EP0559651B1 - Hydraulically driven machine with two drive cylinders - Google Patents

Abstract

In a hydraulically driven machine with two vertical differential cylinders as drive cylinders arranged at a mutual lateral distance on side cheeks of the machine frame, the control components are two follow-up control valves with an electrically controlled predetermined reference position for an upper tool and a mechanical actual position indication which have a locking basic setting and two alternative through-flow settings which are allocated to the fast and load forward action directed downwards on the workpiece and the return action directed away from the workpiece. A piston area of the drive cylinder pistons limit an upper drive pressure chamber and a smaller piston area limits a lower drive pressure chamber of the drive cylinders through which axially passes the piston rod concerned to which the upper tool is connected. In fast forward action the weight of the upper tool and the piston is used as the driving power, whereby the pressure medium flows into the upper drive pressure chamber via inlet valves. In load forward action high pressure is applied to the upper drive pressure chambers. On return action the pressure in the upper drive pressure chambers is relieved via pressure-controlled outlet valves.

Description

The invention relates to a hydraulically driven machine for processing a sheet metal workpiece by cold forming, according to the preamble of claim 1. Such a machine is known from US-A-2 211 692.

Such machines are known, for example, as bending machines, in which the sheet metal workpiece is arranged between a lower tool and an upper tool which can be moved up and down relative to the latter, for the drive of which two drive cylinders, which are arranged at a lateral distance from one another and are designed as differential cylinders, are provided, the pistons of which have piston rods emerging from the cylinder housings with which the upper tool is connected to compensate for slight differences in stroke of the pistons via uniaxial joints, the joint axes of which are perpendicular to the plane marked by the central axes of the drive cylinders, the ratio F 1 / F 2 of the surfaces F 1 and F 2 the piston, with which they each delimit an upper drive pressure chamber against a lower drive pressure chamber axially penetrated by the piston rod, through their common or alternative - valve-controlled - pressurization and / or - relief of the workpiece directed rapid feed movements and load feed movements, and rapid return movements directed away from the workpiece can be controlled, has a value between 5 and 20, with the upper, large-area limited drive pressure space entering in rapid feed operation Post-flow valve and the lower drive pressure chamber is connected to the reservoir of the pressure supply unit via a control element with adjustable throttling effect, the lower and the upper drive pressure chamber are jointly pressurized in load-feed operation and the lower drive pressure chamber is pressurized in high-speed retraction mode, while the upper drive pressure chamber is relieved of pressure and for controlling the piston movements and their synchronization, an electrohydraulic control unit working with an electrically controllable position setpoint specification and comparing the actual position of the pistons or the upper tool connected to them with the setpoint specification is provided.

In such known bending machines - as a drive energy source - for the movement control of the upper tool in rapid feed mode, in which it moves towards the workpiece to be machined, the considerable total weight of the upper tool and drive cylinder piston is used, which in typical cases amounts to 10 kN. This weight is sufficient to achieve a sufficient pressure in the downward rapid feed operation of the two drive cylinders of the machine - with the upper drive pressure chamber of the respective drive cylinder relieved of pressure in their lower drive pressure spaces, below which the pressure medium is sufficiently quick in the usual dimensions of the line cross sections and the control valves the lower drive pressure chambers of the drive cylinders can flow away in order to achieve the required speeds of the rapid feed movements and to be able to regulate them by suitable throttling of the outflow cross-sections, without having to "help" the rapid downward movements of their pistons by increasing the pressure in the upper drive pressure chambers of the drive cylinders, due to the area ratio F₁ / F₂ would require higher delivery rates of the pump of the pressure supply unit.

In machines of this type, in which the drive cylinders are arranged at a relatively large distance from one another which is determined by the width of the workpieces to be machined, the synchronization of the two drives is particularly critical, since small differences in the stroke of the drive cylinder pistons to which the upper tool is connected in a form-fitting manner form one Tilting of the pistons in the cylinder housings and / or damage to the pistons and / or the upper tool itself can result, even if this is articulated to the piston rods of the drive cylinders in order to enable small deflection movements.

In order to counter such dangers, the positions of the drive cylinder pistons must therefore be monitored very closely and be comparable with one another. For this purpose, hydraulic drives (Bosch, Rexroth) for bending machines are usually provided with measuring systems implemented using glass scales, which measure the absolute values of the positions of the pistons at any moment enable capture. Such measuring systems are very expensive and also susceptible to temperature influences and / or environmental conditions in general.

Furthermore, 4/3-way solenoid valves with oppositely operating proportional magnets are generally used in known bending machine drives to control the movement of the drive cylinders, the deflection strokes of such solenoid valves being controlled by the solenoid current. However, this requires that, in the event that a fault occurs in the electrical supply system, additional safety devices must be provided which ensure that the drive cylinders cannot "fall off" if such an electrical fault occurs, for example, in the rapid feed operation of such bending machines occurs. Special pressure-controlled fall protection valves are provided as safety devices, which are also electrical, i.e. can be piloted via solenoid valves.

The total electro-hydraulic effort required for this is considerable.

In addition, the processing of the electronic position data supplied by the position measuring systems, by means of which the setpoint specification must be corrected, takes considerable time, so that the control loops known for the control of the bending machines are relatively slow, ie limited Values in particular of the rapid feed and rapid retraction speeds have to be accepted in order to enable a sufficient synchronization of the piston positions, which of course results in longer cycle times and thus reduces the economic efficiency of the use of such machines.

The object of the invention is therefore to improve a machine of the type mentioned in such a way that not only is the electro-hydraulic effort required to control its movement reduced, but also shorter cycle times of processing can be achieved.

This object is achieved by the features mentioned in the characterizing part of claim 1.

Due to the follow-up valves provided as part of the position control loops for the two drive cylinders, designed as 3/3-way valves, which work with pulse-controlled position setpoint specification and mechanical position actual value feedback, with the position setpoint If electrical AC or stepper motors are used, which allow an incremental position setpoint specification, the motion control of the drive cylinders can be easily achieved by controlling the electric motors at the same frequency, the mechanical actual position value feedback ensuring that the overrun Control valves quickly enough in the sense of the respective direction of movement can be controlled. Provided the pressure supply unit has sufficient performance, the two follow-up control loops can be operated with optimally high loop gain. The permanent connection of the high-pressure outlet of the pressure supply unit to the lower drive pressure chamber of the respective drive cylinder provides effective protection against dangerous "falling" of the drive cylinder pistons or the upper tool of the machine, at least in the event of malfunctions in the electrohydraulic control unit. In the event of a general power failure, the respective overrun control valve automatically returns to its blocking state due to the mechanical feedback of the actual position value, so that an adequate safeguard against "falling" of the upper tool is also guaranteed. The amount of pressure medium displaced from the upper drive pressure chambers in the rapid retraction mode of the drive cylinders can flow back via the outflow valve to the reservoir of the pressure supply unit without an excessive pressure having to be built up in the lower drive pressure chambers of the drive cylinders, which leads to a favorable low overall power consumption of the invention Machine contributes. It is therefore completely sufficient if the high-pressure pump of the pressure supply unit is designed for the maximum delivery capacity provided in accordance with claim 2, it being advantageous from the point of view of minimizing the electrical connection power if, as provided in accordance with claim 3, the high-pressure pump of the Pressure supply unit is designed as a pressure-controlled pump, for example a pump, the flow rate of which is inversely proportional to the outlet pressure generated at its outlet. Alternatively or in combination with this, it can also be sufficient or additionally advantageous if a pressure relief valve designed according to claim 4 is provided, which can also be advantageous as a safety measure in combination with a pump regulated according to claim 2.

