EP0339247B1 - Hydraulic steering device for the drive control of a double-acting hydraulic cylinder - Google Patents

Abstract

In a hydraulic control device for the drive control of a double-acting hydraulic cylinder with a relatively large driving face and a relatively small counterface on its driving piston, a direction- and- motion-control valve is provided in the form of a follow-up control valve which operates with electrically controllable desired position value stipulation and mechanical actual position value feedback. The pressure supply unit provides two supply pressures PN and PH of different magnitude. A pressure-controlled pressure changeover valve is provided which switches over to high supply pressure as required if the load increases. Also provided is a face changeover valve, likewise pressure-controlled, which, after the pressure changeover valve has switched over to higher supply pressure, switches over from differential operation of the hydraulic cylinder to pressurisation on one side of its piston, on the larger piston face of the latter, or, if the requirement for propulsive force has decreased again, after the pressure changeover valve has for this reason switched back to lower supply pressure, for its part switches back to differential operation of the hydraulic cylinder. <IMAGE>

Description

The invention relates to a hydraulic control device for the drive control of a double-acting hydraulic cylinder, which is provided as a drive element for the tool of a processing machine, with which a workpiece can be subjected to a punching or a shaping cold deformation, and with the others, in the preamble of patent claim 1 mentioned generic features.

Such a known control device is in connection with a hydraulic drive device, in which a linear hydraulic cylinder designed as a differential cylinder is provided as the drive element, in which the area ratio F _A / F _{G of} its larger drive area F _A and its smaller drive or counter surface F _{G is} about 1/3, described in DE 37 35 123 A1.

The rapid feed movements of the hydraulic cylinder piston and the tool, by means of which the latter is fed to the workpiece and the workpiece is partially also machined, take place in a feed operation in which both surfaces F _A and F _{G of} the hydraulic cylinder piston, the first mentioned via a directional control valve , the second on a pressure controlled Valve element of a surface switch valve to be pressurized. If the force that the piston develops here - in differential operation - is not sufficient for penetrating machining of the workpiece, the surface switch valve is controlled by the pressure in the smaller drive pressure chamber of the hydraulic cylinder as soon as this pressure exceeds a threshold value of a predetermined amount is lower than the maximum value of the output pressure of the pressure supply unit, switched to its position assigned to the load feed operation, in which the smaller drive pressure chamber of the hydraulic cylinder is now depressurized and only the larger drive pressure chamber via the directional and movement control valve is connected to the pressure outlet of the pressure supply unit, at which a high pressure with a typical value of 200 bar is provided. So that the surface switching valve, when the need for feed force is only slightly greater than the force that can be reached in differential operation, does not simply switch back and forth repeatedly, which in unfavorable cases not only delays the work process, but even could lead to the piston "stopping", the surface switch valve is designed so that it only switches back to differential operation after the need for feed force has become lower by a predetermined safety margin than that in differential operation of the Hydraulic cylinder maximum achievable feed force.

The hydraulic actuating device according to the older patent application DE-A-37 35 123.0 works satisfactorily insofar as in numerous cases in which the feed force which can be achieved in the differential operation of the hydraulic cylinder is approximately sufficient and only in rare cases has to be switched over to the operation with one-sided pressurization of the hydraulic cylinder piston , favorable short cycle times can be achieved. In order to be able to take advantage of such short cycle times, which is particularly advantageous for punching thin steel sheets and yet has sufficient power reserves available for processing "thicker" steel sheets, the area ratio of the hydraulic cylinder must be chosen to be relatively large, but this has the consequence that when the surface switching valve has responded, a correspondingly large increase in the feed force which can now be achieved is achieved, which, even while the tool is being pierced, can already lead to considerable acceleration of the hydraulic cylinder piston, which can then be achieved by switching the surface switching valve back in Whose function position assigned to the differential operation of the hydraulic cylinder has to be caught again, which can lead to considerable impacts, the more the "easier" the tool could pierce the workpiece and accordingly a "runaway" of the drive cylinder can occur before the intercepted by switching back to differential operation.

Although it could be considered to mitigate such shocks by using a controllable valve as a direction or movement control valve is used. However, as the only measure, this would not be able to contribute appreciably to a gentler execution of a work cycle of the aforementioned type, not even if a follow-up control valve with mechanical position actual value feedback were used, since the control frequency of such a valve is still relatively low would be the period of time within which such vibrations can occur and would therefore be practically unavoidable.

It could also be considered to use a hydraulic cylinder with only one drive surface and one counter surface instead of a hydraulic cylinder, which has two drive surfaces, one of which can be switched on or off to increase or decrease the feed force, and the pressurization of these work surfaces via a caster -Control control valve. However, this would require an extremely complex design of the hydraulic cylinder itself, which could then no longer be designed as a standard unit, for whose use only a suitable control periphery is required.

The object of the invention is therefore to improve a control device of the type mentioned in such a way that its use in combination with a simple differential cylinder as the drive element of a hydraulic drive device with a simple overall structure, low-vibration operation of the same can be achieved, if necessary, even if with the machine equipped with the drive must be operated with a fast work cycle sequence.

This object is achieved by the features (e) to (h) mentioned in the characterizing part of patent claim 1.

• Auslegung des Druckversorgungs-Aggregates dahingehend, daß Versorgungsdrücke auf verschiedenem Niveau, dem niedrigen Niveau P_N und dem hohen Niveau P_H bereitgestellt werden,
• selbsttätige Umschaltung auf das bedarfsgerechte Versorgungs-Druckniveau mittels eines druckgesteuerten Umschaltventils, das sehr schnell umschaltet,
• Umschaltung des Flächen-Umschaltventils jeweils erst dann, nachdem das Druck-Umschaltventil, sei es im Sinne einer Erhöhung des Versorgungs-Druckniveaus oder einer Erniedrigung des Versorgungs-Druckniveaus umgeschaltet hat, sowie
• in Kombination hiermit, stufenlos bedarfsgerechte Einstellung des im größeren Antriebsdruckraum des Antriebs-Hydrozylinders herrschenden Betriebsdruckes P_A mittels eines Nachlauf-Regelventils, das mit elektrischer Positions-Sollwert-Vorgabe und mechanischer Positions-Istwert-Rückmeldung arbeitet, wird insgesamt ein weitgehend erschütterungsfreier "sanfter" Ablauf der Arbeitszyklen erzielt, da in dieser Kombination der Maßnahmen das Nachlauf-Regelventil, weil es, dank der genannten Druckumschaltung, im Ergebnis mit kleineren Verstellhüben seiner Durchfluß-Ventilelemente auskommt, "schneller" reagieren und damit höhere Regelfrequenzen ermöglichen kann, was wiederum der Vermeidung ruckartiger Kolben- und Werkzeugbewegungen zugute kommt. Die erfindungsgemäße Steuereinrichtung vermittelt daher auch bei rascher Arbeitszyklenfolge einen komfortablen und geräuscharmen Betrieb der Maschine.

Through the combination of measures that can be easily implemented, namely

• Design of the pressure supply unit in such a way that supply pressures are provided at different levels, the low level P _N and the high level P _H ,
• automatic switchover to the demand-based supply pressure level by means of a pressure-controlled switchover valve that switches very quickly,
• Switchover of the area switchover valve only after the pressure switchover valve has switched, either in the sense of an increase in the supply pressure level or a decrease in the supply pressure level, and
• in combination with this, infinitely variable adjustment of the operating pressure P _A prevailing in the larger drive pressure chamber of the drive hydraulic cylinder by means of a follow-up control valve that works with electrical position setpoint specification and mechanical position actual value feedback, overall a largely vibration-free "gentle" Work cycle is achieved because in this combination of measures, the overflow control valve, because thanks to the pressure switch mentioned, it results in smaller ones Adjustment strokes of its flow valve elements, "react faster" and thus enable higher control frequencies, which in turn helps to avoid jerky piston and tool movements. The control device according to the invention therefore conveys comfortable and quiet operation of the machine even in the case of a rapid work cycle sequence.

In addition, the "interception" of the drive cylinder piston in the last phase of the machining of the workpiece only occurs when the pressure supply is switched back to the low pressure level P _N , which also facilitates this interception.
If one assumes that the control device according to the older patent application, depending on the application, can be equipped with a follow-up control valve as a directional control valve, be it for reasons of simple controllability of the movement sequence when using a CNC control, this results For a technical additional effort to be compared with this, which is necessary to implement a control device according to the invention, only that the pressure supply unit must be designed for two output pressure levels and a pressure changeover valve arrangement must be present in order to be able to use these different output pressure levels as required. From a technical point of view, however, this additional effort is low, so that the control device according to the invention can be regarded as "simple" and the overall costs incurred for the drive device and the control device are not significantly increased by this additional technical effort, especially since it was made possible by the control device according to the invention , low vibration Operation has a wear-reducing effect and therefore slightly higher investment costs can be compared with significantly lower operating costs, which more than compensate for additional investment.

In the design specified by the features of claim 2, the pressure changeover valve arrangement can be realized by means of a simple check valve, which, as it were, mediates the decoupling of the pressure outputs of the pressure supply unit and by means of a simple pressure-controlled 2/2-way valve the features of claim 3 a typical design and configuration is given as a slide valve, which can be equipped with a return spring adjustable bias to adjust the restoring force and adaptation to the desired switching threshold.

In the design of such a pressure changeover valve specified by the features of claim 4, the lower outlet pressure of the pressure supply unit can be used in a simple manner to generate the restoring force that is required for setting the changeover threshold.

The preferred design of the pressure changeover valve according to claim 5 has the advantage that no resilient resetting elements are required to set the pressure changeover valve to the required pressure changeover threshold. There is also at least one wear part that is otherwise exposed to considerable loads.

