US3766771A - Press and drive mechanism therefor - Google Patents

Press and drive mechanism therefor Download PDF

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Publication number
US3766771A
US3766771A US00049495A US3766771DA US3766771A US 3766771 A US3766771 A US 3766771A US 00049495 A US00049495 A US 00049495A US 3766771D A US3766771D A US 3766771DA US 3766771 A US3766771 A US 3766771A
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Prior art keywords
axis
slide
path
coupler
crank
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English (en)
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S Spachner
P Desai
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EW Bliss Co Inc
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Gulf & Western Ind Prod Co
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Assigned to GULF & WESTERN MANUFACTURING COMPANY reassignment GULF & WESTERN MANUFACTURING COMPANY ASSIGNMENT OF ASSIGNORS INTEREST. Assignors: GWM LIQUIDATING COMPANY, A DE CORP.
Assigned to GWM LIQUIDATING COMPANY reassignment GWM LIQUIDATING COMPANY CHANGE OF NAME (SEE DOCUMENT FOR DETAILS). EFFECTIVE JANUARY 5, 1976. Assignors: GULF & WESTERN MANUFACTURING COMPANY
Assigned to E.W. BLISS COMPANY, INC., A DE CORP. reassignment E.W. BLISS COMPANY, INC., A DE CORP. ASSIGNMENT OF ASSIGNORS INTEREST. Assignors: GULF & WESTERN MANUFACTURING COMPANY
Assigned to BARCLAYS AMERICAN/BUSINESS CREDIT, INC. reassignment BARCLAYS AMERICAN/BUSINESS CREDIT, INC. SECURITY INTEREST (SEE DOCUMENT FOR DETAILS). Assignors: E.W. BLISS COMPANY
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    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B21—MECHANICAL METAL-WORKING WITHOUT ESSENTIALLY REMOVING MATERIAL; PUNCHING METAL
    • B21D—WORKING OR PROCESSING OF SHEET METAL OR METAL TUBES, RODS OR PROFILES WITHOUT ESSENTIALLY REMOVING MATERIAL; PUNCHING METAL
    • B21D24/00—Special deep-drawing arrangements in, or in connection with, presses
    • B—PERFORMING OPERATIONS; TRANSPORTING
    • B30—PRESSES
    • B30B—PRESSES IN GENERAL
    • B30B1/00—Presses, using a press ram, characterised by the features of the drive therefor, pressure being transmitted directly, or through simple thrust or tension members only, to the press ram or platen
    • B30B1/10—Presses, using a press ram, characterised by the features of the drive therefor, pressure being transmitted directly, or through simple thrust or tension members only, to the press ram or platen by toggle mechanism
    • B30B1/14—Presses, using a press ram, characterised by the features of the drive therefor, pressure being transmitted directly, or through simple thrust or tension members only, to the press ram or platen by toggle mechanism operated by cams, eccentrics, or cranks

