US3857272A

US3857272A - Counterstroke hammer

Info

Publication number: US3857272A
Application number: US00369549A
Authority: US
Inventors: B Voitsekhovsky; V Nikolaev; G Shoikhet
Original assignee: Individual
Current assignee: Individual
Priority date: 1971-08-26
Filing date: 1973-06-13
Publication date: 1974-12-31
Anticipated expiration: 1991-12-31

Abstract

A counterstroke hammer in which two rams are moving towards one another until they come into contact along guides scured to a frame. The rams are accelerated by a drive provided with a plunger and a cylinder moving with respect to the frame and both rams in the direction of their movement. The drive cylinder is connected to one of the rams through tie-rods to effect simultaneous interaction when an acceleration force is imparted. The cylinder and its ram is then disconnected from the system in order to ensure the ram movement free from the action of the drive. The drive plunger is adapted to contact the other ram during the transmission of the acceleration force and is disconnected from said contact after the acceleration, for effecting the ram movement free from the drive action. The impact interaction of the rams takes place with the drive disconnected and hence the kinetic energy thus obtained is damped, while the kinetic energy of the cylinder and plunger is damped independently of the rams.

Description

United States Patent [1 1 Voitsekhovsky et a1.

1 1 COUNTERSTROKE HAMMER [76] Inventors: Bogdan Vyacheslavovich Voitsekhovsky, ulitsa Akademicheskaya, kottedzh, 2; Valentin Pavlovich Nikolaev, ulitsa Maltseva l, kv. 8; Grigory Yankelevich Shoikhet, ulitsa Pravdy, l, kv. 34, all of Novosibirsk, U.S.S.R.

[22] Filed: June 13, 1973 [21] Appl. No.: 369,549

[30] Foreign Application Priority Data June 13, 1972 U.S.S.R 796429 Aug. 26, 1971 U.S.S.R 687061 [52] US. Cl. 72/407, 1007264 [51] Int. Cl B2lj 9/12 [58] Field of Search 72/407, 453; 100/264, 269 R [56] References Cited UNITED STATES PATENTS 2,863,343 12/1958 Steinfort 100/264 3,036,538 5/1962 Ottestad.... 100/264 3,115,676 12/1963 Quartullo. 100/269 R 3,735,631 5/1973 Schmo11 72/453 FOREIGN PATENTS OR APPLICATIONS 647,910 7/1937 Germany 100/264 [451 Dec. 31, 1974 Primary Examiner-C. W. Lanham Assistant Examiner-Gene P. Crosby Attorney, Agent, or Firml-lolman & Stern [57] ABSTRACT A counterstroke hammer in which two rams are moving towards one another until they come into contact along guides secured' to a frame. The rams are accelerated by a drive provided with a plunger and a cylinder moving with respect to the frame and both rams in the direction of their movement. The drive cylinder is connected to one of the rams through tie-rods to effect simultaneous interaction when an' acceleration force is imparted. The cylinder and its ram is then disconnected from the system in order to ensure the ram movement free from the action of the drive. The drive plunger is adapted to contact the other ram during the transmission of the acceleration force and is disconnected from said contact after the acceleration, for effecting the ram movement free from the drive action. The impact interaction of the rams takes place with the drive disconnected and hence the kinetic energy thus obtained is damped, while the kinetic energy of the cylinder and plunger is damped independently of the rams.

12 Claims, 6 Drawing Figures PAIENTEDUEEBI I914 3. e57, 272

SHEET 3 OF 5 COUNTERSTROKE HAMMER BACKGROUND OF THE INVENTION The present invention relates to improvements in hammers, and particularly, to counterstroke hammers, such as are intended for stamping and pressing workpieces, especially when a high impact energy is required.

Known in the art are counterstrokehammers of di- 1 verse constructions. The term counterstroke hammer denotes a hammer wherein rams in applying force to a workpiece being processed move towards each other.

In known hammers both rams are accelerated by a common drive. One of the rams is secured to the frame, whereinto is built a hydraulic drive cylinder, while the other ram is made as a plunger travelling inside the cylinder.

In such hammers the equality of pulses of both rams before their interaction is preserved, which makes possible the reduction of loads imposed on the hammer foundation (bedplate).

A disadvantage of such a hammer is that the loads (including wave loads) arising during the mutual impact of both rams are taken up by the hammer frame, which essentially is a part having a complex shape, subject to bending and having stress acceptors. This limits the energetic capabilities of the hammer due to processing imperfections encountered in the manufacture of high-strength and largeweight forgings, such as said frame.

Well known is another construction of a counterstroke hammer, wherein both rams are provided with a common acceleration drive. In such a hammer the acceleration drive cylinder is rigidly connected to the hammer frame, while the cylinder plunger is connected to one of the rams. The other ram is suspended on flexible bands interconnecting said rams. The acceleration force of the drive in such a hammer is transmitted by one of the rams ina direct way, while to the other ram through said flexible bands passing via special rollers fixed on the frame in order to-change the application of force by l80.

A disadvantage of said hammers lies in the fact that in order to ensure equal pulses of both rams before the impact, their masses should be strictly equal, which condition is hardly realized when the hammer dies are changed frequently. Another disadvantage is low stability of the flexible bands, subjected to excessive dynamic loads, when the rams rebound after a mutual impact, and to multiple alternating loads due to rolling around the rollers. The resistance to wear and long service life of the bands cannotbe substantially increased due to the availability of complex shock-absorbing devices at points of fixing of the flexible bands.

