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Publication numberUS3532174 A
Publication typeGrant
Publication dateOct 6, 1970
Filing dateMay 15, 1969
Priority dateMay 15, 1969
Publication numberUS 3532174 A, US 3532174A, US-A-3532174, US3532174 A, US3532174A
InventorsDiamantides Nick D, Hinks William L
Original AssigneeDiamantides Nick D, Hinks William L
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Vibratory drill apparatus
US 3532174 A
Abstract  available in
Images(9)
Previous page
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Claims  available in
Description  (OCR text may contain errors)

Inventors Nick I). Diamantides [561 2517 14th St., Cuyahoga Falls, 44223; William L. Rinks, 2449 Kensington Ave., 1 1 12 498 Bath, Ohio 44210 l365l Appl. No. 850,682 29463l4 Filed May 15, 1969 295335l Patented Oct. 6, 1970 3,307,641 3,371,726

References Cited UNITED STATES PATENTS 10/1914 Vanes 4/1938 Heaston........... 7/1960 Nast 9/1960 Bodine 3/1967 Wiggins 3/1968 Bouyoucosu Primary Examiner-Nile C. Byers, .lr. AttorneyN. D. Diamantides VIBRATORY DRILL APPARATUS l75/56X l75/56X l75/56X 175/56 175/56 l75/56X ABSTRACT: The subject matter of this invention is a rock 58 claimsza Drawing Figs drill apparatus whose working member is driven, both, to a 0.8. CI..... 175/56 high frequency longitudinal vibration and to indexing through Int. Cl. E2lb 7/04 rotation, and which is powered by the pressurized fluid Field of Search 175/56, customarily used in removing the rock debris from the drilled 107; 253/22, 33 hole.

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s Q ,0 I f nu i w 11 I40. I2a L 40 E- 12b /l4b l3b I I O l Patented Oct. 6, W70

Sheet INVENTORS N. D. DIAMANTIDES W.L. HINKS Patented Oct. 6, 1970 3,532,114

Sheet 2 of 9 N. D. DIAMANTIDES k BY w.1 HJNKS 345 Patented Oct. 6, 1970 Sheet INVENTORS N. D. DIA MANTIDES BY W. L. HINKS Patented 0a.. 6, 1970 3,532,174

Sheet {5 of9 33E INVENTORS 515 Be. v N. 0. DIAMANTIDES BY W.L. HINKS Patented 0a,. .6, 1970 3,532,174

- Sheet 6 019 V 5,, gas ii INVENT 22 N. D. D/AMANT S BY W. L'. HINKS Patented Oct. 6, 1970 Sheet 7 INVENTORS N. D, DIAMANTIDES W.L. H/NKS Patented Oct. 6, 1970 Sheet of 9 W. l- HINKS VIBRATORY DRILL APPARATUS This application is a continuation of Ser. No. 734,048 filed June 3, 1968, now abandoned.

This invention refers in general to the field of rock penetrating tools, and in particular to a rock drill.

It is well known to those versed in this art that rock penetration for the purposes of ground sampling, mining, quarrying, oil exploitation, and tunnel construction faces many technical and economic problems dependent both on the geological forplicity of handling and repairing.

A further object of the invention is to provide a rock drill requiring a minimum of peripheral and auxiliary equipment and a limited on-the-site rig structure.

Still another object is to provide a rock drill whose function .rests on a rotary-vibratory movement of its working head, and

is powered by a compressed fluid source.

A further object is a rock drill in which vibratory movement is imparted to the working head of the drill by a mechanism involving no solid parts sliding on one another, but, instead, allowing a minimum number of such parts to move with respect to each other by means of a special deformable connecting member.

Additional objects of the invention will become apparent upon reading of the following specification, considered and interpreted in the light of the accompanying drawings in which:

FIG. I is a schematic drawing showing the four principal components of their rock drill and their preferred mechanical interconnection.

FIG. 2 is a cross-sectional view of one configuration of the rock drill equipped with a tricone type bit and showing the arrangement of laminate bearings connecting the principal drill components.

FIG. 2a is a partial cross-sectional view of the vibrator and head showing the use of Bellville type springs in place of laminate bearings.

FIG. 2b is a partial cross section of the vibrator and head showing the use of a resilient elastomer ring in place of laminate bearings.

FIG. 3 is a cross-sectional view of the vibrator and head, equipped with a fish tail or drag type bit, illustrating fluidic oscillator outlets in their association with the annular piston chambers for the generation of the fluid forces that set the vibrator and head in reciprocating motion.

FIG. 4 is a cross-sectional view of the vibrator and head illustrating the fluidic oscillator arrangement and its feedback paths emanating from the piston chambers.

FIG. 4a is a cross-sectional view of the head and vibrator illustrating the use of individual laminate bearings at each piston chamber.

FIG. 5 is a schematic diagram of the fluidic oscillator aiding the explanation of the basic principles of its operation.

FIG. 6 is a cross-sectional view of the vibrator and head showing a mechanical oscillator arrangement and its association with the annular piston chambers for the generation of the fluid forces that set the vibrator and head in reciprocating motion.

FIG. 6a is a detailed drawing of an alternate arrangement of the jet nozzles and passages of the mechanical oscillator and vibrator.

FIG. 6b is a detailed drawing of still another arrangement of jet nozzles and passages of the mechanical oscillator and vibrator.

FIG. 60 is a detailed drawing showing an alternate arrangement of spent fluid paths through the mechanical oscillator, vibrator and head.

FIG. 6d is a detailed drawing showing the connection of the mechanical oscillator to the vibrator and head by means of laminate bearings and a viscous damper respectively.

FIG. 6e is a detailed drawing showing still another connection of the mechanical oscillator to the vibrator and head with the positions of the laminate bearings and viscous damper of the previous figure reversed.

FIG. 7 is a cross-sectional view of the vibrator showing the incorporation of a fluidic oscillator for imparting a reciprocating motion to the mechanical oscillator.

FIG. 8 is a cross-sectional view of the vibrator and head showing an arrangement of the fluidic oscillator conducive to introducing a motional feedback to the function of the oscilla- I01.

FIG. 9 is a cross-sectional view of the vibrator and head showing an arrangement whereby reaction jets are used to generate the vibratory forces required by the present drill.

FIG. 10 is a cross-sectional view of the rock drill showing a mechanical interconnection between head, vibrator, and drill stem different from that of FIG. 1.

FIG. 11 is a cross-section view of the rock drill showing an arrangement of its parts whereby the bit is attached to the vibrator and the stern adapter to the head.

FIG. 12 is a partially cross-sectional elevated view of the vibrator and head illustrating the incorporation of a hydraulic wheel for the generation of the torque that rotates the head.

FIG. 12a is a plane cross-sectional view of the vibrator and head along the line Ila-12a of FIG. 12.

FIG. 13 is a partially cross-sectional elevated view of the vibrator and head illustrating a special laminated bearing arrangement for implementing a stepped rotation of the head and bit as well as axial vibration of it.

FIG. 14 is a cross-sectional view of the rock drill showing still another mechanical interconnection between head, vibrator and drill stem.

FIG. 15 is a schematic diagram of the equivalent mechanical circuit of the arrangement of the :rock drill components shown in FIGS. 2, 3, 4, 8, and 9.

FIG. 16 is a schematic diagram of the equivalent mechanical circuit of the arrangement of the rock drill components shown in FIG. 10.

FIG. 17 is a schematic diagram of the equivalent mechanical circuit of the arrangement of the :rock drill components shown in FIGS. 6 and 7. 7

FIG. 18 is a schematic diagram of the equivalent mechanical circuit of the arrangement of the :rock drill components shown in FIG. 14.

FIG. 19 is a typical force-displacement characteristic associated with the crushing of a rock by drill.