Definitions

  • the disclosure is directed to a long stroke extrusion press provided with an improved mechanical drive arrangement.
  • the drive arrangement includes a linkage having a first drive 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 various links are arranged to provide desirable kinematic and dynamic characteristics by developing a selected coupler curve at the pivot point between the first and second links.
  • the present invention is directed toward the press art and, more particularly, to an improved press and drive mechanism.
  • the invention is especially suited for use in high tonnage, comparatively long stroke extrusion presses and will be described with particular reference thereto; however, it will be appreciated that the invention is capable of broader applications and could be used in a variety of different types and sizes of presses.
  • a press of this type can develop only 35 to 40 percent of its rated tonnage at a point 6 inches above the bottom of its stroke when crank torque is at its rated maximum. Loading to greater values at this point will result in overloading the crank, and overloading the gear reduction train. If a clutch is used in the drive train, the clutch will also be overloaded. And, if the press is directly driven (i.e., no flywheel or other energy storing device is used between the motor and reduction gear train), the press motor will be overloaded as well. Consequently, to obtain the required working force of 200 tons at the top of the 6 inch working stroke, at least a 500 toncapacity press drive would have to be used. This would, of course, be extremely uneconomical since the press would be capable of developing much more than the required 200 tons throughout the major portion of its stroke.
  • the described press would be capable, at a point near the bottom of its stroke, of developing forces in the range approaching 400 percent of its rated tonnage.
  • the rule of thumb is that the frame must be capable of withstanding three times the rated press tonnage. Consequently, in the subject example, the person wishing to have a press capable of producing a 200 ton working stroke throughout a length of 6 inches would be forced to use a press drive rated at 500 tons and equipped with a 1,500 ton capacity frame. The undesirable nature of this is self-evident.
  • Point, or line contact mechanisms such as cams or rollers can develop such characteristics in a press.
  • Such mechanisms cannot be used for highload applications because of the extremely high pressures developed at contact points.
  • Hydraulically operated presses, cam driven presses and drag link presses have been developed to provide a slow-down function in a drawing press. Also, it has been suggested to use two separate drive trains selectively operated by separate friction clutches. Some of these prior presses included a crank directly above the slide so that the slide axis was through the pivot point of the crank. This concept resulted in a drastic swing of the pitman and a corresponding increase in the side thrust on the slide. In addition these concepts in press drive resulted in a toggle action at bottom dead center of the press which could generate high stresses on the crank pin and the crank journal structure at this slide position at nominal crank torque.
  • the subject invention provides a press and drive linkage which meets all of the above-noted requirements, has no load limitations, and permits'highly desirable kinematic and dynamic slide characteristics to be obtained. This is accomplished by a unique combination of linkages, a unique output path created by this linkage and an advantageous relationship of the output path to the path followed by the ram of the press. By using this invention a wider variety of drawing press concepts can be accomplished without substantial reduction of the total production rate.
  • a press which includes a frame and a slide member carried by the frame and mounted for reciprocal movement between first and second positions.
  • 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.
  • the press can have a large number of highly desirable characteristics generally defined by the output curve of the upper pivot axis of the link connected to the slide.
  • presses constructed in accordance with the invention can provide characteristics previously economically obtainable only in hydrualic presses, if at all.
  • the above described invention is further defined by the output curve which it generates at the pivot axis between the first link (connected to the slide) and the drive linkage associated therewith.
  • This curve defines the movement of the slide, or ram, as a function of time and it is generally called a coupler curve or coupler path.
  • the lever system generates a coupler curve or path which allows the ram to move rapidly toward and away from the bot tom dead center (BDC) of the ram except for the portion of the ram movement that accomplishes the drawing operation.
  • BDC bot tom dead center
  • the coupler curve or path creates a distinct slow-down in the velocity of the ram or slide.
  • the coupler curve has a unique shape and disposition with respect to the path of the slide driven thereby.
  • a primary object of the present invention is the provision of a 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 long stroke presses and creates a slow-down of the slide in the drawing portion of the press cycle.
  • a further object of the invention is the provision of a mechanically driven press capable of producing relatively uniform slide forces throughout a comparatively long working stroke for a constant crank'torque.
  • a further object is the provision of a press having relatively low crank input torque requirement during the working stroke, permitting use of a clutch of relatively low torque rating.
  • a still further object is the provision of a press which, when compared to prior mechanically driven press, permits use of a lighter weight frame construction.
  • Yet another object is the provision of a mechanical press which is capable of performing metal forming operations which in the past required the use of hydraulically driven presses.
  • Another object is the provision of a lever system for driving a power press, which system generates a coupler curve disposed with respect to the slide path in a manner to reduce side thrusts, reduce ram speed in the work portion of the press stroke and increase the speed of the ram between work portions of this stroke.
