Search Images Maps Play YouTube News Gmail Drive More »
Sign in
Screen reader users: click this link for accessible mode. Accessible mode has the same essential features but works better with your reader.

Patents

  1. Advanced Patent Search
Publication numberUS3817139 A
Publication typeGrant
Publication dateJun 18, 1974
Filing dateJan 29, 1973
Priority dateJan 29, 1973
Publication numberUS 3817139 A, US 3817139A, US-A-3817139, US3817139 A, US3817139A
InventorsDesai P, Spachner S
Original AssigneeGulf & Western Industries
Export CitationBiBTeX, EndNote, RefMan
External Links: USPTO, USPTO Assignment, Espacenet
Press and drive mechanism therefor
US 3817139 A
Abstract
A billet or slab shearing press is disclosed which includes an improved mechanical drive arrangement. The drive arrangement includes a linkage system including a first link pivotally connected between the press slide and a second link which is connected to a driven crank. A third link is pivotally connected between the press frame and the second link. The several links are arranged to provide desirable kinematic and dynamic characteristics for a shearing press by developing a selected coupler curve at the pivot point between the first and second links, which coupler curve defines the movement of the slide, and accordingly the shearing blade, through the stroke of the slide.
Images(5)
Previous page
Next page
Description  (OCR text may contain errors)

United States Patent [191 Desai et a1.

PRESS AND DRIVE MECHANISM THEREFOR Inventors: Prakash Dhirubhai Desai, Raleigh,

NC; Sheldon Arthur Spachner, Havertown, Pa.

Gulf & Western Industries, Inc., New York, NY,

Filed: Jan. 29, 1973 Appl. No.: 327,453

Assignee:

References Cited UNITED STATES PATENTS 1/1926 Pels 83/632 June 18, 1974 3,158,057 11/1964 Hartman 83/617 3,766,771 10/1973 Spachner et al. 72/450 Primary ExaminerDonald R. Schran Attorney, Agent, or Firm-Meyer, Tilberry & Body [5 7] ABSTRACT A billet or slab shearing press is disclosed which includes an improved mechanical drive arrangement. The drive arrangement includes a linkage system including a first link pivotally connected between the press slide and a second link which is connected to a driven crank. A third link is pivotally connected between the press frame and the second link. The several links are arranged to provide desirable kinematic and dynamic characteristics for a shearing press by developing a selected coupler curve at the pivot point between the first and second links, which coupler curve defines the movement of the slide, and accordingly the shearing blade, through the stroke of the slide.

1 Claim, 8 Drawing Figures ale-317L139 PATENTEDJun 18 m4 sum 1 at 5 PATENTEDJumm 31817339 SHEET 2 [1F 5 PKTENTEDJumm 3.817339 sum 3 a? 5 FIG. 3.

80 so so I so PATENTEDJuu 18 m4 sum u m5 5 FIG. 4.

1 PRESS AND DRIVE MECHANISM THEREFOR The present invention relates to presses and, more particularly, to an improved billet or slab shearing press drive mechanism.

Presses have been provided heretofore for shearing relatively thick slabs or billets of metal ranging, for example, between two to six inches in thickness. Such a billet or slab places a relatively constant load on a shearing blade during most of the working stroke thereof. Heretofore, constant force has been generated during the working stroke by use of hydraulic drives or slider-crank mechanisms. While such hydraulic drives and slider-crank mechanisms individually have certain advantages, both have disadvantages which are overcome by the present invention in a manner whereby a reduction in blade cost may be realized and a reduction in maintenance cost is achieved without any sacrifice in performance or reliability.

Hydraulic drive systems are capable of developing a constant force over the entire shearing stroke of a press, but when designed to generate a required shearing force, they are not physically capable of generating higher forces. In this respect, if a 2,000 ton shearing force is required, only a machine frame designed for a load of 2,000 tons can be used. Moreover, large shears of the character being considered herein make only a few strokes per minute. Accordingly, in operation l ge amounts of power are intermittently required. The most economical way of meeting this requirement is through the use of an energy accumulator which stores energy obtained from a relatively low-horsepower motor then releases the energy in a short time period. Hydraulic drive systems utilize a pressure vessel or accumulator partially filled with compressed gas. Liquid is pumped into the accumulator to further compress the gas. The accumulator is valved, and when the valve is opened the pressurized gas expels the fluid in the accumulator, thus to supply needed liquid to the hydraulic drive at a high rate of speed. The hydraulic accumulator system is relatively expensive and, moreover, maintenance costs are relatively high. Accordingly, the principle saving in the use of a hydraulic drive system in a shearing press is in frame cost since the frame of the shear can be designed for loads which do not greatly exceed the required shearing force.

