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Publication numberUS3795168 A
Publication typeGrant
Publication dateMar 5, 1974
Filing dateJan 4, 1973
Priority dateJan 4, 1973
Publication numberUS 3795168 A, US 3795168A, US-A-3795168, US3795168 A, US3795168A
InventorsFriedland M, Lose J, Spachner S
Original AssigneeGulf & Western Ind Prod Co
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Low-impact four-bar press
US 3795168 A
Abstract
A low-impact four-bar press is provided having a drive mechanism which will cause the press slide to move relatively slowly throughout a metal working portion of the press stroke and relatively rapidly on the return and advance portions of the stroke. The drive arrangement includes a linkage system comprised of a first link pivotally connected between the press slide and a second link which is connected to a driven crank. A third, lazy or constraining link is pivotally connected at one end thereof to the second link and at the other end thereof to the press frame at a point between the crank axis and the path of the slide. The arrangement of the several links is such as to provide desirable kinematic and dynamic characteristics by developing a particular coupler curve at the pivot point between the first and second links.
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Description  (OCR text may contain errors)

United States Patent [1 1 Spachner et al.

Mar. 5, 1974 LOW-IMPACT FOUR-BAR PRESS Inventors: Sheldon A. Spachner, l-lavertown;

John G. Lose, Lansdowne; Martin Friedland, Flourtown, all of Pa.

Gulf & Western Industrial Products Company, Salem, Ohio Jan. 4, 1973 Assignees Filed:

'Appl, No; 321,009

US. Cl 83/617, 83/628, 74/38, 100/285 Int. Cl B26d 5/08 Field of Search 83/602, 617, 628; 100/283, 100/285; 74/38 References Cited UNITED STATES PATENTS 4/1940 Biggert, .lr 83/628 UX 7/1951 Klocke 100/283 X 6/1952 May 100/283 X Primary Examiner-Willie G. Abercrombie Attorney, Agent, or Firm-Meyer, Tilberry & Body [5 7 ABSTRACT A low-impact four-bar press is provided having a drive mechanism which will cause the press slide to move relatively slowly throughout a metal working portion of the press stroke and relatively rapidly on the return and advance portions of the stroke. The drive arrangement includes a linkage system comprised of a first link pivotally connected between the press slide and a second link which is connected to a driven crank. A

third, lazy or constraining link is pivotally connected at one end thereof to the second link and at the other end thereof to the press frame at a point between the crank axis and the path of the slide. The arrangement of the several links is such as to provide desirable kinematic and dynamic characteristics by developing a particular coupler curve at the pivot point between the first and second links.

36 Claims, 11 Drawing Figures ll 1'" 1 n ill I PATENTED 51974 SHEET 1 [IF 7 FIG. I

PATENIED MR 5 I974 SHEEI 8 0f 7 In. Ill" I PATENIEU 3.795.168

saw 5 or 7 FIG. 8

TEB 3.795.168

' SNEEI 8 0f 7 [BEGIN woRK STROKE I 4 LOW IMPACT 3 PRESS 4 /8 SLIDER 5 CRANK RREss I5 IO 5 0 5 l0 l5 Fl 6 H suns VELOCITY (IN/SEC.)

[BEGIN \NC K STROKE 4 LOW IMPACT 3 PRESS m 5 O: (I)

l SLIDER CRANK 8 g PRESS 3 2 42 22 ol 8l 6l ll 2lb 8 lo CRANK TORQUE (TON-IN.)

FIG. 9

PATENTED 74 SHEET 7 0F 7 LOW IMPACT PRESS AmmIOz: mxomkw VEO SLIDER -CRANK PRESS SLIDE VELO FIG.

LOW-IMPACT FOUR-BAR PRESS DISCLOSURE The present invention relates to the press art, and more particularly, to an improved mechanically driven press.

The invention is particularly suited for metal punching and stamping operations and will be described with reference thereto. it will be appreciated, however, that the invention is capable of broader application and could be used for many other types of press operations.

A majority of the mechanically driven presses in service today are used for metal-cutting operations such as punching or stamping, and for shallow metal drawing operations. The majority of these presses are rated between and 100 tons capacity and operate at a high speed and through relatively short strokes. Presses of this character generate considerable noise. In this respect, for example, 10 35 ton presses operating in a press room, punching steel, will generate a noise level in excess of 90 decibefs. Such a noise level makes conversation impossible. Moreover, and more importantly, high noise levels have deleterious effects on plant workers. In this regard, recent studies have illustrated that continuous exposure to levels of noise generated by punching and stamping operations causes permanent hearing loss to press operators. Recently, federal legislation has established maximum permissible factory noise levels.

The problem of noise generated by punch and stamping press operation has for some time received consideration by press builders. Heretofore, however, the approaches taken by the press builders to reduce noise have not been effective. The reason for this is that the press builders have concentrated their efforts on reducing the noise of the press structure itself, such as by providing sound insulating flywheel covers, oil immersed gearing, and the like. Almost any mechanical press, when in proper adjustment, will not generate objectionable noise when it is running idly and is not being employed to punch or stamp metal. Thus, the above mentioned efforts of the press builders simply serve to further reduce a noise level which was not initially objectionable in and of itself.

