US3464501A

US3464501A - Automatic pneumatic impact hammer

Info

Publication number: US3464501A
Application number: US673121A
Authority: US
Inventors: Eugene T Ward
Original assignee: Allied Steel and Tractor Products Inc
Current assignee: Allied Steel and Tractor Products Inc
Priority date: 1967-10-05
Filing date: 1967-10-05
Publication date: 1969-09-02
Anticipated expiration: 1986-09-02
Also published as: DE1583806C3; GB1205428A; DE1583806B2; DE1583806A1

Description

Sept. 2, 1969 a. T. WARD 3,464,501

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Sept. 2, 1969 E- T. WARD AUTOMATIC PNEUMATIC IMPACT HAMMER Filed Oct. 1967 United States Patent 3 464 501 AUTOMATIC PNEUMATIC IMPACT HAMMER Eugene T. Ward, Highland Heights, Ohio, assignor to Allied Steel & Tractor Products, Inc., a corporation of Ohio Filed Oct. 5, 1967, Ser. No. 673,121 rm. c1. B23q 5/06; 1321c 5/08 US. Cl. 173-17 2 Claims ABSTRACT OF THE DISCLOSURE This invention relates to improvements in pneumatic hammers and, more particularly, to large pneumatic hammers mounted on and manipulated by construction machinery, such as a tractor equipped with a backhoe boom, and the like.

In construction work, demolition work for new construction, mines, quarries, mills, in fact nearly everywhere that the well-known air-driven jack hammers are and have been employed, there has been a long standing need for impact hammers which are more massive than those which can be handled by one or two men and yet which are more flexible and manipulatable than rather difficultly mobile power hammers, such as mobile pile drivers and the like. Such large manipulatable impact hammers have been proposed for mounting on the booms or dip-sticks of tractor-mounted backhoes, grading machinery, frontend loaders, or the like, by which the hammers tool bit can be manipulated between vertical to horizontal planes or otherwise positioned for the optimum effect of its impact. With such machine-manipulated hammers, weighing in the order of 1,000 pounds, the tool can be driven, with compressors of standard sizes for manual jack hammer crews, so as to deliver energy at the rate of 300,000 foot/pounds per minute. In many demolition jobs, as in demolishing reinforced concrete pavements, foundations, etc., such hammers, especially if made according to this invention, will out-perform crews operating to 25 conventional jack hammers. One reason for the high performance of such large impact hammers is not so much the rate of delivering energy as it is the amount of energy delivered in a single blow at a fairly high rate of speed, i.e., in the order of 1,000 foot/pounds per blow (at the rate of 300 per minute).

The very massiveness of such machine-mounted high impact hammers and the consequent difliculty of controlling them heretofore militated again-st their general use and caused frequent and expensive breakdowns when used. Heretofore, an operator of a construction machine for manipulating such a massive impact hammer has his hands full in either operating the hammer or in operating the machine to position the hammer. Heretofore the most successful of such massive impact hammers were operated by valves which were manually opened by the operice ator after the hammer had been positioned and closed when the tool had broken through or to allow the operator to reposition the hammer. Not only was the efficiency reduced because the operator had to perform his duties of positioning the hammer and operating it in succession, but many breakdowns, often attributed to other causes, were due to the fact that the operator had no feel of the load on the hammer bit, and the hammer would operate when there was no resistance load on the bit, even though the hammer might theoretically be ported, according to previous constructions, to stop when the hammer bit encountered no resistance.

It is an object and advantage of this invention to provide an impact hammer which is automatically operated by the positioning of the hammer by the operator and which automatically ceases operation when the hammer tool no longer encounters resistance, either due to the fact that the operator has removed the hammer from the load or the tool has broken through or otherwise completed its work at the moment. More efficient operation is thereby attained and breakdowns are minimized.

Other objects and advantages of this invention will be apparent from the following claims and the detailed description, and drawings of one embodiment in which:

FIGURE 1 is a small diagrammatic elevation showing an impact hammer made according to this invention mounted for manipulation on the boom of a backhoe equipped tractor.

