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Publication numberUS3866693 A
Publication typeGrant
Publication dateFeb 18, 1975
Filing dateJun 11, 1973
Priority dateJun 11, 1973
Also published asCA1007143A1, DE2428128A1
Publication numberUS 3866693 A, US 3866693A, US-A-3866693, US3866693 A, US3866693A
InventorsCentury Bernard A
Original AssigneeAllied Steel Tractor Prod Inc
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Vibratory impact hammer
US 3866693 A
Abstract
An impact force producing apparatus that is described in connection with one particularly suitable use on a backhoe vehicle although it is usable on other vehicles and apparatus. A frame is provided which can be attached directly to the backhoe. A heavy ram is flexibly mounted to the frame in such a manner that it may freely oscillate relative to the frame. Attached to the ram is a hydraulic motor coupled with an eccentric weight to provoke vibrational movement in the ram. The relative inertias and locations of the components, the resilient mountings and the driving force of the eccentric weight are interrelated to create a vibratory motion in the ram which provides a unidirectional impact on a tool positioned on the frame and drives the ram near a resonance of the system. Guides may be provided to eliminate secondary harmonics and extraneous vibrations and movements.
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Description  (OCR text may contain errors)

United States Patent 1191 Century VIBRATORY IMPACT HAMMER [75] Inventor: Bernard A. Century, Cleveland Heights, Ohio [73] Assignee: Allied Steel & Tractor Products Incorporated, Cleveland, Ohio 22 Filed: June 11, 1973 21 Appl. No.: 368,463

[ Feb. 18, 1975 Primary Examiner-David H, Brown Attorney, Agent, or Firm-Lyon 8: Lyon [5 7] ABSTRACT An impact force producing apparatus that is described in connection with one particularly suitable use on a backhoe vehicle although it is usable on other vehicles and apparatus. A frame is provided which can be attached directly to the backhoe. A heavy ram is flexibly mounted to the frame in such a manner that it may freely oscillate relative to the frame. Attached to the ram is a hydraulic motor coupled with an eccentric weight to provoke vibrational movement in the ram. The relative inertias and locations. of the components, the resilient mountings and the driving force of the ec centric weight are interrelated to create a vibratory motion in the ram which provides a unidirectional impact on a tool positioned on the frame and drives the ram near a resonance of the system. Guides may be provided to eliminate secondary harmonics and extraneous vibrations and movements.

32 Claims, 13 Drawing Figures 1 VIBRATORY IMPACT HAMMER This invention relates to a hydraulic hammer. More specifically, this invention is directed to an improved vibratory impact hammer employing a rotating eccentric weight to establish vibrational movement.

Vibratory impact hammers have been developed which operate on a variety of principles. Some of these hammers have employed the use of one or more rotating eccentric weights to develop useful vibration that may be employed to break pavement, compact earth, drill holes and perform other similar functions. The more successful of these hammers, driven by rotating eccentric weights, have employed a multiple number of such weights constrained to rotate with respect to one another so that the resulting centrifugal force of the weights in an oscillating force directed along a single line that is in line with the impact motion of the ram. The use of such a system of weights avoids the generation of oscillations which are perpendicular to the desired impact motion of the hammer. In suggested systems employing a single eccentric, such as in US. Pat. Nos. 1,386,329 and 1,410,010, the vibrations perpendicular to the impact direction have generally been restrained or controlled through the use of rigid guides which force the impacting member to oscillate along a single path. In such devices, the forces generated by the single eccentric weight perpendicular to the impact direction are transmitted to the support frame and do not provide useful impact motion but rather they cause imdesirable vibrations. Thus, even the aforementioned patents teach the advisability, if not actually the necessity of employing double counter-rotating eccentrics to eliminate the lateral vibrations.

These vibratory hammers employing one or more eccentric weights to drive the impact member have also generally. relied upon the energy created during each revolution for the succeeding impact. As a result, the relative size of the unbalanced weight system must be large in order that substantial energy can be transmitted to the ram for each blow. Further, secondary spring arrangements are often employed in an attempt to store and redistribute the energy to the rebounding ram between strokes. These secondary spring arrangements cause further vibrational problems and do not operate efficiently to redirect energy back into the ram. Because of the above design features, vibratory impact hammers have traditionally been very noisy, hard to hold and heavy.

The present invention contemplates the use of a single eccentric weight strategically positioned on a ram for producing useful vibratory motion without requiring substantial lateral restraints. The resulting motion induced in the ram by the rotating eccentric weight includes both lateral and longitudinal oscillations at the center of gravity of the ram but is limited to oscillations in the longitudinal or impact direction at the striker plate of the ram. A tool placed to receive impacts from the ram at the striker plate is not subjected to lateral forces and lateral relative motion of the ram which would otherwise result in high wear and energy losses.

The lateral force components of the rotating eccentric weight are allowed to induce lateral vibrations in the ram which are restrained only by flexible, resilient mounts. The resulting motion in both the lateral and longitudinal directions operates to impart a maximum longitudinal impact to the tool at the striker plate of the ram without requiring driving mechanisms of relatively large capacity. This motion of the ram cooperates with the flexible, resilient mounts to store the rebounded energy not transmitted to the work through the tool and to store the continual imput from the eccentric weight for delivery to the tool on succeeding impacts. Because of this continuous storage and delivery of energy to the tool, the driving mechanism need not be of sufficient capacity to induce in one cycle the total power neces sary to provide the impact required for that cycle. Instead, the required additional input to the system during each cycle need only be sufficient to replace the work expended by the tool and lost through mechanical inefficiencies in the system.

The hammer also employs a vibration range near the natural resonance of the system to enhance the energy transfer and storage within the system. By properly relating frequency with the unrestrained operation of the spring and ram system, the overall efficiency of the hammer can be maximized. Further, the relative size of the eccentric weight to the size of the ram can be reduced which makes possible the use of a larger and quieter operating ram. This proper combining of operating and design parameters also results in less vibration of the support vehicle and longer hammer life.

The present invention also includes strategically positioned constraining means which operate to attenuate undesirable harmonic and random vibrations. These constraining means do not interfere with the normal os cillator'y motion of the ram and hence do not detract from the operating efficiencies and advantages of the system.