The noise level which is objectionable results primarily from the metal cutting or stamping operation or other metal working operation which the press performs. It is known that the noise generated by a metal cutting or punching tool is proportional to the impact velocity of the tool against the metal workpiece. if the tool strikes the workpiece at a high velocity, much more noise is generated than when the tool strikes the workpiece at a low velocity. For example, a slow speed hydraulic press generates an insignificant noise level when producing a metal working operation. Such a press, however, is capable of making only a few strokes per minute while operating at a low velocity and, accordingly, a considerable loss in production results. Such a low production rate is basically unacceptable.

The high impact velocity, high speed presses heretofore known, in addition to the foregoing disadvantages, present a problem with regard to tool life. In this respect, impact between a tool and metal workpiece is the governing factor in tool life. Since high production rates are desirable from such presses it has been assumed heretofore that high tool impact is unavoidable. In view of this consideration, the approach of industry has been to develop tool steels which are relatively tough, or impact resistant. Although tools have been developed as a result of this effort which are extremely tough and permit a high production rate, it remains that increased tool steel toughness has been obtained at the expense of tool steel wear. In other words, it is not possible to obtain both maximum wear properties and maximum toughness in the same tool steel.

The solution to the foregoing problems facing the industry is a press drive which will cause the press slide to move relatively slowly through the metal working portion of the press stroke and extremely rapidly through the advance and return portions of the stroke. Moreover, for maximum utility, the drive should possess a high power transmission efficiency, as well as being simple and inexpensive to build, operate and maintain. Such a press drive advantageously reduces the press operation noise level and greatly reduces tool impact. Thus, the press user is provided with a safer plant environment for his press operators, and realizes increased profits from longer tool life.

The present invention advantageously overcomes the disadvantages of high speed presses heretofore known, including the disadvantages specifically pointed out above, and in this respect provides a high speed press having mechanical drive linkage by which low noise level press operation and increased tool life are achieved. The drive linkage is capable of providing the desired velocity characteristics for the press while reducing torque requirements during the working portion of the stroke, in comparison with the torque requirements of conventional mechanical slider-crank driven presses. Moreover, the drive linkage is adapted for use with punching and stamping presses of 10 to 200 tons capacity operating with any stroke or length that might be required by a customer.

In accordance with the present invention there is provided a press including a frame, a slide reciprocable within the frame and mechanical drive means for imparting movement to the slide in a manner whereby the slide has a higher velocity during return and approach portions of the stroke and a lower velocity during the work performing portion of the stroke than high-speed mechanical presses heretofore known. More particularly, the drive mechanism includes a first link pivotally connected at one end thereof to the press slide and pivotally connected at the other end thereof to one end of a second link having its other end pivotally interconnected with the press crank. A third link is pivotally connected at one end thereof to the press frame and at the other end thereof to the second link. The third link defines a lazy or constraining link, and the several links cooperate to impart reciprocable motion to the slide through the first link in response to rotation of the press crank 360. Moreover, the link arrangement is such that the velocity of the slide is extremely low during the work performing portion of the press stroke and is relatively high during the return and approach portions of the stroke. The desired and advantageous velocities are achieved by a unique combination of the linkages which provides a unique output path which has a particular and advantageous relationship to the path followed by the slide of the press. The output path, or coupler curve, is defined by tracing the path of travel of the pivot point between the first and second link components of the linkage assembly.

By use of the subject invention, a substantial reduction may be made in the size and strength of the crank and the gear drive therefor. In this respect, the torque requirements of these components is reduced relative to presses heretofore known which are capable of operating at the same production rate. Thus, a press constructed in accordance with the present invention is economically compatible with high speed presses heretofore known and, at the same time, provides for a considerable reduction in noise level and an increase in tool life in comparison with such prior presses.

In accordance with the present invention, the linkage system generates a coupler curve or path which allows the press slide and thus the tool carried thereby to move rapidly toward and away from the work during advance and return portions of the press stroke while providing for the tool to slow down considerably during the work performing portion of the stroke to achieve a low impact between the tool and workpiece. Thus, the press can operate at a high speed and maintain a desirable high production rate and, at the same time,

achieve low noise level generation and increased tool life.

It is accordingly an outstanding object of the present invention to provide a press having a mechanical drive arrangement which permits highly desirable kinematic and dynamic slide characteristics to be achieved.

Another object is the provision of a press of the character described having a high production rate and which generates lower noise levels than mechanical presses heretofore known having the same production rate.

A further object is the provision of a press of the above character having a high production rate and in which the life of the press tool is longer than that of presses heretofore known having the same production rate.

Yet another object of the present invention is the provision of a press of the above character in which the press tool is reciprocated at a high velocity during return and approach portions of the press stroke and at a low velocity during the work portion of the stroke, thus to provide for low impact forces between the press tool and workpiece.

A further object of the present invention is the provision of a press having a drive arrangement which is especially suited for short stroke presses and creates an impact energy reduction during the work portion of the stroke of the slide of the press by providing for a slide velocity reduction during the work portion of the stroke.

Still a further object of the invention is the provision of a mechanically driven press capable of producing relatively uniform slide forces throughout a comparatively short working stroke and under a constant crank torque.

Yet another object is the provision of a press having a relatively low crank input torque requirement during the work portion of the press stroke.