FIG. 2 is an enlarged side elevation of the hammer shown in FIG. 1.

FIG. 3 is an end view, taken from the line 33 of FIG. 2.

FIG. 4 is a detailed front elevation of the hammer shown in FIG. 2, partly broken away to show the positioning of the significant elements when the hammer is in operation.

FIG. 5 is a view similar to FIG. 4, but showing one positioning of the significant elements when the hammer is operative.

FIG. 6 is an enlarged detail of the mating shoulders of the tool shank band and the tool retainer, taken from FIG. 4.

FIG. 1 shows a hammer 10 made according to this invention mounted on the boom 2 of a vehicle 3 representing, diagrammatically, a tractor equipped with a back-hoe mechanism in which the back-filling bucket has been replaced by the hammer 10. As indicated above, the particular type of vehicle upon which hammers made according to this invention may be mounted is purely a matter of choice determined by the available vehicles and the jobs to be done. In practice such vehicles have varied from universal road-grading machines having a boom arrangement permitting substantially fully universal angular movement of the attachment on the end of the boom to simple fork-lift trucks With the hammers mounted in place of the lifting forks. The cylinder of the hammer 10 is connected by the air-line 4 to an adequate source of operating compressed air, not shown, but usually a portable motor-driven compressor unit as commonly used on construction jobs.

As evident from FIGS. 2 to 5, the main structural elements of the hammer 10 comprise, in the embodiment shown, an elongated cylinder block 11 and end members comprising a cylinder intake head 12 and a tool retainer head 13, all, except as the intermediate portion of the cylinder block is relieved at its corners, essentially rectangular in horizontal cross-section. These members are held in longitudinal alignment by four tie-bolts 14 extending through openings drilled longitudinally through the corners. While not directly relevant to the principal feature of this invention, it should be pointed out that the tie-bolts 14 and the openings in which they are re ceived are preferably carefully machined to provide a close sliding fit between the bolts and the aligned openings; this, together with the setting of equal tension on the tie-bolts by means of the pinned end nuts and cap nuts 16, aids in maintaining, during operation, alignment between the cylinder block 11 and its

end members

12 and 13. This alignment is also obtained by the boss and recess fit as shown between the block 11 and retainer 13 in FIGS. 4 and 5. The tie-bolts 14 are preferably of high-tensile steel, machined and polished to remove all nicks between the enlarged threaded portions which might initiate a fracture; these bolts should not only bear, without deformation beyond their elastic limit, the full longitudinal force of the hammer whenever the tool suddenly breaks through substantial resistance but also the wedging or side loads which the tool encounters during operation of the hammer.

The entire hammer 10 is carried by the support plates 17 having drilled holes 18 for bolting to a boom 2. The plates 17 carry pads 19 which closely fit in the ways 20 cut in the cylinder block 11. The block 11 is thus removably locked between the plates 17 by means of the clamping studs 21 and the dowel studs 22, the latter extending through the plates 17 and pads: 19 into the block 11.

As indicated in FIGS. 4 and 5, the cylinder block 11 is provided with a central longitudinal bore 25, counterbored at its lower end to receive a tool bushing 26. Parallel to the bore 25 in the wall of the block 11 is one or more return passageways 27 connecting the valved intake head 12 to a lower return port 28 opening into the bore 25. The bore 25 is ported to the atmosphere in its upper portion by an impact exhaust port 29 and near its lower end by one or more return-disactivating ports 30 drilled through the bushing 26 and through so much of the pad 19 and plate 17 that, as shown in FIGS. 4 and 5, the positioning of these latter members otherwise might block and thereby effectively alter the position and capacity of the ports 30.