The outflow valve provided in the context of the electrohydraulic control unit can be designed as a solenoid valve controlled according to the features of claim 5, which is particularly advantageous in terms of simple control.

A structurally simpler and also cheaper because of the lower technical complexity, a sufficiently functional control of the outflow valve can also be realized if it is designed as a pressure-controlled valve in accordance with the features of claim 6, which is also very advantageous from a safety point of view, with the features of claim 7 a particularly expedient design of the outflow valve is specified.

By a compensation device with the functional features of claim 8 is largely avoided that there are elastic deformations of the machine frame that this with increasing feed force the drive cylinder is experiencing to an increasing extent, can have an effect on incorrect position feedback and can therefore in particular lead to undesired extensions of the cycle times.

Such a compensation device can be implemented in a simple manner according to the construction principle according to the features of claim 9 and in a special embodiment according to that of claim 10. The design of the compensation device according to claim 12 has the advantage of significantly less influence on the hydraulic rigidity of the drive compared to that of claim 11, but requires a somewhat increased technical effort to implement.

By means of an encoder system acting according to claim 13, the machine can be largely secured against damage which could result from an inexactly synchronous feed of the two drive cylinder pistons.

If such an encoder system conveys the function provided according to claim 14, so that its output signals are a direct measure of the so-called tracking error, ie the distance by which the actual position value lags behind the set position setpoint, then this encoder system is in combination with the setpoint - Presetting system also as electronic Position actual value measuring system can be used, which provides a CNC-controlled machine with the electronic information required for this about the actual position of the upper tool. The output signals of such a tracking error measuring system can also be advantageously used for a type of control of the position setpoint specification provided in accordance with claim 15 and thus gentle operation of the machine can be achieved.

Finally, the features of claim 16 provide a structurally simple safety device which, seen on its own, provides an effective safeguard against such resulting dangerous situations in the statistically predominant number of malfunction cases of the machine.

Fig. 1

eine schematisch vereinfachte Vorderansicht einer erfindungsgemäßen Maschine mit zwei, teilweise im Schnitt dargestellten Antriebszylindern,

Fig. 2

eine Seitenansicht der Maschine gemäß Fig. 1, in Richtung des Pfeils II derselben gesehen,

Fig. 3

ein Druckschaltbild des einem der beiden Antriebszylinder der Maschine gemäß Fig. 1 zugeordneten Teils einer zur Steuerung seiner Arbeitsbewegungen vorgesehenen elektrohydraulischen Steuereinheit mit einem als 3/3-Wege-Ventil ausgebildeten Nachlauf-Regelventil,

Fig. 4

Einzelheiten des Nachlauf-Regelventils gemäß Fig. 3 in vereinfachter, halbschematischer Schnittdarstellung;

Fig. 5

eine Kompensationseinrichtung zum Ausgleich von betriebsbedingten Aufweitungen des Maschinengestells.

Further details and features of the invention emerge from the following description of a special exemplary embodiment with reference to the drawing. Show it

Fig. 1

2 shows a schematically simplified front view of a machine according to the invention with two drive cylinders, partially shown in section,

Fig. 2

2 is a side view of the machine according to FIG. 1, seen in the direction of arrow II thereof,

Fig. 3

a pressure circuit diagram of one of the two drive cylinders of the machine assigned to FIG. 1 Part of an electrohydraulic control unit provided for controlling its working movements with a follow-up control valve designed as a 3/3-way valve,

Fig. 4

Details of the follow-up control valve according to Figure 3 in a simplified, semi-schematic sectional view.

Fig. 5

a compensation device to compensate for operational expansion of the machine frame.

1, 2 and 3, generally designated 10, is a machine of the type in which, by means of a force-controlled lowering of an upper tool 11 relative to a lower tool 12 fixed to the machine frame, a workpiece 13, e.g. a steel sheet, which rests on the lower tool 12, is pressed into a desired shape, which is determined by the design of the lower tool 12 and the upper tool 11, which is complementary in cross section to the lower tool.

The machine 10 comprises a machine frame, generally designated 14, with two C-shaped side walls 17 and 18, which are arranged symmetrically with respect to a vertical transverse center plane 16 of the machine frame 14 and which, as shown in FIG. 1, are arranged at a lateral clearance b from one another and through a base part 19, as it were a base of the machine frame 14, are connected to one another, which extends between the lower, horizontal transverse legs 21 and 22 of the overall C-shaped side walls 17 and 18, which are integrally connected to upper ones via vertical, columnar yoke legs 23 and 24 , horizontal transverse legs 26 and 27 of the C-shaped side walls 17 and 18 are formed contiguously.

On these upper transverse legs 26 and 27, a hydraulic drive cylinder 31 or 32 is fixedly mounted with the vertical course of the respective central longitudinal axis 28 or 29, the pistons of which are labeled 33 and 34, respectively, emerging from the cylinder housings 36 and 37 downwards Have 38 and 39 respectively.

These drive cylinders 31 and 32 are designed as double-acting differential cylinders, in which an upper drive pressure chamber 43 and 44 against a lower drive pressure chamber 46 and 47, respectively, by means of a piston flange 41 or 42 sealed against the cylinder housing 36 or 37 by means of a piston-fixed ring seal is delimited so as to be pressure-tight and movable, by means of which, by means of an electrohydraulic control unit, which is designated in total by 48, it acts upon the outlet pressure P of a pressure supply unit, which is designated in total by 49, and connects it to its - pressure-less - reservoir 51 directed towards the workpiece 13 in rapid downward movements of the upper tool 11 articulated with the piston rods 38 and 39 of the two drive cylinders 31 and 32, load feed movements of the same, by means of which the deformation of the workpiece 13 is achieved, and rapid retraction movements of the upper tool 11 in the sense of lifting the same from Workpiece 13 and return to an upper end position are controllable.

• des Betrages F₁ der "oberen" Kolbenfläche 52 bzw. 53, welche die axial bewegliche Begrenzung des oberen Antriebsdruckraumes 43 bzw. 44 des Antriebszylinders 31 bzw. 32 bildet,
• des Betrages F₃ der Querschnittsfläche der Kolbenstange 38 bzw. 39 und
• des Betrages F₂ der kreisringförmigen Kolbenfläche 54 bzw. 56, welche die axial bewegliche Begrenzung des unteren Antriebsdruckraumes 46 bzw. 47 des jeweiligen Hydrozylinders 31 bzw. 32 bildet und deren Differenz F₁-F₃ der "lichten" Querschnittsfläche des Zylindergehäuses 36 bzw. 37 und der Querschnittsfläche F₃ der jeweiligen Kolbenstange 38 bzw. 39 entspricht, sowie
• der Beträge der maximal ausführbaren Kolbenhübe identisch ausgebildet.