Accordingly, it is an object of the present invention to provide an improved vibratory impact hammer.

A second object of the present invention is to provide a vibratory impact hammer which advantageously allows lateral vibratory motion of the ram and at the same time provides linear impacting motion against an impact tool.

Another object of the present invention is to provide a vibratory impact hammer which employs a single eccentric to drive a ram against a tool where the ram is free tooscillate perpendicular to the impact direction.

A further object of the present invention is to provide a vibratory impact hammer which employs resonant motion of an impacting member to advantageously store and transmit impact energy.

Still another object of this invention is to provide a vibratory impact hammer that employs a ram of a relatively large mass which may be vibrated through a comparatively small stroke and yet produce impact forces comparable to conventional devices of great size, cost or power consumption.

Another object of the present invention is to provide a freely oscillating vibratory hammer having directional constraints to attenuate undesirable harmonic and random vibratory motion.

Thus, a vibratory impact hammer is provided which employs natural vibratory motion to create a quiet, convenient and efficient system. Further objects and advantages will become apparent from the description herein.

FIG. 1 is a side view of a vibratory impact hammer of the present invention.

FIG. 2 is a cross-sectional front view taken along line 2-2 of FIG. 1 illustrating the ram with the eccentric weight assembly exposed.

FIG. 3 is a cross-sectional side view taken along line 3-3 of FIG. 2 illustrating the ram with one side plate removed.

FIG. 4 is a bottom view of the device taken along line 44 of FIG. 2 with the tool in cross-section.

FIG. 5 is a cross-sectional bottom view taken along line 5-5 of FIG. 2 illustrating the tool holder and tool retainer pin.

FIG. 6 is a cross-sectional bottom view taken along line 6-6 of FIG. 2 illustrating the striker plate and lower bracket of the ram.

FIG. 7 is a cross-sectional side view taken along line 77 of FIG. 2 illustrating the ram guide.

FIG. 8 is a front view with a portion of the system broken away illustrating the ram on a downstroke with the tool engaged.

FIG. 9 is a front view of the system illustrating the ram on an upstroke with the tool engaged.

FIG. 10 is a schematic side view of the hammer with a first side plate removed to illustrate the free motion of the ram.

FIG. 11 is an exemplary frequency-stroke response curve for a linear spring-mass system with damping.

FIG. 12 is an exemplary frequency-stroke response curve for a nonlinear spring-mass system with damping.

FIG. 13 is an exemplary time-displacement curve for the present system under a nonlinear preload.

Turning specifically to the drawings, a frame, generally designated 10 is illustrated. The frame 10 consists of two side plates 12 and 14. The side plates 12 and 14 may be identical for ease of fabrication and are oriented in parallel planes. Support members 16 and 18 extend between the plates 12 and 14 to add overall frigidity to the frame 10. The plates 12 and 14 are also held in spaced parallel orientation by the tool holder described below. Boom pins 20 and 22 extend perpendicularly through the parallel plates 12 and 14 near the upper edge of the extended mounting area 24 of the frame 10. The boom pins 20 and 22 are employed to engage the backhoe or other heavy duty articulated arm used to carry and place the system. The boom pins 20 and 22 are of sufficient length to extend outwardly from both side plates 12 and 14. Cotter pins 26 restrain the boom pins 20 and 22 from being withdrawn in a first direction from the frame 10. Klik pins 28 may be employed to prevent withdrawal of the boom pins 20 and 22 in a second direction from the frame 10. The klik pins 28 are employed so that the boom pins 20 and 22 may be easily taken out or repositioned. Collars 30 are welded to the side plates 12 and 14 about the boom pin locations to provide added tear and bearing strength to the frame 10 where the hammer is mounted to the backhoe.

Located between the side plates 12 and 14 is the ram assembly generally designated 32. The ram assembly 32 generally includes a ram 34 resiliently mounted on the frame 10 and a driven eccentric weight 36 for inducing vibration in the ram 34. Due to the fact that the preferred shape of the ram 34 is rather irregular, although basically of a modified l-beam shape, it is most conveniently cast of iron or steel, such as ductile iron. The ram 34 fits entirely between the side plates 12 and 14 and has the sides or open portions of the generally I-beam shape facing the side plates.

As an example of a typical embodiment of the present invention for illustrative purposes and without limiting the scope of the invention, the ram assembly 32 may weigh approximately 360 lbs including all of the sprung components constrained to move with the ram 34 such as the eccentric weight 36 and the hydraulic motor. The movement of inertia about an axis parallel to the axis of the eccentric weight 36 and located through the center of gravity of the ram assembly 32 in the present embodiment is approximately 16,000 lbs. in. The eccentricity (the sum of the products of the weight of each eccentric weight and the distance from the center of gravity of each eccentric weight to its respective axis of rotation) is approximately 69 inch pounds in the present embodiment. Thus it will be seen that the weight of the eccentric is very small compared to the weight of the entire ram it causes to vibrate.

An eccentric weight housing 38 is formed in the ram 34 above the center of gravity thereof. The eccentric weight housing 38 includes a cylindrical cavity 40 for receiving the eccentric weight 36. Circular flanges 42 and 44 extend inward from the cylindrical inner wall of the housing 38 to receive bearing housings 46 and 48. The bearing housings 46 and 48 are cylindrical and are sized to fit into the circular openings in the flanges 42 and 44. Bolts 50 and 52 may be employed to fix the bearing housings 46 and 48 in the flanges 42 and 44 of the eccentric weight housing 38. Annular rings 54 and 56 extend outwardly from the cylindrical bearing housings 46 and 48 to engage the circular flanges 42 and 44 through which the bolts 50 and 52 are positioned. This placement of the bearing housings 46 and 48 establishes the axis of the eccentric weight 36 relative to the ram 34. The distance from the center of gravity to the eccentric axis in the present embodiment is approximately 3% inches.