Referring now to the preferred mechanical configuration of FIG. 1, it will be seen that the rock drill of the present invention comprises in combination the stern adapter 10, which is the lower end part of the drill stem 10a conventionally used in rock drilling; the vibrator 20; the head 40; and the bit 50. Details of their mechanical interconnection may be seen in FIGS. 3 and 6, in which two of the drills configurations are shown and in FIG. 2 which applies to both. In all three cases the massive vibrator 20 is attached resiliently in the axial direction to the stem adapter 10 by means of the interposed laminate elastomeric bearings or bushings 11. A telescoping arrangement, with annular bearings is preferred, as shown. The stem adapter 10 is indicated exterior to the top of vibrator 20, but obviously the top of vibrator 20 could be made the exterior part and the stem adapter 10 interior, with the annular bearings 11 interposed between. Two axially separated bearing assemblies are shown, but there could be instead one long bearing assembly. The total amount of rubber in shear must be able to sustain the axial load applied by the stem. Similarly, the massive head 40 is attached in an axially resiliently stiff manner to the vibrator 20 preferably by means of the laminate elastomeric bearings 12a and 12b. One or all bearings 11, 12a, 12b may be of the rubber sleeve, instead of laminate type. In both configurations the sections 13 and] 13b of the vibrator 20 have a polygonal cross section (in a plane perpendicular to the longitudinal axis) as do the respectively opposed sections 14 and 14b of the head 40 and stem adapter 10. By contrast, the laminate bearing 12a may be made conical and split in two halves radially to facilitate assembling the vibrator-head system. The polygonal cross sections, in fact any noncircular shape, enable torque to be transmitted from the stem adapter 10 to the vibrator 20 and from the latter to the head 40 effectively without windup. A round cross section except for one or more flat sides could accomplish the same purpose. The bearings 11, 12a, and 12b have the additional function of serving as seals between the various sections of the flow path of the pressurized drill fluid. Finally, the bit 50 is affixed rigidly or in a drivable manner to the head 40 by means of the tapered threaded sections 51 (bit not shown in FIG. 6). The bit is shown to be of the tricone type in FIG. 2 and of the fish tail or drag type in FIG. 3. Obviously, a coupler could be interposed between head and bit for the purpose of increasing mass or length or inserting compliance or damping between head and bit. in the same two FIGS. 2 and 3 a shoulder 5 is shown interior to central channel 7 of the stem adapter 10 that serves as a mechanical stop against accidental overloading of the bearings 11.

In the aforementioned designs shown in FIGS. 2. 3, 4, and 6, the massive vibrator is interior to the massive head 40. However, as disclosed later these relative positions may be interchanged. Since either configuration may be used, it is deemed appropriate to use the terms inner massive member and outer massive member when referring to both configurations collectively.

The laminate elastomeric bearings constitute an important member of the preferred configuration on account of their proven unique mechanical behavior. A bearing of this kind consists of a stack of thin metal laminates interleaved and adhered together by thin alternating layers of elastic rubber or other rubberlike material. Such a layer of rubberlike material bonded between metal laminae can withstand high compressive loads applied by the metal layers, it being sufficiently thin as to be restrained from substantially flowing sidewise by its adhesion to the metal. Elastomeric bearings, however, are capable of a deformation in shear parallel to the laminae accompanied by shear stress proportional to the deformation. Thus the elastomeric bearing behaves in a soft springlike fashion in a plane parallel to the laminae, and its spring rate may be adjusted as desired through proper selection of the dimensions of the stack and the characteristics of the elastomer. This spring rate, coplanar with the laminae, is little effected by any compressive forces that may be applied perpendicularly to the layers.

The foregoing mode of bearing operation is the one under which the bearings 11 between stem adapter 10 and vibrator 20 are indicated to function with layers parallel to the longitudinal axis. However, this is not necessarily the case with the bearings 12a and 12b between vibrator 20 and head 40, for the latter bearings have to be much stiffer axially than the bearings 11 and yet able to withstand the severe fatigue stresses necessitated by the present drill design. To accomplish this within small dimensions and without the need for a very stiff elastomer, the laminate bearings 12a, 12b are set with their load faces at an angle with respect to the motion 2: this produces not only shear stresses parallel to the laminae but, also, compression perpendicular to the laminae, with the net result that the equivalent spring constant is increased, by orders of magnitude if necessary, as needed for a high frequency oscillating system. The sloped bearings 12a and 1212 will generally be preloaded against one another so that, within the oscillation amplitude experienced, neither may become completely decompressed. An opposingly preloaded pair of bearings is thus provided placing the vibrator in compression. The vibrator will be subjected to tension instead if the sloping angles of both bearings is reversed.

It should be emphasized that by their nature, elastomeric bearings are capable of prolonged life with proper design,

without lubrication and without fatigue and failure, withstanding severe vibratory deformations under considerable static loading. Regarding the two constituents of the laminate elastomeric beaiings itshotlld belipderstood that the term metafas used herein encompasses any ofi'f'ir 'iBM of metals, or eyen a nonmetallic material, that is characterized by a high compression resistence, as well as high tensile strength; similarly, when the term elastomeric or rubberlike" is used, it is intended to include any and all rubber base products as well as all equivalent resilient products that could be used and which are characterized by a relatively low resistance to shear.

The basic mode of operation of the preferred configuration of our drill involves conventional rotation of the drill stem 10, FIGS. 2 and 3, in the direction of the arrow 1, by applying sufficient torque to it at the surface opening of the well, and, at the same time excitation of the vibrator 20 into axial reciprocating motion indicated by the double arrow 2. This latter motion is imparted to the vibrator by a hydraulic actuator at the expense of some of the energy stored in the pressurized fluid, this fluid normally being sent downhole for the purpose of clearing the hole of the rock debris. Such a fluid may be natural gas, air, water,,oil, or drilling mud as is the common practice in rock drilling. The hydraulic actuator is preferably of the piston type, comprising multiple flanges 35 and 41 formed into the vibrator 20 and head 40 respectively, whereby fluid under pressure in alternate chambers 34a, 34b, enclosed between 35 and 41, may exert an axial force as will be described. The flanges 41 are indicated to be formed, for ease of assembly, as part of an inner shell 40a within the outer shell 40b of the head 40. The inner shell 40a may be split lengthwise into two halves to make assembly possible, and a threaded retainer ring 40c, with a conical load face to the bearing 12a, is used to hold the assembly together under preload. Other assembly means are possible.

Pressurized fluid is fed into the vibrator 20 by means of the inlet 21, is ported to the actuator by oscillatory means to be described, and is discharged out of the vibrator through the outlet 22. The fluid then enters a central hole 52 of the bit 50, or such other jets or courses as may be provided in the bit. As explained in a later paragraph, shortly after the vibrator is set in motion, oscillatory power is transmitted to the head 40 by means of the elastomeric bearings 12a, 12b, and by the hydraulic force generated by the pressurized fluid, provided that certain dynamic conditions are met; conditions involving the bearing resilience, the amounts of friction between moving parts, the magnitude of the masses of the stem, head, and vibrator, and the crushing strength of the particular rock. These conditions are investigated in a later part of this specification. The imparting of oscillatory power to the head 40 causes it, and the attached bit 50, to vibrate axially in the direction of the arrow 3, oppositely to the instantaneous direction of the vibrator mass 20, alternately compressing one of the bearings or 12b while simultaneously decompressing the other. On account of this vibration an alternating pressure component is added to the average pressure applied by the drill weight to the rock surface at the bottom of the hole. Thus, the rotary-vibratory character of the present drill is established by the combination of the vibratory movement 3 and the rotary movement 1. Because the masses of the vibrator 20, head 40 and bit 50 are purposely made small by comparison with the mass of the stem 10a and stem adapter 10, and because the elastomeric bearings 11 are designed to have the response of a soft spring to forced vibration, the stem is substantially prevented from participating in the oscillation of the rest of the members, thus preventing loss of energy and wear of the topside rig parts and equipment.

The stem adapter 10 is the lower end member of the drill stem 10a, FIG. 1, that is made up conventionally of sections of steel or aluminum pipe connected end-to-end by means of a threaded end section, and has a fourfold function; it carries the drilling fluid from the well surface to the bit 50 through a central channel 7; it forms the inner wall of the well annulus i that provides the return path for the drilling mud and rock debris; it helps provide static load on the bit necessary to effect rock penetration; and it transmits the torque necessary to effect rotation of the drill as a whole supplied by the engine at the well surface.

It would be possible to use some other types of bearing spring seals 12a, 12b than the laminated rubber-metal type preferred. For both the top and bottom location, a nested stack of the springy thin dishes or arcuate washers known as Bellville springs could provide the high stiffness over a short range that is needed and simultaneously provide the needed alignment and sealing action and could be arranged to transrnit torque by means of keying or polygonal shaping. This alternative is illustrated in FIG. for the bottom location only, where a stack of eight such Bellville springs are indicated as items 82.