The features of claim 6, on the one hand, provide a structurally simple design of the surface changeover valve and dimensioning regulations for the design of control surfaces of a control valve piston and the clear cross-sectional area of the valve channel of a seat valve of the surface changeover valve, when observed, a high reliability of the function Control is achieved.

Finally, the features of claims 7 to 11 indicate special designs and dimensions of the pressure changeover valve and the surface changeover valve, the drive hydraulic cylinder and the pressure supply unit, which have proven to be particularly advantageous in combination with the control device according to the invention.

Fig. 1

ein Hydraulikschema einer erfindungsgemäßen Steuereinrichtung für eine hydraulische Antriebsvorrichtung mit einem Zwei-Kammer-Hydrozylinder als Antriebselement, der mittels eines Flächen-Umschaltventils von Differential-Betrieb auf einen Betrieb mit einseitiger Druckbeaufschlagung seines Kolbens auf dessen größerer Antriebsfläche umschaltbar ist, und

Fig. 2 und 3

verschiedene Funktionsstellungen des Flächen-Umschaltventils der Steuereinrichtung gemäß Figur 1 in, vergrößertem Maßstab.

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

Fig. 1

a hydraulic diagram of a control device according to the invention for a hydraulic drive device with a two-chamber hydraulic cylinder as the drive element, which can be switched from differential operation to operation with one-sided pressurization of its piston on its larger drive surface by means of a surface switch valve, and

2 and 3

Various functional positions of the surface switching valve of the control device according to Figure 1 in an enlarged scale.

The purpose of the hydraulic control device according to the invention shown in FIG. 1, to the details of which is expressly referred to, designated overall by 10, is the need-based pressurization and / or relief of the drive pressure chambers 11 and / or 12 of a double-action system, designated overall by 13 , linear hydraulic cylinder, which in a punching or embossing machine, more generally a processing machine, by means of which a workpiece 14, for example a steel plate, can be subjected to a piercing - punching - or a shaping cold deformation, as a drive element for the tool 16 thereof Machine is provided which, in the course of a working cycle of this machine, executes a rapid feed movement toward the workpiece 14, through which the tool 16 comes into contact with the workpiece 14, thereafter - if necessary, increasing the force acting in the feed direction and reducing the Advance b-speed executes its load feed movement, which mediates the machining of the workpiece 14, and then, after the workpiece 14 has undergone its desired deformation, is brought back again - in a rapid retraction movement - to its basic position assumed at the beginning of the working cycle, whereby this Rapid retraction movement takes place again with reduced force development of the drive element 13 but with increased movement speed of the tool 16.

"Needs-based" in the above sense of this term is intended to mean that both the greatest possible minimization of the drive energy required to carry out the work cycles and a minimization of the cycle times required for the individual machining processes are aimed at, with priority being given to reducing the cycle times.

The hydraulic cylinder 13 is, without restriction of generality, provided as "standing", i.e. with a vertical course of its central longitudinal axis 17 with respect to a horizontally arranged machine table 18, by which a machine frame, not otherwise shown, is to be represented, on which, fixed to the frame, the housing 19 of the hydraulic cylinder 13 is also permanently mounted.

The workpiece 14 resting on the machine table 18 can be fixed to the machine table 18 by means of a holding device (not shown) or relative to it, according to a "processing path", e.g. be numerically controlled, movable.

The hydraulic cylinder 13 is designed as a differential cylinder, the piston, which is designated as 21 and can be displaced up and down within the cylinder housing 19 within the cylinder bore 22, delimits the two drive pressure spaces 11 and 12 in a pressure-tight manner by means of their valve-controlled, common or alternative action the outlet pressure P _N or P _H of a pressure supply unit designated overall by 23 and, if necessary, pressure relief of each of the two drive pressure chambers 11 or 12, the feed and retraction strokes of the piston 21 or of the tool 16 required for machining the workpiece 14 can be controlled as required in the above sense.

The pressure supply unit 23 has a first supply pressure outlet 24, at which a relatively low supply pressure P _{N is} provided, which has a typical value of 60 bar, and a second supply pressure outlet 26, at which a significantly higher pressure P _{H is} provided which has a typical value of 180 bar.

The pressure supply unit 23, based on its special design because it can be assumed to be known, is regarded as sufficiently explained by this — for example — design.

The effective amount of the, according to the representation of FIG. 1, upper drive pressure chamber 11 of the hydraulic cylinder 13 movably limiting - larger - piston area 27 of the piston 21 of the hydraulic cylinder 13 is equal to the cross-sectional area F 1 of the bore 22 of the cylinder housing 19th

gegeben ist.By acting on this drive pressure chamber 11 with the output pressure P _A (P _A = P _N or P _H ) of the pressure supply unit 23, a force acting on the piston 21 in the direction of arrow 28, ie directed toward the workpiece 14, is exerted K 1, the force of which Amount through the relationship

${\text{K₁ = F₁. P}}_{\text{A}}\text{(1)}$

given is.

wobei mit P_G der Druck bezeichnet ist, der sich in Abhängigkeit von der wirksamen Last-Gegenkraft und dem in den größeren Antriebsdruckraum 14 eingekoppelten Betriebsdruck P_A im unteren, kleineren Antriebsdruckraum 12 ergibt, und mit F₂ die wirksame Querschnittsfläche der gegenüber der Zylinderbohrung 22, in welcher der Zylinderkolben 21 mit seiner größeren Kolbenstufe 31 der Fläche F₁ druckdicht verschiebbar geführt ist, durch eine innere Gehäusestufe 32 abgesetzten - engeren - Bohrungsstufe 33 des Gehäuses 19 bezeichnet ist, in welcher die mit der größeren Kolbenstufe 31 fest verbundene, mit diesem z.B. einstückig ausgeführte, kleinere, stangenförmige Kolbenstufe 34 druckdicht verschiebbar geführt ist, an deren unterem, freiem Ende das Werkzeug 16 befestigt ist.By simultaneously or alternatively acting on the lower drive pressure chamber 12 of the hydraulic cylinder 13 as shown in FIG. 1, a force K 2 acting in the direction of arrow 29, ie in the opposite direction, is exerted on its piston 21, the amount of which by the relationship

${\text{K₂ = F₁. P}}_{\text{H}}{\text{- (F₁ - F₂). P}}_{\text{G}}{\text{= F₁. P}}_{\text{H}}{\text{- F₃. P}}_{\text{G}}\text{(2),}$

where P _G denotes the pressure which results as a function of the effective load counterforce and the operating pressure P _A coupled into the larger drive pressure chamber 14 in the lower, smaller drive pressure chamber 12, and with F₂ the effective cross-sectional area of the cylinder bore 22, in which the cylinder piston 21 with its larger piston stage 31 of the surface F 1 is guided so as to be pressure-tight, by means of an inner housing stage 32 offset - narrower - bore stage 33 of the housing 19, in which the one with the larger piston stage 31 firmly connected, for example in one piece executed, smaller, rod-shaped piston stage 34 is slidably guided pressure-tight, at the lower, free end of the tool 16 is attached.

With F₃ the effective amount of the substantially circular "differential area" 36 is referred to, on which the coupled into the lower drive pressure chamber 12 pressure P _G (P _N ≦ P _G <P _H ) on the cylinder piston 21 in the sense of generating the - upwards directed - force K₂ acts.

bzw.

${\text{K}}_{\text{3H}}{\text{= F₁ . P}}_{\text{H}}{\text{- F₃ . P}}_{\text{G}}\text{ (4)}$

gegeben.If the hydraulic cylinder 13 is used in differential operation, ie if its two drive pressure chambers 11 and 12 are acted upon by the outlet pressure P _N or P _{H of} the pressure supply unit 23, the maximum for the feed and work feed of the tool 16 is maximum exploitable force K _3N or K _3H , which acts in the direction of arrow 37 parallel to arrow 28, according to the amount in each case through the relationships:

${\text{K}}_{\text{3N}}{\text{= F₂. P}}_{\text{N}}\text{(3)}$

respectively.

${\text{K}}_{\text{3H}}{\text{= F₁. P}}_{\text{H}}{\text{- F₃. P}}_{\text{G}}\text{(4)}$

given.

In differential operation of the hydraulic cylinder 13, only the cross-sectional area F₂ of its smaller piston stage 34 is effective as the drive surface of its piston 21.

Before the control device 10 provided for the drive control of the hydraulic cylinder 13 is explained in terms of construction and circuit technology with reference to its functional components, the essential functions of the control device 10 are discussed in advance:

In the rapid feed, in which the tool 16 experiences its feed movement toward the workpiece 14, which is directed downward according to FIG. 1, the hydraulic cylinder 13 is operated in differential mode, the pressure supply initially taking place via the low-pressure outlet 24 of the pressure supply unit 23 .

The pressure that can be built up in this operation of the hydraulic cylinder 13 in its drive pressure chambers 11 and 12 is sufficient in cases in which the workpiece 14 has a relatively small thickness in order to generate the force required for the deformation of the workpiece 14, so that this, as it were "Rapid feed operation" can be edited.

For larger thicknesses of the workpiece, at which the output pressure of the pressure supply unit 23 that can be provided at the low-pressure outlet 24 is no longer sufficient to achieve the required (punching) deformation of the workpiece 14, the control device 10 transmits a switchover by responding to a pressure changeover valve 39 Operating pressure supply of the hydraulic cylinder 13 to the high-pressure outlet 26 of the pressure supply unit 23, at which a pressure P _{H can be} provided in the typical design thereof, the maximum amount (approx. 180 bar) is significantly higher, for the exemplary embodiment it is assumed 3 times higher than the output pressure P _N which can be provided at the low pressure outlet 24 of the pressure supply unit 23 and which may be around 60 bar. Even after switching to the pressure supply at a higher pressure level, the hydraulic cylinder 13 continues to operate in differential mode, ie in the rapid feed mode. The feed force that can be used for machining the workpiece 14 is now higher in accordance with the ratio P _H / P _{N of} the outlet pressure levels at the two pressure outlets 24 and 26 of the pressure supply unit 23.