The eccentric weight 36 is circular in shape as can be seen in FIG. 3 and is rotated about an axis defined by bearings 58 and 60 held within the bearing caps 46 and 48. Axles 62 and 64 extend inward from the bearings 58 and 60 to engage the eccentric weight 36. A hole 66 is provided parallel to the centerline of the eccentric weight through one corner thereof to receive the axles 62 and 64. The axles 62 and 64 are each of circular cross-section. One end of each of the axles 62 and 64 is received by the inner races of the bearings 58 and 60. The inner ends of the axles 62 and 64 are positioned within each end of the hole 66 in the eccentric weight 36. The axle 62 and 64 fit into the hole 66 with an interference fit in order that torque may be transmitted through the axle 64 to drive the eccentric weight 36. A collar 70 is provided at the center of each of the axles 62 and 64. Each collar 70 is larger than the inner races of the bearings 58 and 60 and the hole 66 of the eccentric weight 36. The collars 70 center the eccentric weight between the bearings 58 and 60.

A hydraulic motor 72 is mounted to the bearing cap 48 and in turn the eccentric weight housing 38 by fasteners 52. The hydraulic motor 72 has a fluid inlet 74 and a fluid outlet 76 through flexible hoses which are capable of withstanding the vibration of the impact member 34 to which the motor 72 is mounted. The motor shaft 78 includes splines which engage splines 79 in a hole 80 located through the axle 64 for torque transmission. To counterbalance the motor 72, a counterweight 82 is provided on one end of the eccentric weight housing 38 opposite the hydraulic motor 72. The counterweight 82 is fixed to the eccentric weight housing 38 through the bearing housing 46 by fasteners 50. When the ram 34 is vibrated by the action of the rotating eccentric 36, the counterweight 82 will act to negate the unbalance loading of the motor 72. This is accomplished by selecting a counterweight so that the center of gravity of the motor 72 and counterweight 82 considered together is along the centerline of the ram 34. Semi-circular cuts 84 are made in the side plates 12 and 14 to accommodate the motor 72 and the counterweight 82. The cuts 84 are significantly larger than the motor 72 and the counterweight 82 in order that the vibration induced in the ram 34 will not cause either the motor 72 or the counterweight 82 to impact the side plates 12 and 14.

The ram 34 extends upward from the eccentric weight housing 38 to form a mounting bracket 86. The mounting bracket 86 includes two flat surfaces 88 and 90 for positioning mounts thereon. Extending downwardly from the eccentric weight housing 38 on the ram 34 is a second mounting structure 92. The mounting structure 92 also includes two mounting surfaces 94 and 96 for attaching mounts. Four web 98, 100, 102 and 104 extend down from the eccentric weight housing 38 along either side of the mounting structure 92. At the lowermost end of the ram 34, the webs 98, 100, 102 and 104 and the mounting structure 92 terminate at a lower bracket member 106.

The position ofthe axis of the eccentric weight 36 on the ram 34 is such that a node with respect to horizontal motion of the ram 34 is created at the bottom of the lower bracket member 106. A striker plate 107 made of hardened steel is positioned beneath bracket 106 to provide an impact surface for transmitting energy from the vibrating ram 34 to an impact target. The striker plate 107 is bolted to the lower bracket member 106 at four positions by bolts 108. By having the striker plate 107 at a horizontal node on the ram assembly 32, purely vertical motion will occur at that point. The elimination of all horizontal motion at the center of the striker plate sufface during impact with target prevents wear of the striker plate 107 and the target, eliminates lateral shock loadings of the ram assembly 32 and the target and increases the total available impact energy directed longitudinally into the target.

To predict the proper position for the striker plate 107 in relation to the center of gravity, the following steps may be employed. The center of gravity of the complete ram assembly 32 is first located. The moment of inertia about an axis parallel with the axis of the eccentric weight 36 and located through the center of gravity is next determined. Finally, the distance d, between the center of gravity and the striker plate 107 may be expressed in terms of the distance d the distance between the center of gravity and the axis of the eccentric weight 36, as

where:

I moment of inertia of the ram assembly 32 about an axis through the center of gravity and parallel to the axis of the eccentric weight 36 in lbs. in.*;

w weight of the ram assembly 32 in lbs.;

d, distance from the center of gravity of the ram assembly 32 to the axis of the eccentric 36 in inches.

ds then becomes the proper position for the striker plate 107. The above method for determining the horizontal node does not take into account the sum of the forces of the resilient mounting of the ram assembly 32 to the frame 10. The sum of the resilient mounting forces has been found to not greatly affect the position of the horizontal node due to such factors as the relativly soft nature of the mounts, the operation of the hammer is far above the natural frequency of the resilient mounts, and the mounts furthest from the horizontal node are substantially more flexible than the mounts closest to the horizontal node.

The ram 34 is resiliently mounted to the frame 10 by resilient mounting means which are designed to allow relatively free movement of the impact member 34 with respect to the frame 10. These resilient means may include resilient elements 109 and 110 associated with the mounting bracket 86 and resilient elements 112 and 114 associated with the mounting structure 92. The resilient elements 109 and 110' are identical to prcvent unbalanced loading of the ram assembly 32. The elements 109 and 110 are comprised of cylindrical rubber blocks and are bonded at one end to plates 116 and 118. These plates may be positioned on the flat mounting surfaces 88 and of the mounting bracket 86 and bolted thereto by fasteners 120. The other end of the elements 109 and are bonded to plates 121 and 123 which are in turn bolted to the side plates 12 and 14 through spacers 1 22 and 124, respectively. The resilient elements 112 and 114 are also identical to one another and are of cylindrical construction. One end of each of the elements 112 and 114 is bonded to a mounting plate 126 and 128. Each of the mounting plates 126 and 128 is then bolted to one of the mounting surfaces 94 and 96 of the mounting structure 92 by fasteners 130. The other ends of each of the blocks 112 and 114 are bonded to plates 131 and 133 which are bolted by bolts 136 to the side plates 12 and 14 through spacers 132 and 134.

