Pilot operated servo valve

A pilot-operated servo valve design for efficient mounting in a control block has at least three main-stream ports. An opening into a control sleeve for a first main-stream port is disposed opposite an end surface of a main control piston. The main control piston has a pressure-equalizing surface disposed in a pressure-equalizing chamber, which is fluidically coupled by a pressure-equalizing duct in the main control piston to the first main-stream port. A return spring urges the main control piston toward a first axial end stop to clearly define a safety position. The valve exhibits good dynamic properties.

BACKGROUND OF THE INVENTION 
The invention generally relates to hydraulic servo valves and specifically 
to a pilot-operated servo valve with at least three main-stream ports for 
mounting into a control block. 
Pilot-operated electrohydraulic servo valves of twin- and multi-stage 
design with more than two main-stream ports are used, e.g. as four-way 
valves to control the position, speed, and/or force in hydraulic cylinders 
for linear movement, or position, rotation speed and/or torque in 
hydraulic motors for rotary movement. In either case the hydraulic device 
has two displacement chambers, each chamber being coupled to one of the 
main-stream ports. 
These four-way servo valves are conventionally designed as plate-stack 
valves. A main control valve for the main stage is fitted either directly 
into a valve housing or into a control sleeve which in turn is inserted 
into the housing. The openings of the main-stream ports are typically 
arranged symmetrically relative to the likewise symmetrical main control 
piston. The main control piston is hydraulically actuated by applying 
hydraulic pressure to its two end surfaces, one in each of two control 
chambers defined by end caps flange-mounted onto opposite sides of the 
valve housing. The control chambers are connected via control bores to a 
pilot servo valve. Return springs bias the main control piston to a 
centered position. 
There are various known designs for mounting valves in control blocks. For 
example, there are block mounted servo valves with high flow rates, but 
these valves have only two main-stream ports and are designed as seat 
valves. Screw-in block mounted valves with four main-stream ports are also 
known, but are designed as directional switching control valves and employ 
direct magnetic actuation. 
An issue of particular importance to the practical use of servo valves is 
that of safety in the event of breakdown or fault in the electrical drive 
system or in the pilot servo valve. Such faults must not result in an 
undefined position of the main control piston and thus in uncontrollable 
movements of the hydraulic device, such as closing movements in presses. 
Known multi-stage servo valves of the plate-stack design are constructed 
with an additional, electrically-actuated directional control or clearance 
valve disposed between the pilot servo valve and the hydraulic control 
chambers of the main control piston. In the event of a fault, this 
directional control valve reverts to a spring-biased, center position, in 
which the connection to the pilot servo valve is interrupted and the 
control chambers of the main control piston are fluidically coupled. The 
main control piston is thus biased by two compression springs to a 
centered position between two spring plates abutting the housing. To 
achieve well defined behavior of the cylinder movement when the main 
control piston is centered, the valve control edges must use positive 
overlapping (that is, the axial distance between the control edges 
assigned to a port is greater than the port's axial extent), at least in 
the direction of the pressure source. As compared to designs using 
zero-overlapping (that is, the axial distance between the control edges is 
equal to the port's axial dimension) of the four control edges between 
pressure source, working ports and tank return circuit, such control edge 
positive overlap has serious drawbacks as to the positioning accuracy of 
the cylinder in position-control devices and when the valve is used for 
pressure regulation. 
There is therefore a need for a pilot-operated servo valve that can be 
effectively block mounted and that has a clearly-defined safety position 
without sacrificing the good dynamic properties available with 
zero-overlapping control edges. In particular, there is a need to provide 
these properties in a servo valve of the construction described above, 
i.e., with a control piston slidably mounted in a control sleeve that has 
axially spaced openings for at least three main stream ports, the piston 
having first and second control edges controlling flow through hydraulic 
connections between the first and second, and second and third main-stream 
ports, respectively, and in which the piston's movement is controlled by a 
pilot valve that selectively supplies pressurized fluid to at least one of 
two control chambers that act on opposing first and second actuating 
surfaces of the piston, the pilot valve in turn being in a control loop 
that takes input from a position transducer coupled to the piston. 
SUMMARY OF THE INVENTION 
These needs are met by the servo valve of the invention, which can be 
integrated in a space-saving manner into a control block, has a clearly 
defined safety position, and has good dynamic properties. 