And finally, known are hammers provided with a common acceleration drive of both rams, built-in directly into the bodies of said rams. The cylinder of such a drive is rigidly connected to one of the rams, while the rod is likewise connected to the other ram. These hammers ensure equal energy pulses of the rams before their mutual impact, and in addition, the frame is relieved of dynamic impact loads. However, a disadvantage of such hammers is that in order to effect the acceleration, the energy accumulator made as a compressed air cylinder is to be disposed in the cavity of one of the rams, wherein it is subjected to excessive dynamic loads. Besides, moreover, with such a construction of the hammer it is difficult to make the impact bodies of a simple shape, which is extremely imprortant for increasing the energy and hardness of blow.

In recent years hammers with individual acceleration drives for each of the rams found wide application. Such hammers usually comprise two cylinders built into the hammer frame. The cylinder cavities communicate regularly with the working medium source through shut-off devices. Due to rigid coupling of the rods or plungers with the rams the force is exerted on the rams through said rods (plungers).

A merit of such hammers is that the frame does not take up impact overloads during stamping jobs, since the pulses are damped at mutual impacts of the rams,

and the frame takes up the acceleration forces only. Such an advantage can be put into'practice only under the condition of preserving strict equality of pulses of both rams before they come into contact. However, the presence of individual acceleration drives of both rams hinders the observance of said condition, since asymmetry of hydraulic resistances of the drives fails to ensure the synchronous movement of the rams. In order to balance the energy pulses of the rams, special synchronizers are used, such as made in the form of screws disposed with their opposite ends into the rams and mechanically connecting said rams. Under such conditions the synchronizers are subjected to excessive dynamic loads, especially during rebounds of the rams, and also, the shape of the rams is complicated while the reliability is reduced due to said screws being liable to failure.

The constructions of high-energy hammers herein-' above described have a number of common disadvantages.

Since the acceleration force acting on the rams by the moment of their impact interaction is not relieved, they compress the stamped workpiece while it is in the hot condition, the time of contact of the hot workpiece with the surface of the die being determined by the speed with which the rams are engaged and the speed of the mechanism which forces the rams to occupy their extreme initial positions. In all high-energy impact hammers heretofore known this time constitutes about 1 sec, which substantially reduces the durability of dies.

A disadvantage of the known construction of ham mers is also the incorporation of rigid guides in order to preclude possible angular displacements owing to eccentric impacts of the rams. This is due to the fact that the plunger and cylinder rigidly connected to the rams do not allow for angular displacements owing to beak-tight sealings incorporated therein. Since under actual conditions the impact is always eccentric, the rams are subject to some rotary energy following the impact. This is caused by the fact that during the mutual interaction, especially during idle impacts the rams fail to compensate for the clearance in the guides, intended for effecting a reciprocating motion. A more rigid impact is understood to be such an impact whereat the damping of the kinetic energy takes place over a lesser section of the way during the mutual impact. Such an impact is the idle impact whereat no plastic deformations take place and the kinetic energy of the rams'is damped due to resilient deformation only during the mutual impact of the rams. The value of the rotary energy is proportional to the square of the eccentricity of application of the impact pulse, while this energy is damped during the interaction of the rams and guides. The more rigid are the guides, the more is the force developed during said interaction, the more is the wear of the guides, which makes the operation of the hammers considerably more expensive. One more disadvantage is caused by the rigid connection of the acceleration drive parts (plunger or cylinder) to the rams. It is known that during the mutual impact of the rams stress waves of considerable intensity propagate in the rams. At points of transition from one crosssection to the other, such as from the rod or cylinder to the ram, the stresses increase many fold due to the impression of straight and reflected waves, These disadvantages become especially evident when the rigidity and energy of the impact are increased.

SUMMARY OF THE INVENTION It is an object of the present invention to provide a counterstroke hammer wherein the impact energy and simultaneously the rigidity, are increased.

Another object of the invention is to increase the reliability of the hammer operation.

A further object of the present invention is to provide an acceleration drive of a simpler construction.

Among other objects there shall be noted the isolation of acceleration drive of the hammer rams from shock-wave and inertia loads arising during the mutual impact of the rams.

Yet another object of the invention is to reduce wear on dies by reducing to a minimum the time the two rams are in contact with each other.

Still another object of the invention is to provide higher eccentricity during the mutual impact and hence to extend the impact processing possibilities of various workpieces.

A further object of the invention is to increase the resistance to wear of the ram guides due to reducing considerably the interaction of rams and guides when the rotational kinetic energy is damped.

Yet another object of the invention is to reduce the effect of impact forces on the frame foundation.

In addition, an object of the invention is to simplify the shape of the rams.

These and other objects are attained in a counterstroke hammer, comprising a frame, two rams moving along guides mounted on the frame until they come into interaction in the zone of processing a workpiece and return to their initial positions, an acceleration drive for the rams made as a pair of reciprocating moving parts cylinder-plunger, and a mechanism for returning the rams after their stroke to their initial positions, wherein, according to the invention, both parts of the drive are adapted to move relative to the frame and both rams in the direction of the rams movement, while the value of mutual relative displacement of both parts of the drive is less than the total distance travelled by the rams from their initial position until they come into interaction, and one of the rams is connected to one of said moving parts of the drive through tie-rods made to interact simultaneously with said parts at the instant of imparting the acceleration force to said ram and subsequent disconnecting the ram-drive moving part system in order to ensure the ram movement free from the drive action following the acceleration, the other moving part of the drive being adapted to contact the other ram at the instant of imparting .the acceleration force and adapted to be disconnected from said ram after the acceleration, resulting in that the impact interaction of the rams is effected with the drive being disconnected, the kinetic energy obtained during their acceleration being damped, while the kinetic energy of both parts of the drive being damped during their interaction independently of the rams.