All of the resilient elements 109', 110, 112 and 114 are oriented with respect to the ram assembly 32 and the side plates 12 and 14 so that they will flex in a shearing motion under the vibrations induced by the rotating eccentric weight 36. In the aforementioned typical embodiment, the static spring constants for the upper resilientelements 109 and 110 in this shearing direction are 225 lbs/inch each and the lower resilient elements 112 and 114 have static spring constants in this shearing mode which are 700 lbs/inch each. The total spring constant for the mounting means measured by forcing the ram assembly 32 downwardly toward the striker plate end of the ram 34 with respect to the frame 10 is 1,850 lbs/inch. The resilient elements 109, 110, 112 and 114 are constructed of rubber blocks. Rubber used in applications such as this exhibit a dif' ferent effective spring constant related to dynamic motion than the actual static spring constant. The total dynamic spring constant for the present mounting means at the frequencies at which the present system is to be driven has been empirically determined to be approximately 2,750 lbs/inch.

The motion of the ram assembly 32 with respect to the frame 10 is schematically illustrated in FIG. 10. The motion schematically illustrated constitutes the unrestrained vibration of the ram assembly 32 when positioned in the resilient elements 109, 110, 112 and 114 and driven by the eccentric weight 36. No other guide means need exist to obtain the illustrated free motion. As a result of the inertia considerations previously discussed and the resilient mounting system, a circular path 137 is described by the center of gravity of the ram assembly 32. At the striker plate 107, the horizonta] node makes the ram motion purely longitudinal as illustrated by straight line 138 which is the locus of a point at the center of the striker plate during the vibration of the ram assembly 32. The diameter of the circular path 137 at the center of gravity of the ram assembly 32 is equal in length to the length of the straight path 138 at the lower resilient elements 112 and 114, the ram follows an eliptical path 140 having the major axis thereof along the centerline of the ram 34. This major axis of the eliptical path 140 is equal in length to the diameter of the circular path 137 and the length of the straight line 138. At the upper resilient elements 109 and 110, the ram 34 describes an eliptical path 142 having its major axis perpendicular to the centerline of the ram 34. This eliptical path 142 has a minor axis equal to the length of the straight path 138.

The above-described motion of the ram 34 closely follows the motion which may be predicted by considering only the inertia characteristics of the ram assembly 32. Only inertia considerations were used to predict the proper position for the striker plate 107 to have it coincide with a horizontal node. The effect of the resilient elements 109, 110, 112 and 114 on this motion is minimized because of of the relative spring constants of the upper resilient elements 109 and 110 and the lower resilient elements 112 and 114. The upper resilient ele' ments 109 and 110 are substantially more flexible than the lower resilient elements and consequently resist the lateral motion of the ram 34 less than do the lower resilient elements 112 and 114. Thus, the difference in flexibility and the difference in travel of the two sets of elements cooperate to result in a restoring force system which lends itself to vibratory motion having a node in the horizontal oscillation profile near the striker plate 107. As a result, the inertia considerations are predominent in establishing the horizontal node at the striker plate 107.

The oscillations of the ram 34 which assume the above-described motion because of the inertia and spring characteristics of the hammer are induced by the rotation of the eccentric weight 36. The eccentric weight 36 is physically symetrical with respect to a plane parallel to the side plates 12 and 14 and further create a driving force which is located in that parallel plane. These symetrical characteristics, along with the symetry of the overall hammer configuration, result in motion of the ram 34 which remains parallel to the side plates 12 and 14. Further, the oscillations induced by the rotating eccentric weight 36 have an operating frequency equal to the driving rotational frequency of the eccentric weight. Thus, the symetry and speed of the driving eccentric weight 36 also determine the nature and speed of the ram motion.

The stroke or dynamic displacement of the ram 34 during free vibration of the ram assembly 32 within the frame depends on the mass of the ram 34, the force generated by the rotating eccentric weight 36, the resilient elements 109, 110, 112 and 114, and the frequency of the driving eccentric weight 36. The size and spring characteristics of the system are fixed at the time of manufacture; but the frequency may be varied during operation of the unit. The natural frequency of the freely vibrating ram in the present embodiment is ap proximately 490 cycles per minute. This frequency is of little import to the impact capabilities of the hammer. However, large and potentially destructible strokes may be experienced at this natural frequency which should therefore be avoided. The hammer is to be run at more than twice this natural frequency and need only pass quickly by this range. In the present embodiment, the free vibration of the ram has been established at 1,600 cycles per minute to provide a moderate stroke which will not be destructive to the ram or resilient elements. 7

To control the free vibration frequency of the ram 34, a flow divider 144 is provided. The free vibration of the ram 34 at a frequency such as L600 cpm does not require a great deal of power. Consequently, the rotating eccentric 36 may be driven by the motor 72 without a high demand on the hydraulic input. High pressure is not required by the motor 72 under such circumstances. However, to operate the hammer under loaded as well as unloaded conditions, a capacity for high flow and high pressure must be provided. Consequently, the available hydraulic input is larger than is required to drive the freely vibrating ram 34. The flow divider 144 is introduced into the hydraulic system as a means for limiting the amount of hydraulic fluid which can be distributed to the inlet 74 of the pump 72 regardless of the system pressure. In the present system, the flow is limited by the flow divider 144 to 12 gpm. In addition to the flow divider 144, a pressure control in the form of a pressure relief valve 145 is provided in the hydraulic fluid supply line to the motor 72. For this particular embodiment a pressure relief setting of 1,700 psi has been found to be satisfactory and any fluid in excess of that which is required to maintain that pressure is diverted into the return line from the motor by this pressure relief valve 145. Thus, in the suspended condition of the apparatus, the motor 72 uses the entire 12 gpm and the pressure automatically drops downstream of the flow divider 144 to a level which is just sufficient to drive the eccentric weight 36 which is far below the setting'of valve 145. Flow control, rather than pressure control, for the unloaded frequency pro vides a further advantage by retaining the momentum of the hydraulic system at a level where it will operate when loaded. Consequently, no initial acceleration of the system is required when the ram 34 is loaded. Also, this allows the operator to leave the unit running in the free suspended condition while respositioning the tool at a new point or even relocating the backhoe vehicle, which is either undesirable or impossible with many devices.

