum of shoulder areas 60 and 62 is less than the area of shoulder 50. Shoulder area 66 is greater than shoulder area 68. Shoulder 62 is subjected to constant exhaust pressure by connection through fluid passages 34, 36, and 38 to exhaust pressure volume 18. Shoulder 60 is subjected to exhaust pressure or drive pressure depending on the positions of valve 14 and piston 12. Shoulders 66 and 68 are subjected to constant supply pressure by connection through fluid passages 26 and 28 to supply pressure volume 20. Shoulder 50 is subjected to varying drive pressure.

[0013] FIG. 1 represents the positions of piston 12 and valve 14 just at the moment of impact against drill steel 78. During the impact stroke, chamber 76 has been connected to supply pressure volume 20 through passages 26, 28, 42, and 44; drive pressure has been substantially equivalent to supply pressure; and piston 12 has been moving to the right. Valve 14 is also moving to the right and edge 74 has just crossed body edge 70, breaking the connection between fluid passage 28 and fluid passage 44 and hence disconnecting drive pressure chamber 76 and supply pressure volume 20. Piston shoulder 48 has crossed body edge 46, opening a fluid connection between chamber 76 and exhaust pressure volume 18 through fluid passage 40. Fluid exits exhaust pressure volume 18 through passage 32. Drive pressure is now substantially equivalent to exhaust pressure.

[0014] FIG. 2 represents the beginning of the piston return stroke. Supply pressure acting on area 72 creates a larger force than exhaust pressure acting on area 48, so the net force on piston 12 is to the left and piston 12 begins to move leftward. Momentum continues to carry valve 14 to the right until valve shoulder 68 contacts body shoulder 70. As piston 12 moves to the left, shoulder 48 passes body edge 46 and breaks the fluid connection between chamber 76 and exhaust pressure volume 18. However, fluid continues to exit chamber 76 into exhaust pressure volume 18 in the clearance between piston 12 and valve 14 and through passages 38, 36, and 34.

[0015] FIG. 3 represents the beginning of the valve return stroke. Piston edge 56 has crossed valve edge 58, breaking the fluid connection between chamber 76 and fluid passage 38. Fluid trapped in chamber 76 pushes on valve shoulder 50 and causes valve 14 to be carried to the left with piston 12. Fluid to the left of valve shoulders 60 and 62 exits to exhaust pressure volume 18 through passages 38, 36, and 34.

[0016] FIG. 4 represents the deceleration portion of the piston return stroke. Valve edge 74 crosses body edge 70, establishing a fluid connection between chamber 76 and supply pressure volume 20 through passages 26, 28, 42, and 44. Drive pressure is now substantially equivalent to supply pressure, and supply pressure acting on shoulder 48 creates a larger force than supply pressure acting on shoulder 72 so the net force on piston 12 is to the right and piston 12 begins to decelerate. The leftward motion of valve 14 is arrested by contact between valve shoulder 62 and valve seat shoulder 64. Piston 12 continues moving to the left until its momentum is overcome by the net force to the right, at which time the piston reverses direction and the impact stroke begins.

[0017] FIG. 5 represents the piston impact stroke. The net force generated by supply pressure acting on shoulders 50, 60, 66, and 68 exceeds the net force generated by exhaust pressure acting on shoulder 62, and valve 14 is held against shoulder 64. Shoulders 48 and 72 on piston 12 are both exposed to supply pressure, but shoulder 48 is substantially larger than shoulder 72 so piston 12 accelerates rapidly to the right.

[0018] FIG. 6 represents initiation of the valve switch just before impact. Shoulder 52 on piston 12 crosses over edge 54 on valve 14, closing off the connection between chamber 76 and fluid passage 44. Simultaneously, shoulder 48 on piston 12 crosses edge 46 on body 10, opening the connection between chamber 76 and exhaust pressure volume 18 through passage 40. Drive pressure is now substantially equivalent to exhaust pressure. Momentum carries piston 12 to impact with drill steel 78. Valve 14 is exposed to exhaust pressure acting against shoulders 50, 60, and 62, and supply pressure acting against shoulders 66 and 68. Since shoulder 66 is larger than shoulder 68, the net force on valve 14 is to the right and it begins to move away from contact with shoulder 64. Shoulder 74 on valve 14 crosses shoulder 70 on body 10, closing fluid passage 42 and breaking all connection between supply pressure volume 20 and chamber 76 without regard to the position of piston 12. Impact occurs and the cycle repeats.

[0019] The simplicity and reliability of the operating cycle are obvious from the description. The common disadvantages of both concentric and spool type valves are eliminated. Fluid porting and control of valve switching do not require long drilled holes, so manufacturing is easier and fluid flow is more predictable, and the elimination of impact between piston and valve results in a more predictable valve motion. The short connecting passages between the piston and valve and the high pressure and low pressure reservoirs permit realization of the maximum benefit from the reservoirs.

CONCLUSION, RAMIFICATIONS, AND SCOPE

[0020] The preferred embodiment as described herein includes the improvements in combination, i.e. the high pressure and exhaust pressure fluid volumes, and the simplified valve mechanism. However, the individual improvements are not limited to use in this particular combination; an alternative embodiment may contain any one of the improvements. For example, the high pressure and exhaust pressure reservoirs could be used with a different type of valve mechanism, or the simplified valve mechanism could be used with nitrogen-charged accumulators.

[0021] Thus the scope of the invention should be determined by the appended claims and their legal equivalents, rather than by the examples given.