The two drive cylinders 31 and 32 are in terms of their characteristic data

• the amount F 1 of the "upper" piston surface 52 or 53, which forms the axially movable boundary of the upper drive pressure space 43 or 44 of the drive cylinder 31 or 32,
• of the amount F₃ the cross-sectional area of the piston rod 38 or 39 and
• of the amount F₂ of the annular piston surface 54 or 56, which forms the axially movable limit of the lower drive pressure chamber 46 or 47 of the respective hydraulic cylinder 31 or 32 and the difference F₁-F₃ the "clear" cross-sectional area of the cylinder housing 36 or 37 and the Cross-sectional area F₃ corresponds to the respective piston rod 38 or 39, and
• the amounts of the maximum executable piston strokes are identical.

In a typical design of the bending machine 10, the maximum operating pressure Pmax that the pressure supply unit has Can provide 49, 300 bar, and the two drive cylinders 31 and 32 are dimensioned so that they develop a maximum feed force of 500 kN at this output pressure Pmax. The total weight of the moving parts of the bending machine 10 - essentially its upper tool 11 and the pistons 38 and 39 of the two drive cylinders 31 and 32 articulated to it - is in a typical dimensioning 10 kN. The delivery capacity of the pressure supply unit 49 is adapted to this dimensioning of the drive cylinders 31 and 32 in such a way that in the load-down feed operation, ie in the operating state of the bending machine 10 in which the workpiece 13 is deformed, a maximum value of the feed speed 30 mm / s is possible. Furthermore, in a typical design of the bending machine, the area ratio F 1 / F 2 of the large piston surface 52 or 53 acted upon in the load-down operation of the drive cylinders 31 and 32 by the high outlet pressure P of the pressure supply unit 49 to the smaller, annular piston surface 54 or 56 of the flange 41 or 42 of the piston 33 or 34 of the drive cylinder 31 or 32 the value 10/1, so that there is a maximum speed of the rapid upward movement of 300 mm for the rapid upward operation of the bending machine 10 / s results.

For a more detailed explanation of the type of motion control provided for the two drive cylinders 31 and 32, reference should now be made to the relevant hydraulic circuit details 3, in which the part 48 ′ of the electro-hydraulic control unit 48 provided for the movement control of the drive cylinder 31 of the bending machine 10 according to FIG. 1 is shown, with the part of the electro-hydraulic control part provided for the movement control of the right drive cylinder 32 Control unit 48 is identical. 3, which comprises a high-pressure pump 57 with an outlet check valve 58 and a pressure relief valve 59 with an adjustable response threshold, connected on the suction side to the reservoir 51 of the pressure supply unit 49, is used both for the left drive cylinder 31 and for the right drive cylinder 32 the bending machine 10 is used.

The central functional element of the subunit 48 'of the electrohydraulic control unit 48 is a follow-up control valve, designated overall by 61, for the structural and functional explanation of which reference is also made to the relevant details in FIG. 4.

The overrun control valve 61, by means of which the direction and speed of the possible movements of the piston 33 of the drive cylinder 31 can be controlled - rapid feed and load feed movements directed towards the workpiece 13 and rapid upward movements directed away from the workpiece 13 - , its function is a 3/3-way valve, which is connected to the high pressure outlet 62 of the pressure supply unit 49, first (P-) supply connection 63 and a second, connected to the unpressurized reservoir 51 of the pressure supply unit 49 (T-) supply connection 64 and a single control output 66, which on the one hand via a check valve 67, which is due to relatively higher pressure at the control output 66 of the Follow-up control valve 61 is acted upon in the opening direction and, on the other hand, is connected to the control connection 69 of the drive cylinder 31 via a throttle 68, which is connected in parallel with this and designed as an orifice, via which the upper drive pressure chamber 43 of the drive cylinder 31 can be pressurized or depressurized Reservoir 51 of the pressure supply unit 49 can be relaxed. The lower, ring-cylindrical drive pressure chamber 46 of the drive cylinder 31 is permanently connected to the high-pressure outlet 62 of the pressure supply unit 49 via its control connection 71.

The overrun control valve 61 has a neutral position O, a blocking position, in which its control output 66 is blocked both against its P supply connection 63 and against its T supply connection 64.

The rapid and the load feed operation of the drive cylinder 31 is assigned the functional position of the follow-up control valve 61, which is denoted by I in FIGS. 3 and 4, in which the P supply connection 63 of the valve 61 with its control output 66 and thus also with the Control port 69 of the drive cylinder 31 hydraulically is connected while the T supply connection 64 is blocked against the control output 66.

From the workpiece 13 - after its bending processing - rapid upward movements of the piston 33 of the drive cylinder 31 leading away is assigned the functional position of the follow-up control valve 61, designated II in FIGS. 3 and 4, in which the control output 66 of the follow-up control valve 61 is connected to its T supply connection, but is blocked off from the P supply connection 63. In this functional position II of the follow-up control valve 61, in which the lower drive pressure chamber 46 of the drive cylinder 31 is acted upon by the high output pressure of the pressure supply unit 49 and the upper drive pressure chamber 43 of this drive cylinder 31, at least via the throttle 68 and the follow-up control valve 61 with the unpressurized one Storage container 51 of the pressure supply unit 49 is connected, the piston 33 of the drive cylinder 31 moves upward at rapid speed.

The overrun control valve 61 is - for the purpose of explanation - according to the semi-schematic representation of FIG. 4 provided as a slide valve, the "piston" 72 of which is represented by the 3/3-way valve symbol. It is designed as a proportional valve, which, seen from its blocking basic position O, with an increasing displacement of its valve piston 72 to the "left", ie in the sense of utilizing the functional position II - actuation of the drive cylinder 31 in the upward direction - releases an increasingly larger cross section of the flow thread 73, via which pressure medium can flow out of the upper drive pressure chamber 43 of the drive cylinder 31 to the unpressurized reservoir 51 of the pressure supply unit 49, and with increasing displacement of its valve piston 72 to the "right", ie transition to functional position I - Actuation of the drive cylinder 31 in the downward direction - releases increasingly larger cross sections of the flow path 74, via which pressure medium can flow from the pump 57 and also from the lower drive pressure chamber 46 of the drive cylinder 31 to the upper drive pressure chamber 43 of the drive cylinder 31.

In order to be able to control the follow-up control valve 61 explained in this way in its various functional positions O and I or II in the sense of the sequence of the bending process, the following functional and control elements are further provided:

The housing 75 of the follow-up control valve 61, which is only indicated schematically, has a block-shaped central section 76 with a central bore 77, in which a hollow shaft 79 is mounted, which can be rotated about its central longitudinal axis 78 and displaceable along this axis Figure 4 right end, which protrudes from the housing 75, is provided with a drive pinion 80, with which the output gear 81 of an electric motor 82 meshes, and provided at its other end with an internal thread 83 is in meshing engagement with a threaded spindle 84, which is also rotatable about the central axis 78, but is axially immovably mounted on the housing 75 of the follow-up control valve 61, from which it projects with a free end portion 84 ', on which in turn a pinion 86 is arranged, which is in meshing engagement with a rack 87 which, as can be seen in FIG. 3, is coupled in motion to the upper tool 11 and carries out its downward and upward movements, the rack 87 extending parallel to the central longitudinal axis 28 of the drive cylinder 31 extends and is arranged in the immediate vicinity of the same.