An advantage of the construction hereinabove de scribed is that the rams in the course of their interaction during the time they are disconnected from the acceleration drive, being in a free movement, do not exert impact overloads on the elements of the drive. Both parts of the drive (the cylinder and the plunger group) interact independently of the rams. Accelerations, and hence dynamic overloads, exerted on both parts of the drive do not depend on the forces arising in the zone of processing the workpiece, but are determined only by the characteristics of the hydraulic drive (in case such is used in the system) and can be calculated beforehand. Under such conditions the reliability of operation increases sharply. It should be borne in mind that such a construction of the hammer makes possible to simplify the shape of the rams and to exclude sharp transitions from one cross-section to another wherein stresses develop, and hence, to increase the permissible energy and rigidity of the impact.

The construction hereinabove described is preferred also due to the fact that in the course of free movement of the rams the time they are in contact is reduced to a minimum and does not depend on the speed of action of the mechanisms which part the interacting bodies and is determined only by the energy of impact and resilient properties of these bodies. The rams will part after the impact like billiard balls, while the mechanisms bring the rams to their initial positions only,

in the construction hereinabove described each tierod can be made as a stem rigidly connected with one end to one of the rams and having a thicknened portion on the other end, the cylinder having an eye through which is passed said stem which transmits the acceleration force to the relevant ram in interacting with its thickened portion with the eye.

The eye is preferably connected to the cylinder through a rocker arm which rocks relative to the axis perpendicular to the direction of movement of the rams, the eye being articulated to the rocker arm.

The fixing point of the tie-rod to the rams should be preferably disposed adjacent to the plane passing across said ram through-the centre of gravity.

I It is advantageous to arrange both parts of the drive coaxially with the direction of movement of the rams and to articulate a long dowel on the drive plunger thrusting into the ram in order to transmit the acceleration force.

[n this case the ram accelerated by the plunger should have a tapered bore to accept the dowel, owing to which the acceleration force becomes applied in the direction of movement of the rams, free from eccentricity relative to its centre of gravity.

As the mechanisms for returning the rams to their initial positions, use can be made of hydraulic jacks known per se disposed in pairs symmetrically with each ram, the latter having lugs whereinto rods of the hydraulic cylinders are adapted to thrust. Said lugs should be disposed in a piane passing through the rams centre of gravity, and perpendicularly to the direction of its movement.

Most favourable conditions of operation of said tierods and long dowel are provided when the ratios of masses of the rams and drive parts are taken from where m and m are masses of the drive parts,

M and M are masses of the rams interacting with masses m and m respectively. Owing to such ratios of masses the relative speed at which the rams strike each other when they return to their initial positions following an interaction with the corresponding parts of the drive is minimal.

Each ram should be secured to two hinged supports, each support lying in a plane perpendicular to the direction of movement of the rams and disposed substantially medially of the distance travelled by the centre of gravity of the corresponding ram. The guides are twosupport cantilevered beams.

Owing to such a construction the beams become resilient, and taking into consideration that the rams are in a free flight being independent from the drive, torsional oscillations of the rams can be allowed, whose amplitudes substantially exceed those encountered in known hammers. Hence it is possible to substantially reduce the loads when the rams come into interaction with the guides and thus to increase the latters resistance to wear.

Moreover, this makes possible to increase the permissible eccentricity of application of the impact pulse, i.e., to make the requirements imposed on workpieces less stringent.

The ends of cantilevered portions of the guides should be provided with shock-absorbers locating the rams in their initial positions.

BRIEF DESCRIPTION OF THE DRAWINGS The nature of the invention will be clear from the following description taken in conjunction with the accompanying drawings in which:

FIg. l is a front view of a counterstroke hammer, according to the invention;

FIG. 2 is a schematic vertical section plane passing through the rams and drive and parallel to the drawing plane, and also, a simplified hydraulic feed diagram of the acceleration drive, the guides and hydraulic jacks being omitted for clarity;

FIG. 3 is one of the embodiments of the rams connection to their acceleration drive;

FIG. 4 is a section view taken along line IV-IV of FIG. 1;

FIG. 5 is a section view taken along line V-V of FIG. 4, the rams acceleration drive, tie-rods and hydraulic jacks being omitted for clarity;

FIG. 6 is a section view taken along line VIVI of FIG. 4 passing through the hydraulic jacks and rams of the counterstroke hammer, and a simplified hydraulic feed diagram of jacks, the guides with stop shockabsorbers and acceleration drive with tie-rods being omitted for clarity.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS The counterstroke hammer can be provided with either vertically moving rams or with horizontally moving rams, representing the so-called vertical and horizontal embodiments of the hammer.

In both embodiments the general principles of design and operation are similar, and therefore further de scribed will be a counterstroke hammer with vertically arranged rams.

To a fixed frame 1 (FIG. I) mounted on a foundation 2 are secured guides 3 along which travel a top ram and a

bottom ram

4 and 5, respectively. In their initial positions, i.e., from which the

rams

4 and 5 start moving towards each other, they should be pressedto stop on shock-absorbers 6. The pressing force of the upper ram 4 in its initial position should exceed somewhat its weight, while the lower ram 5 is pressed by gravity and its weight should be partially balanced by an additional force. The ratios of forces pressing the

rams

4 and 5 to the shock-absorbers 6 should be taken such that the forces (F) pressing the rams to said shock-absorbers are approximately equal:

where N, is the force applied to the ram 4;

N is the force applied to the ram 5; G and G are weights of

rams

4 and 5, respectively. wherefrom The value of force F depends on the friction forces obstructing the movement of the

rams

4 and 5 along the guides 3. In practice it is recommended to take N5 2 and N4 (T From their initial positions the

rams

4 and 5 are accelerated towards each other due to the acceleration drive 7 (FIG. 1, 2) provided with two reciprocating parts, plunger 8 and hydraulic cylinder 9. Both parts are mounted so as to move with relation to the frame 1 and rams 4 and 5 in the direction of movement of the latter during acceleration.