Yet another object is the provision of a mechanical press having a linkage system for driving the press slide, which linkage system generates a coupler curve providing for reduced slide velocity during the work portion of the press stroke and increased slide velocity during return and advance portions of the press stroke in a manner whereby the press can be operated at a high speed to provide a high production rate while maintaining a noise level considerably lower than and a tool life considerably longer than high-speed mechanical presses heretofore known.

These objects in part will be obvious and in part more fully pointed out hereinafter in conjunction with the description of the accompanying drawings in which:

FIG. 1 is a plan view partially in section, illustrating a preferred embodiment of press structure in accordance with the present invention;

FIG. 2 is a front elevation, partially in section, of the press illustrated in FIG. 1, the section being taken along line 2--2 in FIG. 1;

FIG. 3 is a side elevation, partially in section, of the press device, the section being taken along line 33 in FIG. 2;

FIGS. 4-7 are schematic diagrams illustrating a drive linkage arrangement within the present invention at various points in a complete cycle of rotation of the crank component;

FIG. 8 is a schematic view illustrating the linkage shown in FIGS. 4'7 and illustrating the coupler curve or path generated thereby;

FIG. 9 is a chart illustrating certain operating characteristics of the present invention;

FIG. 10 is a chart illustrating other operating characterisitics of the present invention; and

FIG. 11 is an enlargement of a portion of the chart of FIG. 10.

Referring now to the drawings in greater detail wherein the showings are for the purpose of illustrating the preferred embodiments of the invention only and not for the purpose of limiting the same, FIGS. 1-3 illustrate a low-impact 25-ton capacity press having a 3- inch stroke and a A inch working stroke. In other words, the press slide has a linear movement from top dead center to bottom dead center of three inches and the portion of the stroke within one quarter inch from bottom dead center is the working portion of the stroke. The press shown is comprised of a frame assembly A, a drive unit B, and a mechanical drive linkage assembly C.

Frame assembly A may be of a variety of constructions and configurations, and in the preferred embodiment it is illustrated as a vertical frame comprising a base portion 10 supporting bed means 12, and an upper portion 14 housing drive unit B and drive linkage assembly C. More particularly, upper portion 14 of the frame assembly includes front wall 16 sidewalls l8 and 20 and intermediate walls 22 and 24 which are substantially parallel to front wall 16 and extend between and are welded or otherwise secured to sidewalls 18 and 20. The top of the press housing may be closed or open and, in the embodiment illustrated, the frame assembly is opened at the upper end thereof.

Drive unit B is comprised of an electric motor 26, flywheel 28, main drive pinion 30 and main drive gear 32. Pinion 30 is keyed to a shaft 34 for rotation therewith, and shaft 34 extends through walls 22 and 24 and is supported for rotation relative thereto by suitable bearing means 36 and 38, respectively. Flywheel 28 is mounted on shaft 34 for rotation relative thereto and is adapted to be rotated by motor 26 through a pulley 40 attached to the motor shaft and a plurality of V-belts 42 disposed in grooves provided therefor in pulley 40 and flywheel 28. Flywheel 28 is adapted to be selectively coupled to shaft 34 to impart rotation thereto by a suitable clutch and brake mechanism 44 indicated generally by broken lines in FIG. 1. Pinion meshes with and drives main gear 32 in response to rotation of shaft 34. Main gear 32 is keyed to shaft portion 46 of a crank element indicated generally by the numeral 47 and which defines one of the components of the mechanical drive linkage assembly C. Shaft portion 46 extends through openings in walls 22 and 24 and is supported for rotation relative thereto by suitable bearing means 48 and respectively. Thus, it will be apparent that motor 26 operates to drive flywheel 28 continuously and that brake and clutch mechanism 44 can be selectively actuated to couple flywheel 28 to shaft 34, thus to impart rotation to shaft 34 and drive pinion 30. Rotation of pinion 30 drives main drive gear 32 to impart rotation to shaft portion 46 of crank 47, whereby the crank is rotated about its axis 51 at a lesser speed than the rotation of pinion 30. The rotational speed of crank 47 is, of course, dependent on the ratio of pinion 30 to main drive gear 32. The means for selectively controlling the brake and clutch mechanism 44 is not illustrated and may be comprised of any suitable means manually controlled by the press operator or automatically controlled in response to press operation.

Crank 47 of the mechanical drive linkage assembly C includes shaft portion 46, mentioned above, and an axially aligned shaft portion 52 axially spaced from shaft portion 46 and interconnected therewith by a radially offset crank shaft portion 54 and crank arms 56. Each crank arm 56 has one end thereof integrally associated with shaft 46 or 52 and the other end thereof integrally associated with offset shaft portion 54. Shaft 52 extends through an opening in front wall 16 and is sup ported for rotation relative thereto by suitable bearing means 58. The outer end of shaft 52 may carry an indicator 59 which operates as a stroke indicator. The mechanical drive linkage further includes a connecting link 60 referred to herein as a drag link, and a pair of constraining links 61 and 62 disposed on opposite sides of connecting link 60. One end of connecting link 60 is recessed to receive offset shaft portion 54 of the crank arm and is pivotally interconnected with shaft portion 54 by a cap 64 suitably attached to the link body. Suitable bearing means 65 are provided between shaft portion 54 and arm 60 and cap 64. The other end of connecting link 60 is provided with spaced apart leg portions 66 between which the upper end ofa slide arm 68 is disposed. Slide arm 68 is pivotally interconnected with legs 66 by means of a cylindrical pin 70 extending through the legs and through an opening in the arm defined by a recess in the arm and a cap 71 suitably attached to the arm. Suitable bearing means 72 are provided between pin 70 and each of the legs 66, and hearing means 74 is provided between pin 70 and arm 68 and cap 71.