Another possibility is to use a resilient elastomer with shear modulus so large that a direct use of it as an annular bearing spring seal without the need for interlaminations of metal could provide the needed high axial stiffness. In such a case, the angled load faces 13a, 13b, 14a, 14b would still be desirable, producing a combination of compression and shear loading, and obviating the need for bonding of the elastomer against the shear load. A view is shown in FIG. 2b of the bottom bearing seal 83 only with large-angled load faces. Some bulging of the elastomer may occur due to preloading plus relative displacement, and partial containment of it is provided at 84 and 85 by the cooperating shape of the parts 40 and 20 respectively. A polygonal shape of the load faces may again be used to facilitate torque transfer.

The definition of a bearing or bearing spring seal as used herein between the head 40 and vibrator 20 refers to an as sembly or group of related parts that provides the functions of alignment, some axial compliance, and a more or less complete sealing property, unless context infers otherwise.

An alternate method of providing bearing spring seal combinations is shown in FIG. 4a, in which a plurality of laminate bearings, rather than only a pair of bearings (12a, 12b as in FIGS. 2, 6), may be provided at any or all flanges 41 and 35 of the head 40 and vibrator 20 respectively, in the annular gap between head and vibrator. That is, the bearings 12c, 12d, 12c, lZf take the place of the usual clearance gap existing between the said flanges in FIGS. 2 and 6. This plurality of bearings can provide sealing between each piston and cylinder combination of the fluid actuator, as well as aligning the head 40 and vibrator 20 and providing a stiff axial spring effect between them. This has the advantage of preventing wear between the coacting flanges and walls due to an abrasive drilling mud, for instance.

Each bearing [20, 12d, l2e, 12f may be split to allow assembly, as may the shells 40a, as previously discussed. The threaded retaining ring 40c holds the assembly together under preload. Each bearing may be made with sloped faces to provide high axial spring force as shown in FIG. 4a, or some could be made with faces parallel to the longitudinal axis. The bearing faces may also be made with circular or polygonal cross sections as previously discussed.

In FIG. 40, bearings 12c and 12d are sloped oppositely to bearings 12c and 12f, thereby providing for preloading of the one type against the other. Other combinations of sloping of the individual bearings are obvious, together with the split shells corresponding to 40a that will permit assembly. Any of the several types of bearings discussed could be employed.

Another variation of bearing and actuator means involving only one bearing would be provided in the case of a long rubber or laminated elastomeric bearing sleeve interposed in the annulus between two long telescoped sections corresponding to members of vibrator 20 and head 40. At one end of this elastomeric sleeve, an actuator could be formed by a series of flanges alternately affixed to head and vibrator as already described relative to FIG. 2. The clearance gap between the last (farthest from hearing sleeve) flange and coacting wall part of the actuator would then simply be open to the outside of the drill. That is, the entire configuration would be as in FIG. 2 except that bearing 12b, say, would be much longer axially, and not sloped, and bearing 12a would be omitted.

I-Iaving indicated the basic mechanical configuration and mode of operation of our drill, we proceed now to describe the internal components of the device and to develop a great depth of detail concerning the operation of the various forms of the invention.

Clearly a very important component of the present rock drill is the vibrator 20 where the vibratory power is generated through a self-sustaining action. Its internal design involves either no moving solid parts whatsoever, or a single moving member whose motion requires no sliding of solid surfaces on one another. The two cases are illustrated in FIGS. 2, 3, 4 on the one hand and 6 on the other. The :nonmoving parts case, pictured in FIGS. 2, 3, and 4, is based on the alternating out put of a fluidic oscillator 23, which, in turn, is based on the principle of momentum exchange explainable with the aid of FIG. 5. A pressurized fluid, such as the drilling fluid in this case, is brought into the source chamber 24 through the inlet opening 21 and proceeds into the interaction chamber 26 from which it starts streaming into the output duct, say, 2712. Once this happens, the fluid jet stream 28 attaches itself to the duct wall at 28b. However, part of the fluid flow through the duct 27b is diverted by the geometry of the device into the feedback tube 30b and through the cavity 31b reaches the port 2% after some delay. This diverted flow entering the interaction chamber 26 imparts momentum to the jet 28 at 28b perpendicularly to the direction of flow, and causes the jet 28 to switch over to the duct 270 where the process is repeated. Thus the flow of pressurized fluid oscillates between the ducts 27a and 27b with a frequency set by the mechanical resistance and length of the feedback tubes 30a, 30b, and by the mechanical capacitance of the cavities 31a, 31b. Additional feedback delay could be obtained if needed by cascading additional fluidic amplifiers in the return paths. As a matter of fact, the two flows out of 27a, 27b may be made of different durations by making the two feedback paths (30a, 31a) and (30b, 31b) different in shape and/or size from each other. The oscillators geometric configuration shown in FIG. 5 is one of a great number of possible configurations, and is used here only by the way of explanation of the fluidic oscillator in general. Accordingly, the term fluidic oscillator as used herein means a fluidic device whose output flow is switched automatically, and without the interference of moving solid parts, other than the vibrator 20 and head 40, between two output ducts by means of a feedback path for each duct, that accepts part of the output flow or energy at some downstream point of the output duct (the duct end point included) and delivers it, delayed, at a point upstream at the interaction chamber 26.

Returning now to FIG. 3 we observe that the fluidic oscillator 23 is formed inside the solid mass of the vibrator 20. In practice the vibrator 20 may be cast in two parts, each of a half-cylinder shape, for the easy forming of the oscillators cavities and channels, and the two parts welded or otherwise connected together.

The oscillators duct 27a communicates with the annular piston chamber 340 by means of the branches 320, while the duct 27b communicates with the annular piston chamber 34b by means of the branches 32b. The chambers 34a. 34b are formed by the flange 35 of the outer surface of the vibrator 20 and by the flanges 41 of the inner surface of the head 40. When pressurized fluid from the oscillator acts upon the chamber 340, it tends to force the vibrator 20 downward and the head 40 upward; by contrast, when pressurized fluid acts upon the chamber 34b, the directions of these forces are reversed. The clearance gaps 36 are made wide enough that fluid leakage out of the chambers 34a, 34b through them is allowed. The fluid to mechanical energy transfer efficiency is thereby reduced but some leakage path is nevertheless necessary; both for purging the respective chambers 34a, 34b during reversal and because without it the oscillator will become choked and the switching of the flow between its two ducts 27a and 27b impossible. This leakage fluid is siphoned through the passages 37 into the outlet opening 22 and forms the main flow of drilling fluid descending into the drill bit and discharging through the bit passages. Although only one pair of piston chambers is shown in the drawing, it should be obvious that more such pairs may be employed in a parallel battery along the length of the vibrator, and so may other fluidic oscillators be used in parallel arranged to act in concert.

In the design variation shown in FIG. 4 the fluidic oscillator feedback paths 30a and 30b emanate from the piston chambers 34a and 34b respectively. Thus the feedback action is coupled to the relative motion between vibrator and head 40. Since, in turn, the latter motion is strongly influenced by the crushing strength of the particular rock encountered by the bit, the oscillator's switching becomes dependent upon the rock characteristics. This is a desirable feature in that it makes the oscillation continuously self-adjusting as the nature of the penetrated strata changes. The starting points a, 25b of the feedback paths a and 30b can be either in close proximity to or away from the corresponding duct 27a, 27b end points, or even placed somewhere in the gap 36 between vibrator and head; in the latter position the feedback process may include a judicious valving by the movement of flange 41. In addition to this variance in the fluidic oscillator paths FIG. 4 also illustrates a different usage of the discharge passages 37. The latter are now placed on the walls of the head and discharged the fluid spent in the chambers 34a, 34b directly into the wells annulus where lower pressure of the drill fluid prevails. In this case, however, a conduit 60, originating at inlet opening 21, has to be added through the actuator 20 to supply the flow necessary at the cutting surfaces of the bit. A restricted open ing or nozzle is provided within the bit body, directed in a conventional manner so that substantial part of the hydraulic pressure of the fluid is converted into velocity through which cleansing of the bit and removal of the rock debris from the well bottom is effected. This restriction is such that necessary pressure for operation of the fluidic oscillator is maintained. The actuator embodiment shown in FIG. 6, like that of FIGS. 2, 3, and 4, features the battery of piston chambers 34a, 34b formed between vibrator 20 and head 40 and fed respectively by the passage pairs 37a, 37b. However, the place of the fluidic oscillator is now taken by the pipe-shaped mechanical oscillator carried within the main conduit or tubular cavity 6 of the vibrator 20 preferably by means of the elastomeric bearings, springs and seals 15 and caused to oscillate axially relative to the vibrator 20 through a small distance shown by arrows 4 by means later described. Other support and sealing means are possible, as by two plain bearings, with springs arranged to center the oscillator 60 when at rest. Such a plain bearing could also perform the sealing function required. Here, a support member is defined as having associated sealing and centering properties. Jet nozzles 61 through the lateral wall of the oscillator face the passage pairs 37a, 37b over a small clearance gap, the oscillator itself being always filled by the pressurized fluid. Depending on the relative axial position of the vibrator 20 and oscillator 60, and because the inner openings 38a and 38b of the passages 37a, 37b respectively are offset axially with respect to one another, pressurized drill fluid from the jet nozzles 61 is forced into one or the other piston chambers 34a, 34b alternately while the spend fluid exits simultaneously from the other. finally leaving outlet port 22 via ports 25d. This alternate action forces the head 40 and vibrator 20 to reciprocate in opposite directions as in the embodiment shown in FIG. 3. It should be noted that these reciprocal motions along with the arrangement of the jet nozzles 61 opposite the openings 38a, 38b result in automatic valving of the pressurized fluid into the chambers 3411, 34b. and that the sense of the generated forces is such as to cause selfsustaining oscillation.