The opening into the control sleeve for the first main-stream port is 
disposed opposite a first end of the main control piston, while the 
openings for the other main-stream ports are disposed to the side of the 
main control piston. The main control piston incorporates a stop surface, 
which, by interaction with a corresponding counter-stop surface, 
mechanically defines a safety end position of the main control piston. A 
return spring urges the main control piston towards this end position. 
The servo valve preferably includes a pressure-equalizing chamber 
fluidically coupled by a pressure-equalizing duct in the main control 
piston to the first main-stream port and in which a pressure-equalizing 
surface on the main control piston is disposed to hydrostatically oppose 
the first piston end-surface. 
The servo valve's control sleeve is inserted directly into a stepped bore 
in the control block. The control block has lateral block bores for the 
second, third and additional main-stream ports. However, the design 
affords great flexibility for arrangement of the block bore for the first 
main-stream port. The block bore for the first main-stream port can be 
disposed, for example, in direct axial extension of the stepped bore for 
the control sleeve, which has not previously been possible in the case of 
traditional pilot-operated servo valves with more than two main ports. The 
need for bridgings in the control block between individual openings into 
the control sleeve is also eliminated. This design affords a more compact 
construction of the control block than is possible with traditional servo 
valves. Even in more complex hydraulic control systems, the servo valve 
according to the invention, together with various additional valves, for 
example two-way built-in valves, can be integrated in a space-saving 
manner into a control block. Direct mounting in the cylinder cover of 
larger cylinders is likewise possible. 
A safety position of the servo valve is clearly, mechanically defined in 
the first axial end position of the main control piston by the direct 
butting contact of the main control piston against the sleeve, with the 
return spring urging the main control piston directly towards this 
position. Even where there is zero-overlapping of the control edges, the 
behavior of the valve in this safety position is clearly defined, which is 
not possible in the case of traditional, middle-centered servo valves. 
The asymmetrical hydrostatic loading of the main control piston is 
compensated for by corresponding dimensioning of a pressure-equalizing 
surface. This hydrostatic compensation reduces the required main control 
piston actuating forces, allowing the actuating surfaces in the control 
chambers to be smaller. This results in smaller control-oil volumes, which 
means that shorter correction times are obtained for a given size pilot 
valve. 
The valve according to the invention is preferably a four-way valve, with a 
fourth main-stream port opening in the control sleeve, an auxiliary 
connecting chamber connected via a cross-bore of the main control piston 
to the pressure-equalizing duct of the main control piston, third and 
fourth control edges on the piston controlling flow through hydraulic 
connections between the third and fourth main-stream port openings and 
between the auxiliary connecting chamber and the fourth main-stream port 
opening. This design does not require bridgings in the control block. 
In a preferred embodiment, the second end of the main control piston is 
introduced, axially sealed, into the pressure-equalizing chamber to 
present a pressure-equalizing surface on the second end of the piston. 
This allows a more compact valve structure than is possible with an 
annular pressure-equalizing surface (although the latter embodiment is not 
precluded). 
A hydrostatic over-compensation of the servo valve may be achieved by 
sizing the pressure-equalizing surface to have a greater axial area than 
that of the first piston end-side. Whenever the first main-stream port is 
pressurized, a correcting force therefore acts to urge the main control 
piston towards the first main-stream port and supplements the biasing 
force of the return spring. 
Preferably, the first main-stream port is coupled to a pump and thus forms 
a pump port, the second main-stream port is coupled to a first 
displacement chamber of an energy-consuming unit and thus forms a first 
working port, the third main-stream port is coupled to a tank and thus 
forms a tank port, and the fourth main-stream port, where present, is 
coupled to a second displacement chamber of an energy-consuming unit and 
thus forms a second working port. In this design, the pump port can be 
introduced axially into the control sleeve, and the tank port can be 
located between the first and second working ports. However, other 
assignments of the main-stream ports are also possible. 
As referred to herein, a "pump" is a hydraulic pressure source or line, a 
"tank" is a vessel or a line without significant counter-pressure, and an 
"energy-consuming unit" is for example a hydraulic rotary or linear drive 
system. 
The control edges of the main control piston preferably exhibit a 
zero-overlapping. This gives excellent positioning accuracy, where the 
valve is used in a position-control circuit of a hydraulic cylinder, and 
excellent dynamic behavior, where the valve is used for 
pressure-regulating purposes. Since the valve is not middle-centered in 
its safety position, but has an axial end position, the zero-overlapping 
of the control edges has no adverse effect on the behavior of the valve in 
its safety position. 