A slider-crank driven shear of the character being considered herein also requires the use of an energy accumulator. The energy accumulator for such a press is much simpler and less expensive than a hydraulic accumulator system. In this respect, the mechanical energy accumulator is simply a flywheel designed driven to store sufficient energy for performance of the working or shearing stroke without a great loss of speed. Further, the mechanical slider-crank mechanism is extremely simple and reliable in comparison with a hydraulically driven press. The slidencrank drive mechanism, however, disadvantageously requires a frame of extremely high strength. The cost of providing such strength in a frame ofi'sets the advantages of a slidercrank mechanism over a hydraulically driven press having a relatively complex power supply. With regard to the frame requirements for a slider-crank driven shearing press, if a 2,000 ton shearing force is required over a 6 inch working or shearing stroke, the machine must have a frame that can withstand a force of at least 21,000 tons in order to prevent damage to the shear in the event of accidental overloading. In contrast to this, as mentioned above, a hydraulic shearing press having a 2,000 ton shearing force requirement requires a frame rated at a little more than 2,000 tons to provide the same factor of safety in operation. Another disadvantage related to the use of a slider-crank drive mechanism is that the latter has a relatively low mechanical advantage at the beginning of the working or shearing stroke, whereby the crank component of the mechanism must be designed to withstand an extremely high torque. More particularly, for a slider-crank shearing press having a 2,000 ton force requirement, the crank would have to be designed to withstand a torque of 18,000 ton-inches.

Other mechanically driven shearing presses have been designed in an effort to obtain the advantages of slider-crank presses while reducing the frame strength requirements. In this respect, a mechanically operated shearing press has been provided which operates to progressively shear a billet or slab of metal. Progressive shearing is achieved by making a series of short shearing strokes through the billet or slab. A slider-crank mechanism is employed to drive the shear. If such a shearing press has a 2,000 ton shearing force require ment over a distance of 6 inches, as in the foregoing examples, the press frame only has to have a strength to resist 8,000 tons of force as compared to the 21,000 tons of force which must be resisted by the frame of a conventional slider-crank shearing press. Such a progressive shearing press, however, is extremely complex and expensive. In this respect, an additional mechanism must be provided to accurately reposition the slidercrank mechanism and slide after each short shearing stroke. Since the slider-crank mechanism and the slider are relatively heavy, and since shearing must be accomplished in a short time, the mechanism for adjusting these components must be capable of exerting a high force in a short period of time to reposition the slidercrank mechanism. Further, the positioning mechanism must also provide a strong locking action for the slidercrank mechanism during a shearing operation. Thus, the additional mechanism required is expensive and introduces maintenance problems. Still further, it will be appreciated that requiring several strokes of the slide mechanism to achieve billet shearing may result in an increase in shearing time required and, consequently a loss in production rate.

The present invention provides a mechanically driven billet or slab shearing press which overcomes the disadvantages of shearing presses heretofore known including the disadvantages pointed out in particular above. Thus, a shearing press is provided by the present invention in which the crank torque and frame strength requirements are greatly reduced to in turn achieve a reduction in weight, size and cost of the shearing press. Further, the press of the present invention employs a minimum number of moving parts, thus to reduce manufacturing and maintenance costs, and to increase reliability of operation.

The foregoing advantages of the shearing press of the present invention are achieved by employing a mechanical linkage drive arrangement wherein three links are used to convert rotary crank motion to linear slide motion, rather than a single link as is employed in conventional slider-crank shears. The particular linkage arrangement of the present invention provides for reducing the peak crank torque by a factor of 3 in a shearing press having a 2,000 ton force requirement. Further, frame strength is greatly reduced and, in this respect, in a shearing press having a 2,000 ton force requirement, the frame strength is reduced by a factor of 2.23 relative to the corresponding slider-crank shearing press. More particularly, a 2,000 ton shear made in accordance with the present invention requires a frame strength of 9,400 tons as compared to the strength of 21,000 tons required for the corresponding slidercrank shearing press.