If this increased feed force is not sufficient to machine the workpiece 14, with the result that the operating pressure in the drive pressure chambers 11 and 12 of the hydraulic cylinder 13 exceeds a value which is increased by a predetermined amount of e.g. Is 20 bar lower than the maximum outlet pressure at the high-pressure outlet 26 of the pressure supply unit 23, then a surface switching valve 42, which is controlled by the pressure prevailing in the larger drive pressure chamber 11 of the hydraulic cylinder 13, responds, whereby the hydraulic cylinder 13 detects the pressure required for the rapid feed. Operated differential operation is switched to a load-feed operation, in which only the larger, larger drive pressure chamber 11 of the hydraulic cylinder, as shown in FIG. 1, is connected to the high-pressure outlet 26 of the pressure supply unit 23, the other, lower, Drive pressure chamber 12 of the hydraulic cylinder 13, however, is relieved of pressure toward the - unpressurized - tank 43 of the pressure supply unit 23.

In this load feed operation, the maximum usable feed force with which the tool 16 is driven is given by the relationship (4). The feed speed is reduced compared to the rapid feed operation by the ratio F₂ / F₁ of the effective areas F₂ and F₁ of the smaller piston stage 34 and the larger piston stage 31 of the hydraulic cylinder piston 21.

As soon as the workpiece 14 is machined - pierced in the selected explanatory example - which has the consequence that the operating pressure P _A in the larger drive pressure chamber 11 of the hydraulic cylinder 13 drops drastically, this will be Surface changeover valve 42, if it had previously addressed, as well as the pressure changeover valve 39 switched back to their initial or basic positions associated with the rapid feed operation, so that the "last" part of the machining stroke of the workpiece 14 is again in the "fast" rapid feed operation can take place.

After the machining of the workpiece 14 has taken place, the hydraulic cylinder 13 is switched to rapid retraction mode, in which the smaller drive pressure chamber 12, which is designed as an annular space, is connected only to the low-pressure outlet 24 of the pressure supply unit 23 and the larger drive pressure chamber 11 to the pressureless tank 43 of the pressure supply unit 23 is relieved of pressure.

The control of the infeed and working stroke as well as the retraction movements of the hydraulic cylinder piston 21 or the tool 16, which is firmly connected to the latter, according to the stroke and speed is carried out by means of an electro-hydraulic follow-up control valve which is of known construction and function works with electrical, eg stepper motor-controlled position setpoint specification and mechanical position actual value feedback. Overrun control valves, which can be used in the context of the control device 10 as a stroke and movement control valve 44, are known, for example, from DE-PS 20 62 134 or DE 36 30 176 A1, on their content with regard to the structure and function of such overrun Control valves including their stepper motor control and electronic control of the same.

Such follow-up control valves are, according to their basic structure, designed as 4/3-way valves, which, however, can also be used as 3/3-way valves with the hydraulic circuit periphery of the hydraulic cylinder 13 shown in FIG. 1 .

Accordingly, it can be considered sufficient if the follow-up control valve 44 is only explained below on the basis of its function and the diverse design options suitable for achieving these functions are not specifically addressed.

• a) durch eine Ansteuerung des Schrittmotors 46 in der einen seiner beiden möglichen Drehrichtungen, z.B. im Uhrzeigersinn, der in der zur Fig. 1 durch den Dreh-Richtungspfeil 47 veranschaulicht ist, gelangt das Nachlauf-Regelventil 44 aus seiner dargestellten Grundstellung O, in welcher der größere Antriebsdruckraum 11 sowohl gegen die Druckausgänge 24 und 26 des Druckversorgungs-Aggregats 23 als auch gegen dessen Tank 43 abgesperrt ist, in seine Funktionsstellung I, in welcher der größere Antriebsdruckraum 11 des Hydrozylinders 13, je nach der Funktionsstellung des Druck-Umschaltventils 39, entweder an den Niederdruckausgang 24 des Druckversorgungs-Aggregats 23 oder an dessen Hochdruck-Ausgang 26 angeschlossen, jedoch gegen den Tank 43 des Druckversorgungs-Aggregats 23 abgesperrt ist. Diese Funktionsstellung I des Nachlauf-Regelventils 44 ist dem "Vorwärts"-Betrieb des Hydrozylinders 13 zugeordnet, in welchem die Eil-Vorschub-Bewegung, der Arbeits-Vorschub und ggf. die Last-Vorschub-Bewegung des Werkzeuges 16 sowie dessen Weiterbewegung bis in seine "untere" Endstellung erfolgen. Die Ansteuerung des Schrittmotors 46 geschieht dabei inkremental, d.h. durch eine Folge 48 von Ausgangsimpulsen 49 eines - programmierbaren - elektronischen Steuergeräts 51. "Inkremental" bedeutet hierbei, daß, wann immer der Schrittmotor 46 mit einem dieser Ausgangsimpulse 49 angesteuert wird, der Anker des Schrittmotors eine Drehung um einen definiert vorgegebenen Winkelbetrag erfährt, mit dem auch ein bestimmter Bruchteil des Hubes des Kolbens 21 des Hydrozylinders 13 verknüpft ist. Durch Vorgabe der Zahl der Impulse 49, mit denen der Schrittmotor 46 angesteuert wird und deren Frequenz ist somit der Weg einstellbar, den der Kolben 21 des Hydrozylinders 13 bzw. das Werkzeug 16 zurücklegt, sowie die Geschwindigkeit, mit der die "Vorwärts"-Bewegungen des Werkzeuges 16 erfolgen. Bei Gleichheit von Soll- und Ist-Position, die mittels der,insgesamt mit 52 bezeichnet, mechanischen Rückmeldeeinrichtung erfaßt wird, gelangt das Nachlauf-Regelventil 44 in dessen - dargestellte - Grundstellung O.
• b) Die Rückzugsbewegungen des Hydrozylinderkolbens 21 und des mit diesem fest verbundenen Werkzeuges 16 werden auf analoge Weise dadurch gesteuert, daß der Schrittmotor 46 mit einer Impulsfolge 53 angesteuert wird, durch welche eine Antriebssteuerung des Schrittmotors 46 in dem durch den Drehrichtungspfeil 55 repräsentierten Gegenuhrzeigersinn erfolgt, wodurch das Nachlauf-Regelventil 44 in seine Funktionsstellung II gelangt, in welcher der obere, seiner Querschnittsfläche nach größere, Antriebsdruckraum 11 des Hydrozylinders 13 mit dem - drucklosen - Tank 43 des Druckversorgungs-Aggregats 23 verbunden, jedoch gegen dessen Druck-Ausgangsseite 24,26, abgesperrt ist.
• c) Je größer, sowohl im "Vorwärts"- als auch im "Rückwärts"-Betrieb des Hydrozylinders 13 die Abweichung des Werkzeuges 16 von seiner Soll-Position ist, desto größer ist auch der Querschnitt des Durchfluß-Strömungspfades 54 bzw. 56, über den - in der Funktionsstellung I - die Druckeinspeisung in den Antriebsdruckraum 11 erfolgt bzw. - in der Funktionsstellung II des Nachlauf-Regelventils - ein Abströmen von Druckmedium zum Tank 43 des Druckversorgungs-Aggregats 23 hin erfolgt.
  Wann immer Gleichheit zwischen Soll- und Ist-Position des Werkzeuges 16 gegeben ist, nimmt das Nachlauf-Regelventil 44 seine Grundstellung O ein.

The functional properties of the follow-up control valve 44 used in the control device 10 according to FIG. 1 are, more specifically, the following:

• a) by controlling the stepper motor 46 in one of its two possible directions of rotation, for example clockwise, which is illustrated in FIG. 1 by the direction of rotation arrow 47, the overrun control valve 44 comes from its illustrated basic position O, in which the larger drive pressure chamber 11 is shut off both against the pressure outlets 24 and 26 of the pressure supply unit 23 and against its tank 43, into its functional position I, in which the larger drive pressure chamber 11 of the hydraulic cylinder 13, depending on the functional position of the pressure changeover valve 39, either connected to the low pressure outlet 24 of the pressure supply unit 23 or to its high pressure outlet 26, but is shut off against the tank 43 of the pressure supply unit 23. This functional position I of the follow-up control valve 44 is assigned to the "forward" operation of the hydraulic cylinder 13, in which the rapid feed movement, the work feed and possibly the load feed movement of the tool 16 and its further movement take place into its "lower" end position. The control of the stepper motor 46 is done incrementally, ie by a sequence 48 of output pulses 49 of a - programmable - electronic control unit 51. "Incremental" means that whenever the stepper motor 46 is controlled with one of these output pulses 49, the armature of the stepper motor undergoes a rotation by a defined predetermined angular amount, with which a certain fraction of the stroke of the piston 21 of the hydraulic cylinder 13 is also linked. By specifying the number of pulses 49 with which the stepper motor 46 is actuated and their frequency, the path that the piston 21 of the hydraulic cylinder 13 or the tool 16 travels and the speed with which the "forward" movements are thus adjustable of the tool 16. If the target and actual positions are the same, which is detected by means of the mechanical feedback device, denoted overall by 52, the overrun control valve 44 reaches its - shown - basic position O.
• b) The retraction movements of the hydraulic cylinder piston 21 and the tool 16 firmly connected to it are controlled in an analogous manner in that the stepper motor 46 is controlled with a pulse train 53 through which the stepper motor 46 is driven in the counterclockwise direction represented by the direction of rotation arrow 55, as a result of which the follow-up control valve 44 reaches its functional position II, in which the upper, cross-sectional area larger, the drive pressure chamber 11 of the hydraulic cylinder 13 with the - unpressurized - Tank 43 of the pressure supply unit 23 connected, but against its pressure output side 24,26, is blocked.
• c) The greater the deviation of the tool 16 from its desired position, both in the "forward" and in the "backward" operation of the hydraulic cylinder 13, the greater the cross section of the flow-through flow paths 54 and 56, respectively which - in the functional position I - the pressure is fed into the drive pressure chamber 11 or - in the functional position II of the run-on control valve - there is an outflow of pressure medium to the tank 43 of the pressure supply unit 23.
  Whenever the desired and actual positions of the tool 16 are identical, the overrun control valve 44 assumes its basic position O.