In one embodiment, the control edges are disposed so that when the main 
control piston is in its safety position, the first working port is 
connected to the tank port, while the second working port is connected via 
the pressure chamber to the pump port. 
In another embodiment, the control edges are disposed so that when the main 
control piston is in its safety position, the second working port is shut 
off from the pump port and is coupled to the tank port. 
In another embodiment, the control edges are disposed so that when the main 
control piston is in its safety position, the first and second working 
ports are shut off from both the pump port and the tank port. 
In a further embodiment, a clearing valve is connected between the pilot 
valve and the main stage. When the clearing valve is relieved into a 
spring-biased basic position, for example in the event of an emergency 
shut-down signal or fault signal, the main control piston is urged by its 
return spring, and preferably by additional hydraulic pressure forces, 
into its safety end position. 
With the above-described piston geometry, a hydraulic cylinder can thus, 
for example, either be stopped by shutting off the working ports or 
depressurized by connecting the working ports to the tank. This prevents 
uncontrolled travel of the cylinder to an end position when the control 
electronics fail in the machine control system or even in the pilot valve 
itself. 
In another embodiment, the servo valve can be pilot controlled by a simple 
three-way pilot valve. This is accomplished by directly coupling the 
second control chamber (which contains the second piston actuating 
surface, oriented to urge the piston toward the safety end position) to a 
constantly unpressurized tank line. The first control chamber (which 
contains the first actuating surface oriented to urge the piston away from 
the safety end position) is coupled to the working port of the pilot 
valve. 
In one embodiment, the position transducer of the main control piston is a 
path-measuring system with an electrical output and is integrated with the 
pilot valve into a closed control loop. In another embodiment, the servo 
valve has a mechanical feedback loop, using a three-way pilot slide valve 
extending axially from the second end of the main control piston. This 
pilot slide valve has a pilot pressure port, a pilot tank port, a pilot 
working port, and a slide piston. A measuring spring connects the slide 
piston axially to the main control piston and an actuating magnet, acting 
proportionally to an electric signal, is connected mechanically to the 
slide piston. The positioning of the main control piston is thus effected 
in a closed position-control circuit until force equilibrium between the 
magnetic force and the measuring-spring force is achieved. 
With an additional clearing valve connected in the pilot-control system, 
additional cut-out safety can also be achieved with a three-way pilot 
valve. The pilot pressure port is in this case directly coupled via the 
clearing valve to the pilot tank port, and the pilot working port is 
coupled either to the pilot tank port or to the pilot pump port, depending 
on the slide piston's position. When the clearing valve is relieved, the 
main control piston travels, as described above, into its first axial end 
position.

DETAILED DESCRIPTION 
FIG. 1 presents a longitudinal sectional view through a first embodiment of 
a servo valve 3 embodying the principles of the present invention. A 
control sleeve 5 is mounted in a stepped bore 2 of a control block 1. A 
main control piston 6 is mounted in control sleeve 5 for sliding axial 
movement. The illustrated servo valve 3 is a four-way servo valve, having 
a pump port P, a tank port T, and first and second working ports A, B. 
Pump port P is fluidically coupled to a pressure line (i.e., a source of 
pressurized hydraulic fluid, not shown). Tank port T is fluidically 
coupled to an unpressurized line (not shown). Working ports A and B are 
fluidically coupled to first and second displacement chambers, 
respectively, of an energy-consuming unit (e.g., a hydraulic linear or 
rotary drive system) (not shown). 
A first control block bore 50 for pump port P opens, as a coaxial extension 
of step bore 2, into pump port opening 50' of control sleeve 5. Three 
block bores 51, 52, 53 in control block 1 for tank port T, first working 
port A, and second working port B, respectively, are disposed transversely 
to step bore 2 and open out laterally into axially spaced, annular 
channels 51', 52', 53' in the outside surface of control sleeve 5, which 
in turn communicate with the interior of control sleeve 5 via 
circumferentially spaced openings therethrough. For purposes of 
description herein each annular channel and its corresponding 
circumferentially spaced openings are collectively referred to as a tank 
port or working port "opening." 