  • Still another object is the provision of a press of the general type described which will considerably reduce the cost of deep-drawing and extrusion presses operating at current production rates.
  • An additional object is the provision of a mechanically driven press which permits a slow working stroke and rapid return stroke.
  • FIG. 1 is a side elevation, having portions broken away, showing a preferred embodiment of a press formed in accordance with the present invention
  • FIG. 2 is a front elevation of the press of FIG. 1;
  • FIG. 3 is a cross-sectional view taken on line 33 of FIG. 1;
  • FIG. 4 is a front view of the drive linkage used in the preferred embodiment
  • FIGS. 5-8 are schematic diagrams showing the drive linkage at various points in a complete cycle of the embodiment shown in FIGS. 1-4; I
  • FIG. 9 is an enlarged schematic view illustrating the linkage shown in the prior figures and the coupler curve or path generated thereby;
  • FIG. 10 is an enlarged view illustrating the coupler curve or path created by the linkage system of the previous figures.
  • FIG. 11 is a schematic view illustrating the lever system in a position on the coupler curve wherein a maximum transmission angle occurs
  • FIG. 11a is a schematic view illustrating another operating characteristic of the present invention.
  • FIG. 12 is a schematic view showing the lever system on the coupler curve in a position wherein a minimum transmission angle occurs
  • FIG. 12a is a schematic view illustrating another operating characteristic of the present invention.
  • FIG. 13 is a chart illustrating certain operating characteristics of the present invention.
  • FIG. 14 is a chart illustrating other operating characteristics of the present invention.
  • FIG. 15 is a chart illustrating still further operating characteristics of the present invention.
  • FIG. 16 is a schematic view illustrating another embodiment of the present invention.
  • FIG. 17 is a representative device used for the purpose of mathematical analysis.
  • FIGS. 1 and 2 show the over-all arrangement of a long-stroke extrusion press formed in accordance with the 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 and configurations; however, in the preferred embodiment it is shown as a vertical frame formed from weldment components. Specifically, the frame comprises a main base unit which houses and supports a vertically positioned billet container 10a. Base unit 10 is mounted so its top surface is flush with the floor l1 and two pairs of generally rectangular box-like upright members 12 and 14 extend vertically therefrom. The uprights l2 and 14 are joined at their upper ends by a crown assembly 16.
  • the crown assembly 16 is a three-part structure formed from rectangular box sections 18, 20 and 22.
  • the crown assembly 16 supports the drive linkage assembly C and is arranged so that the parting lines between the sections 18, 20 and 22 coincide with the main pivot points of the drive linkage to facilitate construction and assembly.
  • the main structural components of the frame are joined by six vertically extending tie rods. As shown, pairs of large diameter tie rods 24 pass upward through each of the uprights 12. Single, smaller diameter tie rods 26 pass upward through the front uprights 14. Referring to FIGS. 1 and 3, it will be noted that the large diameter rods 24 are positioned an equal distance on opposite sides of the main drive linkage pivot points identified by reference letters D and 0. As will be apparent hereinafter, the tie rods 24 carry the main press reaction forces. The smaller diameter rods carry a minor portion of the reaction forces and function mainly to resist transverse slide reaction forces.
  • the tie rods 24, 26 are preferably shrink fitted during assembly of the frame. Additionally, the various mating surfaces between the press frame components are keyed at several points (such as shown at location 28 in FIGS. 1 and 2) so as to lock the frame rigidly in transverse directions.
  • the uprights l2 and 14 are positioned so as to define the slideway or path of movement for the main press slide or ram 114.
  • Conventional gibs l2 and 14' are provided generally at the inner corners of the uprights for guiding the slide through its vertical path.
  • the press is equipped with the usual billet handling and shear mechanisms, and the ad- DRIVE ASSEMBLY B
  • the press drive assembly B could, of course, be mounted on the frame assembly A, in the preferred embodiment, it is carried on a separate support frame or stand 30 which is positioned adjacent the press.
  • the main power source for the press comprises an electric motor 32. As shown, motor 32 is connected through a coupling 34 with a drive shaft 36 that supports a large diameter flywheel 38. The flywheel 38 and shaft 36 are rotatably supported in bearings or pillow blocks 40.
  • a conventional heavy duty, fluid actuated clutch 44 is mounted on flywheel 38 and, when actuated, functions to drivingly connect the flywheel with the main press drive shaft 42 through the shaft 46.
  • the shaft 46 is, as shown, suitably supported in pillow blocks 48 carried on frame 30.
  • the drive train for the press is carried in a two-part housing or gear box 52 which extends outwardly from the back of the press.
  • the drive shaft 42 extends transversely through the gear box 52 and has a conventional brake 50 mounted at its outer end. Suitable bearings carried along the parting line of the housing 52 support the drive shaft 42.
  • the drive train in the preferred embodiment is a gear train having a ratio of 150 to 1. As shown, it includes a first main pinion gear 54 keyed or otherwise positively connected to the drive shaft 42.
  • the pinion gear 54 drivingly engages a second gear 56 which is keyed to a horizontally extending shaft 58 carried in conventional hearings in the drive housing 52.
  • a second pair of pinion gears 60 are also keyed to shaft 58 and positioned on each side of the gear 56.
  • Gears 60 are drivingly engaged with two relatively large diameter gears 62 which are keyed to an elongated drive shaft 64 which extends transversely across the back of the press and outwardly on both sides thereof.
  • Two main drive pinions 66 are positioned at the outer ends of shaft 64.