To obtain useful impact energy from the hammer, a target must be positioned within the path of the freely vibrating ram 34. In this manner, the vibrational energy of the ram 34 may be transmitted to the target during each vibration at impact. Naturally, the vibrational characteristics of the hammer alter greatly when the target is interposed into the free path of the ram 34. The target is illustrated in the present embodiment as a tool 146. The concepts and operating features of the present invention would not be compromised if the ram 34 was designed to impact directly on the material which is ultimately to be worked by the impact forces of the hammer. One such alternate use would include a tamping foot located at the horizontal node of the ram. The tamping foot could then be used to directly compact earth or perform other similar operations.

The tool 146 of the present embodiment is merely a conduit through which the impact energy may be transmitted from the ram 34 to the material to be worked. The tool 146 includes an anvil surface which receives the impacts from the ram 34. A generally cylindrical body 150 extends downwardly from the anvil surface 148. A working point 152 at the lower end of the tool body 150 is illustrated in FIG. 1.

The tool 146 is constrained to move longitudinally with respect to the frame 10 by tool holder 154. The tool holder 154 is rigidly fixed between the side plates 12 and 14 of the frame 10 by fasteners 156. Eight such fasteners 156 are illustrated in the present embodiment. The tool holder 154 is comprised of a cast block of material having a hole 158 located therethrough for receipt of the tool 146. The hole 158 through the tool holder 154 has a diameter which is sufficiently larger than the diameter of the tool so that the wall of the hole 158 will not interfere with the motion of the tool 146. At either end of the hole 158 the diameter thereof is increased for receiving upper and lower bushings 160 and 162. The upper bushing 160 and the lower bushing 162 provide a sliding fit for the tool body 150 extending therethrough. The upper bushing 160 is positioned within the hole 158 of the tool holder 154 by forcing the bushing 161) into the hole 158 until it reaches the end of the increased diameter within the hole. A set screw 164 is then threaded into the tool holder 154 to engage the upper bushing 160. The upper bushing 160 is thereby retained in position. The lower bushing 162 is placed into the hole 158 in the tool holder 154 from the bottom. The lower bushing 162 is restrained from moving upward because of the reduced diameter at the center of the hole 158. The lower bushing 162 is restrained from moving downwardly and out of the tool holder 154 by a tool stop plate 166 bolted to the lower surface of the tool holder 154.

The cylindrical tool 146 is retained within the tool holder 154 by a tool retainer pin 168 which cooperates with a notch 170 in the tool body 150. The tool retainer pin 168 extends laterally across the tool holder 154 in such a manner that it partially extends through the hole 158 in the tool holder 154 as best seen in FIG. 5. Extensions 172 and 174 extend from the ends of the tool retainer pin 168 to engage two retainer plates 176 and 178. The retainer plates are positioned on either side of the tool holder 154 and are fastened thereto by bolts 180. The notch 170 allows free travel of the tool 146 in the hole 158 along the length of the notch 170. when the hammer is lifted off the work, the tool 146 will fall downwardly until the tool retainer pin 168 interferes with the upper end of the notch 170. The notch 170 is of sufficient length so that the tool 168 may drop completely out of the path of the rotating ram 34. The tool retainer pin 168 is not designed to retain the tool 146 when loaded. The notch 170 extends down the tool body 150 far enough so that tool retainer pin 168 will not contact the lower edge of the notch 170 when the tool is forced upward against the ram 34. To limit the travel of the tool 146 into the path of the ram 34, a collar 182 is provided along the cylindrical body 1511. The collar 182 is positioned to engage the tool stop plate 166 at a predetermined preload position. Normal operation of the hammer is conducted with the tool 168 forced upward to engage the collar 182 with the tool stop plate 166.

The limited travel of the tool 146 creates two advantageous conditions. First, the tool is automatically retracted from impacting engagement with the ram 34 when the hammer is lifted from the work. This automatic separation of the tool prevents eventual destruction of the hammer caused by the ram impacting the frame. Elaborate mechanisms are quite often required in other impact hammers to insure that the ram will not destroy the frame when driven in an unloaded condition.

The second advantageous condition provided by the limited travel of the tool 146 is the automatic establishment of a preselected preload on the ram regardless of the downward force exerted against the hammer by the backhoe. The backhoe may be operated by forcing the frame 10 toward the work until the tool 146 moves far enough to engage the collar 182 with the tool stop plate 166. Any further downward pressure by the backhoe would be directly transmitted to the work because the collar 182 interferes with any further upward motion of the tool 146. Thus, the preload position of the tool 146 is determined by the position of the collar 182 on the tool body 150. The tool 146 is shown in the preloaded position in FIG. 9.

The preloaded position of the tool 146 establishes the extent to which the tool interferes with the path of the vibrating ram 34. If the selected preload is excessive, the eccentric weight 36 would not be capable of lifting the ram 34 from the tool to provide a subsequent impact. This condition would occur when the rotating eccentric weight 36 is unable to overcome the downward forces of the resilient elements 109, 110, 112 and 114 created by forcing the tool 146 and the ram 34 upward relative to the frame 10. When an insufficient preload is established for the hammer, the ram 34 will begin to lose velocity in the impact direction before the striker plate 107 of the hammer 34 contacts the anvil surface 148 of the tool 146. In this condition, the maximum available impact would not be realized.

A range of preloadings are therefore available with any specific hammer configuration. Only a change in the distance from the collar 182 to the anvil surface 148 is required. By placing the tool 146 at a position where the resilient elements are in the relaxed condition at the point of impact, all of the impacting force is derived from the momentum of the ram 34 at that in stant. When the tool is positioned so that the resilient elements are still pulling downwardly on the ram 34 at the point of impact, the downward :force of the resilient elements will be added to the force resulting from the developed momentum of the ram 34 at the point of impact.