The value of mutual relative displacement of the plunger 8 and hydraulic cylinder 9 is determined by the distance 1 (FIG. 2) from a circular lug 10 made on the plunger 8 to a mushroom ll rigidly connected to the hydraulic cylinder 9. The value of distance 1 should be less than the value 1 of total distance of the

rams

4 and 5 from their initial positions till interaction.

The top ram 4 is adapted to interact with the hydraulic cylinder through two tie-rods l2. Said tie-rods are intended to provide simultaneous interaction of the hy draulic cylinder 9 and ram 4 at the instant an acceleration force is transmitted to said ram and ensure coupling of the ram-cylinder system in order to provide for the movement of the ram free from the drive action after acceleration. As can be seen in FIG. 2, the tierods are made as rods 13 provided with screw thread 14 on one end and a lug 15 on the other end. The tierods 12 are arranged in symmetry with respect to the ram 4. In the construction described two tie-rods are used, but their number can be increased as required, provided that they will be arranged symmetrically.

Each tie-rod 12 is rigidly connected to the ram 4' through the medium of said threads provided in the lugs 16. The tie-rods are adapted to interact with the hydraulic cylinder 9 through eyes 17 made integral with cylindrical sleeves 18 connected through highes 19 on a rocker arm 20 arranged to rock relative to a pin The lugs 15 of the tie-rods 12 are made larger in size than holes 22 in the eye 17.

Thus, when the hydraulic cylinder 9 moves downwardly the eye 17 is adapted to interact with the lug l and imparts acceleration force to the ram 4. If the hydraulic cylinder 9 is brought to a stop at any time during the acceleration of the top ram 4, the eyes 17 come to a stop too. The rods 13 will slip in the holes 22, while the lugs will move inside the cylindrical sleeves 18 without disturbing the downward movement of the ram 4 free'from the drive action.

In order to increase the reliability of operation of the tie-rods 12 the rods 13 thereof are made of a plurality of parallel bars, and breakage of a portion of the bars will not lead to an emergency and can be easily traced. Finally, the rocker arm is also made up of several parallel strips, and breakage of a portion of the strips will not cause any emergency either.

To reduce oscillations in the horizontal plane of the rigid connection assembly of the tie-rods 12 with the ram 4 the lugs 16 should be arranged in a plane perpendicular to the direction of movement of the ram 4 during acceleration and passing through its centre of gravity.

FIG. 3 presents a schematic illustration of another version of connecting the tie-rods and cylinder of the acceleration drive of a counterstroke hammer. Here the tie-rods 12 also have lugs 15 and a screw thread 14 on opposite ends. As can be seen in FIG. 3, the tie-rods 12 are rigidly connected to the cylinder 9, their lugs 15 being passed through the eyes 17. With the downward movement of the cylinder 9 the eye 17 comes into interaction with the lug l5 and imparts the acceleration force to the ram 4. If the cylinder 9 is brought to a stop, the tie rods, and hence the lugs 15, also come toa stop, while the eyes 17 sliding along the rods of ties 13 do not interfere with a free movement independent of the drive action. 7

The acceleration force is transmitted to the bottom ram 5 (see FIG. 2) from the plunger 8 with the help of a long dowel 23 which is articulated with one of its ends in the plunger 8, and thrusts with the other end into the end face of said ram during acceleration. The dowel 23 can rock in the hinge joint relative to the plunger 8. The value of the dowel 23 turn angle is determined based on the condition of the allowable misalignment and allowable angle turn of the bottom ram 5 relative to the drive 7.

Thus, during the upward movement of the plunger 8 (see FIG. 2), the dowel 23 comes in contact with the ram 5 and transmits the acceleration force thereto from the plunger 8. It shall be pointed out that should the plunger 8 be stopped at any instant during the acceleration, the dowel 23 will come to a stop too, while the ram 5 will continue to move upwardly, free from the action of the drive 7. A tapered bore 24 is made on the end face of the ram 5 facing the acceleration drive 7 This tapered bore 24 performs the function of a catcher for the dowel 23 of the plunger 8 at the instant the ram 5 and the drive 7 come into contact and ensures the movement of the dowel 23 along the axis passing through the centre of gravity of the ram 5.'

The guides 3 (FIG. 4) are arranged in pairs in proximity to the ram ribs, being fabricated of rolled stock such as rails or I-beams.

Each of said guides includes a pair of beams with consoles'25 (FIG. 5).

Each of the beams 3 is articulated to the hammer frame 1 through a pair of hinged supports 26 (FIG. 5).

One of said supports 26 lies in a plane perpendicular to the direction of movement of the rams during their acceleration and crossing substantially'medially all the distance travelled by the centre of gravity of the ram 4, while the other support 26 also lies in a plane perpendicular to the direction of movement of the rams during acceleration and crossing substantially medially the total distance travelled by the centre of gravity of the ram 5. Each support 26 allows for the guide 3 to be turned around an axis lying in the plane perpendicular to the direction of movement of the rams during accel eration. In other directions the connection of the guides is rigid. Such a fixing of the guides 3 provided for an increased yield when loaded, allows for angular oscillations of the rams at substantial amplitudes and makes possible to absorb the rotary energy of the rams with moderate forces exerted on the surfaces of contact of the rams with the guides.

The illustrated hammer has eight guides 3. However, a hammer having a number of guides less than eight, such as four guides is possible. Some guides are provided with somewhat elongated console portions 25 whereto are secured shock-absorbers 6.

Hydraulic jacks 27 and 28 (FIG. 6) are provided respectively for the

rams

4 and 5 in order to return them to their initial positions following their mutual impact, press them to the shock-absorbers and hold them in their initial positions. The force developed by said hydraulic jacks can be determined by the formulas (2) and (3).