In FIG. 6, large ports 250 are provided through the vibrator wall to allow pressure equalization in the chambers next to the bearings 12a and 1212. It would be possible instead to omit the top and bottom flanges 41, which would result in the bearings 12a and 12b respectively themselves forming more or less functional flanges against which pressure forces may act. However, since the bearings 12a, 12b are joined to the vibrator 20 at their inside radii, some of the force transferred to them will be transferred to the vibrator rather than all to the head 40, rendering said bearings less effective than the flange 41 for force production. However, if by design the outside bearing faces and mating load surfaces of head 40 at 14a, 14b were placed closer to the central longitudinal axis than is the case in FIG. 6, then more effective flanges would result, fuctionally. Thus, any piston chamber, as in FIG. 4a, for instance, may be defined to be more or less effective for force production as the functional flange and piston flanking it are effective in that sense.

Restricted outlet paths 25h, FIG. 6, may be provided to maintain fluid circulation and prevent it from stagnating and possibly forming solid deposits within piston chambers.

The end plug and port device 62 has the function of separating the high pressure region within oscillator 60 from the relatively low pressure region outside the oscillator and in the port 22, while still providing access between the latter two through the port 25d. A passage 64 is necessary as a connection between the high pressure area and the enclosed volume 65 to equalize pressure on the top and bottom of the mechanical oscillator 60 and bearings 15. Otherwise, a considerable downward steady load on the oscillator 60 could occur. In the method shown, the plug 62 is separated from the oscillator 60 by a clearance gap 63 (drawn exaggerated) that is made as narrow as possible to prevent flow loss.

In some cases of design, it may be desirable to have a means of providing an additional damping effect on the motion of the mechanical oscillator 60 beyond what would otherwise exist, i.e., provide a force opposing its velocity. This can affect its relative phase of oscillation with respect to the vibrator 20, thus affecting the efficiency and extent of fluid power transfer. One means of providing such damping is readily available by making use of the enclosed volume 65 and passage 64, which can be made small enough to result in significant pressure drop across its length as the fluid is forced back and forth through it due to the piston action of the bearing 15 and bottom of the mechanical oscillator 60.

Another means of blocking flow from the bottom of mechanical oscillator 60 and providing an exit port for spent fluid is shown in FIG. 60, in which high pressure fluid has direct access to both ends of oscillator 60, and spent fluid is allowed to exit through the ports 25c into a space above the bottom bearing 12b and then out through the exit port 67.

There are shown in FIGS. 6a and 6b several other ways of arranging the jet porting corresponding to the typical view within the circled area of FIG. 6. In both of these, there are two jet nozzles 61a and 61b, to take the place of the single nozzle 61 at every location on FIG. 6. Nozzle 61a separately discharges only into passage 37a, while nozzle 61b discharges only into passage 37b in both FIGS. 6a and 6b, wherein the oscillator 60 is shown instantaneouly positioned so that nozzle 61b is aligned with its passage 37b, while nozzle 61a is cut off. The difference is that the flow passages in FIG. 6b are not crossed, necessitating a relative phasing of motion of oscillator 60 that differs by l from that in FIGS. 6 and 6a. In both new cases, however, the passages and their respective nozzles are parallel, necessitating less directional change of the fluid, reducing wear on the passages in the case of some drilling muds which may be abrasive.

Other variations in porting arrangements may be visualized without difficulty.

The initiation of relative oscillation and valving by oscillator 60 depends upon the ever present small turbulance in the flowing drill fluid and upon the mechanical vibration of the bit moving over the uneven rock face, both these factors being reflected in a jitter type movement of the mechanical oscillator 60. Once this jitter causes the first valving and'the vibrator 20 is forced into its initial strokes, the oscillator 60 by design does not follow the actuators motion. This effect is achieved by making the natural frequency of the oscillator 60 very low compared to the relatively high natural frequency of the vibrator. To satisfy this condition the equivalent spring constant of the elastomeric bearings is made relatively low by comparison with the equivalent spring constant of the laminate bearings 12a and 12b. Although the first few strokes may be rather weak resulting in only partial valving, the energy imparted by them to the vibrator forces it into a movement of larger and larger amplitude and, therefore, into more and more complete valving and the concomitant actuator energy buildup, with the head and bit masses finally involved in the oscillation as in the previous case.

p A combination of the already explained fluidic and mechanical means for inducing and/or maintaining the oscillatory motion of the bit is illustrated in FIG. 7. As seen there, the arrangement of the mechanical oscillator 60, suspended by means of the elastomeric bearings 15 within the vibrator 20, is identical with the arrangement already discussed in connection with FIG. 6. However, the mechanical oscillator 60 is set into a reciprocating motion by an hydraulic actuator, as by a source of oscillating fluid acting on a piston, preferably by a small auxiliary fluidic oscillator 70 which is formed within the body of the sleeve 77 inserted and rigidly affixed within the body of the vibrator 20. The source chamber 71 of this auxiliary oscillator is fed the pressurized drilling fluid through the auxiliary inlet 72. The output ducts 73a and 73b alternate in supplying the annular piston chambers 74a and 74b with the pressurized fluid which acts alternately upon one side or the other of the flange or piston 75 causing a reciprocating motion of the mechanical oscillator 60. The drain channels 76 serve to carry the spent fluid away from the piston chambers 74a, 74b into the low pressure space 78 between oscillator 60 and vibrator 20, and thence eventually into the well hole. Feedback paths 80a, 80a, with capacitance cavities 81a, 8111, or other means of delaying the feedback signal if needed, are provided between the reaction chamber 79 of the fluidic oscillator and each of the output ducts 73a, 73b. Obviously more piston chambers acting in parallel with those described are possible. The fluidic oscillator drive is provided here to positively establish relative motion of the mechanical oscillator 60 and vibrator 20 at about the resonant frequency of the vibrator itself, so that the startup of vibrator oscillation does not have to depend upon chance vibration causing the start of oscillatory valving. Such startup with the previous simpler configuration of FIG. 6 would probably become increasingly more difficult as the bearing springs 12a, 12b are designed to be stiffer for higher resonant actuator frequency. Addi' tionally, the present configuration of FIG. 7 could maintain higher amplitudes of relative oscillation between oscillator 60 and vibrator 20, providing better valving while allowing smaller oscillation amplitudes between vibrator 20 and head 40 and the latter with respect to the rock.

It is indicated with respect to FIG. 7 that the forced oscillation ofthe mechanical oscillator 60 is parallel to the drills lon' gitudinal axis. However, the functions indicated could be served as well by a torsional oscillation instead, with the jet nozzles 61 or 61 and 61b and receiving passage pairs 3711, 37/2 suitably respositioned for coaction torsionally over a small angle. Conical laminate bearings 15 would suffice to allow such motion while preventing axial motion, and a rotational vane type fluid actuator would suffice for oscillatory torques in place ofthe piston actuator shown.

The auxiliary fluidic oscillator 70 could be designed to be primarily effective in starting up the vibratory motion of the system rather than also being responsible for maintaining it. That is, when the growing amplitude of vibration of the vibrator 20 reaches a certain point, the spring or inertial forces applied to the mechanical oscillator 60 in FIG. 7 due to the vibration (e.g., through the bearing or support springs 15) could become adequate to continue its oscillation in the manner already described with respect to FIG. 6. This could result if the system is designed so that actuation forces on the piston 75 of FIG. 7 become relatively weak under this condition compared to the other forces on the mechanical oscillator 60, including support spring and inertial forces. This could be a consequence of the sizing of the fluidic oscillator 70 and its associated piston 75. The character of the steady state vibration of the mechanical oscillator 60 and that of the whole system could still be influenced beneficially by the continued action of the forces of piston 75, however; e.g., by providing a damping action.