In accordance with the present invention there is provided a billet or slab shearing press capable of shearing metal billets or slabs which are in a range of approximately 2 to 6 inches in thickness. The press includes a frame and a slide member carried by the frame and mounted for reciprocable movement between first and second positions. The slide, of course, carries a shearing blade. Drive means are provided for reciprocating the slide member between the first and second positions. The drive means includes a first link member having first and second end portions and having its first end portion pivotally connected to the slide member. A crank member is rotatably mounted in the frame and a connecting link member has opposite end portions pivotally connected to the crank and the second end portion of the first link member. A constraining link member has one end portion pivotally connected to the frame and another end portion pivotally connected to the connecting link member intermediate the end portions of the connecting link member. During each cycle of operation the crank rotates 360 and the pivot axis defined by the connection between the second end portion of the first link member and the connecting link member travels along a path referred to herein as a coupler curve or coupler path. The coupler path is indicative of the movement of the slide as a function of time and is defined by the lengths and arrangement of the several links of the system.

It is a primary object of the present invention to provide a billet or slab shearing press and mechanical drive arrangement which permits highly desirable kinematic and dynamic slide characteristics to be obtained.

Another object is the provision of a press having a mechanical drive arrangement which is especially suited for slab and billet shearing presses and which creates a slowdown of the slide during the shearing or working portion of the slide stroke.

Yet another object of the present invention is the provision of a mechanically driven billet and slab shearing press capable of producing substantially constant slide forces throughout the working or shearing portion of the stroke for a constant crank torque.

A further object is the provision of a billet and slab shearing press having a relatively low crank input torque requirement during the shearing or working portion of the stroke, whereby a clutch having a relatively low torque rating may be used.

Still a further object of the present invention is the provision of a slab and billet shearing press which, when compared to prior mechanically driven presses, permits use of a lighter weight frame construction.

Still another object of the present invention is the provision of a linkage system for driving a billet and slab shearing press, which system generates a coupler curve disposed with respect to the slide path in a manner to reduce side thrust, reduce slide speed during the shearing portion of the slide stroke and increase the slide speed during advance and return portions of the slide stroke.

A further object is the provision of a billet and slab shearing press which will considerably reduce the cost of such presses operating at current production rates.

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

FIG. I is an end elevation view, partially in section, of a shearing press made in accordance with the present invention;

FIG. 2 is a side elevation, partially in section, of the press of FIG. I, the section being taken along line 2-2 in FIG. ll;

FIG. 3 is a detail plan view of the drive system for the press of FIGS. l and 2;

FIGS. 44' are schematic diagrams illustrating the drive linkage at various points in a complete cycle of the embodiment shown in FIGS. l-3; and

FIG. 8 is a schematic illustration of the coupler curve generated by a drive linkage arrangement within the present invention.

Referring now to the drawings wherein the showings are for the purpose of illustrating a preferred embodiment of the invention, and not for the purpose of limiting the same, FIGS. 1 and 2 illustrate a slab or billet press formed in accordance with the present invention. The press shown is comprised of a vertically extending main frame assembly A, a drive unit B, and a mechanical drive linkage assembly C.

Frame assembly A could be of a variety of constructions. In the preferred embodiment, however, it is shown as a vertical straight-side steel frame. The frame is comprised of a base or bed portion 10, two uprights l2, and a crown section 14. Base 10 rests on a floor or the like and is suitably secured thereto. Uprights 12 ex tend vertically from base I0 and are joined at their upper ends by crown portion M.

Base T0 is provided with run out table guide means 16 and wear blocks or bolster plates 18 for transfer and support of a workpiece W which is fed through the press for shearing thereof into suitable lengths. The bed portion of the frame may also be provided with suitable means indicated generally at 20 for aligning slabs fed through the press. Mechanical work transfer equipment, not illustrated, is disposed in bed cavities which further provide for scrap removal from the press.