  • the drive pinions 66 each engage a separate, large diameter bull gear 68.
  • the bull gears 68 each include an annular gear section 70 which is welded to two spaced plates 72 joined to a central hub 74.
  • the central hub 74 is keyed or otherwise positively connected to the outer ends of two stub shafts 76 carried inthe crown of the press.
  • the stub shafts 76 are mounted on the parting line between the crown sections 18 and 20 and are provided with suitable sleeve bearings 78.
  • the inner ends of stub shafts 76 are keyed to two link members 80 which carry a transversely extending wrist pin or shaft 82. This arrangement, in effect, forms a main drive crank having a throw equal to the distance between the center line 0 of stub shafts 76 and the center line a of wrist pin 82.
  • the linkage assembly includes a first link member 84 which is pivotally'connected to the pin.
  • the cylindrical portion 86 is provided with a slot 87 which receives link 84 and is connected thereto by a pair of pins 87a.
  • a pair of oscillating or constraining links 88' are pivotally mounted at one end to a horizontally extending shaft 92 which is carried in the top of the frame be-. tween the crown sections 20 and 22. It will be noted that the center of shaft 92 (identified with a reference letter D) is vertically aligned with the center 0 of the stub shafts 76. Additionally, a vertical line passing through these two centers (i.e., 0 andD) is parallel to the line or path of movement of slide 114.
  • the outer ends of the constraining links 88 are connected to the cylindrical end portion 86 of link 84 by split caps 90. As can be seen in FIG. 1, this provided a pivotal connection between links 88 and 84 with an effective center of pivot at the center b of cylindrical portion 86.
  • a relatively large diameter shaft or pivot pin 94 Extending transversely through the cylindrical portion 86 at a distance L from its center is a relatively large diameter shaft or pivot pin 94.
  • This pin extends outwardly on the opposite sides of the cylindrical portion 86 and is connected by caps 95 to the bifurcated upper end of the main slide drive link 98.
  • the lower end of the drive link 98 is, of course, pivotally connected to the slide assembly 104 which moves along a slide axis F, as shown in FIGS. -9.
  • this connection is provided by a pin 100 which extends between a pair of bracket members 102 which extend upwardly from a horizontal plate member 103 carried in the slide assembly 104.
  • the slide assembly is generally conventional, its construction should be briefly noted. As shown, the previously mentioned member 103 is adjustably connected to the slide 104 through a motor driven slide adjusting screw 106. This screw passes through a threaded opening in a member 108 received in the lower end of a sleeve 110 which forms a slideway or housing for member 103. Additionally, as can be seen, sleeve 110 is closed at its lower end by a plate 112 which carries the main ram or slide assembly 114.
  • the assembly 114 includes an outer housing 116 which encloses main pressure blocks 118 and 120.
  • the necessary extrusion ram 122 and piercer 123 is releasably connected to the lower plate 124 which is in turn positively connected to the housing 116.
  • FIGS. 5-8 These figures diagrammatically show the link assembly as it passes through one complete cycle.
  • the various elements of these figures are identified with the same reference numerals used to identify the corresponding elements in FIGS. l4 however, because the elements are shown only diagrammatically, a prime suffix is added to avoid confusion.
  • FIG. 5 shows the linkage when the slide connection point E is at the top dead center.
  • the crank is rotating in the clockwise direction and with the particular linkage layout shown, the top dead center occurs when the crank is at approximately 56rotation.
  • the press is designed to have a 131 inch total stroke and to be capable of a 36 inch working stroke against a constant load resistance of 3,600 tons (extrusion stroke) and a 41.75
  • 0 to a distance 5 ft. 0 to b distance equals 7.5 ft. b to C distance equals 2.5 ft. b to D distance equals 10 ft. C to E distance equals 15 ft. 0 to D distance equals 10 ft. 0 to F distance equals 8 ft. a to C distance equals 10 ft.
  • Member 84' a to b distance equals 16.53 in.
  • Member 84' b to C distance equals 7.27 in.
  • Member 88 b to D distance equals 28.33 in.
  • Member 98 C to E distance equals 42.72 in.
  • O to D distance equals 32.11 in. 0 to F distance equals 16.48 in. a to C distance equals 23.80 in.
  • the angle between a line from O to D and slide path F is 7, and the distance 0 to F is measured at a right angle to this line between axes O and D.
  • FIG. 9 is a schematic representation of a drive system proportioned in accordance with EXAMPLE II.
  • various lengths and axes are identified in accordance with the following legend:
  • a major axis through the coupler curve is defined as the longest distance between two points in the curve.
  • the minor axis is defined as the longest line between two points on the coupler curve and perpendicular to the major axis.
  • the minor and major axes intersect at a point P.
  • the second section of the coupler curve includes the largest portion of the major axes. This section is on the side of slide path F opposite to the crank 80. This second section relates to the faster motion of the slide and corresponds to the upper movement of the slide.
  • This rapid velocity is enhanced by offsetting the crank on the other or opposite side of the path F. In this manner, the offset crank creates a rapid motion when the axis C is in the second section. This rapid motion occurs during both movement of the slide toward and away from the top dead center (TDC).
  • the second section of the coupler curve has a vertical height m and a horizontal length n.
  • the height m is less than the length n.
  • the first section of the couplercurve has a height a and'a width b.
  • the height a is greater than the width b. Since the width b is a contributing factor to the side thrust of the slide during the power of drawing stroke, it should be maintained at a minimum.
  • the ratio of a to b' is generally above l.5:l.0.
  • the ratio of m to n is greater than 1.0 and preferably greater than about 1.5 11.0. This relationship of the two section dimensions of the coupler curve provides the rapid upswing and high power development in the downswing or working stroke with a minimum of side thrust.