Constraining links 61 and 62 each have one end thereof pivotally interconnected with the press frame and the other end thereof'pivotally interconnected with connecting link 60. More particularly, one end-of constraining link 62 is pivotally associated with wall 22 of the press frame by means of a pin 76 projecting from wall 22, and the corresponding end of link '61 is pivotally associated with wall 16 of the press frame by apin 78 projecting from wall 16.'Pins 7'6 and 78 are coaxial, and bearing means 80 and '8-1 are interposed between pins 76 and 78 and the corresponding link. The other end of each of the links 61 and 62 is pivotally associated with connecting link 60 by means of coaxially disposed pins 82 and 84 which project from link 60 through suitable openings in the corresponding constraining link. Suitable bearing means 86 and 88 are disposed between pins 82 and 84 and the corresponding constraining link.

Slide arm 68 is associated at its lower end with a vertically reciprocable slide mechanism 90. Slide mechanism 90 includes a slide member 92 having guide bars 94 and 96 associated with guide way means 98 and 100 on front wall 16 and wall 22, respectively, which slide member guide bars and guide way means cooperate to guide slide member 92 during vertical movement thereof. The slide mechanism further includes a connecting member 102 having a cup-shaped upper end provided with socket defining components 104 and 106 which together cooperate with a ball component 108 on the lower end of arm 68 to define a ball and socket joint or pivot axis between arm 68 and slide 92. Connecting member 102 is rigidly fastened to slide 92 in any suitable manner. Ball 108 is adjustably mounted on slide arm 68 and in this respect may, for example, include a threaded shaft 110 adapted to project axially into arm 68 in threaded engagement with a threaded bore in arm 68. Thus, by rotating ball 108 the effective length of slide arm 68 can be varied. Ball 108 may be locked in an adjusted position thereof by any suitable means such as a split collar (not illustrated) and a cooperating threaded sleeve 112 which surrounds the collar and is threadedly associated with the lower end of arm 68. In a well known manner, sleeve 112 cooperates with the underlying spilt collar to radially compress the collar about arm 68 to lock ball 108 against displacement relative to arm 68. The lower end .of the shaft portion 110 of ball 108 may be provided with tool pads 114 to facilitate adjustment of the ball by means of a wrench or the like.

It will be noted, with reference to FIGS. 2and 3, that the press slide is adapted to reciprocate along a vertical path having an axis indicated by line 116. It will be noted further that the pivot axis 51 of the crank is fixed relative to the press frame as is the pivot axis of the ends of constraining links 61 and 62 mounted on frame walls 16 and 22 by corresponding pins 78 and 76. It will be further noted that crankshaft axis 51 and the axes of constraining link pins 76 and 78 are parallel and spaced apart with the axes of pins 76 and 78 being disposed below the horizontal plane of the crank axis. Moreover, the constraining link axis defined by the axes of pins 76 and 78 is disposed between slide path 116 and the crank axis. Rotation of the crank about its axis imparts, through the linkage mechanism, reciprocating movement to slide assembly 90 toward and away from press bed 12. Slide 92 of the slide assembly in use will be provided with a suitable tool such as apunch 93, for example. The particular four-bar linkage arrangement provides for the slide to have a low velocity during the work portion of the stroke thereof, whereby the tool 93 engages a workpiece supported on bed 12 with a low impact velocity during punching drawing or other work performed thereon.

FIGS. 4-7 of the drawing diagramatically illustrate one embodiment of a drive linkage assembly adapted to be employed in a press having a three inch total stroke such as, for example, a press of the character illustrated in FIGS. 1-3. FIGS. 4-7illustrate the linkage assembly as it passes through one complete cycle corresponding to rotation of the crank 360. For purposes of discussion of these FIGURES, the slide arm 120 defines a first link member having first and second end portions with the first end portion pivotally connected to a slide at a first axis A. The crank member 122 is rotatably mounted on the press frame at a second axis B and a second link member 124 has one end thereof pivotally connected to the crank arm at a third axis C and the other end thereof pivotally connected to the second end portion of slide arm link i20 at a fourth axis D. A third or constraining link member 126 has one end thereof pivotally connected to the press frame at a fifth axis E and the other end thereof pivotally connected to link 124 at a sixth axis F. The first axis A between line 120 and the slide is adapted to reciprocate along a path 128. It will be noted that fixed axis E between the constraining link 126 and the frame is disposed below fixed axis B between the crank arm and frame. Further, it will be noted that fixed axis E is located horizontally between slide path 128 and crank arm axis B. FIGS. 4-7 illustrate the positions of the various link components resulting from counterclockwise rotation of crank arm 122 360 in increments of 90. More particularly, when the linkage assembly is in the position illustrated in FIG. 4, slide arm axis A is in approximately the top dead center position thereof indicated A. As the crank arm rotates 90 counterclockwise to the position thereof illustrated in FIG. 5, slide arm axis A descends or advances toward the bottom dead center position thereof and after 90 rotation of the crank is positioned approximately at point A During rotation of the crank 90 from the position illustrated in FIG. 5 to the position illustrated in FIG. 6, slide arm axis A descends to the position A corresponding substantially to the bottom dead center position thereof. During rotation of the crank arm from the position of FIG. 6 to the position of FIG. 7, slide arm axis A enters the ascent or return portion of the stroke and moves to position A. Further counterclockwise rotation of the crank arm returns the linkage assembly to the position illustrated in FIG. 4, whereby the linkage assembly has completed one full cycle and the slide mechanism associated with slide arm axis A has travelled through one full stroke.