When the tool collar 182 establishes the point of impact at the relaxed position of the resilient elements 109, 110, 112 and 114, a substantially linear frequency-stroke response is developed. Such a response is illustrated in FIG. 11. When a substantial preload is retained on the resilient elements at the point of impact, a nonlinear frequency-stroke response is experienced. Such a nonlinear frequency-stroke response curve is illustrated in FIG. 12. Practical considerations must dictate the frequency-stroke response curve most desirable for the conditions under which the hammer is to be used. This frequency-stroke response relationship is controlled by the selection of the allowed tool movement upwardly against the ram 34. Because the size of the stroke is more critically related. to frequency in the linear condition, hammers which must be used with a variety of work should not be preloaded for a linear frequency-stroke response. Instead, the more gradual stroke change with respect to frequency of the nonlinear frequency-stroke response system should be employed because of the difficulty in maintaining a constant frequency when the resiliency of the work varies. Further, greater element life is experienced where the length of the stroke is not allowed to become excessive.

The vibrational motion of the ram 34 is significantly altered by the imposition of the tool 146 into the paths of the vibrating ram. When substantial resistence to motion is provided by the tool placed on the work, the time-displacement curve for the vertical oscillations of the ram 34 varies from sinesoidal to a cut-off wave motion similar to that shown in FIG. 13. If the impact of the ram 34 on the tool 46 were completely elastic and if the too] were placed at the midpoint in the vertical travel of the ram, it is believed that the natural frequency of the system, operating under linear conditions, would be double the free natural frequency of the ram 34. Because the present system envisions a nonlinear preload condition, the natural frequency is not located at twice the natural frequency of the free moving ram 34. This is, the natural frequency of the ram of the present embodiment is not 980 cpm which is double the free natural frequency of the ram 34, 490 cpm. Instead, the system response more closely approximates the shape of the curve in FIG. 12. The operating frequency may therefore be in a broad range above the nonlinear natural frequency. The present system has been found to operate best between 1,150 and 1,350 cpm when loaded. This places the operating zone of the present system along the portion of the nonlinear curve marked 184 in FIG. 12. As with the free vibration mode of the system, the symmetry of the system causes the motion of the ram 34 to lie in a plane parallel to the two side plates 12 and 14. 7

The frequency range of operation of the hammer when loaded is directly dependent upon the input pressure to the motor 72 and the hardness and resiliency of the work. The ram 34 will be driven at the same frequency as the driving eccentric weight 36. The pressure relief control valve 145 is provided wnst eamw qfl irfl wh sr .lioarwrate as a control on the working frequency of the system. As described above, when the ram 34 is vibrating freely, the input is controlled by the maximum flow allowed through the flow divider. The pressure downstream of the flow divider 144 is reduced to equalize the power requirement of the motor 72 in this unloaded condition. When the tool 146 is brought into impacting engagement with the ram 34, a substantial amount of power is required to keep the system vibrating. Consequently, more power must be delivered to the eccentric weight 36. As the eccentric weight 36 provides greater and greater resistence to the motor 72, the pressure will increase at the inlet 74. With the control system of the present embodiment, the relief pressure setting is reached on the valve 145 before equilibrium is established in the system. Consequently, the flow to the motor 72 is reduced because of the loss of fluid through the relief pressure system. Equilibrium is established with the flow of hydraulic pressure to the inlet 74 reduced below that amount provided by the flow divider 144. The remainder of the flow through the divider 144 escapes through the pressure relief system to the outlet passage to maintain the selected pressure at the inlet 74.

At equilibrium, the frequency at which the eccentric weight 36 can be rotated under a preselected hydraulic pressure is achieved. Various modifications to the device and its controls can cause changes in the frequency of the loaded system, such as the pressure setting of valve 145, the magnitude of precompression caused by tool 146, a flow rate of a different order of magnitude to cause operation near a different mode of resonant frequency, etc. With all other factors remaining the same, by changing the pressure setting on the relief valve 145, the system may be driven toward or away from the natural frequency of the system. In this way, the frequency and the responsive stroke can be regulated through regulation of the pressure of the hydraulic fluid delivered to the motor 72. The pressure setting for the present embodiment has been established at 1,700 psi. This gives an operating range of l,150 to 1,350 cpm under varying work conditions. Under such conditions, the stroke is around 0.4 inches. As can be seen in FIG. 12, a point may be reached where the frequency becomes so low that the stroke response of the system will jump from the upper curve to the lower curve. In such a case, the resulting stroke is significantly less and the system becomes inefficient and ineffective. Consequently, it is advantageous to keep the stroke response of the system above the discontinuity in the nonlinear frequency-stroke response curve.

The unconstrained motion of the ram 34 provides advantageous operation when the total force of the backhoe presses directly along the centerline of the too]. However, such conditions cannot always be met. Further, it is often desired that the hammer be employed to pound at an angle to the vertical. Moreover, variations in the resistance encountered by the tool 146 can cause momentary load changes. When any such unbalanced conditions are imposed upon the vibrating hammer, undesirable harmonic and random vibrations are experienced. Such vibrations are undesirable because they tend to interrupt the steady state vibration of the ram 34 and create large vibrations which can damage the resilient elements. Constraining means are strategically positioned on the frame 10 to cooperate with the vibrating ram 34 to attenuate such undesirable harmonic and random vibrations. In order that such unwanted vibrations may be eliminated without detracting from the operating efficiencies of the normal vibrating mode of the ram 34, the constraining means cooperate with the vibrating ram 34 near the horizontal node thereof. This constraining means may be provided by two guides generally designated 186 and 188 attached to the frame 10 and positioned on either side of the vibrating ram 34 at the horizontal node.

The two guides 186 and 188 are identical and only guide 186 will be described. The guide 186 includes a base plate 190 which is fixed to the side plate 12 by fasteners 192. A key 194 extends into a slot 196 provided on the tool holder 154 to insure that the guide 186 will be unable to rotate from its position. The base plate 190 has a vertical channel 198 extending the height of the base plate 190. Flanges 200 are provided on either side of the base plate 190 along the inner face thereof. The flanges 200 extend inwardly to form two lips overhanging the channel 198. The channel 198 is open at the bottom. At the top, a resilient snubber 202 is provided as a means for resiliently closing the upper end of the channel 198. The resilient snubber 202 is bonded to a thin plate 204 which extends outwardly from the snubber 202 to cover the top of the base plate 190. An angle member 206 is positioned over the thin plate 204 to securely support the snubber 202. The angle memher includes a leg 208 which extends downwardly to cover a portion of the channel 198 and provide lateral support to the resilient snubber 202. Two fasteners 210 hold the angle member 206 rigidly to the base plate 190.