The pressure changeover valve 39 is designed as a pressure-controlled 2/2-way valve, which has the function as soon as the operating pressure P _A in the larger drive pressure chamber 11 of the hydraulic cylinder 13 exceeds a threshold value P _A1 , which, for the purpose of explanation, is 90% of the - lower - supply pressure P _N provided at the low pressure outlet 24 of the pressure supply unit 23 is assumed, reached or exceeded, is switched from its previously assumed blocking position 0 to a flow position I in which the high pressure outlet 26 of the pressure supply unit 23 is now connected to the supply pressure (P) connection of the follow-up control valve 44 is connected.

Between the supply pressure connection 57 of the follow-up control valve 44 and the low pressure outlet 24 of the pressure supply unit 23, a check valve 58 is connected, which is held in its blocking position by higher pressure at the supply pressure connection 57 than at the low pressure outlet 24 of the pressure supply unit. Through this check valve 58, the operating pressure P _N provided at the low pressure outlet 24 of the pressure supply unit 23 is transmitted to the supply pressure connection 57 of the follow-up control valve 44 while the pressure changeover valve 39 is in its blocking position.

When pressure is supplied to the hydraulic cylinder 13 from the high-pressure connection 26 of the pressure-supply unit 23, the check valve 58 prevents pressure from being coupled over from the high-pressure outlet 26 of the pressure-supply unit 23 to its low-pressure outlet 24.

The pressure changeover valve 39 is designed in the special embodiment shown as a slide valve, in the housing 60 of which two bore stages 59 and 61 of different diameters are introduced, which, merging into one another, are offset from one another by an inner, radial housing stage 62 and each have an end end wall 63 or 64 of the housing are completed.

The piston, which is designated overall by 66, is displaceably guided with an end flange 67 and 68, respectively, in the smaller bore step 59 in diameter or in the larger bore step 61 in diameter, these end flanges 67 and 68 being movable in a pressure-tight manner Form boundaries each of a control pressure chamber 69 and 71, which are closed by the end end walls 63 and 64 fixed to the housing.

The control pressure chamber 69 of the pressure changeover valve 39, which is smaller in diameter, is - permanently - connected to the low pressure outlet 24 of the pressure supply unit 23.

The control pressure chamber 71 of the pressure changeover valve 39, which is larger in diameter, is connected to the working connection 72 of the follow-up control valve 44 which is connected to the larger drive pressure chamber 11 of the hydraulic cylinder 13.

Unter Vernachlässigung geringfügiger Reibungsverluste wird daher der Ventilkolben 66 in seine - in ausgezogenen Linien dargestellte - mit minimalem Volumen seines größeren Steuerdruckraumes 71 verknüpfte Grundstellung gedrängt, wenn und solange der in diesen - größeren - Steuerdruckraum 71 eingekoppelte und gleichzeitig in dem größeren Antriebsdruckraum 11 des Hydrozylinders 13 herrschende Betriebsdruck P_A kleiner ist als der mit dem Wert 1,1 dividierte Wert P_N des am Niederdruck-Ausgang 24 des Druckversorgungsaggregates 23 bereitgestellten - niedrigeren - Versorrgungsdruckes P_N, der permanent in den kleineren Steuerdruckraum 69 des Druck-Umschalt-Ventils 39 eingekoppelt ist, das heißt, wenn gilt:

${\text{P}}_{\text{A}}{\text{≦ P}}_{\text{N}}\text{/ 1,1 (6)}$

Solange der Ventilkolben 66 seine hierdurch bedingte Grundstellung 0 einnimmt, ist ein ringspaltförmiger Eingangsdruckraum 79 des Druck-Umschaltventils 39, in den der am Hochdruckausgang 26 des Druckversorgungsaggregates 23 bereitgestellte - höhere - Versorungsdruck P_H eingekoppelt ist, gegen einen - ebenfalls ringspaltförmigen - Ausgangsdruckraum 81 des Druckumschalt-Ventils 39 abgesperrt, der an den Versorgungsdruck-Anschluß 57 des Nachlauf-Regelventils 44 angeschlossen ist.The diameter of the larger end flange 68 is followed by a piston stage 73 corresponding to the diameter of that of the smaller bore 59 of the valve housing 58, by means of which the valve slide 66 is also shown in FIG which is guided so that it is displaceable in a pressure-tight manner according to the smaller housing bore 59. This piston stage 73 is fixedly connected by means of a rod-shaped piston intermediate piece 74 to the end flange 67 of the valve piston 66 which also corresponds to the diameter of the smaller bore step 59, the piston 66 being embodied in one piece overall. The end flanges 67 and 68 each have short support projections 77 and 78, pointing towards the end end walls 63 and 64 of the valve housing 58, in the direction of the central longitudinal axis 76 of the pressure changeover valve 39, by means of which the piston 66 in its functional positions O and I corresponding positions either on the one, according to FIG. 1 "lower" end wall 64 or on the other, according to FIG. 1 "upper" end wall 63 of the valve housing 58 is supported centrally. In a special design of the pressure changeover valve 69, the effective cross-sectional area f₂ of the larger control end flange 68 of its piston is 10% larger than the effective cross-sectional area f₁ of the smaller control end flange 67 of the valve piston 66, so that:

$\text{f₂ = 1.1 f₁ (5)}$

Neglecting slight frictional losses, the valve piston 66 is therefore pushed into its basic position - shown in solid lines - linked to the minimal volume of its larger control pressure chamber 71, if and as long as the - larger - control pressure chamber 71 is coupled into the larger drive pressure chamber 11 of the hydraulic cylinder 13 and at the same time prevailing operating pressure P _{A is} smaller than the value P _N divided by the value 1.1 of the low-pressure outlet 24 of the pressure supply unit 23 Provided - lower - supply pressure P _N , which is permanently coupled into the smaller control pressure chamber 69 of the pressure changeover valve 39, that is, if:

${\text{P}}_{\text{A}}{\text{≦ P}}_{\text{N}}\text{/ 1.1 (6)}$

As long as the valve piston 66 assumes its basic position 0 caused thereby, an inlet pressure chamber 79 of the pressure changeover valve 39 in the form of an annular gap, into which the - higher - supply pressure P _H provided at the high pressure outlet 26 of the pressure supply unit 23 is coupled, against an - also annularly shaped - outlet pressure chamber 81 des Shut off valve 39, which is connected to the supply pressure port 57 of the follow-up control valve 44.

The inlet pressure chamber 79 of the pressure changeover valve 39 is - seen in the basic position of the valve piston 66 shown in solid lines - fixed to the housing through the smaller bore step 59 of the valve housing 58 and - axially - movable through the mutually facing, annular, inner end faces 82 and 83 of the smaller end flange 67 of the valve piston 66 and the piston stage 73 adjoining its larger end flange 68.

The outlet pressure chamber 81 of the pressure changeover valve 39 is fixed to the housing in the axial direction and radially on the outside by an annular groove 84 made in the smaller bore 59 of the valve housing 58 and radially on the inside by the cylindrical outer surface 86 of the smaller end flange 67 of the valve piston 66.

gilt, der Ventilkolben 66 in seine der Offen-Stellung des Druck-Umschaltventils 39 entsprechende Funktionsstellung I, in welcher die die zylindrische Mantelfläche 86 des kleineren Endflansches 67 des Ventilkolbens 66 gegen dessen kreisringförmige, innere Stirnfläche 82 absetzende Steuerkante 87 innerhalb der lichten Weite der den Ausgangsdruckraum 81 gehäusefest begrenzenden Gehäuse-Ringnut 84 liegt und somit der - in axialer Richtung gleichsam "verschobene" - Eingangsdruckraum 79 des Druck-Umschaltventils 39 mit dessen Ausgangsdruckraum 81 in kommunizierender Verbindung steht und damit der Hochdruck-Ausgang 26 des Druckversorgungsaggregates 23 an den Versorgungsdruck-Anschluß 57 des Nachlauf-Regelventils 44 angeschlossen ist.If the operating pressure P _A in the larger drive pressure chamber 11 of the hydraulic cylinder 13 exceeds the value P _N / 1.1, which will be the case when the tool 16 strikes the workpiece 14 and - in the differential operation of the hydraulic cylinder 13 - the outlet pressure provided at the low-pressure outlet 24 of the pressure supply unit 23 no longer "begins to suffice" in order to allow the tool 16 to pierce the workpiece 14, because the relationship:

${\text{P}}_{\text{A}}{\text{≧ P}}_{\text{N}}\text{/ 1.1 (7)}$

applies, the valve piston 66 in its open position of the pressure changeover valve 39 corresponding functional position I, in which the cylindrical outer surface 86 of the smaller end flange 67 of the valve piston 66 against its annular, inner end face 82 is offset by the control edge 87 within the clear width of the Output pressure chamber 81 housing-bounding housing annular groove 84 is located and thus the - as it were "shifted" in the axial direction - input pressure chamber 79 of the pressure changeover valve 39 communicates with its outlet pressure chamber 81 and thus the high-pressure outlet 26 of the pressure supply unit 23 to the supply pressure Connection 57 of the follow-up control valve 44 is connected.