This threaded spindle 84 and the rack 87 are the functionally essential elements of a mechanical feedback device, by means of which the position of the upper tool 11 is "reported back" to the relief control valve 61. In the housing 75 of the follow-up control valve 61 is axially displaceable, but non-rotatably, a generally designated 88 yoke-shaped valve actuating member, which has two parallel yoke legs 89 and 91, which are parallel to the central longitudinal axis 78 of the follow-up valve. Control valve extending guide rod 92, which passes through a radially lateral guide bore 93 of the block-shaped, central housing part 76, are firmly connected to one another and are axially supported on the opposite sides of the valve piston 72 via an actuating pin 94 and 96, wherein this support of the yoke legs 89 and 91 on the actuating pins 94 and 96 or the valve piston 72 is fully positive.

The two yoke legs 89 and 91 are aligned with each other, with the central longitudinal axis 78 of the valve housing 75 coaxial bores 97 and 98, the diameter of which is slightly larger than the outer diameter of the hollow shaft 79, so that this with a sufficient for their smooth rotation by this game Bores 97 and 98 of the yoke legs 89 and 91 of the valve actuating member 88 can pass through.

The valve actuation member 88 is axially free of play between radial driving flanges 102 and 103 of the hollow shaft 79 via ball bearings 99 and 101, which impart smooth rotation of the hollow shaft 79 relative to the valve actuation member 88.

The electric motor 82 is designed as a motor with a reversible direction of rotation, for example as a stepper motor or as an AC motor, that is to say as a pulse-controlled motor, which receives output pulses from an electronic control unit 100 received at a first supply connection 105 in the clockwise direction represented by the arrow 104 and by one second supply connection 106 received output pulses of the electronic control unit 103 is driven in the counterclockwise direction represented by arrow 107, the clockwise direction being indicated by arrow 108 is related. For the electric motor 82 it is further assumed that its armature, which is connected to the driven gear 81 in a rotationally fixed manner, with each control pulse which it receives at one of its two supply connections 105 and 106, in one or the other direction of rotation, a rotation corresponding to a constant - small - incremental angle of rotation executes such as an angle of 3.6 °, ie that a hundred pulses are necessary for the armature of the electric motor 82 to rotate through 360 °. The electric motor 82 is used in the follow-up control valve 61 for position setpoint specification control, which is achieved by - program-controlled - sequence of the control pulses which are fed to the electric motor 82 at its one supply connection 105 or at its other supply connection 106.

In the context of the sub-unit 48 'of the electrohydraulic control unit 48 assigned to the left drive cylinder 31, a downstream valve 109, which is connected between the control connection 69 leading to the upper drive pressure chamber of the drive cylinder 31 and the reservoir 51 of the pressure supply unit 49, is provided, via which the piston moves rapidly during downward movements 33 of the left drive cylinder 31 - and the upper tool 11 of the bending machine 10 - pressure medium can flow from the reservoir 51 of the pressure supply unit 49 into the upper drive pressure chamber 43. This post-flow valve 109, which is illustrated to illustrate its function as a springless check valve, is of a construction known per se, so realizes that it responds to minimal pressure differences between the upper drive pressure chamber 43 of the drive cylinder 31 and the reservoir 51 of the pressure supply unit 49 and assumes its blocking position when the pressure in the upper drive pressure chamber 43 of the drive cylinder 31 is slightly greater than in the reservoir 51 of the pressure supply unit 49 and changes to its open position as soon as the pressure in the upper drive pressure chamber 43 of the drive cylinder 31 is slightly less than the pressure in the “unpressurized” reservoir 51 of the pressure supply unit, where “unpressurized” is intended to mean equality with the surrounding atmospheric pressure.

Between the control port 69 communicating with the upper outlet pressure chamber 43 of the drive cylinder 31, i.e. the common connection point of the post-flow valve 109, the throttle orifice 68 and the check valve 67 connected between the control outlet 66 of the follow-up control valve 61 and the control port 69 of the drive cylinder 31, the by a higher pressure at the control outlet 66 of the follow-up control valve 61 than at the control connection 69 in the opening direction and is otherwise blocked and the pressure tank 51 of the pressure supply unit 49, which is depressurized in the above-described sense, is further connected to an outflow valve 111 which is in the rapid retraction operation of the piston 33 Drive cylinder 31 an outflow of pressure medium from the upper drive pressure chamber 43 of the drive cylinder 31 enables directly to the reservoir 51 of the pressure supply unit 49.

The outflow valve 111 is designed as a pressure-controlled 2/2-way valve, the basic position 0 of which is its blocking position, into which it is urged by the valve spring 112. Its second functional position I is a flow position, in which the control connection 69 of the drive cylinder 31 via the flow path 13 of the discharge valve 111 in addition to the flow path leading to the reservoir 51 via the throttle orifice 68 and the follow-up control valve 61, which is designated as a whole by 114 the reservoir 151 of the pressure supply unit 49 is connected. The outflow valve 111 has a first, schematically indicated control pressure chamber 116, which is acted upon by the pressure prevailing at the control outlet 66 of the follow-up control valve 61, and a second, likewise schematically indicated, control pressure chamber 117, which is connected to that in the upper drive pressure chamber 43 of the drive cylinder 31 prevailing pressure is applied. In a special design of the discharge valve 111, the control piston surfaces of the discharge valve 111 provided as a slide valve, which delimit the control pressure spaces 116 and 117, are coordinated with one another in such a way that, assuming the same pressures in the two control pressure spaces 116 and 117, the force resulting from the pressurization of the valve as shown 3 "right" control pressure chamber 117 results and the piston of the outflow valve represented by the circuit symbol 111 tries to urge in its flow position, is equal to the sum of the restoring force of the valve spring 112 and the force resulting from the pressurization of the first control pressure chamber 116, which is "left" as shown in FIG. 3, which urge the piston of the discharge valve 111 into its basic position . With this design of the outflow valve, a minimal pressure drop across the throttle 68 is sufficient to switch the outflow valve 111 to its flow-through position I in the rapid retraction mode of the piston 33 of the drive cylinder 31.

To explain the function of the bending machine 10 described so far in terms of its construction, a typical working cycle of the same will now be considered, that with an upper end position of the pistons 33 and 34 of the two drive cylinders 31 and 32 - and thus also the upper tool 11 of the bending machine 10 Rapid downward operation of these functional elements begins, which is then followed by the load feed operation of drive cylinders 31 and 32, which conveys the deformation of the workpiece, and returns to the assumed starting position with a rapid retraction stroke. It is assumed here that the pressure supply unit 49 is permanently in operation and that the high pressure pump 57 of the pressure supply unit 49 is designed as a regulated pump which only maintains the pressure in the lower drive pressure chambers 46 and 47 of the two drive cylinders 31 and 32 that is required to the pistons 33 and 34 and with them the upper tool 11 in the upper end position hold and in this operating state of the bending machine 10 to compensate for any leakage oil loss that may occur.