  • the side thrust is generally in the range of less than 10 percent of the total amount of load created on the slide.
  • the major axis of the coupler curve intersects the slide path F at an obtuse angle k. This angle should be greater than about 110 and preferably greater than about 120 to assure that the coupler curve moves in a generally vertical inclined direction with respect to the path F.
  • the angle q shown in FIGS. 11 and 12 is called the transmission angle of the lever system.
  • the transmission angle q is the greatest when the ram or slide is moving toward the bottom dead center (BDC) position. This is when the best force transmission characteristics is required.
  • BDC bottom dead center
  • the angle q is the smallest when the fourth axis C is moving in the second section of the coupler curve. This is when the motion transmission is to be the most efficient in the illustrated embodiment.
  • the angle q should not decrease substantially below 40 to avoid high slide acceleration, and rapid changes of slide acceleration.
  • the coupler curve is illustrated as being divided into first and second sectionsby the slide path F, as previously described.
  • the crank 80 when the crank 80 overlies the second link 84, the fourth axis C is'well within the first section of the coupler curve. This causes the fourth axis-to move slowly in a vertical direction as it .moves through the bottom dead center (BDC), shown on FIG. 1.1a.
  • BDC bottom dead center
  • the overlying of the crank and the second link conditions the-fourth axis for a slow drawing operation. Indeed, this'fourth axis is well within the first section of the coupler curve and is immediately ready to commence its downswinginto the bottom dead center position.
  • FIG. 12a another characteristic of the-present invention is illustrated.
  • the crank 80 and'the third link 88 have two positions. in which they are parallel.
  • crank 80 rotates about an axis which is offset from the slide path F.
  • the link 88 swings about an upper axis D in a downward arcuate path.
  • the axis 0 of the crank is vertically below the pivot D of link 88 so that the link 88 is driven substantially along a path corresponding generally to the coupler curve of the lever system.
  • the major portion of the coupler curve as defined by the larger length of the major axis of one side or the other of slide path F, is on the side of the slide path away from the crank axis. It is also tilted upwardly and away from the crank axis and the slide member of the press.
  • the axis b is closer to the axis 0 than it is to the axis a. In this manner, an efficient force transmission is accomplished during the latter part of thedownward thrust acting upon theslide.
  • FIG. 13 there is shown a chart illustrating a comparison between the operating characteristics of the present invention and the operating characteristics of a press constructed in accordance with the prior art and-adapted for use in a drawing operation. Assuming that the maximum drawing speed for steel is approximately 90 ft/min, and the press is to have a 16 inch stroke with a drawing stroke of approximately 6 inches, during the last 6 inches of the power stroke, the velocity of the slide should not exceed approximately 90 ft/min.
  • the solid line shows a press constructed in accordance with the present invention.
  • the dashed line indicates a'press constructed in accordance with the prior art.
  • the crank speed is reduced to 21 rpm so that velocity in the last 6 inches will not exceed approximately 90 ft/min. It is noted that the up and down movement of the slide follows generally along the same velocity curve. This is not true with the present invention.
  • the crank can be increased to rpm and still not exceed generally ft/min in the drawing stroke.
  • a press constructed in accordance with the present invention has its slider, or ram, moving from bottom dead center (BDC) to top dead center (TDC), it circumscribes the upper curve a having a maximum velocity of approximately 500 ft/min.
  • the velocity curve of a press constructed in accordance with the present invention takes the form of curve b having a first section I wherein the velocity increases to approximately 240 ft/min and then decreases to approximately 90 ft/min at the 6 inch position. Thereafter, curve b goes into a second section II wherein the velocity generally fluctuates in the range of 90 ft/min. Thereafter, the velocity curve for a press constructed in accordance with the present invention gradually decreases from approximately 90 to ft/min to zero in the section of curve b labeled III. It can be readily seen by this curve that a press can be operated substantially more rapidly when incorporating the present invention.
  • the curve a represents a priorpress for creating a minimum tonnage of approximately 200 over the total working stroke of the 6 inch drawing operation. It is seen that the actual tonnage of such a press exceeds 1,700 tons at bottom dead center. If the prior press were reduced to a maximum tonnage of approximately 800 tons it would follow a tonnage curve b during the drawing stroke. It is seen that a major portion of this curve is substantially below 200 tons, the desired rating of this particular example. When a press is constructed in accordance with the present invention, the tonnage curve follows curve 0.
  • FIG. 15 another operating characteristic of the present invention is illustrated.
  • This chart compares the torque of the crank during the last'ap proximately 6 inches of the drawing. operation of a press constructed in accordance with the prior art as shown in curve a and a press constructed in accordance with the present invention as shown in curve b.
  • a peak torque of approximately 1,670 ton-inches is required .in a 200 ton-6 inch draw press so that'200 tons force can be created'when the slide is approximately at the6 inch location.
  • a peak torque of approximately 1,670 ton-inches is required .in a 200 ton-6 inch draw press so that'200 tons force can be created'when the slide is approximately at the6 inch location.
  • crank drive can be reduced in size and strength.
  • FIGS. 13-15 illustrate generalized operating characteristics of a press constructed in accordance with the present invention. Various other presses contructed in accordance with the present invention would vary these curves but not their general import.