The

hydraulic jacks

27 and 28 comprise respectively

hydraulic cylinders

29 and 30 with

rods

31 and 32. Said hydraulic cylinders are rigidly connected, by means of

flanges

33 and 34, to a bedplate 35, which in its turn is connected to the frame 1 through uprights 36. It should be taken into consideration that in FIG. 6, being a crosssection of FIG. 4 along the line VI-VI, one of the jacks of each ram can be seen. Each of the rams 4 and 5- comes into'interaction with a pair of hydraulic jacks, since FIG. 4 illustrates two hydraulic jacks 27 adapted to interact with the top ram 4 (the hydraulic jacks 28 are not shown in this Figure). The hydraulic jacks 27 (and hence hydraulic jacks 28) are arranged in symmetry with respect of the rams inorder to preclude the development of force moments 7 applied to the rams.

Said forces are transmitted to the rams by the rods of the hydraulic cylinders through bosses made on the rams. FIG. 4 illustrates the bosses made on the ram 5 (those made on the ram 4 are not shown), while FIG. 6 shows an interaction of the rod 31 of one of the hydraulic jacks 27 with the boss 37 of the ram 4 and the interaction of the rod 32 with the boss 38 of the ram 5. The

bosses

37 and 38 are arranged in the planes perpendicular to the movement of the

rams

4 and 5 during their acceleration and passing through the centres of gravity of said rams. Such positions of the

bosses

37 and 38 ensure the minimum amplitude of displacement of the points of contact of said bosses with the

rods

31 and 32 during angular oscillations of the

rams

4 and 5.

This feature makes it possible to simplify the fastening means of the hydraulic jacks on a bedplate 35, since in this case insignificant displacement of said point of contact is possible due to elastic deformation of the

hydraulic cylinders

29 and 30 and the

rods

31 and 32.

FIG. 6 illustrates a simplified hydraulic diagram of feeding

hydraulic jacks

27 and 28. The hydraulic cavities of the

cylinders

29 and 30 are interconnected through the medium of by-passes 39. The hydraulic jacks are fed from a hydraulic accumulator 40 along

pipelines

41, 42, 43 and 44.

The feed control of the hydraulic jacks is effected by means of

valves

45 and 46. The design of the latter is not shown in the Figure, but valves of any construction can be used. The valve 45 can attain two positions, in one position the pipeline 41 communicates with the pipeline 42 while the passage to the pipeline 43 is shut off. In the other position the pipeline 41 communicates with the pipeline 43, while the passage to the pipeline 42 is shut off.

The valve 46 has two positions as well. In one of the positions the pipeline 43 communicates with the pipeline 44, while the passage to drain pipeline 47 is shut off. In the other position the pipeline 41 communicates with the pipeline 47, while the passage to the pipeline 44 is shut off.

In the working cycle the rams 4 and are normally moving in opposite directions, either in the direction wherein they come into contact, or vice versa. Therefore, the following flow rate will be established in the pipelines 41 and 42:

Q ai 4 32 5 where S, is the total flow cross-section of rods 31,

S is the total flow cross-section of rods 32;

V, and V are speeds of

rams

4 and 5 respectively.

The flow cross-section of the

pipelines

41 and 42 and of the valve 45 can be selected such as to ensure minimum losses during the flow Q The cross-sections S and S must be reduced to a minimum by means of increasing the fluid working pressure. The reccommended fluid working pressure is 150 to 200 bar.

During idle when it is necessary to close the rams or open them at low speed, the fluid flows along the

pipelines

43 and 44. The speeds during idle time are insignificant, therefore the flow cross-section of the

pipelines

43 and 44 and of the valve 46 are much smaller than the flow cross-section of the valve 45 and the

pipelines

41 and 42.

The hydraulic accumulator 40 is refilled with fluid through a pipeline 48, and with compressed gas from supply sources through a pipeline 49.

FIG. 2 presents a schematic diagram of supply of the hydraulic drive 7. One of the embodiments of the drive construction is described, but it is to be understood that in principle other modifications of the drive can be used as well.

However, in any case the drive should include a builtin hydraulic brake in order to damp the kinetic energy of masses of the plunger 8 and cylinder 9 during their counteraction. The cylinder 9 of the drive is embraced with a circular manifold 50, the latter being made leaktight with respect to the atmosphere with movable seals (not shown in the Figure) sliding along the outer surface of the cylinder 9. The manifold 50 communicates with the hydraulic accumulator 55 through ducts 51 along a pipeline 52 and control valve 53 and pipeline 54 The control valve has two positions.

In one position the pipeline 52 communicates with the pipeline 54, while the drain passage to the pipeline 58 is closed.

In the other position the pipeline 52 communicates with the drain pipeline 58, while the passage to the pipeline 54 is shut off.

In parallel with the valve 53 the pipeline 52 communicates with the pipeline 54,via a non-return valve 59 connected so that when the pressure in the pipeline 52 exceeding that in the pipeline 54 the fluid is bypassed from the hydraulic drive 7 to the hydraulic accumulator 55. The manifold 50 communicates with the interior cavity 60 of the hydraulic cylinder 9 through openings 61. The cavity 60 communicates with the cavity 62 arranged inside the plunger 8 through holes 63. The plunger end 8 facing the cylinder bottom 9 has a hole 64 through which a rod 65 passes, one of its end being connected to the cylinder 9 bottom, while the other end is provided with a mushroom 11. A profiled annular lug 66 is provided in the inner cavity 62 below the holes 63. When the plunger 8 is being withdrawn from the cylinder 9, the mushroom 11 will pass along the lug 66 meanwhile thrusting into the lug 10, thereby shutting off a portion of volume 67 from the cavity 62. The lug 66 is profiled in such a way that when kinetic energy of freely moving masses of the plunger 8 and cylinder 9 is damped due to forcing the fluid through the annular clearance between the mushroom I1 and lug 66 from the volume 67 into the cavity 62, the pressure in the volume 67 remains constant.