Uprights 12 are positioned so as to define the slideway or path of movement for the press slide 22. The uprights are provided with suitable gibbing 24 for guiding slide 22. Slide 22 carries a removable shearing blade 26 having a cutting edge 28 which is inclined relative to the press bed. The bottom surface of slide 22 may be provided with a pair of spring loaded hold-down devices 30 which normally project below the cutting edge 28 of blade 26 to engage workpiece W and hold the latter against wear block 18 during shearing of the workpiece by blade 26. It will be appreciated that a pneumatically or hydraulically actuated hold-down means may be employed rather than the spring loaded devices shown. I

Crown M houses the main drive unit 8 and mechanical drive linkage assembly C. The main drive assembly is best illustrated in FIG. 3 and is comprised of a main motor 32, flywheel M and a gear train leading to main drive pinion 356 and main drive gear 38. Motor 32 is mounted on frame A by means of a suitable support platform 40 extending from one side of the frame. Motor 32 is a standard industrial AC or DC motor and is operable to rotate flywheel 34 by belt or chain drive means 42 which extend around flywheel 34 and between the flywheel and a pulley device 44 attached to the motor shaft. The first gear 46 of the gear train leading to the main drive pinion 36 is keyed or otherwise secured to a shaft 48 which is suitably supported for rotation relative to the frame. Flywheel 34 is mounted on one end of shaft 48 in a manner whereby the flywheel is rotatable relative to the shaft, and clutch means 56 is provided for selectively interconnecting flywheel 34 with shaft 48 so that the shaft is driven by the flywheel. It will be appreciated, however, that the press may be continuously driven as opposed to being selectively intermittently driven and that for the purpose of a continuous drive arrangement flywheel 34 may be keyed or otherwise secured to shaft 46 and the clutch mechanism eliminated.

Pinion 46 is in driving engagement with a gear 52 which is keyed or otherwise secured to a shaft 54 having its opposite ends suitably supported for rotation of the shaft relative to the frame. A second pinion 56 is also keyed to shaft 54 and thus rotates at the same rotational speed as gear 52. Pinion 56 is in driving engagement with a second gear 58 which is keyed or otherwise secured to a shaft 60 which is also suitably supported For rotation relative to the press frame. Main drive pinion 36 is keyed or otherwise secured to shaft 66 for rotation therewith and, accordingly, at a rotational speed which is the same as that of gear 56. Suitable brake means 62 is associated with one end of shaft 60 and is actuable by means not illustrated to apply a breaking force to shaft 60 and thus to the main drive pinion 36. Main drive pinion 36 is in driving engagement with main drive gear 38. Gear 38 is keyed or otherwise secured to shaft means 64, which shaft means is suitably supported for rotation relative to the press frame. it will be appreciated that, through the gear train, main drive gear 38 is rotated at a speed considerably less than the speed of the first pinion 46 of the gear train. The reduction in speed will, of course, depend on the particular gear ratio employed.

Shaft 64 of main drive gear 3 also defines the shaft of a crank member which is included in the mechanical drive linkage assembly C. The crank member includes offset shaft portions 66 rigidly interconnected with shaft 64 for rotation therewith by crank amrs 6. Drive linkage assembly C is further comprised of twin linkage assemblies disposed on opposite sides of main gear 36 and driven simultaneously by the crank member. Each of the linkage assemblies includes a pair of connecting links 70. Corresponding ends of the links 76 are pivotally interconnected with the corresponding shaft 66 of the crank member. Suitable sleeve bearings 72 are interposed between links 76 and shafts 66.. The other ends of connecting links 70 are pivotally interconnected with corresponding slide member links 74. in this respect, each of the slide members 74 is provided with a shaft 76 extending through the corresponding connecting links 70. Suitable bearihg means 76 are interposed between each of the links 70 and the corresponding shaft 76. The lower end of each of the slide member links 74 is pivotally interconnected with slide member 22 by means of pins 86 attached to the slide links and connector assemblies 82 attached to the top surface of slide member 22. Each. connector assembly is comprised of a housing in which pins 84 of links 74 are disposed, and block means 86 and 88 having opposed cooperating recesses which. receive pins 80 and support the pins for pivotal movement relative to slide member 22. Blocks S6 and 66 are separable and are suitably interconnected with one another and with housing 64.