In this functional state of the bending machine 10, the overrun control valves 61 and the outflow valves 111 of the electrohydraulic control unit 48 are in the basic positions shown in FIG. 3, whereby the afterflow valves 109 can be open or closed. These post-flow valves 109 are designed so that their valve bodies (not shown in detail) come into the open position due to the installation position of the valves 109 due to the force of gravity acting on them, but with a slight excess pressure in the upper drive pressure spaces 43 and 44 of the two drive cylinders 31 and 32 are "raised" into the locked position. Starting from the assumed starting position of the pistons 33 and 34 of the two drive cylinders 31 and 32 shown in FIG. 1, their rapid downward feed operation is initiated in that the electric motor 82 is actuated with a pulse train, the frequency of which determines a specific feed rate. Speed is linked in the direction of arrow 118 of Figure 3, which has a typical value of 300 mms⁻¹. Assuming that the feed movement of the pistons 33 and 34 should be programmable in increments of 1 μ and that each increment that is fed to the electric motor 82 corresponds to this incremental increment of the feed movement, this means that the electric motor has a pulse repetition frequency of 300 kHz must be controlled so that the feed rate of 300 mms⁻¹ results.

In the case of the special design of the follow-up control valve 61 explained with reference to FIG. 4, it arrives in its functional position I corresponding to the downward feed operation when the valve actuating element 88 moves to the right, which is the case when the electric motor 82 moves in the direction of the Arrow 108 seen, is driven clockwise. For the purpose of explanation, it is assumed that this is the case when the electric motor 82 is supplied with output pulses from the electronic control unit 103 at its first supply connection 105.

The following description of the movement control of the piston 33 of the drive cylinder 31 also applies analogously to the piston 34 of the "right" drive cylinder 32.

When the - rotational - position setpoint movement of the output gear 81 of the electric motor 82 and thus the counterclockwise rotary movement of the hollow shaft 79 starts, the piston 33 of the drive cylinder 31 is initially at rest, with the result that the threaded spindle 84 is also at rest still "stands". In this constellation, the hollow shaft 79 moves according to FIG. 4 in the direction of the arrow 119 to the right, as a result of which the piston 72 of the follow-up control valve 61 carried by the valve actuating element 88 reaches its functional position I, which is assigned to the downward movement control operation of the piston 33 of the drive cylinder 31, the flow cross-section of the flow path 121 of the follow-up control valve 61 released in this functional position I increasing increasing the further the piston 72 is shifted to the right. The consequence of this is that the two drive pressure chambers 43 and 46 of the drive cylinder 31 are now in communicating connection with one another via the flow path 74 of the overrun control valve, which in turn has the consequence that the piston 33 of the drive cylinder 31 now moves in the direction of arrow 118 - "down" - can move.

The high-pressure pump 57 of the pressure supply unit 49 is designed with regard to the maximum delivery capacity in such a way that its delivery volume Q, based on the time unit, has only a value which is sufficient for the load feed operation of the drive cylinder 31, in which the feed speed is only 1/10 of the speed Feed rate, in the selected explanatory example thus 30 mms⁻¹, to maintain the maximum pressure of 300 bar required for the bending or pressing processing of the workpiece 13 in the upper - large area - drive pressure chamber 43 of the drive cylinder 31. For the selected explanatory example, in which each of the two drive cylinders 31 and 32 develops a maximum feed force of 500 kN, the maximum delivery volume Q of the high-pressure pump 57 is approximately 1 ls⁻¹ or 60 l / min, respectively the cross-sectional area of the upper drive pressure chamber 43 or the effective piston surface 52 of the piston 33, which has the value 167 cm² in order to develop a maximum feed force of 500 kN at a maximum operating pressure of 300 bar.

Since the delivery volume of the high-pressure pump 57 related to the time unit plus the volume of pressure medium related to the time unit, which is displaced in the course of the downward movement of the piston 33 of the drive cylinder 31 from its lower drive pressure chamber 46 and via that in the functional position I of the follow-up control valve 61 open flow path 74 of the same can also flow to the upper drive pressure chamber 43 of the drive cylinder 31, is not sufficient to keep this upper drive pressure chamber 43 completely filled with pressure medium, a slight negative pressure arises in the upper drive pressure chamber 43 in the rapid downward operation of the drive cylinder 31, but it does is sufficient to control the afterflow valve 109 into its open position, so that pressure medium can now flow from the reservoir 51 of the pressure supply unit 49 via the afterflow valve 109 into the upper drive pressure chamber 43 of the drive cylinder 31.

As soon as the downward movement of the piston 33 or the upper tool 11 of the bending machine 10 begins, the pinion 86 is seen in the direction of the arrow 121 of FIG. 4 via the rack 87 of the feedback device 84, 86, 87 performing this movement driven by the arrow 122 symbolized clockwise, with the result that the hollow shaft 79 - because of the meshing engagement of the spindle thread 84 with the internal thread 83 of the hollow shaft 79 - is again pulled in the direction of the arrow 123 of Figure 4 to the left, so that the flow cross section of the flow path 74 of the follow-up control valve 61 decreases again, but this remains in its functional position I in accordance with the setpoint value, which is controlled by the electric motor 82.

After this "settling" of the motion control, which requires only a short period of time, there is a stationary state in which the hollow shaft 79 and the threaded spindle 87 are driven at the same speed, that is to say the set target value of the feed rate and the actual value thereof are identical is reached and the opening cross section of the flow path 74 of the follower control valve 61 is just so large that the follower control valve 61 has a throttling effect, as a result of which a pressure is built up in the lower drive pressure chamber 46 of the drive cylinder 31, the amount of which is G / 2 F₂ corresponds to, where G denotes the total weight of the upper tool 11 and the pistons 33 and 34 of the two drive cylinders 31 and 32 connected to it, which is assumed to be 10 kN in the selected explanatory example. For this "equilibrium pressure" thus results when the value F₁ / 10 is set for F₂, a value of about 30 bar, which is also maintained and thus the Piston 33 of the drive cylinder 31 "stops" when, for example, the position setpoint stalls due to a malfunction of the electric motor 82, ie the hollow shaft 79 also stops, whereby a slight further downward movement of the piston 33 of the drive cylinder 31 is sufficient to to switch the follow-up control valve 61 back into its neutral blocking position 0, in which the lower drive pressure chamber 46 of the drive cylinder 31 is shut off against the control output 66 of the follow-up control valve 61, so that no pressure medium can escape from the lower drive pressure chamber 46 and thus that Locking the piston 33 is ensured.

The transition from rapid downward feed operation to load feed operation takes place in that the frequency with which the output pulses are fed to the "forward" control input 105 of the electric motor 82 is reduced to 1/10 of the frequency at which the Output pulses to control the rapid feed operation are given.

The consequence of the feed speed for load downward operation reduced to 1/10 of the previously decisive feed speed, which is now approx. 30 mms⁻¹, is that the delivery capacity of the high-pressure pump 57 - plus the pressure medium volume displaced from the lower drive pressure chamber 46 which can now flow via the flow path 74 of the run-on control valve located in the functional position I,
it is sufficient to convey enough pressure medium into the upper drive pressure chamber 43 of the drive cylinder 31 and to build up a pressure of up to 300 bar sufficient for the bending or pressing processing of the workpiece 13 therein.