Deviations in the disclosed connections of the fluidic oscillator 70 and piston 75 are believed to he obviously within the spirit of the invention, and furthermore, different kinds of fluid oscillators and actuators to perform the same task are intended to be included. For instance, a reaction type actuator or an unbalanced rotating weight connected to the mechanical oscillator 60 could be used to drive it, as will be discussed further at a later point.

In FIG. 6 the mechanical oscillator 60 is described to be positioned at every instant with respect to vibrator 20 by means of its inertia causing it to be effectively stablized. However, the phase relationship established thereby between the mechanical oscillator and the vibrator may not be optimal.

Another means of relatively positioning the mechanical oscillator 60 is to provide force-transmitting means to it not only from the vibrator 20 but also from the head 40 to the mechanical oscillator 60. A relative phasing relationship wherein the position of the mechanical oscillator 60 is lagged may be provided, for instance, where the mechanical oscillator may undergo a substantially cosinusoidal oscillation of position while that of the vibrator is sinusoidal and that of the head is negatively sinusoidal. This can be done by connecting the mechanical oscillator 60 by an axial spring to one of the massive members (head 40 or vibrator 20) and by a damper to the other (vibrator 20 or head 40). In FIG. 6d, the damper connection is to the head 40, via damping piston 75, while bearing springs 15 provide the spring connection to the vibrator 20 as before. In opposite fashion, in FIG. 62, the damper connection is made to the vibrator 20 via damping piston 75, while the bearing spring 15b, attached between the head 40 and mechanical oscillator 60, has much greater axial stiffness than the bearing 15a, so that the behavior effectively is determined by the former bearing spring 15b.

In FIG. 6d, the multiple passages 25c and 67 provide a way for spent fluid to exit from the annular space surrounding the mechanical oscillator 60, while in 62, passages 67 do the same.

In FIG. 6d the passages 65a provide a means whereby the space 65 may be equalized in pressure with the interior of the mechanical oscillator 60. The passage 65b provides similar equalization of pressure for the space 68, whereby average pressure loads on the mechanical oscillator 60 may be made negligible. The damper piston chambers on either side of piston 75 have restricted outlets formed, for instance, by appropriately-sized gaps 63a, 63b at flanges 75b, 75c and/or by a bypass formed by a gap 63 between the piston 75 and its cylinder wall. When the mechanical oscillator 60 is moved axially relative to the head 40, a pressure drop depending upon relative velocity will be created forcing flow through the gaps mentioned and creating a damping force equal to the pressure drop times the piston area. Multiple damping pistons and flanges may obviously be provided to increase the damping force if desired, in the manner of the main driving actuator. Restricted orifices 64 may be provided to allow a constant flow from the damper piston on either side of piston 75 to the low pressure region via passages 67 in order to prevent fluid stagnation and solidification in the damper piston chambers. If enough flow is so allowed it would be possible to obtain more constant damping characteristics for various oscillatory amplitudes, since the average damping flow would be shifted from zero where the damping characteristics change most rapidly. It is desirable that the mass of mechanical oscillator 60 be made relatively small so that its position is determined mostly by the coaction of the springs 15 and damper.

In Fig. 6e, although the bottom bearing 15b is attached to head 40 and is represented to have high axial stiffness, the bottom bearing support of the mechanical oscillator could as well be attached to the vibrator 20 as in FIG. 6, with an axiallyoperative spring attached between the head 40 and the mechanical oscillator 60.

Also in FIG. 62, the damper piston chambers on either side of piston 75 have restricted outlets formed, for instance, by an appropriately sized gap 63a at the flange 75b and holes 63b,

both outlets communicating with the high pressure region. A

bypass gap could also be provided at 63 as in the former case. f Restricted orifices 64 could be provided for the same purpose as discussed vis-a-vis FIG. 6d. Passages 25c allow equalization of pressure in the space above top flange 41 at the low pressure in the annulus surrounding the mechanical oscillator 60. The mass of the latter is again made relatively small as before.

It would otherwise be possible for the damping flow to occur primarily through the restricted orifices 64 rather than through the gap and holes 63a, 63b, or bypass 63, so that the latter could be omitted. It will then be noticed that these passages 64 correspond to the passages 76 of FIG. 7.

It appears obvious that an auxiliary fluidic oscillator could be provided and joined to the piston chambers as in FIG. 7, with the purpose of providing a starting means. When the mechanical oscillator motion amplitude builds up to a certain point under the influence of the fluidic oscillator-driven mechanical oscillator 60, the increasing damper and spring lSb forces could cause a shift toward the foregoing type of operation in the steady state, wherein the mechanical oscillator motion is primarily controlled by the head 40 and vibrator 20 motions through spring b and the damper respectively.

From the description of the invention presented so far, it is obvious that the function of the drill is based upon the dynam ic interaction between the vibrator 20, on the one hand, and the head 40 and bit 50 on the other. This interaction, which leads to the desired transfer of mechanical energy from the vibrator to the bit and eventually to the rock, is controlled by the magnitude of the two masses of vibrator and head-plus-bit, by the strength of the spring effect of the laminate elastomeric or other type bearings 12a, 12b connecting these two masses, and by the crushing strength of the particular rock on which the bit is working at a given time. These four mechanical parameters make the drill a system of two oscillators coupled to each other and to the rock and, therefore, a system possessing a natural frequency of oscillation. Consequently, as the nature of the rbck changes during the drilling operation so may the natural frequency at a given amplitude change, although the amount of the latter change may not be large. Generally, an inherently oscillatory system operates at maximum amplitude when the frequency of the force actuating it is substantially equal to the system's natural frequency. Fluidic oscillators, like the one used in the present invention to generate this force can be induced to change their natural frequency by various means, one of them being the fluids pressure. We wish, however, to show at this point one particular technique for controlling the natural frequency of the system optimally without outside intervention from the well's surface. The technique is illustrated in FIG. 8 and is based on the use of the acceleration force of the vibrator for imparting momentum to the fluid stream in the cavities of the fluidic oscillator 23 and in a manner effecting the switching of the flow between the ducts 27a and 27b alternately. As shown in FIG. 8 the oscillator 23 is placed transversely within the body of the vibrator 20 with the plane of its feedback paths 30a, 30b parallel to the vibrators longitudinal axis. The source chamber 24 is connected to the main inlet opening 21 through inlet 19. During the first few moments of the oscillators operation its flow is switched between 270 and 27b by the action of the feedback paths 30a and 30b as already described. However, after a number of vibrator strokes at or near the natural frequency of the resonant system, sufficient acceleration is imparted to the vibrator 20 to beneficially affect the function of the oscillator. Specifically, if at a given instant the fluid is streaming through the duct 27a into the piston chamber 34a and the vibrator is moving downward, an upward acceleration is applied to it by the bearing springs I211, 12b as it nears the end of the downward stroke; consequently, an inertial force is applied to the stream at the interaction chamber 26 forcing the stream to switch over to the duct 27b. This brings the pressurized fluid into the piston chamber 34b, thus reversing the force applied on the vibrator 20 and urging it upward. At the end of the upward stroke, the inertial force appearing on the fluid in 26 is pointed upward pushing the stream upward and causing a new switching, and this process will continue ad infinitum. While the duct 27b is operative and fluid begins to flow through the corresponding feedback path 30b, the inertial force acting at 26 is seen to also reinforce the feedback flow through 29b thus adding to the initial cause of switching the main flow from 2712 to 27a. A similar statement may be made when 27a is operative. Since the oscillator flow switching due to acceleration forces is directly dependent upon the motion of the vibrator 20 which in turn is determined by the crushing strength of the rock along with the mechanical parameters of the drill system itself, the system sets its frequency of oscillation automatically and in accordance with the prevailing physical condition of the rock formation that is being drilled. This effect we choose to call motional feedback from the vibrator to the oscillator. While only one combination of a fluidic oscillator 23 and a pair of piston chambers 34a, 34b are shown in FIG. 8, it should be obvious that two or more such combinations may be used within the same actuator 20.