In this functional position I of the pressure changeover valve 39, the hydraulic cylinder 13 is operated at a higher drive pressure level and with a correspondingly increased feed force, although - initially - in differential operation.

After such a pressure changeover has been achieved, the feed force which the hydraulic cylinder 13 can develop is increased by the ratio P _H / P _N.

If the feed force achievable in differential operation of the hydraulic cylinder 13 is not sufficient for the tool 16 to pierce the workpiece 14, the surface switch valve 42, which is designed as a pressure-controlled 3/2-way valve, provides pressure relief for the smaller one , annular drive pressure chamber 12, with the result that now the entire cross-sectional area F₁ of the larger piston stage 31 is used for the feed force development and this thus in cases of high load - large workpiece thickness - down to the value F₁. P _H can be increased. In this operating state of the hydraulic cylinder, which is achieved by automatic switching of the surface switching valve 42, the still usable feed rate is then reduced by the area ratio F / F 1.

Furthermore, this surface switching valve 42 fulfills the function that after it had been switched into its functional position which mediated the pressure relief of the annular drive pressure chamber 12 of the hydraulic cylinder 13 and thereby enabled the use of an increased feed force, only then again in its pressurization of the Annular drive pressure chamber 12 mediating functional position is switched back after the - for example penetrating - machining of workpiece 12 the need for feed force on tool 16 has become a defined minimum amount ΔK lower than the amount of feed force or operating pressure in drive pressure chambers 11 and 12 of the hydraulic cylinder 13, by exceeding which the switching of the surface switching valve 42 in which the pressure relief of the annular drive pressure chamber 12 was triggered position.

On the one hand, this ensures that, as long as possible, the highest possible feed rate of the tool 16 remains usable and, on the other hand, it ensures that after the control device 10 has switched over in the sense of increasing the feed force, it is not "too early" again to a reduced feed force. is switched back ", which could lead to undesired vibrations and, as a result, to a" standstill "of the tool 16.

In order to implement these functions, the surface switching valve 42 is designed in more detail as follows, reference being now made to the details of FIGS. 2 and 3, which show two possible operating positions of the surface switching valve 42, while the Surface changeover valve is shown in FIG. 1 in its basic position corresponding to the non-activated state of the drive device.

The area changeover valve 42 comprises a first valve chamber 88, which is permanently connected to the tank 43 of the pressure supply unit 23 via a relief flow path 89 and is therefore kept depressurized.

This valve chamber 88 is sealed off to the outside by a set screw 91, which forms the end end wall of the valve housing, which is designated as a whole by 90. By turning this set screw 91, the bias of a valve closing spring 92 is adjustable, which engages a centering piece 93, which urges a valve body formed as a ball 94 of a seat valve, generally designated 96, against its valve seat 97, ie into the closed position of this seat valve 96, which is formed by the inner, ie the clear diameter towards the smaller, edge of a conical depression 98, which in turn serves to center the valve ball 94, of an intermediate wall 99 of the valve housing 90. Between this valve seat 97 and a central valve chamber 101 extends a valve channel 102 opening into the central valve chamber 101. The central valve chamber 101 is in constant communication with the annular, smaller drive pressure chamber 12 of the hydraulic cylinder 13 via a first hydraulic control line 103. The central valve chamber 101 is bounded by the one, the diameter smaller bore step 104 of a stepped bore of the housing 90, generally designated 106, whose diameter larger bore step 107 at the other end of the housing 90 is sealed pressure-tight by a housing cover 108 forming the end wall of the valve housing 90 there is.

In the two bore stages 104 and 107 of the stage bore 106, a stage piston, designated overall by 112, is displaceably guided in a pressure-tight manner, each with a piston stage 109 or 111, the smaller piston stage 109 of which forms an axially movable boundary of the central valve chamber 101, and its diameter after larger piston stage 111 on the one hand forms the axially movable boundary of an annular chamber 115 which is axially fixed to the housing by the annular housing stage 113 which mediates between the smaller bore step 104 and the larger bore stage 107, and further forms the axially movable boundary of a control chamber 114, the axial limit of which is fixed to the housing is formed by the housing cover 108. This control chamber 114 is kept in constant communication with the larger drive pressure chamber 11 of the drive hydraulic cylinder via a second hydraulic control line 116.

The stepped piston 112 is urged toward the valve ball 94 by a - slightly preloaded - return spring 117, which is supported on the inside of the housing cover 108, on which it rests in the basic position shown in FIG. 1 with a plunger-shaped, axial extension 118 supports its smaller piston stage 109. The outer diameter of this tappet-shaped extension 118 is significantly smaller than the diameter of the valve channel 102 through which it passes. The smaller piston stage 109 is offset from the larger piston stage 111 by an annular groove-shaped constriction 119, which is penetrated by a transverse bore 121 opening into the annular chamber 115. This transverse bore 121 is in constant communication with the central valve chamber 101 via a central longitudinal bore 122 which penetrates the smaller piston stage 109 and its plunger-shaped extension 108 in the axial direction and one or more transverse bore (s) 123 in the plunger-shaped extension 118.

The smaller bore step 104, viewed in the direction of the central longitudinal axis 100 of the housing 90, is provided in its central region with an annular, radial extension 124 which is connected via a third control or. Pressure supply line 134 is permanently connected to the low pressure outlet 24 of the pressure supply unit 23. The edge formed by the radially inner edge 126 of the groove flank 127 facing the central valve chamber 101, as shown in FIG. 1, forms a control edge fixed to the housing, with which the outer edge 128 of the annular end face 129 of the smaller piston stage 109 delimiting the central valve chamber 101 serves as a movable control edge can cooperate.

In the basic position of the stepped piston 112 shown in FIG. 1, the movable control edge 128 of the stepped piston 112 is in positive overlap with the control edge 126 fixed to the housing, this overlap ΔX₁ corresponding to only a small fraction of the stroke X₁ that the stepped piston 112 from its illustrated basic position out in the opening direction of the seat valve 96, ie can perform in the direction of arrow 131 and only a small fraction of the stroke X₂ that the stepped piston 112 in the opposite direction, i.e. can perform in the direction of arrow 132.

In the illustrated basic position of the stepped piston 112, the annular chamber 124 delimited by the annular groove-shaped extension 124 and the smaller piston step 109, regardless of the overlap ΔX₁ of the movable control edge 128 and the control edge 126 fixed to the housing, is not hermetically sealed against the central valve chamber 101, but stands with it a peripheral edge notch 133 with a small overflow cross-section still in communicating connection, which is canceled, however, when the stepped piston has carried out a small fraction ΔX₂ of its possible stroke in the direction of arrow 131, after which the annular groove-shaped communicating connection with the low-pressure outlet 24 of the pressure supply unit 23 Extension 124 of the smaller bore step 104 is blocked against the central valve chamber 101.

The preload of the valve closing spring 92 is or is set so high that the force with which the valve ball 94 is pressed against the circular valve seat 97 approximately corresponds to the force, e.g. 90% of that force corresponds when the valve ball 94 is pressurized within the circular area delimited by the valve seat 97 with a pressure which corresponds to the maximum outlet pressure of the pressure supply unit 23, which can be provided at its high pressure outlet 26.

Such a high pressure can be injected into the central valve chamber 101 if the tool 16 - in differential operation of the hydraulic cylinder 13 - is acted upon by the high output pressure P _H after the pressure changeover valve 39 has been switched over, and this pressure is also applied to the larger ones via the follow-up control valve 44 Drive pressure chamber 11 of the hydraulic cylinder 13 is coupled.

Assuming a maximum output pressure of the pressure supply unit 23 at the outlet 26 of 180 bar, the pretensioning of the closing spring 92 is accordingly set to a value equivalent to a "closing pressure" of 162 bar.

wobei mit P_H und P_N die Werte des Ausgangsdruckes des Druckversorgungsaggregates 23 an dessen Hochdruckausgang 26 bzw. an dessen Niederdruckausgang 24 bezeichnet sind, die beim gewählten, speziellen Erläuterungsbeispiel im Verhältnis 3/1 zueinander stehen.In contrast, the bias of the return spring 117 is negligible and equivalent to a pressure of only a few, for example 5 bar. F₄ denotes the amount of the circular area bordered by the valve seat 97, within which the pressure on the valve ball 94, which is coupled via the first hydraulic control line 103 into the central valve chamber 101 of the area changeover valve 42, can be built up in the annular drive pressure chamber 12 of the hydraulic cylinder 13 can act, and with F₅ the cross-sectional area of the larger piston stage 111 of the stepped piston 112, which is acted upon by the outlet pressure P _{A of} the follow-up control valve 44, which is also coupled into the larger drive pressure chamber 11 of the hydraulic cylinder, these surfaces are F₄ and F₅ the area switching valve 42 dimensioned so that they satisfy the following relationship:

${\text{F₅ / F₄> P}}_{\text{H}}{\text{/ P}}_{\text{N}}\text{(8th)}$

P _H and P _N denote the values of the outlet pressure of the pressure supply unit 23 at its high-pressure outlet 26 and at its low-pressure outlet 24, which are in the ratio 3/1 to one another in the selected, specific explanatory example.

The annular chamber 124 of the surface changeover valve 42 is connected via a first control line 134 to the smaller control pressure chamber 69 of the pressure changeover valve 39.