• a) Beendigung der Ausgabe von Ansteuerimpulsen, die dem Vorwärts-Steuereingang 102 des Elektromotors 82 zugeleitet werden, nachdem deren Anzahl demjenigen Positions-Soll-Wert entspricht, bei dem die erwünschte Formung des Werkstückes 13 erfolgt sein muß.
• b) Beendigung der Ausgabe von Vorwärts-Impulsen sobald der Auslenkungshub des Ventilbetätigungsgliedes 88 des Nachlauf-Regelventils einen Schwellenwert überschreitet, was ein Indiz dafür ist, daß der Druck im oberen Antriebsdruckraum 43 des Antriebszylinders 31 seinerseits einen Schwellenwert erreicht hat, bzw. ein Indiz dafür, daß der Kolben 33 im Last-Abwärtsbetrieb des Antriebszylinders 31 nicht mehr weiter fahren kann, weil er sich nunmehr - nach Erreichen der Position, in der die Verformung des Werkstücks 13 abgeschlossen ist, über dieses am Maschinengestell 14 abstützt. Zur diesbezüglichen Erfassung des Auslenkungshubes des Ventilbetätigungsgliedes 88 ist beim dargestellten Ausführungsbeispiel gemäß Figur 4 ein Endschalter 124 vorgesehen, der, sobald der Schwellenwert des Auslenkungshubes des Ventilbetätigungsgliedes 88 erreicht ist, ein Signal abgibt, durch das die Ausgabe weiterer Positions-Sollwert-Vorgabeimpulse gestoppt wird.
• c) Begrenzung des Druckes im oberen Antriebsdruckraum 43 des Antriebszylinders 31 mittels eines Druckbegrenzungsventils, als welches das Abströmventil 111 ausnutzbar ist.

The following measures are suitable individually or in combination to end the load feed operation of the drive cylinders 31 and 32:

• a) End of the output of control pulses which are fed to the forward control input 102 of the electric motor 82, after their number corresponds to that desired position value at which the desired shaping of the workpiece 13 must have taken place.
• b) termination of the output of forward pulses as soon as the deflection stroke of the valve actuating element 88 of the follow-up control valve exceeds a threshold value, which is an indication that the pressure in the upper drive pressure space 43 of the drive cylinder 31 has in turn reached a threshold value or an indication thereof that the piston 33 can no longer drive in the load downward operation of the drive cylinder 31, because it is now - after reaching the position in which the deformation of the workpiece 13 is supported - on the machine frame 14. For the relevant detection of the deflection stroke of the valve actuating element 88 is in the illustrated embodiment According to FIG. 4, a limit switch 124 is provided which, as soon as the threshold value of the deflection stroke of the valve actuating element 88 is reached, emits a signal by which the output of further position setpoint input pulses is stopped.
• c) Limiting the pressure in the upper drive pressure chamber 43 of the drive cylinder 31 by means of a pressure limiting valve, as which the outflow valve 111 can be used.

The control of the rapid retraction operation of the two drive cylinders 31 and 32 is achieved in an analogous manner in that - in the case of the left drive cylinder 31 - control pulses are fed to the setpoint control motor 82 at its second supply connection 106 - the upward control connection - whereby the electric motor 82 is driven in a counterclockwise direction 107 and the follow-up control valve 61 reaches its functional position II, in which its control output 66 is connected via the flow path 73 of the valve piston 72 to the unpressurized reservoir 51 of the pressure supply unit 49 and the high pressure pump 57 with all of it Delivery volume works in the lower drive pressure chamber 46 of the drive cylinder 31, the piston 33 of which is thereby moved upwards in rapid upward operation. In this operating state occurs via the throttle 68, via which the control connection 69 of the upper drive pressure chamber 43 of the drive cylinder 31 is connected to the control output 66 of the follow-up control valve 61 Pressure drop, which leads to a switchover of the discharge valve 111 into its flow position I, so that - in rapid upward operation of the drive cylinder 31 - pressure medium can also flow "directly" from the upper drive pressure chamber 43 via the discharge valve 111 to the unpressurized reservoir 51 of the pressure supply unit 49 .

In the context of the follow-up control valve 61, a switch-off element, which acts as an "end switch" and is shown as an inductive transmitter 126, is also provided, which emits an electrical signal for switching off upward control pulses for the electric motor 82 when the deflection stroke of the valve actuating element 88 in the direction of the arrow 123 in turn exceeds a threshold value with which an excessive value of the drive pressure is linked in the lower drive pressure chamber 46 of the drive cylinder 31. The occurrence of excessive pressure in the upward operation is an indication in the majority of the statistically significant operating situations that the drive cylinder 31, e.g. due to uneven feed speed of the pistons 33 and 34 of both drive cylinders 31 and 32 "sticks", which is an overall "dangerous" operating situation for the bending machine 10, which must be ended for safety reasons.

If, during load feed operation of the bending machine 10, its drive cylinders 31 and 32 develop their maximum feed force, 500 kN per drive cylinder, this leads to even with the most stable design of the side cheeks 17 and 18 of the machine frame 14 for a vertical expansion of the side cheeks, even if the vertical distance of the upper cross legs 26 and 27 of the side cheeks 17 and 18 from their lower horizontal cross legs 21 and 22 only by a few tenths of a millimeter would lead to a considerable error of the actual position feedback provided by the threaded spindle 84 of the follow-up control valve 61 and the toothed rack 87, which is coupled in motion to the upper tool 11 of the bending machine, such that the follow-up control valve 61 positions "would see", which have not yet been reached, since the rack 87 is not stretched, with the result that considerable delays in the load feed operation - the deformation operation of the bending machine 10 - would have to be accepted.

In order to avoid such errors of the actual position value feedback as much as possible, the rack 87 of the respective mechanical feedback device of the overrun control valve 61 is provided with a compensation cylinder, designated overall by 127, which, with increasing pressure in the upper drive pressure chamber 43 of the drive cylinder 31, is one of the linked thereto vertical expansion of the respective side cheek 17 or 18 of the machine frame 14 corresponding extension of the rack 87 mediated.

The compensation cylinder 127 is designed as a single-acting linear cylinder, the housing 128 of which has a Ball joint 129 is connected to the upper tool 11 and carries out its downward and upward movements. A pressure chamber 132 is delimited in the axial-vertical direction by the bottom 129 of the cylinder housing 128 and a piston 131 that can be displaced in a pressure-tight manner. The pressure chamber 132 is connected to the control connection 69 of the upper drive pressure chamber 43 of the drive cylinder 31 via a control line 133 that is flexible at least in sections. The piston 131 is urged into the basic position linked to the minimal volume of the pressure chamber 132 of the compensating cylinder 127 by a prestressed coil spring 134, which is supported on the one hand on the piston 131 and on the other hand on an end end wall 136 of the cylinder housing 128 opposite the bottom 129 of the cylinder housing 128. The toothed rack 87 is designed as a free end section of a piston rod 138 which is firmly connected to the piston 131 and emerges vertically through a central bore 137 in the end end wall 136 of the cylinder housing 128.

Bias and spring rate of the coil spring 134 are matched to one another and also adapted to the effective area of the piston 131 of the compensating cylinder 127, that if and for as long as the pressure in the upper drive pressure chamber 43 of the drive cylinder 31 is equal to or approximately the same as that in the reservoir 51 of the pressure supply unit 49 prevailing ambient pressure - atmospheric pressure - is, the connection between the rack 87 and the upper tool 11, seen in the vertical direction, is sufficiently rigid by one to ensure exact feedback of the position of the upper tool 11 in the rapid operating states of the drive cylinder 31 and, on the other hand, to ensure that the displacement of its piston 131 which occurs when the pressure chamber 132 of the compensation cylinder 127 is pressurized corresponds to the amount after the expansion of the side cheeks 17 and 18, which experience this under the development of the feed force of the drive cylinders 31 and 32.