The volume 67 and the braking pressure are selected such that where W is the kinetic energy of masses of plunger 8, cylinder 9 and fluid mass from the hydraulic drive to the accumulator;

P is the braking pressure in the volume 67;

P is the working fluid pressure in the cavity 60;

V is the capacity of the volume 67,

K is a constant depending on the drive construction.

In FIG. 2 the index I, designates the value of possible total relative displacement of the plunger 8 and cylinder 9 until the mushroom l1 thrusts into the lug 10; the index l shows the value of relative displacement of the plunger 8 and cylinder 9 until the hydraulic brake comes into action. Over the distance 1 the

rams

4 and 5 are accelerated by means of the drive 7, while over the distance l l the kinetic energy of the masses of the plunger 8 and hydraulic cylinder 9 and the mass of fluid in the pipelines from the drive to the accumulator are damped.

The total flow area of the ducts 51,

pipelines

52, 54 and valve 53 should be of the same order that the working section of the plunger 8 of the drive 7, designated with the letter S in FIG. 2. The flow cross-section of the non-return valve 59 should be of the same order as the section S In order to reduce the kinetic energy of the fluid in the

pipelines

52, 54 and hydraulic resistances along way of the fluid from the hydraulic accumulator 55 to the manifold 50, the accumulator 55 is disposed in proximity to the drive 7. For this purpose, several pairs of control valves 53 and non-return valves 59 disposed equidistantly along the circumference of the annular mainfold 50 can be used.

As can be seen from the above description, the drive 7 overcomes not only the inertia forces determined by the masses of the rams 4 and during their acceleration, but also the forces required to part and hold them in their initial positions. This force is designated as F. The cross-section S is selected so that with a given value of the fluid pressure P the following condition is met:

PS F.

Based on our experiment the maximum speed of the rams at which they come into action, as regards the requirements placed upon the pressing techniques, should not exceed to 12 m/sec. With such speeds the flow area of the valve 53 is selected to be equal to one third of the working flow section S of the plunger. Under such conditions with adequate approximation the .acceleration time of rams t can be taken approximate to:

t V 2Ml /PS where M 4 9)( s s)/ 4 s 5 s) where M is the top ram mass 4;

M is the bottom ram mass 5;

m is the mass of the cylinder 9 and parts rigidly connected thereto; m is the mass of theplunger 8 and parts rigidly connected thereto;

P is the fluid working pressure;

S is the working section of the plunger 8;

I is the acceleration distance of the

rams

4 and 5.

It shall be noted that along the distance 1 at the drive 7 end to the

rams

4 and 5 is imparted the acceleration force.

The ratio V 2Ml /PS when designing the hammer is selected such that the accelerationtime of the rams 4 and S exceeds the time (1') of action of the control valve 53. Since the rams attain their maximum speed at the end of their acceleration, the values I and 1' can be of the same order 1 r. v

A hammer possessing the values M 1 P and S ensuring the acceleration time in the order of 0.1 see, as the valve 53 having the operating time of the same order, is feasible.

The working fluid is supplied to the accumulator 55 through the pipeline 56, while the compressed gas is fed through the pipeline 57.

It shall be noted that a hammer can be executed with the drive disposed so that the plunger 8 is connected to the ram 4, while the cylinder 9 is connected to the ram 5, i.e.. the drive is inverted. I

The horizontal version of the hammer is distinguished in that the hydraulic jacks 28 are interlinked with the ram 5 in the same way as the hydraulic jacks 27 with the ram 4 in the embodiment hereinabove described. The

hydraulic jacks

27 and 28 perform the function of moving the rams to their initial positions only, the gravity of the rams being taken up in this case by the guides.

In any embodiment of the hammer construction the dies 68 and 69 for processing a workpiece are secured to opposing surfaces of the

rams

4 and 5.

The above hammer operates as follows.

The valve 45 is positioned so that the pipeline 4] communicates with the pipeline 42, while the passage to the pipeline 43 is shut off. The valve 46 is positioned so that the pipeline 43 communicates with'the pipeline 44, while the drain passage to the pipeline 47 is shut off. The

rams

4 and 5 are clamped to the shock absorbers 6 by the

hydraulic jacks

27 and 28.

In the initial position the control valve 53 is in such a position wherein the pipeline 52 communicates with the drain pipeline 58, while the passage to the pipeline 54 is shut off.

To effect an impact the control valve 53 is shifted to a position for the time of acceleration of the

rams

4 and 5, whereat the pipeline 5 2 communicates with the pipeline 54, while the passage to the drain pipeline 58 is shut off. The fluid is supplied into the cavity of the cylinder 60 from the accumulator 55 along the

pipelines

54, 52 and through the hole 51, manifold 50, holes 61. The fluid can pass freely from the cavity of the cylinder 60 through holes 63 into the cavity of the plunger 62. The downward stroke of the ram 4 is accomplished by the static pressure of the fluid exerted onto the cylinder 9 through the rocker arm 20, eyes 17 and tie-rods 12, the upward stroke of the ram 5 is accomplished by the static pressure of the fluid exerted on the plunger 8 through the dowel 23.

The acceleration force is applied until the mushroom 11 compresses the fluid in the braking volume 67, wherein the pressure starts building up quickly. The force exerted by the fluid in the volume 67 on the plunger 8 and cylinder 9 applied in the direction opposite to the acceleration force brakes the plunger 8 and the drive cylinder 9. The lug 66 is profiled so that the pressure in the braking volume is kept constant. This is the most favourable braking duty,since with the preset strength of the braking chamber the braking volume 67 occupies minimum space.