The lower bracket member 106 includes two slider pins 212 and 214. Slider pins 212 and 214 are cylindrical, extend outwardly from the lower bracket 106 of the ram 34 and are centered on either side of the striking surface of the striker plate 107. Because of their location relative to the striker plate 107, the pins 212 and 214 are located at the horizontal node of the oscillating ram 34. Slider pin 212 extends into the guide 186 and slider pin 214 extends into the guide 188. Around each of the slider pins is positioned a slide block 216 and 218. The slide blocks 216 and 218 are also identical. The slide block 216 is generally rectangular and slidably fits within the channel 198. The slider block 216 prevents a lateral motion of the slider pin 212 because of its location within the channel 198. The flanges 200 associated with the base plate 190 extend inwardly to lock the slider block 216 within the channel 198. Further, the slider pin 212 is prevented from substantial upward movement by the contacting of the slider block 216 with the resilient snubber 202. Consequently, the slider pins 212 and 214 and necessarily the bottom of the ram 34 are prevented from performing lateral oscillations. Further, the ram 34 is prevented from exaggerated vertical oscillations which may randomly occur by the snubber 202.

The positioning of the guides 186 and 188 at the horizontal node of the ram 34 allows the ram 34 to progress through its normal oscillations without interference from the guides 186 and 188. Consequently, the efficiencies and advantageous motions of the ram 34 are not interferred with by the guides. Further, the ram 34 does not operate to exert continuing forces against the guides 186 and 188 during vibration of the hammer.

Thus, a vibratory impact hammer is provided which is driven by a single rotating eccentric weight and is allowed to freely oscillate to effect a straight line impact on a tool or other target. The system is designed to pro vide long component life, ease of operation, and high system efficiency.

While embodiments and applications of this invention have been shown and described, it would be apparent to those skilled in the art that many more modifications are possible without departing from the inventive concepts herein described. The invention, therefore, is not to be restricted except by the spirit of the appended claims.

What is claimed is:

1. An impact force producing apparatus, comprising a support frame;

an impacting member resiliently mounted on said frame for relatively free vibrational movement thereof;

a plurality of resilient elements, each resilient element being attached at one position thereof to said support frame and at a second position thereof to said impacting member;

vibration drive means operably mounted on said impacting member, said drive means being positioned to induce straight line motion in said impacting member at one location on said impacting member; and

an impact transmitting surface located on said impacting member at the point on said impacting member where straight line motion is experienced.

2. The apparatus of claim 1 wherein said vibration drive means comprises an eccentric weight rotatably mounted on said impacting member and a motor for driving said eccentric weight.

3. The apparatus of claim 1 including means capable of rigidly attaching said support frame to a heavy-duty articulated arm.

4. The apparatus of claim 1 wherein said impact transmitting surface is located at a horizontal node.

5. The apparatus of claim 1 wherein said apparatus further comprises guide means for attenuating random vibrational disturbances in said impacting member, said guide means being on said support frame.

6. The apparatus of claim 5 wherein said guide means cooperates with said impacting member adjacent said impact transmitting surface to insure straight line motion of said impact transmitting surface.

7. The apparatus of claim 1 wherein there are four of said plurality of said resilient elements the first pair of said resilient elements being located between the center of gravity of the impacting member and the impact transmitting surface, and the second pair of said resilient elements being located on the other side of the center of gravity of said impacting member from said first pair of resilient elements.

8. The apparatus of claim 7 wherein said first pair of resilient elements is substantially more rigid than said second pair of resilient elements.

9. The apparatus of claim 7 wherein said resilient elements are blocks of flexible material.

10. The apparatus of claim 1 wherein said apparatus further comprises a tool slidably mounted to said support frame, said tool being capable: of being positioned in impacting contact with said impact transmitting surface.

11. The apparatus of claim 10 wherein said tool is oriented in said support frame to slidably move along the same direction as the straight line motion of the point on said impacting member.

12. The apparatus of claim 10 further comprising means for restricting movement of said tool beyond a preselected position towards said impacting member, said restricting means including a collar rigidly fixed to said tool capable of engaging said support frame, said collar being placed on said tool to allow said tool to force said impacting member a fixed distance from its relaxed position.

13. A vibratory impact hammer comprising a support frame;

a ram;

means for supporting said ram in said support frame,

said means having a low effective modulus of elasticity in at least one plane of movement of said ram relative to said support frame;

an eccentric weight rotatably mounted on said ram at a distance from the center of gravity of said ram;

means for rotating said eccentric weight to cause vibratory movement of said ram; and

an impact transmitting surface located on said ram at a lateral node in the motion of said ram defined when said eccentric weight is rotated.

14. The apparatus of claim 13 wherein said means for rotating said eccentric weight include a hydraulic motor mounted in said ram and means for controlling the volumetric rate of flow of hydraulic fluid to said motor for controlling the vibration rate.

15. The apparatus of claim 13 wherein said eccentric is rotatably mounted on an axis perpendicular to said plane.

16. The apparatus of claim 13 further comprising a tool slidably mounted on said support frame, said tool being capable of impacting contact with said ram on said impact transmitting surface. a

17. The apparatus of claim 16 wherein said tool is constrained from moving against said ram past a preselected position to establish a preload on said ram. 18. A vibratory impact apparatus comprising a support structure; a ram assembly; 7 suspension means for suspending said ram assembly on said support frame for resilient movement in at least one direction; said ram assembly including a ram, eccentric means for inducing vibration in said ram with movement in at least said one direction and means for driving said eccentric means, said ram assembly having a ratio of eccentricity of said eccentric means measured in inch-pounds to the ram assembly weight measured in pounds of about 1:5; and tool means adapted to be mounted on said support structure and capable of being impacted by said ram during vibration of said ram. 19. The apparatus of claim 18 wherein said ram has a generally l-beam cross-section at least at portions therealong taken in a plane perpendicular to said one direction, and said suspension means comprises rubber mounts extending from said frame to said ram perpendicular to said one direction and between the flanges of and connected to the web of said l-beam cross-section.