In a linkage system for a press of the character of the present invention it is important that the angle which is designated Q in FIGS. ;7, be no less than 40 and preferably be at least 45 for all positions of the crank during 360 rotation thereof. Accordingly, in FIGS. 4-7, the angle Q is the angle between links 124 and 126, and in the press structure of FIGS. l-3 the angle between link 60 and links 61 and 62. If the angle Q is less than 40 at any time during rotation of the crank, linkage slam and excess vibration may develop under high speed press operation, both of which are highly undesirable. Maintenance of a minimum angle of 45, however, will facilitate mechanism balancing and permit attainment of the desired high press stroke rates.

As the linkage assembly of FIGS. 4-7 rotates through one complete cycle, axis D between connecting link 24 and the second end of slide arm link 120 travels along a path 130 illustrated in FIG. 8, which path is referred to herein as a coupler curve or coupler path. Coupler curve 130 is indicative of the velocity of the slide arm axis A during a 360 rotation of the crank and, accordingly, the velocity of the slide during the complete stroke of the slide. The coupler curve has a major axis 132 which is the longest distance between any two points within the curve. In a linkage mechanism within the present invention, the major axis extends transversely of the slide path 128 and generally perpendicular thereto. In FIG. 8, the path of crank arm 122 is designated in 20 increments and the corresponding path of pivot axis D is indicated along the coupler curve. It will be noted with regard to these position indicators that when the crank arm is rotating from approximately the 40 to approximately the 180 position thereof, that slide arm axis A is ascending, and that when the crank arm rotates from approximately the 180 position through approximately the 260 position that the slide arm axis A pases through the top dead center position thereof and begins to descend. The ascent of slide arm axis A is relatively rapid as is indicated by the coupler curve, and the descent occasioned by rotation of the crank arm to about the 260 position thereof is also reasonably rapid. It will be noted, on the other hand, that final downward movement of slide arm axis A from approximately the 280 position of the crank arm through the 0 position and back to appfoximately the 40 position is extremely gradual. Accordingly, the velocity of the slide arm axis and thus the velocity of the slide is very low during the latter portion of rotation of the crank which corresponds to the portion of the total stroke in which the slide passes through the work portion and bottom dead center positions of the stroke.

The lengths of the several links and the points of cor rection therebetween can be varied to achieve the desired dynamic and kinematic characteristics sought for a given press size. In the particular embodiment under consideration, the press is a 25 ton press having a total stroke of 3 inches and a working stroke of A inch. The press is adapted to operate at a continuous speed of 150 strokes per minute. To achieve such press operation and provide for the desired low impact velocity during the working stroke, the preferred lengths for the various linkage components are as follows:

Crank arm 122; B-C distance 4.753 inches Link 124; C-F distance 9.506 inches Link 124; C-D distance 12.477 inches Link 124; D-F distance 2.971 inches Link 126; E-F distance H.883 inches Slide Path Intercept; B-I distance 15.345 inches Slide Path Inclination; Angle M l39.5

Ground Link; distance B-E l4.l64 inches Slide Arm A-D distance 21.237 inches It will be seen from the foregoing link dimensions that certain ratios exist with respect to the distances between axes of the link components of the drive linkage. For example, the distance between axes C and F along connecting link 124 is twice the distance between axes B and C of crank arm 122. Further, the distance between axes C and F along connecting link 124 is more than three times the distance between axes F and D along the connecting link. Moreover, the distance between axes E and F along constraining link 126 is more than twice the distance between axes B and C along the ,crank arm. These ratios may, of course, vary within the present invention, but it should be noted that the particular ratios establish a linkage system capable of producing the desired kinematic and dynamic characteristics of press operation regardless of the particular lengths of the various links in the assembly.

FIGS. 9, 10 and 11 are charts illustrating comparative requirements of a four-bar low-impact press made in accordance with the present invention and a mechanically driven slider-crank press having the same stroke and operated at the same number of strokes per minute, namely a 3-inch stroke and 150 strokes per minute. For example, it will be seen in FIG. 9 that at the beginning of the A inch work stroke a slider-crank press has a crank torque requirement of better than 22 ton inches while the lower impact press of the present invention has a crank torque requirement of approximately 8.25 ton inches. Moreover, during the total A inch work stroke, the low impact press crank torque requirement is substantially below that of the slider-crank press.