Furthermore, the control chamber 114 of the surface changeover valve 42, which is movably delimited by the larger piston stage 111, is connected to the larger control pressure chamber 71 of the pressure changeover valve 39 via a second control line 136.

Furthermore, it is assumed that the area ratio F₁ / F₂ of the drive hydraulic cylinder 13 has the value 2 and that the larger piston area 27 designated by F₁ of the piston 21 of the drive hydraulic cylinder 13 has an amount of 100 cm².

In a typical work cycle, the control device 10, which is specified in more detail both in terms of its basic structure and by means of a special exemplary embodiment, operates more specifically as follows in a typical work cycle:
When the pressure supply unit 23 is switched on, for the start-up of the drive and control device 10 as a whole, the follow-up control valve is - initially - in order to bring the tool 16 of the hydraulic cylinder 13 into a defined starting position - for example in its upper end position controlled in its functional position designated II. As a result, the larger drive pressure chamber 11 of the hydraulic cylinder 13 and the control chamber 114 of the surface switch valve 42 to the tank 43 of the pressure supply unit 23 are relieved, while at the same time the outlet pressure P _N provided at the low pressure outlet 24 of the pressure supply unit 23 is released into the annular groove extension 124 of the housing 90 of the surface switching valve 53, its central valve chamber 101 and its annular chamber 115 and via the first hydraulic control line 103 into the annular drive pressure chamber 12 of the hydraulic cylinder 13 and on the other hand - via the first control line 134 of the pressure changeover valve 39 in its smaller control pressure chamber 69 is coupled.

The larger diameter control pressure chamber 71 of the pressure changeover valve 39, which is connected via the second control line 116 of the pressure changeover valve 39 to the control chamber 114 of the surface changeover valve 42, which through the larger piston stage 111 of the stepped piston 112 of the surface changeover valve 42 is movably limited, is also relieved of pressure to the tank 43 of the pressure supply unit 23, with the result that the pressure changeover valve 39 is held in its - shown in FIG. 1 - basic position, in which, via the check valve 58, the - lower - Output pressure of the pressure supply unit 23 is present on the one hand at the supply pressure connection 57 of the follow-up control valve and on the other hand is directly coupled into the annular radial extension 124 of the surface switch valve 42.

In this operating state of the control device 10 and the follow-up control valve 44, the piston 21 of the hydraulic cylinder 13 first reaches its upper end position, the basic position shown in FIG. 1, during the stepped piston 112 of the surface switching valve 42, which is in total on one of the cross-sectional area F₅ its larger piston stage 111 is acted upon by the outlet pressure P _{N of} the pressure supply unit 23, is pushed into its lower end position, shown in FIG. 2, ie removed from the valve ball 94.

This functional position of the surface changeover valve 42 in combination with the functional position II of the overrun control valve 44 also corresponds to the withdrawal operation of the hydraulic cylinder 13, through which it returns to its starting position after the tool 16 has carried out its working stroke.

In order to initiate its feed operation from the basic position of the hydraulic cylinder piston 21, the overrun control valve 44 is switched to its functional position I by actuating the stepping motor 46 with the “forward” control pulse sequence 49.

As a result, both the larger, upper, drive pressure chamber 11 of the hydraulic cylinder 13 and the control chamber 114 of the surface switchover valve and the larger control pressure chamber 71 of the pressure switchover valve 39 are acted upon by the outlet pressure P _{A of} the follow-up control valve 44, which affects the open state of the flow - Flow path 54 of the follow-up control valve can be regulated as required.

, beim gewählten Erläuterungsbeispiel somit nur geringfügig größer als P_N/2 und daher erheblich niedriger als der am Niederdruckausgang 24 des Druckversorgungs-Aggregats 23 abgegebene Ausgangsdruck P_N, der - zunächst - für die Steuerung der Eil-Vorschubbewegung des Hydrozylinderkolbens 21 ausgenutzt wird. Daher bleiben in einer derartigen Eil-Vorschub-Betriebsphase das Druck-Umschaltventil 39 in seiner dargestellten Grundstellung und das Flächen-Umschaltventil 42 in der in der Figur 2 dargestellten Funktionsstellung, weil der kleinere Steuerdruckraum 69 des Druck-Umschaltventils 39 und gleichzeitig auch die in dieser Funktionsstellung des Flächen-Umschaltventils 42 mit der zentralen Ventilkammer 101 in kommunizierender Verbindung stehende ringförmige radiale Erweiterung 124 des Ventilgehäuses und die Ringkammer 115 mit einem deutlich höheren Druck, nämlich dem Druck P_N beaufschlagt sind, während der Steuerdruckraum 71 des Druck-Umschaltventils 39 und die Steuerkammer 114 des Flächen-Umschaltventils 42, die "lediglich" mit dem Ausgangsdruck P_A des Nachlauf-Regelventils 44 beaufschlagt sind, unter diesem deutlich geringeren Druck, dessen Wert nur weniger größer ist als P_N/2, stehen.In this, the - load-free - rapid feed operation of the piston 21 of the hydraulic cylinder 13 assigned differential operation of the same is the pressure P _A , which must be coupled into the larger drive pressure chamber 11 of the hydraulic cylinder 13, around the piston and the tool 16 in the direction of to move the workpiece 14 just a little larger than the value ${\text{P}}_{\text{N}}\text{· F₃ / F₁}$

, in the selected explanatory example thus only slightly larger than P _N / 2 and therefore considerably lower than the output pressure P _N delivered at the low-pressure outlet 24 of the pressure supply unit 23, which - initially - is used for controlling the rapid feed movement of the hydraulic cylinder piston 21. Therefore, stay in one Rapid feed operating phase, the pressure changeover valve 39 in its basic position shown and the surface changeover valve 42 in the functional position shown in FIG. 2, because the smaller control pressure chamber 69 of the pressure changeover valve 39 and at the same time also in this functional position of the surface changeover valve 42 annular radial extension 124 of the valve housing communicating with the central valve chamber 101 and the annular chamber 115 are subjected to a significantly higher pressure, namely the pressure P _N , while the control pressure chamber 71 of the pressure changeover valve 39 and the control chamber 114 of the surface Changeover valve 42, which are "only" supplied with the outlet pressure P _{A of} the follow-up control valve 44, are under this significantly lower pressure, the value of which is only less than P _N / 2.

As soon as the tool 16 hits the workpiece 14 and thus the feed movement of the tool 16 is opposed to a significant resistance, the follow-up control valve 44 as a result of the now increasing follow-up error between the stepper motor-controlled position setpoint specification and the means the feedback device 52 detected actual position of the tool 16, the flow cross section of the flow flow path 54 increases, with the result that the pressure in the larger drive pressure chamber 11 of the hydraulic cylinder 13 and thus also in the larger control pressure chamber 71 of the pressure changeover valve 39 and in the control chamber 114 of the surface switching valve 42 rises.

Is the achievable feed force K₃, which is given by the relationship (3), sufficient to pierce the workpiece 14, i.e. to carry out the working stroke, this is carried out in the differential operation of the hydraulic cylinder 13 with pressure supply from the low pressure outlet 24 of the pressure supply unit 23. A "jerky" acceleration of the hydraulic cylinder piston 21 after the workpiece 14 has been pierced is not to be feared, since the movement control of the hydraulic cylinder piston 21 and the tool 16 remains guided by the overtravel control and the tool 16 is "passed" too quickly by the control system. can be intercepted sufficiently gently "so that undesirable vibrations of the machine are avoided.

If the feed force achievable in differential operation of the hydraulic cylinder 13 with pressure supply from the low-pressure outlet 24 of the pressure supply unit 23 is not sufficient to allow the tool 16 to pierce the workpiece 14, with the result that the outlet pressure P _{A of} the follow-up control valve 44 "Approaching" more and more the outlet pressure level P _{N of} the low pressure outlet of the pressure supply unit 23, this leads to the moment when the outlet pressure P _{A of} the follow-up control valve 44 reaches the value indicated by the relationship (7) or exceeds, that the pressure changeover valve 39 is "switched" to its alternative position, shown in dashed lines to the basic position shown in FIG. 1, in which the high pressure outlet 26 of the pressure supply unit 23 with the supply pressure connection is now 57 of the follow-up control valve 44, but this is blocked by the check valve 58 against the low pressure outlet 24 of the pressure supply unit 23.

The consequence of this is that the hydraulic cylinder 13 is still used in differential operation, but at an increased level of pressure, both in the larger drive pressure chamber 11 and in the smaller, annular-shaped drive pressure chamber 12, as a result of which - in the area and pressure ratios explained, the maximum force , with which the tool 16 can carry out its working stroke, is now raised to a value which corresponds to three times the value that could be achieved with pressure supply from the low-pressure outlet 24 of the pressure supply unit 23.

For the selected design example, this means that now - instead of a maximum feed force of 30,000 N / 1.1 - a feed force is available, the maximum amount of which is limited by the value 90,000 N, provided the hydraulic cylinder 13 continues to be used in differential operation becomes.