  • slide member 200 is movable in a path F by the first link 202.
  • a second link 204 joins third link 206 with crank 208.
  • a coupler curve is illustrated as curve 210 having a major and minor axis, as illustrated. The numbers along the coupler curve correspond with numbers along the arc of the crank 208. The closer the corresponding points are located on the coupler curve the slower the slide velocity, since the numbers on the crank circle are equal increments of a uniformly rotated crank. As the space increases between points on the coupler curve, the slide velocity is increased.
  • First link 202 is connected between axis a and axis a.,.
  • the second link is connected between axis a and axis a Stationary axis a is the center of the crank 208,.and stationary axis a is the centerof the arc defined by the swinging third link 206.
  • Stationary axis a is the center of the crank 208,.and stationary axis a is the centerof the arc defined by the swinging third link 206.
  • Coupler curve illustrated in FIG. 16 has the general characteristics of the coupler curve defined in accordance with the Examples I and II. Also, the arrangement of the links although somewhat different in species is generically the same as that described in connection with Examples I and II.'
  • FIG. 17 a mathematical relationship drawing is illustrated to define points along a coupler curve 220 inrelation to two orthogonal axes having their origin at the center of the crank. It is not necessary to go into details of the various mathematical representations.
  • the crank is r, the second link is r the third link is r and the spacing between the two stationary axes O, C is r,.
  • the fourth axis is D which is coupled onto an appropriate first link, not shown.
  • a power press comprising a frame; a slide member carried by said frame for reciprocal movement along a generally straight slide path between first and second positions; 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 onto said crank at a third axis, spaced from said second 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 pivotally connected to said frame at a fifth axis and a second end pivotally connected to said second link at a sixth axis, said sixth axis being spaced from said third and fourth axes; said fourth axis defining an elongated generally elliptical couple
  • a power press comprising a frame; a slide member carried by said frame for reciprocal movement along a generally straight slide path between first and second positions; 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 firstaxis; a crank member rotatably mounted on said frame at a second axis; a second link member pivotally connected onto said crank at a third axis, spaced from said second axis and pivotallyconnected to the second end portion of said first link member at a fourth axis, and a third link member having one end pivotally connected to said frame at a fifth axis and a second end pivotally connected to said second link at a sixthaxis, said sixth axis being spaced from said third and fourth axes; said fourth axis defining an elongated' generally elliptical coupler
  • a power press comprising a frame; a slide member carried by said frame for reciprocal movement along a generally straight path between first and second positions; 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 onto said crank at a third axis, spaced from said second axis and pivotally connected to the second end portion of said first link member at a fourth axis, and a third link member havine one end pivotally connected to said frame at a fifth axis and a second end pivotally connected to said second link at a sixth axis, said sixth axis being spaced from said third and fourth axes; said fourth axis defining an elongated generally elliptical couple
  • X is a first orthogonal component of the curve with an origin at said second axis; Y second orthogonal component of the curve with an origin at said second axis; r is the length of said crank; e is the fixed distance between said third axis and said fourth axis; 0 is the variable angle of said crank with respect to said first component; 19 is the variable angle of the line between said third and sixth axes with respect to said first component; and w is the fixed angle between the line between said third and fourth axes and the line between said third and sixth axes; and, said coupler path having a major axis defined by the longest line between any two points in said coupler path and a minor axis defined by the longest line between two points in said coupler path and perpendicular to said major axis, said major and minor axes intersecting at a
  • a power press comprising a frame; a slide member carried by said frame for reciprocal movement along a generally straight slide path between a lowermost and an uppermost position; 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 onto said crank at a third axis, spaced from said second 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 pivotally connected to said frame at a fifth axis and a second end pivotally connected to said second link at a sixth axis, said sixth axis being spaced from said third and fourth axes; said fourth axis defining an elongated generally
  • said first section has a third dimension parallel to said slide path and a fourth dimension perpendicular to said slide path wherein said third dimension is greater than said fourth dimension.
  • said first section has a first dimension parallel to said slide path and a second dimension perpendicular to said slide path, with the ratio of said first dimension to said second dimension being at least 1.5 to l.
  • crank member and third link member have two parallel positions and at least one of said parallel positions occurs when said fourth axis is substantially spaced from both of said first and second ends of said major axis.
  • crank member and third link member have two parallel positions and at least one of said parallel positions occurs when said fourth axis is substantially spaced from both of said first and second ends of said coupler path.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Press Drives And Press Lines (AREA)
US00049495A 1970-06-24 1970-06-24 Press and drive mechanism therefor Expired - Lifetime US3766771A (en)