The chart in FIG. 10 illustrates the relative slide velocities of a slider-crank press and the low impact press of the present invention over the entire 3-inch stroke. It will be noted that the low-impact press has a higher velocity during the return and approach portions of the stroke and advantageously has a much lower velocity during the work portion of the stroke. The velocity relationship during the work portion of the stroke is better illustrated in FIG. 11 wherein the chart is an enlargement of the circled portion of the chart of FIG. 10. It will be seen in FIG. 11 that at the beginning of the work stroke the low-impact press has a slide velocity of approximately 5.5 inch per second while the slidercrank press has a slide velocity of more than 14 inches per second. Moreover, during the work stroke the slide velocity of the low-impact press is continuously lower as the two presses being compared approach the bottom dead center positions of the slides.

The following tabulation illustrates the decreases in slide velocity, impact energy and torque requirements of the low-impact press of the present invention when compared with the slider-crank press. The comparison is at increments of 1/32 inch during the inch work stroke illustrated in the charts of FIGS. 9, l and 1 1 for the two presses.

Distance from Slide Impact Stroke Velocity Energy Torque Bottom in Reduction Reduction Reduction inches 7a [/32 O 75 50 H16 59 83 59 3/32 66 88 66 H8 72 92 72 5/32 73 93 73 3/l6 7| 92 71 7/37 67 88 67 [/4 61 85 61 It will be appreciated from the foregoing that the lowimpact press of the present invention materially reduces slide velocity and impact energy during the work stroke, thus to achieve a considerable reduction in noise level and longer tool life and that these characteristics are achieved with a considerable reduction in the crank torque requirement which facilitates economical construction of the press.

It will be appreciated that by varying the lengths of the several links of the press described that other lowimpact presses can be produced having the desired lowimpact velocity during the work stroke and a high return and approach velocity thus to permit the press to operate at a desirable high stroke per minute rate While maintaining the noise level caused by impact of the press tool with the workpiece at a desired low level.

The present invention has been described in conjunction with certain preferred embodiments; however, various changes in these embodiments, and other embodiments of the present invention, will be obvious to those skilled in the art upon reading and understanding of the foregoing description. It is our intention to include all such embodiments and modifications of the present embodiments within the present invention insofar as they are within the scope of the appended claims.

We claim:

1. A press including: a frame, a slide member carried by said frame for reciprocable movement along a path between first and second positions, drive means for reciprocating said slide member between said first and second positions, said drive means including a first link member having first and second end portions with the first end portion pivotally connected to said slide member, a crank member rotatably mounted in said frame, a connecting link member having opposite end portions pivotally connected respectively to said crank member and the second end portion of said first link member, and'a constraining link member having a first end portion pivotally connected to said frame intermediate said path and the axis of said crank and a second end portion pivotally connected to said connecting link member intermediate said opposite end portions thereof.

2. The press as defined in claim 1, wherein said first end portion of said constraining link member is pivotally connected to said frame on an axis spaced below said axis of said crank member.

3. The press as defined in claim 1, and a second constraining link member parallel to said first constraining link member.

4. The press as defined in claim I, wherein said first end portion of said constraining link member is pivotally connected to said frame at a point intermediate said axis of said crank member and the adjacent end of said path.

5. The press as defined in claim 1, wherein the angle between said constraining link member and said connecting link member is at least 40 at all times during rotation of said crank member 360.

6. The press as defined in claim 1, wherein the distance between the pivot axis of said crank member and connecting link member and the pivot axis of said connecting link member and constraining link member is more than three times the distance between the pivot axis of said connecting link member and constraining link member and the pivot axis of said connecting link member and said first link member.

7. The press as defined in claim 6, wherein the distance between the pivot axes at opposite ends of said constraining link member is more than twice the dis tance between the pivot axis of said crank member and frame and the pivot axis of said crank member and connecting link member.

8.-The press as defined in claim 7, wherein the distance between the pivot axis of said crank member and connecting link member and the pivot axis of said connecting link member and constraining link member is approximately twice the distance between the pivot axis of said crank member and frame and pivot axis of said crank member and connecting link member.

9. In a power press comprising a frame, a slide member carried by said frame for reciprocable movement along a substantially straight slide path between first and second positions relative to said frame, and a linkage drive means for reciprocating said slide member between said first and second positions, the improvement comprising: said linkage drive means including a first link having first and second end portions with said first end portion pivotally connected to said slide at a first axis, a crank member rotatably mounted on said frame at a second axis, a second link having one end thereof pivotally connected to said crank member at a third axis spaced from said second axis and having the other end thereof pivotally connected to said second end portion of said first link at a fourth axis, and a third link having one end thereof pivotally connected to said frame at a fifth axis and the other end thereof pivotally connected to said second link at a sixth axis, said sixth axis being spaced from said third and fourth axes and being closer to said fourth axis than said third axis, said fourth axis defining a generally elliptical coupler path when said crank member is rotated 360 about said second axis, said coupler path having a major axis defined by the longest line between two points inside said coupler path, said major axis extending generally transversely of said slide path.

10. The press as defined in claim 9, wherein said coupler path includes portions disposed on opposite sides of said slide path.

1 l. The press as defined in claim 9, wherein the angle between said second link and said third link is at least 40 at all times during rotation of said crank member.

12. The press as defined in claim 11, wherein the distance between said third and sixth axes is more than three times the distance between said sixth and fourth axes.

13. The press as defined in claim 12, wherein the distance between said fifth and sixth axes is more than twice the distance between said second and third axes.

14. The press as defined in claim 13, wherein the distance between said third and sixth axes is approximately twice the distance between said second and third axes.