As indicated in FIGS. 4 and 5, the bore 25 receives a heavy ram reciprocal in the bore as a free piston. The upper end of the bore 25 is closed by the intake head 12 having an intake port 31 to which an air pressure hose 4 may be connected; the intake head is also provided with suitable valving exhaust ports opening, in this instance, into the circumferential gap 32 between the head 12 and block 11. Within the head 12 is an automatic valve V which admits air under pressure when the free-piston ram 35 is in the upper end of the bore 25 until the ram is driven forward in an impact stroke so as to clear the exhaust port 29. In response to the pressure drop in the upper end of the bore, the valve closes the upper portion of the bore 25 to line pressure and opens it to exhaust through the gap 32 while admitting line pressure to the passageway 27 and port 28; thus, if the tool is under load, as explained below, the ram 35 will be returned to the upper portion of the bore 25 to position the ram for an impact stroke and repetition of the ram-reciprocating cycle. Automatic valves such as the valve V, usually of a poppet and sleeve type, are conventional in the pneumatic hammer art and since at least several of such valves are available and operable, the valve V is indicated only diagrammatically.

The lower end of the cylinder bore 25 is closed by a tool carried by the tool retainer head 13. A tool 40 comprises a shank having an upper portion 41 and a lower portion 42 on either side of an integral enlarged shank band 43. The body of the tool below the shank terminates in an end shaped according to the specific job for which the hammer is to be employed. The pointed end shown is for general demolition work; chisel ends are generally employed for line breaking of pavements and flat ends for compacting or post driving; also, the end may be equipped with a suitably shaped fitting for driving sheet piling and the like. Tools 40, provided with the above-described shank construction, are readily interchangeable in the hammer, either to accommodate it for different jobs or for replacement of worn or broken tools; this is accomplished by simply disconnecting the retainer head 13 from the cylinder block 11, replacing a tool, and then reconnecting the retainer head.

As indicated in FIGS. 4, S and 6, the retainer head 13 is provided with a central bore 45, axially aligned with the cylinder bore 25, and also a counter-bore 46 to receive the shank band 43. It is to be noted, as shown in detail in FIG. 6, that the shank band 43 merges into the lower shank portion 42 by convex and cancave fillets to provide a filleted shoulder 47 which, when the tool 40 is in its lowermost position in the retainer head 13, mates with an oppositely filleted shoulder 48 at the juncture of the counter-bore 46 with the bore 45 on the barrel 49 of the retainer head 13.

Referring to FIGS. 4 and 5, it is to be noted that the axial length of the shank band 43, and the length of the upper shank portion 41 are proportioned, with respect to the depth of the counter-bore 46, so that, when the shoulder 47 of the tool 40 is engaged with the shoulder 48 of the retainer head, an appreciable length of the shank portion 41 is engaged and guided in the closely fitting bushing 26. Likewise, the head 36 of the ram 35 is tapered so as to engage the top of the tool 40 when the tool is in the lowermost position in the retainer 13 while providing a slight clearance with the bushing 26. The length of the upper shank portion 41 is somewhat greater than or at least substantially equal to the length of the bushing so that when the shank band 43 engages the lower end of the bushing 26 as a stop, the upper shank portion 41 extends above and thereby closes the return-disactivating ports 30. The return port 28 is located in the bore 25 so that it is closed by the cylindrical body of the ram 35 when the tool is in its lowermost position in the retainer 13, as shown in FIG. 4. The port 28, however, is opened to the space between the tapered head 36 of the ram when the tool is in its uppermost position and, the ram reciprocates between the solid line and dot-and-dash line positions, as shown in FIG. 5, in the operating cycle of the impact hammer.

The automatic control of the hammer operation by its above-described ports is as follows: With the point of the tool 40 out of contact with the surface to be struck, as it will be when the tool is being moved to its desired position by the operator, the tool 40 and ram 35 will be in the disactivated position shown in FIG. 4. This will be due to gravity and/or the ram-disactivating effects of the ports 30 and the filleted shoulders 47 of the shank band 43 and the mating shoulder 48 of the barrel 49 of the tool retaining head 13. In this disactivated position, the air-line operating pressure is switched by the valve V to the passageway 27, but its port 28 is not only closed by the body of the ram 35, but the ports 30 vent to the atmosphere the leakage from the ports 28, which leakage occurs due to the necessary tolerance and unavoidable wear between the ram 35 and the bore 25 and, in the absence of the ports 30, could otherwise build up to an activating pressure around the ram head 36. As the operator positions the hammer so that the tool point engages the surface to be hammered and moves the hammer toward such surface, the tool is automatically forced to the position shown in FIG. 5, where the tool shank closes the ports 30 and the ram head 36 is moved back to uncover the port 28 and commence the operating cycle of the hammer. This cycle repeats itself rapidly until either the tool breaks through and clear of the material being hammered or the operator, in order to reposition the tool,