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US4949570A 1970-06-24 1970-06-24

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US3766771A true US3766771A (en) 1973-10-23

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US00049495A Expired - Lifetime US3766771A (en) 1970-06-24 1970-06-24 Press and drive mechanism therefor

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US (1) US3766771A (enExample)
JP (1) JPS5141946B1 (enExample)
CA (1) CA947532A (enExample)
DE (1) DE2131013C2 (enExample)
FR (1) FR2100820B1 (enExample)

Cited By (11)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3817139A (en) * 1973-01-29 1974-06-18 Gulf & Western Industries Press and drive mechanism therefor
DE2816429A1 (de) * 1977-08-04 1979-02-15 Gulf & Western Mfg Co Presse
EP0040965A3 (en) * 1980-05-27 1982-03-17 Danly Machine Corporation A power metal-forming press
US4590784A (en) * 1984-07-27 1986-05-27 Nippon Kokan Kabushiki Kaisha Press brake
US4773325A (en) * 1986-11-19 1988-09-27 Hiroyasu Shiokawa Machine press with cam type ram actuator
US5950531A (en) * 1997-05-28 1999-09-14 Apic Yamada Corporation Electric press machine
US7062949B1 (en) * 2002-12-20 2006-06-20 Haulsee Donald R Metal forming press having straight line drive mechanism
US20160107407A1 (en) * 2013-05-28 2016-04-21 Schuler Pressen Gmbh Method for controlling a press with a variable gear ratio
US9365007B2 (en) * 2014-04-09 2016-06-14 Metal Systems Technologies, LLC Four-bar press with increased stroke rate and reduced press size
CN106427007A (zh) * 2016-10-26 2017-02-22 广东工业大学 伺服机械压力机用双驱动肘杆机构及优化方法
CN106584898A (zh) * 2016-10-26 2017-04-26 广东工业大学 一种用于机械压力机的双驱动肘杆机构及优化方法