Each of the linkage assemblies further includes a pair of constraining links 96. Links 96 of each pair have one end thereof interconnected by shaft means 92. Each shaft 92 is supported for rotation relative to the press frame by corresponding support means 94 through which the shaft extends. Suitable bearings 26 are interposed between shafts 92 and support means 94. The other ends of each pair of constraining links 9 6' are pivotally interconnected with an adjacent one of the corresponding pair of connecting links 76. in this respect, constraining links 66 are provided with axially aligned pins S d which project through openings in the adjacent connecting link 76. Sleeve bearings 1166 or the like are interposed between each of the pins 96 and a corresponding connecting link 76. it is to be noted that constraining linlm 96 are pivotally interconnected with connecting links 76 at a point intermediate the opposite ends of the connecting links, which point is closer to the end of the connecting link attached to the slide link than to the end thereof connected to the crank member. it will be further noted that the pivot axis of constraining links W defined by shafts 92 are disposed above and in vertical alignment with the axis of rotation of the crank member defined by shaft 64.

ln operation of the apparatus thus far described, the crank member is rotated counterclockwise, as viewed in MG. 2, whereby offset shaft portion 66 thereof is rotated about the crank axis defined by shaft 64. Movement of shaft 66 imparts movement to connecting link 76 which in turn imparts movement to slide link 74 attached thereto. The movement imparted to slide link 74 causes reciprocation of slide member 22 vertically relative to the press frame, and the speed at which the slide moves in response to constant speed rotation of the crank member is controlled by constraining links Wl. This movement is discussed in more detail hereinafter.

The movement of slide member 22 through the entire stroke thereof will be more readily understood by referring to the schematic diagrams of lF'llGS. 4-7 which diagramatically illustrate the link assembly as it passes through one complete cycle. The various link elements of these figures are identified with the same reference numerals used to identify the corresponding elements in FIGS. 1-3. Since the elements are shown only diagrammatically, however, a prime sufiix is added to avoid confusion.

FM). 4 illustrates the linkage when the slide connection point e is at the bottom dead center position thereof. The crank is rotating in the counterclockwise direction and with the particular linkage position shown, the bottom dead center for point e occurs when the crank is approximately 320 of rotation from vertical. As the crank continues to rotate counterclockwise, the linkage is displaced such that slide connection point e is elevated. en the crank reaches the point corresponding to approximately of rotation from vertical, the linkage is disposed in the position illustrated in FIG. 5 wherein the slide point e is at top dead center.

Continued counterclockwise rotation of the crank causes slide point e to descend, which descent continues until slide point e again reaches the bottom dead center position thereof corresponding to approximately 320 of rotation of the crank. During this movement of slide point e, the linkage sequentially passes through the dispositions thereof illustrated in FIGS. 6 and 7.

During rotation of the crank member counterclockwise from the position of approximately 185 to the position of approximately 15 from vertical, slide connection point e moves downwardly through the work portion of the total stroke of the press, then upwardly. The work portion of the total stroke includes a shearing portion defined by downward movement of the slide connection point e from the beginning of the work stroke to bottom dead center of the slide. During rotation of the crank member counterclockwise from the position of approximately 15 to the position of approximately 185, slide connection point e is moved upwardly and downwardly through the return and advance portions of the total stroke of the press.

In the embodiment under consideration, the shearing press is designed to have a total stroke of 16.0 inches and a working stroke of 6.0 inches against a constant load resistance of 2,000 tons during the shearing stroke. To achieve this, the preferred lengths for the various links of the drive assembly are as follows:

member 68' (crank throw) to a distance 9.25 inches member 90" (constraining link) distance b-d 27.75 inches member 70 (connecting link) distance 0-0 18.50 inches member 70' (connecting link) distance a-b 14.00 inches member 70 (connecting link) distance b-c 4.50 inches member 74' (slide link) distance c-e 34.50 inches 0-d distance 27.50 inches distance g l 1.10 inches A shearing press having the above link length relationships is adapted to operate at a rate of 4 strokes per minute. The maximum torque required over the 6 inch working stroke equals approximately 6,000 ton-inches. Further, the press frame needs to have a resistance without yielding of only 9,400 tons. These particular requirements are considerably lower than those of a conventional slider-crank mechanically driven shearing press capable of performing the same functions. In this respect, such a slider-crank driven press would require 3 times greater torque and a frame 2.23 times stronger. To more fully appreciate the operation of the present invention, reference is made to FIG. 8 which is a schematic representation of a drive system proportioned in accordance with the foregoing dimensional relationships within the linkage system. To facilitate the discussion of FIG. 8 the various links and axes are identified in accordance with the following legend:

First link link 74 Second link link 70' Third link link 90' Crank member crank 68' First axis axis e Second axis axis 0 Third axis axis a Fourth axis axis 0 Fifth axis axis d Sixth axis axis b Referring now to FIG. 8, as crank 68' rotates counterclockwise about axis 0, the fourth axis c moves along a path or coupler curve which defines the movement of the first axis 2 along the slide path f. The configuration of the coupler curve is definitive of the fact that the slide moves rapidly in approaching the working stroke, moves slowly through the shearing portion of the working stroke and then rapidly toward the top dead center position thereof. Rotation of crank member 68 360 imparts reciprocating motion to the slide member to move the latter through one full stroke thereof.

As carnk 68 rotates counterclockwise, third axis a is rotated in a circular path about the second or crank axis 0. The path of third axis a is illustrated in increments of 20 starting at a point vertically above crank axis 0. The movement of fourth axis c along the coupler curve in response to rotation of crank member 68' is illustrated in increments corresponding to the 20 increments of rotation of third axis 0. Further, the reciprocating movement of the slide connection point or first axis e through the full stroke thereof as dictated by the coupler curve is illustrated in increments corresponding to the increments of rotation of third axis a. In order to more clearly demonstrate the controlled movement of the shear slide through the working portion of the stroke, the positions of first axis e at the beginning and end of the work stroke are designated W and W respectively. The beginning and ending points of the work stroke are also identified W and W respectively, on the coupler curve and the path generated by third axis a as it rotates about crank axis 0.

With the foregoing designations in mind, it will be seen that the working stroke begins at approximately of crank rotation counterclockwise from vertical and ends at approximately 15 of crank rotation counterclockwise from vertical. Thus, the working stroke requires rotation of the crank approximately Counterclockwise rotation of the crank from the point corresponding to bottom dead center of the slide to point W corresponding to the beginning of the next working stroke moves slide connection point e upwardly through the top dead center position thereof and thence downwardly towards point W, at which the next working stroke begins. The latter movement of slide connection point e defines the return portion of the work stroke and the return and advance portions of the total stroke.

The spacing of the indicator points along the coupler curve is indicative of the velocity of fourth axis 0 during one complete cycle of rotation of crank member 68. Similarly, the indicator points illustrated along slide path f are indicative of the velocity of slide connection point or first axis e during one full working stroke thereof. The velocity of first axis e is, of course, equal to that of the slide member and the shearing blade carried thereby.

As is illustrated in FIG. 8, the top dead center position of first axis e occurs at approximately 70 of crank rotation from vertical. Further, the bottom dead center position of first axis e occurs at approximately 320 of crank rotation from vertical. The positions of first axis e during movement of the latter from the top dead center position to the bottom dead center position thereof are designated by the indicators appearing on the left side of slide path f. The positions of first axis e during movement thereof from the bottom dead center to the top dead center position thereof are designatedby the indicators appearing on the right side of the slide path. f. The crank member is rotated at constant speed and, therefore, the spacings between the indicators are indicative of the velocity of axis e and thus the shearing slide member. In this respect, larger spacings between two given indicators is indicative of higher slide velocity than smaller spacings.

Considering a full stroke of first axis e beginning at the top dead center position thereof, it will be seen that after axis e passes through the top dead center position the velocity thereof as it approaches the beginning of the working stroke W is relatively high. As axis e enters the shearing portion of the work stroke, however, the velocity thereof is considerably reduced and continues to be reduced as axis e moves through the bottom dead center position thereof. After passing through the bottom dead center position the velocity of axis e is again increased considerably as it moves upwardly toward the top dead center position thereof. It is to be noted that the velocity of first axis e increases substantially immediately upon passing through the bottom dead center position thereof, whereby the return stroke of axis e and the slide member is at an extremely high velocity compared to the velocity during the shearing portion of the working stroke, and is at a velocity which is higher than the velocity thereof during the portion of the stroke in which the axis is advancing toward the shearing or work stroke.