If follows from the description that when both parts of the drive 7 are braked, the

rams

4 and 5 will continue to move freely in the direction of acceleration.

In FIG. 1 the plane wherein the

rams

4 and 5 come into contact is designated us lV-lV. Let as assume that the time required for the rams to come into contact, also, the time of rams counteraction, the time of movement of the rams until coming into contact with respective parts of the drive is equal to t. During this time t the masses of the cylinder 9 and plunger 8 inter act independently of the

rams

4 and 5.

The interaction of the plunger 8 with the cylinder 9 through the braking volume 67, due to unconvertible loss of the energy by the fluid flowing through the annular clearance made between the mushroom 11 and annular lug 66 into the cavity 62 can be considered as absolutely unresilient counteraction. If the ratio of the plunger 8 mass with the parts connected thereto and the cylinder 9 with the parts connected thereto is taken equal to the ratio of the ram 4 mass to the ram 5 mass, their speeds in the laboratory coordinate system become equal to zero as a result of unresilient counteraction of the plunger 8 and cylinder 9. This will be favourable during the counteraction of the drive 7 with the rams and after the latter rebound as a result of resilient impact, since this will reduce the speed of encounter of the drive parts with the rams.

Hence the drive masses are taken basing on:

where m is the plunger mass of the acceleration drive;

M is the ram mass accelerated byforces exerted by the mass m m is the hydraulic cylinder mass;

M is the ram mass accelerated by forces exerted by the mass m By the moment the time t elapses, the control valve 53 is shifted over to another position, wherein the pipeline 52 communicates with the pipeline 58, while the passage to the pipeline 54 is shut off. By the moment the time t elapses, the top ram 4 under the action of rebound speed actuates the cylinder 9 through the tierods 12, eyes 17 and rocker arms 20 and accelerates the cylinder until the latter attains its own speed. At the same time the bottom ram 5 interacts with the plunger 8. The displacement of the plunger 8 and cylinder 9 by the rams causes the pressure in the hydraulic cavity of the drive 7 to rise, since the flow area of the drain pipeline 58 is made specially small. The non-return valve 59 opens as the pressure in the pipeline 52 exceeds that in the accumulator 55.

Further the kinetic energy imparted to the rams after their mutual impact and rebound displaces the plunger 8 relative to the cylinder 9. Some portion of the fluid is returned to the accumulator, while another portion is drained. Final displacement of the rams to their initial positions is performed by the

jacks

27 and 28, the fluid from the drive being drained through the pipeline As the

rams

4 and 5 reach the shock-absorbers 6 the hammer is prepared for a subsequent stroke.

It shall be added that the pressure in the drain pipeline 58 is kept higher than the atmospheric one in order to preserve the fluid in the hydraulic system which is liable to flow out under gravity.

As can be seen from the description of operation, the impact of the rams takes place under conditions free from the drive action. This, on the one hand, excludes the action of impact overloads exerted on the most complicated unit of the hammer, i.e., the acceleration drive. On the other hand, it makes possible to make the shape of the rams extremely simple, executing them as prisms free from sharp transitions from one section to another, thereby increasing their stability to wave impact overloads and allow for rigid impacts to be performed. Thirdly, since the forces of acceleration during mutual impact of the rams are removed, and only the forces returning the rams to their initial positions are applied, the time of contact of the rams is minimal, being determined by the time of double path of stress waves along the ram lengths. With the length of rams 2500 mm this time is in the order of see. And, finally this makes possible the allowances of substantial angular oscillations of the rams to be effected with high amplitudes. This does not affect the leak-tightness of the acceleration drive, since the latter is free from the angular oscillations experienced by the rams. It is known that during the stamping of even coaxial parts the rams fail to exhibit central mutual impact, the more so that the central mutual impact cannot be effected during such processes as briquetting of loose materials. This is due to the fact that the workpieces exhibit uneven preheating, lubrication, unaccurate installation of the workpieces in the die, clearances between the guides and other factors.

It is clear from the consideration of the physical na ture of the phenomenon that since the rams feature out-of-centre mutual impact, they possess some rotary energy W after their impact irrespective of the construction of the guides. A notion that rigid guides prevent twisting ofthe rams appears erroneous, since during their mutual impact (especially during rigid impacts) the rams have no time to-eliminate the clearance relative to the guides. The rotary energy is proportional to the square of eccentricity of application of the impact pulse.

This rotary energy is damped when the rams and the guides strike against each other. The less resilient are the guides (i.e., rigid) the higher the forces applied during the counteraction and the more is the wear of the guiding surfaces.

Considering the two types ofthe guides having rigidity K and k (size n/m) it can be shown that the ratio of forces exerted onto the guides P, and P with the same value W is determined by the equation Hence, if the rigidity of the guides is reduced fold, the force applied to the guides can be reduced 10- fold. This would improve the efficiency of the guides and increase the allowable eccentricity during mutual impact.

When the rigidity of the guides is reduced, considerable angular oscillations of the rams after their parting are inevitable. However, these are precisely the oscillations that are allowable in the construction hereinabove described since in order to increase the resilience and reduce the rigidity the guides are made as two-support beams having consoles.

Moreover, independent from the impact of the rams the interaction of the plunger and cylinder during the free movement allows for calculating loads imposed on the acceleration drive and makes them independent from the stamping techniques, due to the hydraulic brake incorporated in the acceleration drive. This increases substantially the long service life of the drive and improves its operating conditions, since no accidental loads are imposed on the drive parts.