manipulates the necessary controls to lift the hammer so as to clear the tool.

The key feature of the above described automatic control of the hammer operation is a positive disactivation of the hammer when the tool breaks clear or is lifted out of contact by the machine operator. In the first place, if the hammer operates with no load on the tool, the violent vibration of the hammer itself and the boom of the machine which positions it makes the hammer very diflicult to position as well as frightening to the operator-the latter not being without cause. Wholly apart from the ability of the boom or other associated parts of a machine used to manipulate the hammer to withstand violent vibrations, the hammer itself is designed to withstand, with an ample factor of safety, the full impact of the ram 35 upon the tool 40 when the latter is subject to no resistance load, but not unlimited no-load driving of the ram upon the hammer. That is, the energy of a no-load impact of the ram 35 upon the tool 40 can normally be safely taken up by the retainer head 13 and the block 11 acted upon through the bolts 14 and intake head 12, in sum, by the mass of the entire hammer. But if the hammer were designed to take up and dissipate the energy delivered by the ram 35 over a prolonged period of noload conditions without rupture of some element of the hammer, its mass with respect to the impact that could be delivered under load would be so great as to make the hammer both unwieldy to operate and uneconomical to make.

The successful positive disactivation of a hammer as above described, upon one no-load impact of the ram 35 upon a tool 40 depends upon several factors, two of which, preferably in combination, are not readily apparent or calculatable but which can usually be determined empirically. These factors-mot necessarily in the order of their importance-are: (a) the contour of the mating shoulders 47 and 48 with respect to the mass of metal in the retainer head 13 and, particularly, the barrel 49; (b) the distance between the effective closing edges of the upper portion 41 of the tool shank and the port 30, when the tool is in the disactivated position shown in FIG. 4, with respect to the resiliencies under a full noload impact, of the ram 35, the hammer 10, and the tiebolts 14 plus, to a certain degree, the resilience of the intake head 12 and block 11. Discussing the latter factor first-in designing any specific model of a hammer made according to the invention to avoid continued operation under no-load conditions, it has been found necessary to visualize the hammer as though its elements were made of rubber. That is, if the port is of insufficient diameter and/or located too close to the edge of upper tool shank portion 41, under no-load impact the ram will bounce back from its impact on the tool so as to uncover the return port 28 while, at the same time, the tool 40 will bounce back from the impact of its band 43 upon the shoulder of the barrel 49, closing the disactivating ports 30. Under such conditions, the hammer will thereby continue operating, until the air supply is cut off, as though there were no provision for a no-load disactivation of the ram 35. Once the above cause of such ineffectiveness of the disactivating port is discovered, the condition is usually cured by empirically lengthening the counter-bore 46 so as to drop the shank until trial and error demonstrates that the edge of the upper portion 41 will not bounce into closing relation with the port 30. In regard to the contour of the

shoulders

47 and 48, because these shoulders provide a stop which must resist the full impact of the ram 35 when the shoulder 47 is in engagement with the shoulder 48, normal design principles would dictate that there be a minimum filleting of the bottom of the counter-bore 46 and the upper edge of the bore of the barrel 49 and the corresponding surfaces of the band 43 and lower shank 42. The purpose of minimizing such filleting would be to avoid the wedging action which would be expected to arise from such mating fillets and the consequent jamming and/or splitting of the barrel 49; also, the