The above velocities are imparted to first axis 2 as a result of the particular lengths, arrangement and cooperation of the various links of the linkage drive assembly. In this respect, as previously explained, crank 68' rotates about axis which is laterally offset from slide path f. Constraining link 90' pivots about an axis d in a downward arcuate path. Axis 0 of the crank and axis d of the constraining link are in vertical alignment, whereby the constraining link is driven along a path substantially corresponding to the coupler curve of the linkage system. Connecting link 70, by being connected to crank 68 is thrust to and fro laterally in response to rotation of the crank. The constraining function of link 90' on connecting link 70' resulting from their interconnection at axis b provides an efficient motion and force transmission operation for slide link 74'. Because of link dimensions and orientation during the work stroke, high force can be generated some distance from point e through use of nominal crank torque.

It will be further noted with regard to FIG. 8 that the coupler curve is divided into first and second sections by the intersection of slide path f therewith. The first section corresponds substantially to movement of first axis e through the shearing portion of the work stroke defined by movement of axis e along the coupler path from point W to bottom dead center, and the second portion corresponds substantially to the movement of first axis e from the bottom dead center position thereof through the top dead center position thereof and back toward the point W,, at which the shearing portion of the work stroke begins. Accordingly, the second section corresponds to the faster motion of the first axis and shearing slide, while the first section corresponds to the slower movement of the first axis and slide. The disposition of the coupler curve with respect to slide path f is also indicative of side thrust imposed on the slide during movement thereof through the total stroke. In this respect, when fourth axis 0 is disposed in the second section of the coupler curve at a point corresponding substantially to the top dead center position of first axis e, side thrust is imposed on the slide due to the angular relationship between slide links 74' and slide path f. On the other hand, when fourth axis 0 is either in the position corresponding substantially to the bottom dead center position of first axis e, or the posi tion indicative of the beginning of the shearing stroke, slide link 74' is in substantial alignment with slide path f, whereby little or no side thrust is imposed on the slide. Moreover, it will be appreciated that as fourth axis 0 moves through the first section of the coupler curve corresponding generally to the shearing stroke that very little side thrust is imposed on the slide during the shearing operation. In fact, in the particular embodiment of shearing press being considered, the maximum side thrust generated by the linkage during the shearing stroke is 111 tons. This figure is many times below the maximum side thrust which would be realized with a slider-crank press having the 2,000 ton working force.

Another characteristic of the linkage system of the present invention is the angle q between connecting link and constraining link Angle q is referred to as the transmission angle of the lever system. When the transmission angle is approximately 90 the most efficient force transmission takes place. When the angle is relatively small, the most efficient motion transmission is achieved. As illustrated in FIGS. 4-7, the transmission angle q is the greatest when the shear slide is moving through the working stroke toward the bottom dead center position. This, of course, is when the best force transmission characteristics are required. The transmission angle q is the smallest when the fourth axis 0 is moving through the second section of the coupler curve which is indicative of movement of the shear slide toward the top dead center position. This, of course, is when the most efficient motion transmission is desired. in accordance with the preferred embodiment of the present invention, the transmission angle q should not be less than 40 in order to assure smooth linkage operation.

As many possible embodiments of the present invention may be made and as many possible changes may be made in the embodiment herein set forth, it is to be distinctly understod that the foregoing description is to be interpreted merely as illustrative of the present invention and not as a limitation thereof.

We claim:

1. in a billet or slab shearing press comprising a frame, a slide member carried by said frame for reciprocable movement along a generally straight slide path between first and second positions defining a total stroke for said slide member including advance, work and return portions, said work portion including a shearing portion, said slide carrying a shearing blade, and a linkage drive means for reciprocating said slide member and shearing blade between said first and second positions with a predetermined velocitytime relationship for said blade during the shearing portion of the work portion of the total stroke to sever a slab or billet or substantial thickness disposed across said slide path and beneath said blade, the improvement comprising: said linkage drive means including slide link means having first and second end portions, said first end portion being pivotally connected to said slide member at a first axis, a crank member rotatably mounted on said frame at a second axis laterally spaced from said slide path in a first direction, connecting link means pivotally connected to said crank member at a third axis spaced from said second axis and pivotally connected to said second end portion of said slide link means at a fourth axis, and constraining link means having one end pivotally connected to said frame at a fifth axis and a second end pivotally connected to said connecting link means at a sixth axis between said third and fourth axes, said fifth axis being spaced above and in vertical alignment with said second axis, said second axis being spaced from said slide path a distance greater than the distance between said second and third axes, the distance between said third and fourth axes being approximately two times the distance between said second and third axes, the distance between said fifth and sixth axes being approximately three times the distance between said second and third axes, the distance between said third and sixth axes being approximately three times the distance between said fourth and sixth axes, said fourth axis defining an elongated generally elliptical coupler path when said crank member is rotated 360 about said second axis at substantially constant speed, said coupler path having major and minor axes intersecting in a point spaced laterally of said slide path in the direction opposite said first direction, said slide path intersecting said coupler path to define a first section of said coupler path on the side of said slide path in said first direction and generally indicative of low velocity movement of said slide member through the shearing portion of the work portion of the total stroke and a second section generally indicative of high velocity movement of said slide member through said return and advance portions of the total stroke.

Patent Citations
Cited PatentFiling datePublication dateApplicantTitle
US1569569 *Nov 8, 1924Jan 12, 1926Henry PelsMetal-cutting machine
US3158057 *Nov 20, 1961Nov 24, 1964Interstate Bakeries CorpGuillotine type cinnamon roll cutter having epicycle gearing means connected to the cutter blade
US3766771 *Jun 24, 1970Oct 23, 1973Gulf & Western Ind Prod CoPress and drive mechanism therefor
Referenced by
Citing PatentFiling datePublication dateApplicantTitle
US4013003 *Jan 19, 1976Mar 22, 1977L. Schuler GmbhPress with toggle joint drive mechanisms
US4031792 *Jul 12, 1976Jun 28, 1977Alpha Industries, Inc.Hammer ram press
US4257295 *Feb 26, 1979Mar 24, 1981Eubanks Engineering Co.Multiple purpose cutter apparatus
US5095618 *Apr 12, 1990Mar 17, 1992Utica Enterprises, Inc.Mechanical force enhancer
US5125307 *Nov 13, 1990Jun 30, 1992Emhart Inc.Cropping mechanism for surface mount placement machine
US5839338 *Sep 25, 1996Nov 24, 1998Tcholakov; Stoil MetodievWire steel rope cutter machine
US5919015 *Apr 16, 1997Jul 6, 1999Utica Enterprises, Inc.Mechanical drive for a blind spline broaching machine
US6080097 *Dec 8, 1997Jun 27, 2000Ranpak Corp.Cushioning conversion machine with single feed/cut motor
US6176818Dec 11, 1998Jan 23, 2001Ranpak Corp.Cushioning conversion machine cushioning conversion method and method of assembling a cushioning conversion machine
US6311596Aug 20, 1993Nov 6, 2001Ranpak Corp.Cutting assembly for a cushioning conversion machine
US6826980 *Nov 5, 2001Dec 7, 2004George SchmidegDrive and control systems for high speed intermittent motion generations, control and applications
US7022059 *Jul 3, 2001Apr 4, 2006Tokyo Kikai Seisakusho, Ltd.Chopper folder for rotary press
US7062949 *Dec 19, 2003Jun 20, 2006Haulsee Donald RMetal forming press having straight line drive mechanism
US7721636 *Jun 19, 2007May 25, 2010Sdi CorporationHole puncher
EP0356564A1 *Sep 2, 1988Mar 7, 1990Gräbener Pressensysteme GmbH & Co. KGMechanical press with double crank drive
WO1998013176A1 *Jul 16, 1997Apr 2, 1998Tcholakov Stoil MWire steel rope cutter machine
Classifications
U.S. Classification83/617, 100/282, 83/632, 83/626, 83/452, 83/635, 83/456
International ClassificationB30B1/00, B30B1/14
Cooperative ClassificationB30B1/14
European ClassificationB30B1/14
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 A DE CORP
Free format text: ASSIGNMENT OF ASSIGNORS INTEREST.;ASSIGNOR:GULF & WESTERN INDUSTRIES INC;REEL/FRAME:004204/0809
Dec 23, 1983AS02Assignment of assignor's interest
Owner name: E.W. BLISS COMPANY INC A DE CORP
Owner name: GULF & WESTERN INDUSTRIES INC :