15. The press as in claim 10, wherein said slide path substantially bisects said major axis of said coupler path.

16. A power press comprising a frame, a slide member carried by said frame for reciprocable movement along a generally straight path between first and second positions relative to said frame, and a linkage drive means for reciprocating said slide member between said first and second positions with a predetermined velocity-time relationship, said linkage drive means including a first link member having first and second end portions with said first end portion pivotally connected to said slide at a first axis, a crank member rotatably mounted on said frame at a second axis, a second link member pivotally connected at one end thereof to said crank member at a third axis and pivotally connected to the second end portion of said first link member at a fourth axis, and a third link member having one end thereof pivotally connected to 'said frame at a fifth axis and a second end pivotally connected to said second link member at a sixth axis, said sixth axis being spaced from said third and fourth axes and being closer to said fourth axis than said third axis, said slide member moving from said one position thereof through said second position and back to said one position thereof in response to rotation of said crank member 360 about said second axis, said third link member constraining said fourth axis to follow a generally elliptical coupler path in response to rotation of said crank member, said coupler path having a major axis defined by the longest distance between two points within said path, said major axis extending substantially transverse to said slide path, said movement of said fourth axis along said coupler path being definitive of movement of said slide from said first position through said second position and back to said first position thereof in a manner whereby said slide member moves with high velocity during a substantial portion of movement thereof and at a much lower velocity during the remaining portion of the movement thereof.

17. The press as defined in claim 16, wherein the distance between said third and sixth axes is more than three times the distance between said sixth and fourth axes.

18. The press as defined in claim 17, wherein the dis tance between said fifth and sixth axes is more than twice the distance between said second and third axes.

19. The press as defined in claim 18, wherein the distance between said third and sixth axes is approximately twice the distance between said second and third axes.

20. A power press comprising a frame, a slide member carried by said frame for reciprocable movement along a generally straight path the length of which defines a stroke and which stroke includes a return and advance portion and a working portion, and a linkage drive means for reciprocating said slide member through said return and advance portion and said working portion, said linkage drive means including a first link member having one end thereof pivotally connected to said slide member, a crank member rotatably mounted on said frame and having a crank arm rotatable about the axis of said crank member, a second link member having one end thereof pivotally connected to the second end of said first link member and having the other end thereof pivotally connected to said crank arm on an axis spaced from said crank axis, and a constraining link member having one end thereof pivotally connected to said frame and the other end thereof pivotally connected to said link member at a point intermediate the opposite ends of said second link member, said slide member moving through said stroke thereof in response to rotation of said crank member 360, said constraining link member operating during rotation of said crank member to constrain movement of said slide member through said stroke thereof in a manner whereby said return and advance portion of said stroke is at a higher velocity than said working portion of said stroke.

21. The press as defined in claim 20, wherein said end of said constraining link pivotally connected to said frame is connected to said frame at a point intermediate the path of said slide and the axis of said crank member.

22. The press structure as defined in claim 21, wherein said constraining link pivot point is disposed below said axis of said crank member.

23. The press of claim 22, wherein said constraining link constrains the pivot axis between said first and second link members to follow a generally elliptical coupler path definitive of high velocity return and advance portions of said stroke and low velocity work portion of said stroke.

24. The press structure as defined in claim 22, wherein the angle between said second and third link members is greater than 40 at all times during rotation of said crank member 360.

25. The press as defined in claim 24, wherein the distance between the pivot axis of said crank member and connecting link member and the pivot axis of said connecting link member and constraining link member is more than three times the distance between the pivot axis of said connecting link member and constraining link member and the pivot axis of said connecting link member and said first link member.

26. The press as defined in claim 25, wherein the distance between the pivot axes at opposite ends of said constraining link member is more than twice the distance between the pivot axis of said crank member and frame and the pivot axis of said crank member and connecting link member.

27. The press as defined in claim 26, wherein the distance between the pivot axis of said crank member and connecting link member and the pivot axis of said connecting link member and constraining link member is approximately twice the distance between the pivot axis of said crank member and frame and the pivot axis of said crank member and connecting link member.

28. In a low-impact metal cutting power press comprising a frame,'a slide member carried by said frame for reciprocable movement along a generally straight path the length of which defines a stroke for the slide member and which stroke includes a short working portion and a relatively longer return and advance portion, the improvement which comprises: a linkage drive means for reciprocating said slide member through said working portion at a low velocity to achieve low-impact force between a tool carried by said slide member and a workpiece engaged by said tool during said work portion of said stroke, said linkage drive means including a first link member having one end thereof pivotally connected to said slide member, a crank member rotatably mounted on said frame and having a crank arm rotatably 360 about the axis of said crank member, a second link member having one end thereof pivotally connected to the other end of said first link member and having the other end thereof pivotally connected to said crank arm on an axis spaced from said crank axis,

and a constraining link member having one end thereof pivotally connected to said frame and the other end thereof pivotally connected to said second link member at a point intermediate the opposite ends of said second link member, said constraining link member operating during rotation of said crank member to constrain movement of said slide member through the stroke thereof so that said slide member moves through said work portion of said stroke at a low velocity to achieve low-impact force between said tool and workpiece.

29. The improvement of claim 28, wherein said end of said constraining link pivotally connected to said frame is connected to said frame at a point intermediate the path of said slide and the axis of said crank member.