shoulders

47 and 48 would then engage in planar contact perpendicular to the direction of the impact and thus, in normal theory, best positioned to resist such impact. In practice, however, it was discovered that such normal sound design of the

shoulders

47 and 48 with a minimum of filleting apparently promoted the bounce of the tool and the hammer under full no-load impact; this in turn required not merely lengthening the depth of the counter-bore 46, as described above, but also redesigning the cylinder block 11 and upper tool shank 41 in order to achieve suflicient spacing (especially in some models designed for higher operating air presures) to allow the port 30 to effectively disactivate the hammer under no load. Instead, it was found that by enlarging the fillets of the

shoulders

47 and 48 up to the maximum which would still provide a projected area perpendicular to the direction of impact, the bounce is apparently decreased, thereby avoiding a need for excessive depth of the counter-bore 46 and/or a lengthened design of the hammer. The decrease in bounce attributable to such full-filleted shoulders as shown in FIG. 6 (i.e., fillets whose radii are approximately half the difference between the radius of the band 43 and the radius of the shank portion 42) may be due to momentary wedging of the mating fillets; however, the cold-working of the

shoulders

47 and 48 during break-in and use indicates that the momentary braking effect of the filleted shoulders may be due to a momentary shrinking of the bore of the barrel 49 around the lower shank portion 42.

I claim:

1, In a pneumatic impact hammer comprising a cylinder block provided with a central bore, a ram reciprocal in said bore, an automatic valve closing the inner end of said bore and which switches operating fluid from ports in the inner and outer portions of the bore so as to automatically reciprocate said ram in said bore, a tool having a shank portion which closes the outer end of said bore and which is reciprocal therein, said tool shank being reciprocal in said bore in response to and between the impact on the shank end of said tool imparted by said reciprocal ram and pressure applied to the outer end of said tool, and means limiting the inner and outer reciprocal movement of said tool shank, said bore being provided with a return-disactivating port located in the portion of said bore traversed by said shank and communicating said more to the atmosphere, said outer end of the ram in the latters outermost position along the bore being of substantially smaller cross-section than the bore at said disactivating port and outward from said disactivating port, whereby, when said tool shank is in its outermost position in said bore, any fluid pressure between the inner end of said shank and the outer end of said ram and tending to force said ram toward the inner end of said bore is exhausted rapidly to the atmosphere through said disactivating port to disactivate a reciprocating pressure on said ram until sufficient pressure is applied to the outer end of said tool to force said tool shank inwardly and close said disactivating port and whereby relief of such pressure on the outer end of said tool terminates reciprocation of said ram in said bore, the length of the bore traversed by said tool shank being greater than the axial length of said disactivating port plus the axial distance said tool shank will be bounced inwardly by a resilient reaction to an outward impact of said ram when there is substantially no opposing pressure on the outer end of the tool, said disactivating port being located inwardly of the innermost distance to which said tool shank will be bounced.

2. An impact hammer according to claim 1 and further comprising a tool retaining head on the outer end of said cylinder block, said tool retaining head having a bore in its outer end for passing the tool and a counterbore extending inward from said last-mentioned bore and aligned with and opening into said cylinder block bore, said tool retaining head presenting an inwardly-facing transverse 7 shoulder at the intersection of its bore and counterbore, said tool having an enlarged annular portion slidably received in said counterbore and engageable with said shoulder on the tool retaining head to limit the outward movement of the tool, the walls of said bore and counterbore in the tool retaining head merging into said shoulder by oppositely curved arcuate fillets of substantially equal radii, and the sum of the radii of said fillets being substantially equal to the difference between the radii of the counterbore and the bore in the tool retaining head, said enlarged annular portion of the tool having curved portions at the radially outward and radially inward extremities of its outer end which have contours that substantially mate With the respective fillets.

References Cited UNITED STATES PATENTS 1,637,192 7/1927 Jimerson 17317 1,713,784 5/1929 Slater 173-17 3,398,801 8/1968 Kotone 173l6 10 ERNEST R. PURSER, Primary Examiner US. Cl. X.R. 173-17