20. The apparatus of claim 18 wherein said vibration inducing means is comprised of an eccentric means rotatably mounted in said ram and a hydraulic motor mounted on said ram and connected to and driving said 9.ntr. 9 n a 21. A vibratory impact apparatus comprising a support structure; a ram assembly; suspension means for suspending said ram assembly on said support frame for resilient movement in at least one direction: 4

said ram assembly including a ram, eccentric means for inducing vibration in said ram with movement in at least said one direction and means for driving said eccentric means, said ram assembly having a ratio of eccentricity of said eccentric means measured in inch-pounds to the ram assembly weight measured in pounds of about 1:5; and

tool means adapted to be mounted on said support stru a reandsaizablq9f e n mpa by sa r a m during vibr ation of said rain.

' 22. The apparatus of claim 18 wherein said means for driving said eccentric means includes a hydraulic mo- 23. The apparatus of claim 18 wherein said suspen- 24. The assembly of claim 18 wherein said ram assembly includes a single eccentric weight mounted on said ram at a distance from the center of gravity of said ram and an impact transmitting surface located on said ram at a lateral node in the motion of said ram defined when said eccentric weight is rotated.

25. The apparatus of claim 18 wherein said tool means is moveably mounted on said support frame for selective movement into or out of position to be impacted by said ram for selective free vibrtion of said ram without impacting said tool means.

26. The apparatus of claim 25 wherein said tool means is moveable to a limit position toward said ram beyond which said tool means will not move toward ram regardless of forces applied, and in said limit position said tool means cause a precompression of said resilient suspension by displacing said ram from the natural free-suspended position of the ram.

27. The apparatus of claim 18 wherein said suspension means has a low modulus of elasticity in a single plane that includes said one direction.

28. The apparatus of claim 27 wherein vibration inducing means is comprised of a single eccentric weight rotatably mounted in said ram and rotating on an axis perpendicular to said plane.

29. The apparatus of claim 28 wherein said eccentric weight is spaced on said ram from the center of gravity of said ram to create a point on said ram experiencing a lateral node during rotation of said eccentric weight, said point being located for impacting said tool means.

30. An impact force producing apparatus, comprising a support frame;

an impacting member resiliently mounted on said frame for relatively free vibrational movement thereof;

a plurality of blocks of flexible material, each block being attached at one position thereof to said support frame and at a second position thereof to said impacting member;

vibration drive means operably mounted on said impacting member, said drive means being positioned to induce straight line motion in said impacting member at one location on said impacting member; and

an impact transmitting surface located on said impacting member at the point on said impacting member where straight line motion is experienced.

31. An impact force producing apparatus, comprising a support frame;

an impacting member resiliently mounted on said frame for relatively free vibrational movement thereof;

a plurality of blocks of flexible material, each said block being attached at one position thereof to said support frame and at a second position thereof to said impacting member;

an eccentric weight rotatably mounted on said impacting member at a distance from the center of gravity of said impacting member;

means for rotating said eccentric weight to cause vibratory movement of said impacting member; and

an impact transmitting surface located on said impacting member at a lateral node in the motion of said impacting member defined when said eccentric weight is rotated.

32. A vibratory impact hammer comprising a support frame;

a ram assembly;

7176s 18 a plurality of blocks of flexible material, each said sembly, and an impact transmitting surface located block being attached at one position thereof to said on said ram at a lateral node in the motion of said support frame and at a second position thereof to ram defined when said eccentric weight is rotated; said ram assembly; and said ram assembly including a ram, an eccentric tool means adapted to be mounted on said support weight rotatably mounted on said ram at a distance structure and capable of being impacted by said from the center of gravity of said ram assembly, ram during vibration of said ram. said eccentric weight having a relatively small eccentricity compared to the weight of the ram asi UNITED STATES PATENT OFFICE CERTIFICATE OF CORRECTION PATENT NO. 3,866,693

DATED I February 18, 1975 INVENTOR(S) BERNARD A. CENTURY It is certified that error appears in the above-identified patent and that said Letters Patent are hereby corrected as shown below:

Column 15 lines 42 through 59, cancel all of claim 21 and insert the following claim:

21. The apparatus of claim 18 wherein said eccentric means includes a single eccentric weight.

Signed and sealed this 6th day of May 1975.

(SEAL) Attest:

C. MARSHALL DANN RUTH C. MASON Commissioner of Patents Attesting Officer and Trademarks

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Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US4257648 *Apr 9, 1979Mar 24, 1981Bodine Albert GNon-resonant cyclic drive system employing rectification of the cyclic output
US4265129 *Apr 6, 1979May 5, 1981Bodine Albert GOrbiting mass oscillator with oil film cushioned bearings
US4616716 *Mar 26, 1984Oct 14, 1986Allied Steel & Tractor Products, Inc.Synchronous vibratory impact hammer
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US8690261 *Dec 31, 2010Apr 8, 2014Zhongsheng TangHydraulic resonant breaking hammer
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Classifications
U.S. Classification173/49, 74/61
International ClassificationE02F5/00, B06B1/10, B25D15/02, B25D17/00, B25D9/12, B25D9/00, B25D11/00, B06B1/16, E02F9/00, B25D17/28, B25D15/00, B25D11/06
Cooperative ClassificationB25D9/12, B25D11/06
European ClassificationB25D9/12, B25D11/06
Legal Events
DateCodeEventDescription
Feb 12, 1987AS07Mortgage
Free format text: NATIONAL WESTMINSTER BANK USA, 175 WATER STREET, NEW YORK, NY., 10038, A NATIONA * ALLIED STEEL & TRACTOR PRODUCTS, INC., A CORP OF DE. : 19861230
Feb 12, 1987ASAssignment
Owner name: NATIONAL WESTMINSTER BANK USA, 175 WATER STREET, N
Free format text: MORTGAGE;ASSIGNOR:ALLIED STEEL & TRACTOR PRODUCTS, INC., A CORP OF DE.;REEL/FRAME:004666/0099
Effective date: 19861230