The variant shown in FIG. 5, to the details of which is now referred to, the compensation device, again designated overall by 127, is structurally and functionally largely identical or analogous to that explained with reference to FIG. 3. To the extent that the same reference numerals are given in FIG. 5 as in FIG. 3, this is intended to indicate the structural and functional identity or analogy and also the reference to the explanation already given with reference to FIG. 3 of the parts designated in this way include.

In the variant of the arrangement of the compensation cylinder 127 representing the compensation device shown in FIG. 5, this is firmly connected via a block-shaped connecting part 141 to the portion emerging from the housing 36 of the drive cylinder 31 with the piston rod 33. An elongated blind bore 142 is made "from above" in this, which is connected to the upper one via an upper, radial transverse bore 143 Drive pressure chamber 43 of the drive cylinder 31 is in communicating connection. In the lower end section 143 'of this elongated blind bore 142, a lower radial transverse bore 144 of the piston rod 33 opens, to which the connection block 141 is connected in a pressure-tight manner and which has a transverse channel 146 which continues the radial transverse bore of the cylinder piston, via a short longitudinal channel 147 into the pressure chamber 132 of the compensation cylinder 127 opens, so that its pressure chamber 132 is communicatively connected to the upper drive pressure chamber 43 of the drive cylinder 31 via the longitudinal channel 147, the transverse channel 146, the transverse bore 144, the longitudinal bore 142 and the upper transverse bore 143.

This "internal" connection, as it were, of the upper drive pressure chamber 43 to the pressure chamber 132 of the compensation cylinder 127 avoids flexible line sections of a control line 133 provided as in the exemplary embodiment according to FIG. 3.

In a further difference from the exemplary embodiment according to FIG. 3 and with regard to the design of the follow-up control valve 61 according to FIG. 4, the follow-up control valve provided in connection with the compensation cylinder 127 according to FIG. 5 has a mechanical feedback element which is connected to the piston 131 of the compensation cylinder 127 is connected non-displaceably, a non-rotatably arranged on the housing 74 'of the follow-up control valve 61' threaded spindle 84 '' which is of sufficient length is so that it can remain in meshing engagement within the entire possible stroke with the internal thread of the hollow shaft 79 coaxially surrounding it of the follow-up control valve 61 ', which is arranged here with its longitudinal axis 78 parallel to the central longitudinal axis 28 of the drive cylinder 31.

Furthermore, the driven gear 81 of the electric motor 82 provided for setting the setpoint value is coupled in motion to the drive gear 80 of the hollow shaft 79 via a toothed belt 148.

Otherwise, the follow-up control valve 61 'is structurally identical to that shown in FIG. 4.

As a safety device that, in the event of a power supply failure, e.g. An inertia-related coupling 139 is further provided, due to the inertia-related further rotation of the setpoint-setting motor 82, which could cause undesired further movement of the pistons 33 and 34 of the drive cylinders 31 and 32, which could be dangerous for an operator in particular in the rapid downward operation of the bending machine which automatically changes into its blocking position mediating the fixing of the rotor of the electric motor 82 in the event of a power failure.

Claims (16)