Families Citing this family (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE202013102487U1 (de) * 2013-06-11 2013-07-03 Karl Marbach Gmbh & Co. Kg Stanzmaschine mit reduzierter dynamischer Verformung durch optimierte Krafteinleitung

Citations (6)

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Publication number Priority date Publication date Assignee Title
GB410366A (en) * 1932-10-19 1934-05-17 Guido Hugo Goetz Improvements in and relating to crank presses
US1966953A (en) * 1931-08-04 1934-07-17 Tubus A G Extrusion press
US2546100A (en) * 1944-02-14 1951-03-20 Clearing Machine Corp Metalworking power press
US2696165A (en) * 1950-10-04 1954-12-07 Otto Hermann May Toggle press
US2781015A (en) * 1952-06-11 1957-02-12 Cleveland Crane Eng Draw press
US3315596A (en) * 1964-07-21 1967-04-25 May Pressenbau G M B H Press

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GB397558A (en) * 1931-12-31 1933-08-31 Frank Humphris Improvements in presses and other pressure-exerting machines
GB424591A (en) * 1933-09-01 1935-02-25 Massey Ltd B & S Improvements in presses and pressing mechanism

Patent Citations (6)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1966953A (en) * 1931-08-04 1934-07-17 Tubus A G Extrusion press
GB410366A (en) * 1932-10-19 1934-05-17 Guido Hugo Goetz Improvements in and relating to crank presses
US2546100A (en) * 1944-02-14 1951-03-20 Clearing Machine Corp Metalworking power press
US2696165A (en) * 1950-10-04 1954-12-07 Otto Hermann May Toggle press
US2781015A (en) * 1952-06-11 1957-02-12 Cleveland Crane Eng Draw press
US3315596A (en) * 1964-07-21 1967-04-25 May Pressenbau G M B H Press

Non-Patent Citations (1)

* Cited by examiner, † Cited by third party
Title
Industrieblatt Stuttqart, January 1959 Germany (Article by Von Dipl. ing. H. Rankers) Pages 17 20. *

Cited By (14)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3817139A (en) * 1973-01-29 1974-06-18 Gulf & Western Industries Press and drive mechanism therefor
DE2816429A1 (de) * 1977-08-04 1979-02-15 Gulf & Western Mfg Co Presse
DE2816429C2 (de) * 1977-08-04 1984-07-19 Gulf & Western Manufacturing Co., Southfield, Mich. Mechanischer Pressenantrieb
EP0040965A3 (en) * 1980-05-27 1982-03-17 Danly Machine Corporation A power metal-forming press
US4590784A (en) * 1984-07-27 1986-05-27 Nippon Kokan Kabushiki Kaisha Press brake
US4773325A (en) * 1986-11-19 1988-09-27 Hiroyasu Shiokawa Machine press with cam type ram actuator
US5950531A (en) * 1997-05-28 1999-09-14 Apic Yamada Corporation Electric press machine
US7062949B1 (en) * 2002-12-20 2006-06-20 Haulsee Donald R Metal forming press having straight line drive mechanism
US20160107407A1 (en) * 2013-05-28 2016-04-21 Schuler Pressen Gmbh Method for controlling a press with a variable gear ratio
US9365007B2 (en) * 2014-04-09 2016-06-14 Metal Systems Technologies, LLC Four-bar press with increased stroke rate and reduced press size
CN106427007A (zh) * 2016-10-26 2017-02-22 广东工业大学 伺服机械压力机用双驱动肘杆机构及优化方法
CN106584898A (zh) * 2016-10-26 2017-04-26 广东工业大学 一种用于机械压力机的双驱动肘杆机构及优化方法
CN106427007B (zh) * 2016-10-26 2018-04-06 广东工业大学 伺服机械压力机用双驱动肘杆机构及优化方法
CN106584898B (zh) * 2016-10-26 2018-05-01 广东工业大学 一种用于机械压力机的双驱动肘杆机构及优化方法

Also Published As

Publication number Publication date
CA947532A (en) 1974-05-21
DE2131013C2 (de) 1983-05-26
FR2100820A1 (enExample) 1972-03-24
JPS5141946B1 (enExample) 1976-11-12
DE2131013A1 (de) 1971-12-30
FR2100820B1 (enExample) 1978-07-28

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