If, in this operating mode, the feed force that can at most be deployed by the hydraulic cylinder 13, now in accordance with the relationship (4), is not sufficient to allow the tool 16 to pierce the workpiece 14, this has, due to the overrun control, in the larger drive pressure chamber 11 of the hydraulic cylinder 13 results in a pressure increase, which initially leads to the "vicinity" of the outlet pressure level at the high-pressure outlet 26 of the supply pressure unit 23. The stepped piston 112 of the surface switching valve 42 is now relieved of pressure, as it were, since it has almost the same pressures both via the central valve chamber 101 and the annular chamber 115 and via the - lower - control chamber 114, which amounts to the high output pressure P _{H of} the pressure supply unit 23 or almost this value P _H correspond, is exposed, and insofar as it is "neutral" pressurized. In this operating state of the surface changeover valve 42, the relatively weak return spring 117 is capable of displacing the stepped piston 112 in the direction of the valve ball 64 and bringing it into contact with the valve ball 94, ie into the position shown in FIG. 1. If the pressure in the larger drive pressure chamber 11 of the hydraulic cylinder 13 continues to rise due to the increasing resistance that the workpiece 14 opposes to the tool 16, then the pressure exerted on the area Frand surrounded by the valve seat 97 is sufficient to counter the valve ball 94 to lift the action of the valve closing spring 92 from its seat 97, as a result of which the communicating connection of the central valve chamber 101, which previously existed via the notch 133, with the groove-shaped extension 124, which is under the high outlet pressure P _{H of} the pressure supply unit 23, is canceled. As a result, the stepped piston 112 now reaches the "upper" end position shown in FIG. 3, in which the annular drive pressure chamber 12 via the central valve chamber 101 and the "pressure" valve chamber 88 arranged "above" to the tank 43 of the pressure supply unit 23 is relieved. The surface switching valve 42 has now "switched". Only the upper, larger drive pressure chamber is now subjected to the high outlet pressure of the pressure supply unit 23 11 of the hydraulic cylinder, which now executes its working stroke mediating the machining of the workpiece 14 in the load feed operation with increased feed force but at a low feed speed. In this load-feed operation, in which the hydraulic cylinder 13 is used with, as it were, "one-sided" pressurization of its piston 21, the maximum feed force in the selected explanatory example is 180,000 N.

If the workpiece 14 has been machined, for example punctured, the pressure in the larger drive pressure chamber 11 of the hydraulic cylinder 13 dropping again, the corresponding pressure drop also occurs in the control chamber 114 of the surface switchover valve 42 and in the larger control pressure chamber 71 of the pressure switchover valve 39, in its smaller control pressure chamber 69, the pressure P _N delivered at the low pressure outlet 24 of the pressure supply unit 23 of - in the selected explanatory example - 60 bar is still coupled.

If the operating pressure P _A regulated as required by the follow-up control valve 44 drops below the lower limit value given by the relationship (7), namely below the value of P _N / 1.1, which is 55 bar in the selected illustrative example switches the pressure changeover valve 39 back to the basic position shown in FIG. 1, with the result that the pressure supply is now switched back to the low pressure outlet 24 of the pressure supply unit 23. The operating pressure P _A , which continues in the larger drive pressure chamber 11 of the hydraulic cylinder 13 is not necessarily "withdrawn" as a result, since the follow-up control valve 44 can - by controlling the flow path 54 - "maintain" an operating pressure of 55 bar or a little less.

Only when the operating pressure P _A , because the resistance, which the workpiece 14 opposes to the tool 16 in the final phase of its machining, has decreased further, has dropped below the value - in the explanatory example 50 bar - at which the step piston 112 of the surface switching valve 42 by the pressurization of its larger piston stage 111 on the surface F₅ is still able to keep the seat valve 96 open against the action of the valve closing spring 92, the surface switching valve 42 comes because the valve closing spring 92 the step piston 112 in the direction again is able to push back to its basic position, again into the closed position shown in FIG. 1, in which the hydraulic cylinder 13 is now operated in differential mode, ie with pressure on its piston 21 on both sides, with at most the low outlet pressure P _{N of} the pressure supply unit 23. As a result, the piston 21 of the hydraulic cylinder 13 is "gently" intercepted and a largely vibration-free and thus also machine-friendly control of the rapid feed, machining and load feed movements of the tool 16 and, as it were, steadily into the rapid withdrawal operation of the hydraulic cylinder 13 transient control of the piston and tool movements achieved.

In the sense of a generalizing formulation of the dimensioning relationships given above for a specific design example, it should apply that the pressure changeover valve is always switched over when the operating pressure P _A , which is currently coupled into the larger drive pressure chamber 11 of the hydraulic cylinder 13, has the value P _N b₁ exceeds or falls below.

With b₁, a factor is designated which is less than 1 and the area ratio f₁ / f₂ of the areas f₁ and f₂ of the end flanges 67 and 68 of the valve slide 66 of the pressure changeover valve 39 corresponds.

Functional values of the parameter b 1 are between 0.85 and 0.95, preferably around 0.9.

The switchover of the surface switchover valve 42 should, after it has been in its position providing the pressure relief of the smaller drive pressure chamber 12 of the hydraulic cylinder 13, at an operating pressure value P _AF which is lower than the value P _N. b₁.

gegeben, wobei mit K_R die Schließkraft der Ventil-Schließfeder 92 des Rückschlagventils 96 des Flächen-Umschaltventils 42 bezeichnet ist.The P _AF value is due to the relationship

${\text{P}}_{\text{AF}}{\text{= K}}_{\text{R}}\text{/ F₅ (9)}$

given, with K _R the closing force of the valve closing spring 92 of the check valve 96 of the surface switching valve 42 is designated.

in der mit ΔP eine Druckdifferenz bezeichnet ist, die einen kleinen Bruchteil von z.B. 10 % des am Hochdruck-Ausgang 26 des Druckversorgungs-Aggregates 23 bereitgestellten - höheren - Versorgungsdruckes P_H entspricht.The relationship applies to this closing force K _R

${\text{K}}_{\text{R}}{\text{= (P}}_{\text{H}}\text{- ΔP) · F₄ (10),}$

in which ΔP denotes a pressure difference which corresponds to a small fraction of, for example, 10% of the - higher - supply pressure P _H provided at the high pressure outlet 26 of the pressure supply unit 23.

wobei b₂ wiederum einen Wert bezeichnet, der kleiner als 1 ist und beispielsweise zwischen 0,85 und 0,95, vorzugsweise um 0,9 beträgt. Unter Berücksichtigung der Beziehungen (9), (10) und (11) ergibt sich somit, daß die Forderung, daß das Flächen-Umschaltventil 42 erst dann - in der Endphase eines Arbeitszyklus - in seine den Differentialbetrieb des Hydrozylinders 13 vermittelnde Funktionsstellung "zurückschalten" soll, nachdem das Druck-Umschaltventil 39 zuvor schon in seine - im Ergebnis - die Druckversorgung des Hydrozylinders 13 aus dem Niederdruck-Ausgang 24 des Druckversorgungs-Aggregates 23 vermittelnde Funktionsstellung zurückgeschaltet worden ist, dadurch erfüllbar ist, daß das Flächenverhältnis F₄/F₅ der vom Ventilsitz 97 des Flächen-Umschaltventils 42 umrandeten Querschnittsfläche F₄ zu der Steuerfläche F₅ des Stufenkolbens 112 des Flächen-Umschaltventils 42 der folgenden Beziehung

genügt, die auch in der folgenden Form geschrieben werden kann

wobei mit a eine - kleine Sicherheitsmarge bezeichnet ist, die einen Wert um zwischen 2 % und 10 %, vorzugsweise einen Wert um 5 % hat.The relationship is equivalent to the relationship (10)

${\text{K}}_{\text{R}}{\text{= b₂ · P}}_{\text{H}}\text{· F₄ (11),}$

where b₂ in turn denotes a value which is less than 1 and is, for example, between 0.85 and 0.95, preferably around 0.9. Taking into account the relationships (9), (10) and (11), it follows that the requirement that the surface switching valve 42 only "switch back" to its functional position mediating the differential operation of the hydraulic cylinder 13 only in the final phase of a working cycle after the pressure changeover valve 39 has previously been switched back to its - as a result - the pressure position of the hydraulic cylinder 13 from the low-pressure outlet 24 of the pressure supply unit 23, the functional position which can be fulfilled is that the area ratio F₄ / F₅ of the Valve seat 97 of the area switching valve 42 bordered cross-sectional area F₄ to the control surface F₅ of the stepped piston 112 of the area switching valve 42 of the following relationship

is sufficient, which is also written in the following form can be

where a denotes a - small safety margin, which has a value between 2% and 10%, preferably a value around 5%.

Claims (11)

1. A hydraulic control system for the drive control of a double-acting hydraulic cylinder provided as the drive unit for the working tool of a processing machine by means of which a work piece, a steel plate for example, is subjected to cold deformation actions such as punching or embossing, where

   a) in the course of a processing cycle the tool performs a fast forward movement towards the work piece, a working stroke in which the work piece is actually deformed, and a fast return movement to bring it back into the starting position for the next processing cycle,

   b) the hydraulic cylinder has a total of two driving pressure spaces that are delimited in a mobile but pressure-tight manner by the different-sized surfaces F₁ and F₂ of its driving piston designed as double-diameter piston,

   (b₁) by means of which it becomes possible, when the driving and or operating pressures derived from the output pressure of a pressure source are applied to both its piston surfaces, to control the feed and working movements of the tool in fast forward operation, and

   (b₂) by means of which it becomes possible, when pressure is applied only to its larger piston surface F₁, while the smaller piston surface is relieved of pressure, to control forward working movements under load that call for a greater forward driving force, and

   (b₃) by means of which it becomes possible, when pressure is applied only to its smaller piston surface F₂, while its larger piston surface F₁ is relieved of pressure, to control the fast return movements of the tool, with

   (c) an electrically-controlled directional control valve by means of which it becomes possible - by switching it into alternative operating positions, one of which causes pressure to be applied to the driving pressure space of the hydraulic cylinder delimited by the larger piston surface F₁ while the other is associated with depressurizing this space - to control the stroke and the speed of the forward and return movements and, further,