What we claim is:

1. A counterstroke hammer comprising: a frame; two rams moving towards each other from their initial positions until they come into interaction in a zone for processing a workpiece and return to their initial positions; guides secured to the frame along which movements of the rams are effected; an acceleration drive for the rams made as a pair of reciprocatingly moving parts, cylinder-plunger, which move relative to the frame and both rams in the direction that the latter move during acceleration, the value of mutual relative displacement of both parts of the drive being less than the total distance travelled by said rams from their said initial positions until they come into interaction; tie-rods intended to ensure simultaneous interaction with one of said rams and one of said moving parts of the drive in order to impart an acceleration force to said ram and further tained during their acceleration being damped, while the kinetic energy of both parts of the drive being damped during their interaction independently of the rams.

2. The hammer as claimed in claim 1, wherein the ratio of the-masses of the rams and drive parts is selected according to the formula:

where m is the mass of one part of the acceleration drive;

M is the mass of the ram accelerated under the action of the mass m m is the mass of the other part of the acceleration drive;

M is the mass of the ram accelerated under the ac tion of the mass m 3. The hammer as claimed in claim 1, wherein the tierods are arranged in pairs, in symmetry with respect to the rams and each tie-rod is made as a stem rigidly connected with one end to the relevant ram and having a lug on the opposite end, while the cylinder is provided with an eye through which is passed said stem transmitting the acceleration force to said ram during interaction of the lug with the eye.

4. The hammer as claimed in claim 1, wherein the rigid connection of the tie-rod with the ram is disposed in a plane perpendicular to the direction of movement of the ram during acceleration of the ram and passing through its centre of gravity.

5. The hammer as claimed in claim 3, wherein the eyes are connected to the cylinder through a rocker arm rocking relative to a pin perpendicular to the direction of movement of the rams, the eye being articulated to the rocker arm.

6. The hammer as claimed in claim 1, wherein the tierods are arranged in pairs in symmetry with respect to the rams and each tie-rod is made as a stem rigidly connected with one end to the cylinder and having a lug on the other end, while the ram is provided with an eye through which is passed said stem transmitting the acceleration force to said ram during interaction of the lug with the eye.

7. The hammer as claimed in claim 3, wherein the stem is a plurality of parallel bars.

8. A hammer as claimed in claim 1, wherein both parts of the acceleration drive are disposed coaxially with the direction of movement of the rams, while the plunger imparts the force to the ram being accelerated through a dowel hinge-connected with one end in the plunger and thrusting the other end into the end face of said ram during the acceleration.

'9. The hammer as claimed in claim 8, wherein the ram accelerated under the action of the plunger is provided with a tapered bore at the drive and to accept said dowel,

10. The hammer as claimed in claim 1, wherein each guide is articulated in two supports, one lying in the plane perpendicular to the movement of the rams and crossing substantially medially all the distance from the initial position to mutual impact, passed by the centre of gravity of one of the rams, while the second support is disposed in the plane perpendicular to the direction of movement of the rams and crossing substantially medially all the distance from the initial position to mutual impact passed by the centre of gravity of the other ram, the guides being made as two-support beams with console portions.

11. The hammer as claimed in claim 8, wherein the ends of console portions are provided with shockabsorbers fixing the initial positions of the rams.

12. The hammer as claimed in claim 1, further comprising a mechanism responsible for returning the rams comprising hydraulic jacks disposed in pairs in symmetry with respect to the rams, said rams having lugs disposed in the plane perpendicular to their movement and passing through their centre of gravity, whereinto said hydraulic jacks thrust in order to impart a returning force to the rams and to hold them in their initial positions.

Claims

1. A counterstroke hammer comprising: a frame; two rams moving towards each other from their initial positions until they come into interaction in a zone for processing a workpiece and return to their initial positions; guides secured to the frame along which movements of the ramS are effected; an acceleration drive for the rams made as a pair of reciprocatingly moving parts, cylinder-plunger, which move relative to the frame and both rams in the direction that the latter move during acceleration, the value of mutual relative displacement of both parts of the drive being less than the total distance travelled by said rams from their said initial positions until they come into interaction; tie-rods intended to ensure simultaneous interaction with one of said rams and one of said moving parts of the drive in order to impart an acceleration force to said ram and further disconnect the ram-drive moving part system to ensure movement of the ram free from the drive action following the acceleration; the other moving part of the drive being adapted to contact the other said ram for transmitting an acceleration force and is disconnected from said contact after the ram is accelerated to perform a movement free from the action of the drive, resulting in that the impact interaction of the rams is effected with the disconnected drive, the kinetic energy obtained during their acceleration being damped, while the kinetic energy of both parts of the drive being damped during their interaction independently of the rams.

2. The hammer as claimed in claim 1, wherein the ratio of the masses of the rams and drive parts is selected according to the formula: m1/m2 M1/M2 where m1 is the mass of one part of the acceleration drive; M1 is the mass of the ram accelerated under the action of the mass m1; m2 is the mass of the other part of the acceleration drive; M2 is the mass of the ram accelerated under the action of the mass m2.

3. The hammer as claimed in claim 1, wherein the tie-rods are arranged in pairs, in symmetry with respect to the rams and each tie-rod is made as a stem rigidly connected with one end to the relevant ram and having a lug on the opposite end, while the cylinder is provided with an eye through which is passed said stem transmitting the acceleration force to said ram during interaction of the lug with the eye.

6. The hammer as claimed in claim 1, wherein the tie-rods are arranged in pairs in symmetry with respect to the rams and each tie-rod is made as a stem rigidly connected with one end to the cylinder and having a lug on the other end, while the ram is provided with an eye through which is passed said stem transmitting the acceleration force to said ram during interaction of the lug with the eye.

9. The hammer as claimed in claim 8, wherein the ram accelerated under the action of the plunger is provided with a tapered bore at the drive and to accept said dowel.

11. The hammer as claimed in claim 8, wherein the ends of console portions are provided with shock-absorbers fixing the initial positions of the rams.