The same principle of motional feedback may be applied to the case of the fluidic oscillator 70 and vibrator 20 of FIG. 7, in which the auxiliary fluidic oscillator 70 is already indicated to be horizontal. In this case, because of the inclusion of an additional component taking part in the oscillation of the overall system (Le. the oscillator 60) different phasing problems may be encountered from the case of FIG. 8. Phasing lag or lead mechanisms may need to be incorporated into the fluid paths in order to make the feedback phasing correct for self-oscillation. Motional feedback switching may be feasible with smaller amplitudes of oscillation in this case because of the smaller dimensions of the auxiliary fluidic oscillator 70.

It is apparent that driving of the mechanical oscillator 60 in the mode described in connection with FIG. 6 involves an imparting of force to it (60) that causes it to oscillate and switch flow to piston chambers 34a or 34b. This imparted force depends upon the motion of the vibrator 20, so that the previ ously-discussed operation of the mechanical oscillator 60 in the system can also be defined as an instance of motional feedback. Other mechanizations or variations of the principle appear obvious.

From the description of the vibrator 20 given so far, it can be seen that the two primary preferred versions of the device relative to obtaining oscillating flow at the piston areas feature either no moving solid parts within the vibrator, or only one such part, and that all motions between vibrator and drill stem or between vibrator and head involve no sliding of solid surfaces against each other. Instead, advantage is taken of the excellent charcteristics of the preferred elastomeric bearings as far as mechanical strength, wear, resilience, and sealing are concerned to accomplish the desired task.

Some other fluid actuator means of providing an oscillating force to cause vibration of the vibrator and head-bit system are possible also. Specifically, as seen in FIG. 9, one way is to direct the output ducts 27a and 27b of the fluidic oscillator 23 to nozzles 22a and 22 directed upwardly and downwardly into the drill hole space, causing alternate upward and downward reaction forces to be exerted upon the vibrator 20.

Another actuating technique is to mount unbalanced rotating weights, driven by fluid turbine means (or by any rotational energy actuator, such as an electric motor), to the vibrator 20, the centrifugal forces of the unbalanced rotating weights causing an axial component of oscillating force on vibrator 20. In the previous art, this has been done with long solid vibrating columns, but in the present device with two discrete masses (vibrator and head-bit) separated by novel bearing spring means as described, a more compact vibrating arrangement can be provided. Contrarotating pairs of equal offset weights geared lzl are preferable, to cancel or minimize lateral forces.

Either of the above two techniques could be used for providing actuation forces on the mechanical oscillator 60 as previously mentioned.

In all variations of the subject rock drill discussed so far, the stem adapter has been connected to the vibrator preferably by means of the laminate bearings 11. However, in all these cases the drill can be made to function equally well if the stem adapter 10 is connected to the head 40 by means of the relatively soft elastomeric bearings 11 (preferably laminated), while the vibrator is connected only to the head 40 through the relatively hard bearing springs 12 as shown in FIG. 10.

Another variation in design is illustrated in FIG. 11, in which the bit 50 is attached to the inner massive member 20 of the drill and the stem adapter 10 is connected to the outer massive member 40, a reversal of the cases described to this point. This rearrangement of connections to the vibratory massive members 20 and 40 results in essentially the same internal features as heretofore discussed except that phasing and flow switching parameter revisions may be required in th cases of motional feedback.

In the embodiments of the invention presented so far only the arrangements causing the axial reciprocating motion 2, FIGS. 2, 3, 4, and 6 were explained. At this point, however, we wish to disclose additional means for implementing the rotary motion 1 as well. These means comprise a hydraulic turbine wheel 42 shown in FIGS. 12 and 12a to consist of a plurality of radially placed turbine blades 43 rigidly attached to the interior wall of the head 40, and of a plurality of feeder ducts 44 provided in the body of the vibrator 20. The feeder ducts 44 are connected to a central conduit (not shown), similar to 6a of FIG. 4, and direct high velocity jets of pressurized drill fluid onto the blades 43, imparting a substantial portion of the mo mentum of their fluid to the blades and, therefore, applying a torque to the head 40. This torque forces the head 40, and bit, to turn on the conventional bearings 39. Although only one turbine wheel 42 is shown in the drawings it should be obvious that a battery of such wheels may be used as well ifthe torque requirements dictate it. Thus the need to turn the whole drill assembly from the surface of the well is eliminated as is the need for the drill stem 10 to extend all the way to the top of the well. This is so because the feeder ducts 44 are directed along radii of the vibrator 20 so that the reaction force of the outgoing jets generates no reaction torque on the vibrator 20. Thus, instead of the rigid drill stem, a flexible stem or power hose 46 connecting the pressurized fluid supply to the inlet 21 through a rotating joint 45 suffices to supply the necessary power to the whole drill, and results in a substantial simplification and economy of the surface equipment. It will be appreciated that the power hose is not required to transmit torque and, therefore, can be made sufficiently flexible to be stored in spools. Furthermore, this embodiment materially reduces torque requirements since there is no rotating lengthy drill stem to be handled. After impinging on the blades 43 and delivering energy to them, the drill fluid is diverted into the cavity 47 in a vortexlike fashion and eventually discharged from the bit passages onto the bottom of the well. In this case the length of the drill stem is only sufficient to provide the static weight required for pressing the bit against the rock for efficient drill action and to control hole deviation from the vertical. Hence the drill stem may comprise only the stem adapter 10.

A modification ofthe means for implementing the aforesaid rotary motion 1 further reduces the torque requirements. The modification consists of the feeder ducts 44 being connected to the branches 32a of the aforementioned fluidic oscillator 23 instead of the ducts being connected to the inlet 21 through a conduit. This causes fluid jets to issue only when the annular piston chamber 34a is active and, therefore, torque to be applied to the head 40 at the time the latter executes its upward stroke partially relieving the force between the bit and the bot= tom of the wellv This, of course, is the most appropriate time to index the bit since the coulomb friction between bit and rock is at a minimum resulting in a reduction of bit wear.

An alternate design for effecting bit rotation through use of mechanical energy released downhole is the configuration shown in FIG. 13 wherein the laminae of the elastomeric bearing 12b are stacked into sections of parallel helical surfaces, instead of being parallel conical surfaces which is the case with the bearing 12a, these sections coacting with the helical faces 12h of the section 13b of the vibrator. Such a bearing when compressed in the axial direction undergoes a deformation parallel to the helical surfaces; this deformation possesses both an axial component and a rotary component of which the former is made to provide the vibratory action against the rock bottom of the well, while the latter will cause rotary displacements of the vibrator and head in opposite directions; the mechanism involved is as follows: in the steady state of the vibratory motion the head 40 and vibrator 20 are moving in opposite directions; therefore, during a down stroke 2dr ofthe latter and upstroke Zuh of the former the compression of the bearing 12b is accompanied by a rotation ofthc vibrator to the left. llv and of the head to the IIght I f/l. Since the head and bit is pressed against the rock at all times the extent of lr/i in degrees of arc will be smaller than the extent of llv; i'.e., lIv=u lrlz=h, and a b,. When the strokes are reversed and the vibrator moves upward, Zur. While the head moves downward, 2(1/1, the bearing deformation will be removed since the compression is relieved. In the course of doing so the vibrator will rotate to the right, lrv=a,, and the head to the left, l/h=b,. with u b for the reason given above. The net result after one oscillation will be an angular displacement u,u,. for the vibrator and Iii-l1, for the head. Since the bearing 12b should be left with no additional stress at the end of any one oscillation than at the beginning, we should have u u,.=l/,.. But during the heads downstroke the increased pressure against the rock makes 11f u,f': when this inequality is substituted in the last equation we obtain /i,." b Thesetwoincqualities mean that there is a net rotation of the head 40 to the left and of the vibrator 20 to the right; the former results in bit indexing, while the latter is accommodated by the rotary joint 45 between vibrator and stem adapter 10.

In the two types of self-rotating rock drill just discussed (FIGS. l2, 13) the stem adapter 10 had no direct connection to the surface. Another approach is to have it attached to a cable from the surface. As shown in FIG. 14, the eyelets 17 rigidly attached to the upper end of the: stern adapter 10 serve for fastening a lift cable through which the weight of the drill is partially counteracted by pulling on the cable as drilling conditions require. The elastomeric bearing ll, besides attenuating the upward transmission of vibration, permits drill weight reduction without lifting the bit off the wells bottom. Means of attaching the cable to the stem adapter 10 other than eyelets may be used. In all other respects the drill assembly remains the same as disclosed previously, with the vibrator 20 being resiliently attached to the head 40 by means of laminate elastomeric bearings 12a and 12b, and with external attachment of the vibrator to the stem adapter 10 as discussed earlier. The respective external-internal relationships of vibrator and adapter could be reversed as well.