   (d) with a surface-reversing valve that, controlled by the pressure prevailing in the larger driving pressure space of the hydraulic cylinder, can be made to switch from an operating position associated with fast forward operation, in which a pressure outlet of a pressure source is connected to the driving pressure space of the hydraulic cylinder delimited by the smaller piston surface, into an alternative positition associated with fast forward motion under a greater load, in which the smaller driving pressure space of the hydraulic cylinder is relieved of pressure, and that - by discharging the pressure from the larger driving pressure space of the hydraulic cylinder - can then be made to switch back into the operating position in which the smaller driving pressure space of the hydraulic cylinder is again connected to the pressure outlet of the pressure source, where

   (d₁) the switching of the surface-reversal valve to fast forward operation of the hydraulic cylinder under load will take place when the operating pressure in the larger driving pressure space of the hydraulic cylinder exceeds a value that corresponds to a large fraction (85%, for example) of the maximum obtainable operating pressure P_H, and where

   (d₂) the subsequent switching of the surface-reversing valve into the operating position associated with fast forward and return movements of the hydraulic cylinder will take place when the operating pressure prevailing in the larger driving pressure space of the hydraulic cylinder understeps a value that corresponds to a substantially smaller fraction (30% to 50%, for example) of the maximum exploitable operating pressure of the hydraulic cylinder,

   characterized by the following features:

   (e) the directional control valve is designed as a - by itself - known follow-up adjusting valve (44) that operates with an electrically-controlled indication of the set value - control being provided, for example, by a step motor - and mechanical feedback (possibly a worm gear)(52) of the actual position and makes it possible to obtain continuous variation of the operating pressure P_A prevailing in the larger driving pressure space (11) of the hydraulic cylinder,

   (f) in addition to a first pressure outlet (24), where the pressure supply is made available at a relatively low pressure P_N, the pressure source also has a second pressure outlet (26), where the pressure supply is provided at a markedly higher pressure level P_H,

   (g) there is provided a pressure-reversing valve arrangement (39, 58) that is controlled by the operating pressure P_A prevailing in the larger driving pressure space (11) of the hydraulic cylinder (13) and which, when and for as long as the operating pressure P_A prevailing in the larger driving pressure space (11) of the hydraulic cylinder (13) remains smaller than a switching threshold corresponding to a large fraction (for example, 85 to 95%) of the output pressure P_N made available at the low-pressure outlet (24) of the pressure source (23), will connect the low-pressure outlet (24) to the pressure supply connection (57) of the follow-up adjusting valve (44) and, alternatively, when and for as long as the operating pressure P_A prevailing in the larger driving pressure space (11) of the hydraulic cylinder (13) remains above this switching threshold, will connect the high-pressure outlet (26) of the pressure source (23) to the said pressure supply connection (57) of the follow-up adjusting valve (44),

   (h) the surface-reversing valve (42)is designed in such a manner that the switching threshold that, on being understepped, triggers the switching back of the surface reversing valve (42) into its operating position associated with the fast operating modes of the hydraulic cylinder (13) is lower than the switching threshold of the pressure reversing valve (39).
2. A hydraulic control system according to Claim 1, characterized in that the pressure reversing valve arrangement (39, 58) comprises a pressure-controlled 2/2-way valve (39) that, for as long as the operating pressure in the larger driving pressure space (11) of the hydraulic cylinder (13) remains lower than its switching threshold, is maintained in a basic position in which the pressure supply connection (57) of the follow-up adjusting valve (44) is cut off from the high-pressure outlet (26) of the pressure source and further, when and for as long as the operating pressure in the greater driving pressure space (11) of the hydraulic cylinder (13) is higher than the switching threshold (b₁ . P_N; 0.5 ≦ b₁ ≦ 0.95), switches into an open position in which the high-pressure outlet (26) is connected to the pressure supply connection (57) of the follow-up adjusting valve (44), and also comprises a non-return valve (58) inserted between the pressure supply connection (57) of the follow-up adjusting valve (44) and the low-pressure outlet (24) of the pressure source that is kept in its closed position for as long as the pressure at the pressure supply connection (57) of the follow-up adjusting valve (44) is higher than the output pressure P_N of the low-pressure output of the pressure source (23).
3. A hydraulic control system according to Claim 2, characterized in that the pressure-reversing valve (39) is designed as a slide valve in which the piston (66) is pushed into its basic position by a return force of pre-set magnitude and has a control end flange (68) that delimits in a mobile manner one side of the control pressure space, the the area f₂ of the surface of the said end flange being so dimensioned that the force that has to be exerted in order to cause the pressure reversing valve (39) to switch into its operating position in which the high-pressure outlet (26) of the pressure source .rm52 (23) is connected to the pressure supply connection (57) of the follow-up adjusting valve (44) calls for a pressure P_A that is given by
   
   ${\text{P}}_{\text{A}}{\text{≧ P}}_{\text{N}}\text{. b₁}$

   

   
   
   where b₁ stands for a coefficient that is smaller than unity (0.85 ≦ b₁ ≦ 0.95) and preferably amounts to 0.95.
4. A hydraulic control system according to Claim 3, characterized in that the valve piston (66) of the pressure reversing valve (39), has another end flange (67) at its end facing away from the control pressure space (71), in which the prevailing operating pressure P_A is the same as the output pressure of the follow-up adjusting valve (44) that is applied to the larger driving pressure space (11) of the hydraulic cylinder (13), the said other end flange constituting the mobile delimitation of a control pressure space (69) of the pressure reversing valve (39) in which there permanently prevails the output pressure P_N made available at the low-pressure outlet (24) of the pressure source (23).
5. A hydraulic control system according to Claim 4, characterized in that the ratio f₁/f₂ between the areas f₁ and f₂ of the end flanges (67 and 68), to which there are applied, respectively, the output pressure P_A of the follow-up adjusting valve and the lower output pressure P_N of the pressure source (23), has the value b₁ and, further, that the valve piston (66) of the pressure reversing valve (39) is designed as a free piston.
6. A hydraulic control system according to any of the previous claims, characterized in that the valve element of the surface-reversing valve (42) that in its open position causes the pressure to become discharged from the smaller driving pressure space (12) of the hydraulic cylinder (13) is designed as a non-return valve (96) that in its opening direction sustains the operating pressure P_A of the smaller driving pressure space (12) of the hydraulic cylinder (13) prevailing in a central valve chamber (101) of the surface-reversing valve (42) and, further, that the force with which a precompressed valve closing spring (92) pushes a valve body (94) of this non-return valve (96) into its blocking position is equivalent to an opening pressure that corresponds to a large fraction b₂ (0.85 ≦ bP₂ ≦ 0.95) of the higher pressure P_H made available at the high-pressure outlet (26) of the pressure source (23) and, further, that the surface-reversing valve (42) comprises as another valve element a slide valve (101, 111, 114, 124) that, for as long as the non-return valve (96) remains in its blocking position, assumes an open position in which the lower output pressure P_N of the pressure source (23) is applied to the smaller driving pressure space (12) of the hydraulic cylinder (13), and, upon the opening of the non-return valve (96) switches into its blocking position, in which the smaller driving pressure space (12) of the hydraulic cylinder (13) becomes cut off from the low-pressure outlet (24) of the pressure source (23), and where the slide of this additional valve element is designed as a step piston (112) that a weakly precompressed return spring (117) pushes into supporting contact with the valve body (94) of the non-return valve (96) and which is thus maintained in an operating position in which even a displacement of the step piston amounting to no more than a small fraction of the opening stroke of the non-return valve or the closing stroke of the slide valve (101, 111, 114, 124) will be quite sufficient to bring the slide valve into its blocking position, in which one side of the step piston (112) becomes depressurized, while the working surface F₅ of its other side, namely the one that delimits the control pressure space (114), in which there prevails the operating pressure P_A of the larger driving pressure space (11) of the hydraulic cylinder, becomes subject to this pressure P_A, and, further, that the ratio F₄/F₅ between this control surface F₅ and the cross section area F₄ surrounded by the valve seating (97) of the seat valve (96), so that within this area F₄ the valve body (96) becomes subject to the pressure prevailing in the smaller driving pressure space 12) of the hydraulic cylinder for as long as the non-return valve (96) remains in its blocking position, satisfies the relationship
   
   ${\text{F₄/F₅ ≦ (b₁ . P}}_{\text{N+a}}{\text{)/(b₂ . P}}_{\text{H}}\text{),}$

   

   
   
   where b₂ designates a coefficient smaller than unity (0.85 ≦ b₂ ≦ 0.95) that defines the amount by which the operating pressure P_A at which the seat valve (96) opens may understep the maximum possible possible operating pressure P_H, while a stands for a small safety margin (say, 2 to 10%).
7. A hydraulic control system according to Claim 6, characterized in that the parameter b₁ has a value between 0.85 and 0.95, preferably close to 0.9, while the parameter b₂ has a value between 0.8 and 0.95, preferably once again close to 0.9.
8. A hydraulic control system according to claim 6 or Claim 7, characterized in that the ratio F₁/F₃ between the cross section area F₁ of the larger driving pressure space (11) of the hydraulic cylinder (13) and the cross section area F₃ of the smaller working surface (12) of the hydraulic cylinder (13) amounts to between 1.5 and 3, and preferably has a value close to 2. .mb7
9. A hydraulic control system according to Claim 8, characterized in that the larger working surface F₁ of the hydraulic cylinder (13) has an area of between 60 and 300 cm², preferably close to 100 cm².
10. A hydraulic control system according to any of the preceding claims, characterized in that the ratio P_H.P_N between the output pressures P_H and P_N of the pressure source (23) has a value between 4 and 2, preferably some value close to 3.
11. A hydraulic control system according to Claim 10, characterized in that the output pressure level at the low-pressure outlet (24) of the pressure source (23) amounts to between 50 bar and 80 bar and preferably has a value close to 60.

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