30. The improvement of claim 29, wherein said constraining link pivot axis is disposed below said axis of said crank member.

31. The improvement of claim 30, wherein said constraining link constrains the pivot axis between said first and second link members to follow a generally elliptical coupler path definitive of high velocity return and advance portions of said stroke and a low velocity work portion of said stroke.

32. The improvement of claim 31, wherein the angle between said second and constraining link members is greater than 40 at all times during rotation of said crank member 360.

33. The improvement of claim 32, wherein said angle is at least 45 at all times during rotation of said crank member.

34. The press as defined in claim 2, wherein said axes of said crank member and said constraining link memslide at an angle of approximately 139.5.

UNITED STATES PATENT OFFICE CERTIFICATE OF CORRECTION Patent No. 3,795, 168 Dated March 5, 1974 Inventor) Sheldon A. Spachner, John G. Lose & Martin Friedland It is certified, that error appears in the above-identified patent and that said Letters Patent are hereby corrected as shown below:

Column l2, line 55, after "stroke" insert said end of said constraining link pivotally connected to said frame being connected to said frame at a point intermediate the path of said slide and the axis of said crank member Column 12, delete lines 56-60 in their entirety.

Column l2,' line 61, change the numeral "21" to 20 Column 14, line 10, after "workpiece" insert said end of said constraining link pivotally connected to said frame being connected to said frame at a point intermediate the path of said slide and the axis of said crank member Column 14, delete lines 11-15 in their entirety.

Column 14, line 16, change the numeral "29" to 28 On the cover sheet, "36 Claims should read 34 Claims Signed and sealed this 17th day of September 1974.

v (SEAL) Attest:

MCCOY M. GIBSON JR. c; SHALL DANN Attesting Officer Commissioner of Patents FORM PO-1OSO (10-69) USCOMM DC 60376.:89 ursr GOVERNMENT PRINTING orncs; I969 0-366-334

Patent Citations
Cited PatentFiling datePublication dateApplicantTitle
US2197391 *May 4, 1937Apr 16, 1940United Eng Foundry CoPress and drive therefor
US2562044 *Mar 12, 1949Jul 24, 1951William KlockeHigh-speed mechanical power press
US2600242 *Feb 28, 1949Jun 10, 1952Hermann May OttoToggle press
Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US3919876 *Nov 11, 1974Nov 18, 1975Du PontToggle press
US4138904 *Jul 20, 1977Feb 13, 1979Verson Allsteel Press CompanyLink drive mechanism for mechanical presses
US5189922 *May 7, 1990Mar 2, 1993Parashikov Peter HForce impulse generator
US5919015 *Apr 16, 1997Jul 6, 1999Utica Enterprises, Inc.Mechanical drive for a blind spline broaching machine
US6826980Nov 5, 2001Dec 7, 2004George SchmidegDrive and control systems for high speed intermittent motion generations, control and applications
US7191526Dec 9, 2005Mar 20, 2007Milwaukee Electric Tool CorporationMovable handle for a power tool
US7308764Jan 24, 2006Dec 18, 2007Milwaukee Electric Tool CorporationPower tool with movable handle
US7497152Jan 24, 2006Mar 3, 2009Milwaukee Electric Tool CorporationMovable handle for a power tool
US8061043Nov 15, 2007Nov 22, 2011Milwaukee Electric Tool CorporationPower tool
US8640346Oct 19, 2011Feb 4, 2014Milwaukee Electric Tool CorporationPower tool
WO1998001244A1 *Jul 7, 1997Jan 15, 1998Jeandeaud Jean ClaudeDevice for moving part of a machine and for exerting a force at the end of its stroke
Classifications
U.S. Classification83/617, 74/38, 83/628, 100/285
International ClassificationB30B1/00, B21D28/00, B30B1/14
Cooperative ClassificationB21D28/002, B30B1/14
European ClassificationB21D28/00B, B30B1/14
Legal Events
DateCodeEventDescription
Jul 12, 1995ASAssignment
Owner name: SHAWMUT CAPITAL CORPORATION, CONNECTICUT
Free format text: SALE/TRANSFER OF SECURITY INTEREST TO A NEW SECURED PARTY;ASSIGNOR:BARCLAYS BUSINESS CREDIT, INC.;REEL/FRAME:007644/0215
Effective date: 19950130
Oct 5, 1988ASAssignment
Owner name: BARCLAYS AMERICAN/BUSINESS CREDIT, INC., CONNECTIC
Free format text: SECURITY INTEREST;ASSIGNOR:E.W. BLISS COMPANY;REEL/FRAME:005880/0330
Effective date: 19880915
Dec 23, 1983ASAssignment
Owner name: E.W. BLISS COMPANY, INC., DE. A CORP OF DE
Free format text: ASSIGNMENT OF ASSIGNORS INTEREST.;ASSIGNOR:GULF & WESTERN INDUSTRIAL PRODUCTS COMPANY A CORP OF DE;REEL/FRAME:004204/0264
Effective date: 19831110
Owner name: E.W. BLISS COMPANY, INC.,, DELAWARE
Free format text: ASSIGNMENT OF ASSIGNORS INTEREST;ASSIGNOR:GULF & WESTERN INDUSTRIAL PRODUCTS COMPANY A CORP OF DE;REEL/FRAME:004204/0264