1. Hydraulically driven machine for cold-work processing, in particular for bending or pressing, a sheet metal part, which is arranged between a bottom tool (12) and a top tool (11) which is displaceable vertically up and down relative to the bottom tool, with there being provided for driving said tools two drive cylinders (31, 32) in the form of differential cylinders, which are disposed laterally spaced apart from one another and whose pistons (33, 34) have piston rods (38, 39), which exit in a downward direction out of the cylinder housings (36, 37) and to which the top tool (11) - for compensating slight stroke differences of the pistons (33, 34) - is connected by monoaxial joints, whose joint axes extend at right angles to the plane marked through the central axes of the drive cylinders, wherein the ratio F₁/F₂ of the areas F₁ and F₂ of the pistons, with which said pistons each delimit a top drive pressure chamber (43, 44) from a bottom drive chamber (46, 47) axially penetrated by the piston rod, through the - valve-controlled - joint or alternative pressurization and/or pressure relief of which drive pressure chambers- rapid feed movements towards the workpiece (13) and load feed movements as well as rapid return movements away from the workpiece are controllable, has a value between 5 and 20, wherein in the rapid feed mode the top drive pressure chamber (43, 44) delimited with the larger area is connected by a suction valve (109) to the unpressurized storage tank (51) of the pressure supply unit and the bottom drive pressure chamber (46, 47) is connected by a control element (48) having an adjustable throttling action to the top drive pressure chamber (43, 44) of the drive cylinder (31, 32), in the load feed mode the bottom and the top drive pressure chamber are jointly pressurized and in the rapid return mode the bottom drive pressure chamber (46, 47) is highly pressurized, while the top drive pressure chamber is relieved of pressure and, for controlling the piston movements and their synchronization, an electro-hydraulic control unit (48) with electrically controllable position setpoint selection is provided, which compares the actual position of the pistons (33, 34) and of the top tool (11) connected to said pistons with the setpoint selection, characterized in that provided in the electro-hydraulic control unit (48) as a control element for each of the two drive cylinders (31 and 32) is a servo-control valve (61) in the form of a proportional valve, which operates with mechanical feedback of the actual value of the position of the piston (33 or 34) and is operable as a 3/3-way valve having a blocking position (O), in which the control output (66) of said servo-control valve (61) connected to the control connection (69) of the top drive pressure chamber (43) of the associated drive cylinder (31 or 32) is blocked off both from the high-pressure output (62) of the pressure supply unit (49) and from the storage tank (51) of said unit, having a first throughflow position I which is associated both with the rapid feed and load feed modes of the drive cylinder (31 or 32) and in which the P supply connection (63) connected to the high-pressure output (62) of the pressure supply unit (49) is connected to the control output (66) of the servo-control valve (61) and said control output is blocked off from the T supply connection (64) of the servo-control valve (61) connected to the unpressurized storage tank (51) of the pressure supply unit (49), as well as having a second throughflow position II which is associated with the rapid return mode of the drive cylinder (31 or 32) and in which the control connection (69) of the top drive pressure chamber (43) of the drive cylinder (31 or 32) is connected to the storage tank (51) and is blocked off from the high-pressure output (62) of the pressure supply unit (49), that the high-pressure output (62) of the pressure supply unit (49) is permanently connected to the bottom drive pressure chamber (46) of the drive cylinder (31 or 32), and that provided in hydraulic parallel connection to the secondary flow valve (109) is a flow-off valve (111) in the form of a 2/2-way valve which, during rapid return movement phases of the piston (33 or 34) of the drive cylinder (31 or 32), may be moved into its operating position I effecting its direct connection to the storage tank (51) and otherwise assumes its - blocking - basic position O.
2. Machine according to claim 1,
   characterized in that
   the maximum delivery quantity per unit of time of the high-pressure pump (57) of the pressure supply unit (49) is at least and approximately equal to the pressure medium quantity which has to be displaced into the top drive pressure chamber (43) of the drive cylinder (31 or 32) in order for the piston (33 or 34) of said drive cylinder (31 or 32) to execute its load feed movement at the appropriate feed rate for said purpose.
3. Machine according to claim 2,
   characterized in that
   the high-pressure pump (57) of the pressure supply unit (49) takes the form of a pressure-controlled pump, whose delivery quantity is inversely proportional to the output pressure generated at its output.
4. Machine according to claim 2 or 3,
   characterized in that
   a pressure relief valve (59) connected in parallel to the high-pressure pump (57) of the pressure supply unit (49) is provided with an adjustable response threshold value of the pressure.
5. Machine according to one of claims 1 to 4,
   characterized in that
   the flow-off valve (111) takes the form of a solenoid valve which onwards from a minimum frequency of electric position setpoint selection pulses, which are produced for controlling rapid return movements of the drive cylinder pistons (33 and 34) and by means of which an electric motor (82) provided for setpoint selection control of the servo-control valve (61) may be activated, is moved into its throughflow position I.
6. Machine according to one of claims 1 to 4,
   characterized in that
   the flow-off valve (111) takes the form of a pressure-controlled valve which as a result of pressurization of a first control pressure chamber (116) is urged into its - blocking - basic position O and as a result of pressurization of a second control pressure chamber (117) is urged into its throughflow position I, that the first control pressure chamber (116) is connected to the control output (66) of the servo-control valve (61) and the second control pressure chamber is connected to the control connection (69) of the top drive pressure chamber (43) of the drive cylinder (31 or 32), that connected between said control pressure chambers and the control output (66) of the servo-control valve (61) are a throttle (68) and, parallel to said throttle, a non-return valve (67) which is held in its blocking position by higher pressure in the top drive pressure chamber (43) than at the control output (66) of the servo-control valve (61), and that the area - which may be acted upon by the pressure prevailing at the control output (66) of the servo-control valve (61) - of the control piston movably delimiting the second control pressure chamber (117) of the flow-off valve (111) corresponds at least and is approximately equal to the area - which is acted upon by the pressure prevailing in the top drive pressure chamber (43) of the drive cylinder (31 or 32) - of the control piston movably delimiting the first control pressure chamber, and that a valve spring (112) which urges the flow-off valve (111) into its - blocking - basic position (O) is provided, whose restoring force is equivalent to a small fraction, e.g. 1/20 to 1/5, of a pressure drop occurring above the throttle (66).
7. Machine according to claim 6,
   characterized in that
   the areas - which may be acted upon by control pressure - of the control pistons of the flow-off valve (111) and the initial tension of the valve spring (112) are tuned to one another in such a way that the flow-off valve (111) is switched over into its throughflow position I as soon as the pressure drop occurring above the throttle (68) exceeds a value of 1 bar.
8. Machine according to one of claims 1 to 7,
   characterized in that
   the position actual-value feedback device (86, 87) of the servo-control valve (61) is provided with a compensation device (127) which effects a raising of the reference position, to which the position feedback is referred, corresponding to the amount of a vertical expansion of the machine frame (14, 23, 24) which said frame experiences under the feed forces arising at least in the load feed mode.
9. Machine according to claim 8, in which the position actual-value feedback device comprises a pinion, which is non-rotatably connected to a feedback spindle of the servo-control valve, and a toothed rack which meshes with said pinion, or a threaded spindle (84'') which is meshed with a hollow-shaft-like spindle nut, by means of whose motor-controlled rotators the position setpoint selection is effected, and simultaneously executes the movements of the piston of the drive cylinder and extends parallel to the longitudinal axis of said cylinder,
   characterized in that
   the toothed rack (87) or the threaded spindle of the mechanical feedback device (86, 87) of the servo-control valve (61) is provided with a compensating cylinder (127) which, as pressure increases in the top drive pressure chamber (43) of the drive cylinder (31 or 32), effects a lengthening of the toothed rack (87) or the threaded spindle (84'') corresponding to the vertical expansion of side pieces (17 and 18) of the machine frame (14) linked with said pressure increase.
10. Machine according to claim 9,
    characterized in that
    the compensating cylinder (127) takes the form of a single-acting hydraulic linear cylinder, whose housing (128) is connected non-displaceably to the piston rod (33 or 34) of the drive cylinder (31 or 32) or in the direct vicinity of the cylinder axis (28 or 29) to the top tool (11), with the base (129) of the cylinder housing (128) and a piston (131) displaceable in a pressure-sealed manner in said housing delimiting movably in an axial-vertical direction a pressure chamber (132) which
       is held in communication with the top drive pressure chamber (43) of the drive cylinder (31 or 32), that the piston (131) is urged by a preloaded restoring spring (134), which is supported at one end on the piston (131) and at the other end on an end wall (136) of the cylinder housing (128) lying opposite to the base (129) of the cylinder housing (128), into the basic position linked with a minimal volume of the pressure chamber (132) of the compensating cylinder (127), and that the toothed rack (87) or the threaded spindle (84'') takes the form of a free end portion of a piston rod (138) firmly connected to the piston (131) and exiting vertically out of the cylinder housing (128), with the initial tension and spring rate of the helical spring (134) and the effective area of the piston (131) of the compensating cylinder (127) being tuned to one another in such a way that the displacement of the piston (131) occurring upon pressurization of the pressure chamber (132) of the compensating cylinder (127) corresponds in its amount to the expansion of the side pieces (17 and 18) which said side pieces experience under the feed force development of the drive cylinders (31 and 32).
11. Machine according to claim 10,
    characterized in that
    the pressure chamber (132) of the compensating cylinder (127) is connected by an, at least in sections, flexible control line (133) to the control connection (69) of the top drive pressure chamber (43) of the drive cylinder (31 or 32).
12. Machine according to claim 10,
    characterized in that
    the housing (128) of the compensating cylinder (127) is firmly connected to the piston rod (33 or 34) of the drive cylinder (31 or 32) by a block-shaped connection part which is provided with a line bore, which communicates, on the one hand, with a line bore axially penetrating the piston (33 or 34) and opening into the top drive pressure chamber (43) of said piston and, on the other hand, with the pressure chamber (132) of the compensating cylinder (127).
13. Machine according to one of claims 1 to 12,
    characterized in that
    the servo-control valve (61) is equipped with a pickup system (124, 126) which produces electrical output signals, which may be used as warning signals and/or as control signals by means of which the machine (10) may be brought to a standstill, when the excursion stroke of the valve body (72) of the servo-control valve (61), in the operating position associated with the feed mode of the drive cylinders (31 and 32) or in the operating position associated with the return mode, exceeds a predetermined threshold value.
14. Machine according to claim 13,
    characterized in that
    the output signals produced by pickups (124, 126) of the pickup system are according to their level and/or frequency a measure of the amount of the excursion of the valve body (72) of the servo-control valve (61) in the operating positions (I and II) of said valve.
15. Machine according to claim 14,
    characterized in that
    at least in the rapid feed mode and in the rapid return mode of the drive cylinders (31 and 32), the output frequency of the position setpoint selection pulses is increasingly reduced with progressive excursion of the valve body (72) of the servo-control valve (61).
16. Machine according to one of claims 1 to 15,
    characterized in that
    the electric motor (82) provided for position setpoint selection control is provided with a coupling (139) or brake which is releasable by electrical control and which, in the event of a power failure or an interruption of the electrical setting into the release position, automatically moves over into a blocking position effecting arresting of the rotor of the electric motor (82).

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