The dynamics of the three-mass system of FIGS. 2, 3 and 4 is modeled in the mechanical circuit diagram of FIG. 15 that in turn leads to the following system of three differential eq uations in the unknown time varying displacements x=x(r). y= yll). 2 20):

In this system M M M are the masses of the stem 10, vibrator 20, and head-bit 40, 50; W W W;, are the corresponding weights; (k k and (n, r are the spring constants of the elastomeric bearings 11 and 12 respectively, and

- the viscous friction coefficients associated with the motion of I M and M f the reciprocating force supplied by the oscillator 23; and x, y, z the displacements of the stem, vibrator, and bit from the bottom of the hole in that order; finally r is the friction coefficient between head-bit and the sides of the hole in all probability nonviscous.

The alternate arrangement of FIG. 10, whose equivalent mechanical circuit is seen in FIG. 16, is correspondingly described by the system of equations in a similar fashion the four-mass system of FIG. 6 is modeled by the mechanical circuit diagram of FIG. 17 from which the following system of differential equations is formed:

in this system M is the mass of the mechanical oscillator 60 and A1,, in, the spring constant of the laminate bearings 15 and the coefficient of the viscous friction encountered by M In all three systems Ws represent the corresponding weights and is the force-displacement curve associated with the crushing strength of the particular rock. Typically, C(z) has the form shown in FIG. 19 with the curve parameters dependent upon the physical characteristics of the rock and upon the geometry of the drill bit. Although the term G(z) is nonlinear, solution and optimization of the above systems is possible using either nonlinear-analytical or computer-based techniques.

Finally, the system of FIG. 14 is represented by the circuit diagram of FIG. 18 from which we obtain:

in all these systems the presence of the terms containing the quantity (z-y) in the equations for M a y/dt and M d"z/dt explains and describes quantitatively the interaction between vibrator and head-and-bit. The term G(z) on the other hand, describes the mechanism of the transfer of mechanical energy from the drill to the rock.

While full and complete disclosure of the invention has been set forth in accordance with the dictates of the patent statutes, it is to be understood that the invention is not intended to be so limited. It will be apparent to those skilled in the art that various changes may be made to the embodiments described herein without departing from the spirit of the invention or the scope of the appended claims. Moreover, it is apparent that the vibratory system resulting could be used in above-ground applications as for tamping or pile-driving purposes, or as a vibration generator in conveyors or other machinery. In this case, the end connections in lieu of stem and bit may be adapted to the particular purposes at hand, so that the words bit" and stem or stem adapter may be broadly construed. Also, it is not necessary that the connecting members taking the place of the bit and stem adapter necessarily be on opposite ends of the assembly. Yet another important application of our vibratory system is in the manually held air hammer where noise and operator fatigue can be substantially lessened. Still other applications include the implementation of vibratory drilling of hard metals or 5 other types of material removal as in mold-carving, dental work, etc.

We claim:

1. A vibratory drill assembly as described including:

an inner massive member at least partially telescoped within an outer massive member, the two said massive members being separated by, and held substantially concentric to each other about a common longitudinal axis by at least one bearing member having sealing properties and stiff axial spring characteristics whereby limited axial displacement of the one massive member with respect to the other results upon the application of axial force between 4 said massive members;

a bit drivingly attached to one of said massive members and a stem adapter attached resiliently in said axial direction to one of said massive members; and

a fluid-operated actuator producing force between said two massive members, including at least one piston functionally formed on the outer surface of the teleseoped portion of said inner massive member and coacting with the inner surface of said outer massive member, and at least one flange functionally formed on the inner surface of said outer massive member and coacting with the outer surface of said inner massive member to provide at least one effective piston chamber to accept pressurized flow cyclically through porting means from oscillatory converting rneans within said inner massive member, said oscillatory converting means acting upon an inlet flow of pressurized fluid supplied through said stem adapter via an internal inlet passage and converting said inlet flow to said cyclic flow, whereby fluid energy may be converted to vibratory energy carried by the resonant system formed by said inner and outer massive members and said springlike behaving bearing members coupling them together, said vibratory energy dissipated by said bit repeatedly striking the material to be penetrated by said drill assembly.

2. The device of claim 1 wherein each one of at least two said bearing members is of the laminated elastomer-metal type as described and has load faces inclined with respect to said longitudinal axis, said bearing members being preloaded against one another.

3. The device of claim 1 wherein each one of at least two said bearing members is make of a stiff resilient elastomeric material as described and has load faces inclined with respect to said longitudinal axis, said bearing members being preloaded against one another.

4. The device of claim 1 wherein each one of at least two said bearing members is composed of a nested stock of thin springy dish-shaped metal washers as described, said bearing members being preloaded against one another.

5. The device of claim 1 wherein at least one of said bearing members has a noncircular cross section in a plane perpendicular to said longitudinal axis, whereby said bearing member may transmit torque without substantial windup.

6. The device of claim 1 wherein said bearing members are two in number, forming an opposed pair on opposite ends of said fluid actuator and sealing same, preloaded against one another.

7. The device of claim 6 wherein said bearing members are laminate elastomer-metal bearings as described and have load faces inclined with respect to said longitudinal axis, being formed of joined flat sections so as to have a polygonal cross section in a plane perpendicular to said longitudinal axis, whereby said bearing members may transmit torque without substantial windup.

8. The device of claim 1 wherein said bearing members are interposed in the annular spaces between said pistons and said coacting inner surface of said outer massive member and between said flanges and said coacting outer surface of said inner massive member, acting to seal said piston chambers from one another.

9. The device of claim I wherein said stem adapter is engaged telescopically to one of said two massive members, said internal inlet passage comprising longitudinal connecting holes within said stem adapter and said massive member, at least one relatively axially-resilient elastomer-containing sleevelike member interposed in the annulus between said telescoping members, connecting and aligning them, sealing between them against said inlet flow of pressurized fluid, and transferring axial load force from said stern adapter to said massive member, whereby limited axial movement of said massive member is possible with respect to said stem adapter through shear strain of said elastomer, resulting in effective isolation of said stem adapter from said resonant system.

10, The device of claim 9 wherein each said elastomer-containing member is a laminated elastomer-metal bearing as described.

H. The device of claim 9 wherein mechanical stop means are provided to prevent axial overstressing of said elastomercontaining sleevelikc member.

12. The device of claim 9 wherein at least one said elastomercontaining member has a noncircular cross section in a plane perpendicular to said longitudinal axis, whereby said bearing member may transmit torque without substantial windup.

13. The device ofclaim 9 wherein each said elastomer-containing member is a laminate elastomer-metal bearing as described, being formed of joined flat sections so as to have a polygonal cross section in a plane perpendicular to said longitudinal axis, whereby said bearing member may transmit torque without substantial windup.

14. The device of claim 13 wherein said stem adapter is telescoped externally to said inner massive member, and said bit is attached to said outer massive member on the opposite end of said drill.

15. The device of claim 13 wherein said stem adapter is telescoped internally to said outer massive member and said bit is attached to said outer massive member on the opposite end ofsaid drill.

16. The device of claim 9 wherein said stem adapter is telescoped externally to said inper massive member.

17. The device of claim 9 wherein said stem adapter is telescoped internally to said outer massive member.

18. The device of claim 1 wherein said oscillatory converting means includes at least one fluidic oscillator, said fluidic oscillators alternately feeding said inlet flow of pressurized fluid into alternate piston chambers respectively through one of a pair of output ducts, each said fluidic oscillator having a source chamber communicating with said inlet passage and an interaction chamber communicating with said source chamber, said pair of output ducts connected to said interaction chamber.

19. The device of claim 18 wherein each of said output ducts is connected to at least one branch ending on the exterior surface of said inner massive member and inside at least one of said alternate piston chambers, and whereby at least one leakage path is provided from each. said piston chamber to a lower pressure region.

20. The device of claim 19 wherein said leakage paths converge and lead to said bit through said inner massive member.

21, The device of claim 19 wherein said leakage paths are orifices on the outer wall of said outer massive member, discharging into the well,

22. The device of claim 18 wherein each said fluidic oscillator has a pair of feedback paths, each of said feedback paths connecting said interaction chamber to a point of one of said output ducts to said interaction chamber.

23. The device of claim 18 wherein each said fluidic oscillator has a pair of feedback paths, each of said feedback paths connecting said interaction chamber to a point on the exterior surfate of said inner massive member, said point lying within the gap between said exterior surface and said coacting flange.

24. The device of claim 18 wherein said output ducts of each said fluidic oscillator are displaced relative to one another in the direction of said longitudinal axis, the fluid flow through said interaction chamber and said output ducts being substantially transverse to said longitudinal axis, whereby axial acceleration of said inner massive member may act favorably toward switching the flow insaid interaction chamber from one slid output duct to the other, resulting in said resonant system sustaining itself in vibration through motional feedback 25. The device of claim 1 wherein said inlet passage connects to an axial tubular cavity within said inner massive member internal to the region of said piston chambers, the wall of said cavity being traversed by at least one pair of passages for each piston, the passages of each pair being connectc l one'to-one to the piston chambers on opposite sides of said IlStOfl, and said oscillatory converting means includes a mechanical oscillator, said mechanical oscillator supported within said tubular cavity by at least a pair of support members that rave fluid sealing means and allow movement in one degrt of freedom, whereby said mechanical oscillator is allOWCll limited oscillatory movement, said sealing means precluding direct access of said inlet flow of pressurized fluid to the inner openings of said passage pairs, said mechanical oscillator being hollow and open to said fluid inlet passage at one end, the other end being blocked against flow, the wall of said mechanical oscillator being travensed by holes forming a plurality of nozzles, said nozzles being; located cooperatively adjacent to said passage pair inlets, whereby outward-extending jets may be formed by said pressurized fluid flowing through said nozzles and impinge upon one or the other inlet of each said passage pair, causing pressure forces to be developed alternately on one side or the other of each said piston, depending upon the relative instantaneous position of said mechanical oscillator with respect to said inner massive member, such porting action causing fluid energy to be imparted to said massive members at about the natural frequency of said resonant system, the spent fluid from said piston cham bers and leakage flowing in the annular space within said tubular cavity external to said mechanical oscillator and out through exit porting means.

26. The device of claim 25 wherein said support members are a pair of laminate elastomer-metal bearings that have annular shape surrounding said mechanical oscillator, and have laminae parallel to said longitudinal axis, whereby limited axial movement of said oscillator is allowed.

27. The device of claim 25 wherein said support members have an associated axial spring constant small enough in mag nitude to make the natural frequency of axial resonance, defln :d by the mass of said mechanical oscillator and said spring constant, low by comparison with the natural frequency of said resonant system, whereby said mechanical oscillator tends to remain virtually stationary as said inner massive member reciprocates, causing said porting action to occur, and resulting in said resonant system sustaining itself in a vibratory state through motional feedback.

28. The device of claim 27 wherein additional damping means are provided at one end of said mechanical oscillator by the action ofa functional piston pumping fluid in and out of an enclosed volume through a restricted passage.

29. The device of claim 25 wherein said degree of freedom is along said longitudinal axis, further characterized by the presence of a small fluid-operated oscillator within said inner massive member (i.e.,any kind of oscillator such as spool-type, diaphragm-type, fluidic type, etc.), said small fluid oscillator having a source chamber communicating with said internal inlet passage, and a small fluid-operated actuator having at least one piston connected to said mechanical oscillator, at least one output duct of said small fluid oscillator being connected functionally to said small actuator to produce an oscillating axial force on said pistons and to urge axial oscillation of same at about the frequency of said resonant system resulting in said porting action by said mechanical oscillator.

30. The device of claim 29 wherein said small fluidoperated oscillator has an interaction chamber communicat' ing with said source chamber, and a pair of output ducts connected to said interaction chamber, said output ducts being displaced relative to one another in the direction of said longitudinal axis, the fluid flow through said interaction chamber and said output ducts being substantially transverse to said longitudinal axis, whereby axial acceleration of said inner massive member may act favorably toward switching the flow in said interaction chamber from one said output duct to the other, said flow operating upon said small actuator resulting in said resonant system sustaining itself in a vibratory state through motional feedback.

' 31. The device of claim 29 wherein the average magnitude of said piston forces becomes smaller than the average magnitude of inertial forces on said mechanical oscillator when the amplitude of oscillation of said inner massive member reaches a certain level, whereby control of oscillation of said resonant system shifts to self-sustainment through motional feedback.

32. The device of claim 25 wherein said degree of freedom is around the longitudinal axis, further characterized by the presence of a small fluid-operated oscillator within said inner massive member, said small fluid-operated oscillator having a source chamber communicating with said internal inlet passage, and a small angular fluid-operated actuator functionally connected to said mechanical oscillator, at least one output duct of said small fluid-operated oscillator being connected functionally to said small actuator to produce an oscillating angular force on said mechanical oscillator and to urge angular oscillation of same at about the frequency of said resonant system, resulting in said porting action by said mechanical oscillator.

33. The device of claim 32 wherein said small fluidoperated oscillator is of the fluidic type, and has an interaction chamber communicating with said source chamber, and a pair of output ducts connected to said interaction chamber, said output ducts being displaced axially relative to one another, the fluid flow through said interaction chamber and said output ducts being substantially transverse to said longitudinal axis, whereby axial acceleration of said inner massive member may act favorably toward switching said flow in said interaction chamber from one said output duct to the other, said flow operating upon said small actuator, resulting in said resonant system sustaining itself through motional feedback.

34 The device of claim 32 wherein said support members are a pair of laminated elastomer-metal bearings that have annular shape surrounding said mechanical oscillator, and have laminae inclined at an angle to said longitudinal axis, whereby limited relative angular movement of said mechanical oscillator is allowed.

35. The device of claim 25 wherein said degree of freedom is along said longitudinal axis, further characterized by the presence of a fluid damper including a piston and a cylinder member, one of said damper members attached to said mechanical oscillator and the other to one of said massive members, whereby a damping force on said mechanical oscillator results as a consequence of relative velocity between it and said massive member; and at least one axially operative spring attached effectively between said mechanical oscillator and the other of said massive members, whereby spring force results on said mechanical oscillator as a consequence of relative displacement between it and said other massive member, whereby the phase relationship of said mechanical oscillator to said inner massive member depends upon said damping and spring forces.

36. The device of claim 35 wherein said support members are attached to and within said inner massive member and at least one of said support members has axial spring properties whereby said axially-operative spring is effectively provided.

37. The device of claim 35 wherein one of said support members is attached within said outer massive member and has axial spring properties much stiffer than those of the other said support member, whereby said axiallyoperative spring is effectively provided.

38. The device of claim 1 and in addition a hydraulic turbine wheel consisting of a plurality of radially positioned turbine blades rigidly attached onto the interior wall surface of said outer massive member, a plurality of feeder duets provided in the body of said inner massive member, said feeder ducts being connected to said internal inlet passage and directing high velocity jets of said pressurized fluid onto said turbine blades, said high velocity jets transferring a substantial part of their momentum to said turbine wheel and urging said outer massive member to rotate, conventional plain bearings inserted between said outer massive member and said bearing members separating said outer and inner massive members, said conventional bearings permitting said rotation of said outer massive member.

39. The device of claim 18 and in addition a hydraulic turbine wheel consisting of a plurality of radially positioned tur bine blades rigidly attached onto the interior wall surface of said outer massive member, a plurality of feeder ducts provided within the body of said inner massive member, said feeder ducts being connected to one duct of said pair of output ducts of said fluidic oscillator and directing high velocity jets of said pressurized fluid onto said turbine blades, said high velocity jets transferring a substantial part of their momentum to said turbine wheel and urging said outer massive member to rotate, conventional bearings inserted between said outer massive member and said bearing members separating said outer and inner massive members, said conventional bearings permitting said rotation of said outer massive member.

40. The device of claim 1 wherein at least one of said bearing members is of the laminate elastomer-metal type consisting of alternate metal and elastomer laminae, said laminae having the shape of parallel helical surfaces, said shape resulting in said bearing members deformationby compression having an axial component parallel to said longitudinal axis and an angular component about same said axis.

41. A vibratory drill assembly as described including:

an inner massive member at least partially telescoped within an outer massive member, the two massive members being separated by and held concentric to each other about a common longitudinal axis by two axially separated laminate elastomer-metal bearing members as described, each said bearing member having load faces inclined to said longitudinal axis and said bearing members being preloaded against one another to provide stiff axial spring characteristics whereby limited axial displacement of the one massive member with respect to the other results upon the application of axial force between said massive members;

a bit drivingly attached to one of said massive members, and

a stem adapter attached resiliently in said axial direction to one of said massive members; and

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Classifications
U.S. Classification175/56
International ClassificationE21B7/00, B06B1/18, E21B7/24
Cooperative ClassificationE21B7/24, B06B1/183
European ClassificationE21B7/24, B06B1/18B