The stepped internal surface of control sleeve 5 can be considered to 
define, in addition to the port openings, a series of chambers and axial 
hydraulic passages or connections between the chambers and the port 
openings. For example, portion 28 of the internal surface of control 
sleeve 5 between pump port opening 50' and first working port opening 52' 
is considered to be a first hydraulic connection that connects the two 
openings. Similarly, a second hydraulic connection 29 connects tank port 
opening 51' to first working port opening 52', a third hydraulic 
connection 30 connects tank port opening 51' to second working port 
opening 53', and a fourth hydraulic connection 31 connects second working 
port opening 53' to an auxiliary connecting chamber 22 disposed coaxially 
within the control sleeve 5. The axial distance between second and third 
hydraulic connections 29 and 30 is much greater than the axial distances 
between first and second hydraulic connections 28 and 29 or between third 
and fourth hydraulic connections 30 and 31. 
Main control piston 6 has a first coaxial piston collar 8, which is 
assigned to working port A and is displaceable axially into first and 
second hydraulic connections 28 and 29, and a second coaxial piston collar 
9, which is assigned to working port B and is displaceable axially into 
third and fourth hydraulic connections 30 and 31. First piston collar 8 
has a first control edge 28' that is assigned to (controls flow through) 
first hydraulic connection 28, and a second control edge 29' that is 
assigned to second hydraulic connection 29. Second piston collar 9 has a 
third control edge 30' that is assigned to third hydraulic connection 30 
and a fourth control edge 31' that is assigned to fourth hydraulic 
connection 31. All four control edges 28', 29', 30', 31' have 
zero-overlapping. 
Main control piston 6 has an axial end surface 12 that is disposed opposite 
pump port P and is therefore always acted on by the supply hydraulic 
pressure. An axial piston bore 18 extends from piston end surface 12, 
through the piston body to piston cross-bores 19, disposed above (in FIG. 
1) second piston collar 9. Piston cross-bores 19 open into auxiliary 
connecting chamber 22 formed in control sleeve 5. Thus, auxiliary 
connecting chamber 22 is constantly fluidically coupled to pump port P and 
therefore operates at the supply hydraulic pressure. Accordingly, main 
control piston 6 selectively connects first working port A (with coaxial 
piston collar 8) and second working port B (with coaxial piston collar 9) 
to pump port P or tank port T. The respective flow rates of hydraulic 
fluid between the working ports and the pump or tank ports is regulated by 
the four control edges 28', 29', 30', 31'. 
The hydraulic pressure on piston end-surface 12 presents an asymmetrical 
axial hydrostatic load on main control piston 6. To equalize the 
hydrostatic forces on main control piston 6, the second, opposite end of 
main piston 6 is disposed in a pressure-equalizing chamber 25, which is 
disposed in a valve cap 40. Pressure-equalizing chamber 25 is maintained 
at the same hydraulic pressure as supply port P by connecting them via 
coaxial piston bore 18, which extends to the second end of main control 
piston 6 and is fluidically coupled via piston cross-bores 20 with 
pressure-equalizing chamber 25. The second end of main control piston 6 
extends into pressure-equalizing chamber 25, axially sealed by a sealing 
insert 7. The portion of main control piston 6 that extends into 
pressure-equalizing chamber 25 is referred to as pressure-equalizing 
protrusion 21. The cross-sectional area of pressure-equalizing protrusion 
21 presents a pressure-equalizing surface (which is indicated in FIG. 1 by 
reference to annular shoulder 13) which hydrostatically opposes piston 
end-surface 12. Full hydrostatic pressure equalization is obtained if 
pressure-equalizing surface 13 is chosen to be equal in area to piston 
end-surface 12. Hydrostatic over-compensation is achieved if 
pressure-equalizing surface 13 is chosen to be greater in area than piston 
end-surface 12. 
Axial movement of main control piston 6 is effected via coaxial actuating 
piston collar 11 by the imposition of appropriate hydraulic pressure on 
its annular first or second actuating surfaces 14, 15. Piston collar 11 
divides the large internal diameter portion at the upper end of control 
sleeve 5 into first and second control chambers 26 and 27, within which 
hydraulic pressure acts on first and second actuating surfaces 14 and 15, 
respectively. Control chambers 26 and 27 are connected via pilot ports to 
working ports A' and B' of a flange-mounted, four-way pilot servo valve 
60. The axial position of main control piston 6 is measured by electrical 
position transducer 63. The output of transducer 63 (i.e., the position of 
main control piston 6) is input into electronic control amplifier 64, 
which compares this actual position information to a desired value, and 
outputs a control signal to pilot servo valve 60, thus forming a closed 
electrohydraulic feedback loop. 
The dimensions of actuating surfaces 14, 15 are selected so that the flow 
forces generated when the control edges 28', 30' or 29', 31', 
respectively, are overflowed are reliably overcome. For a given pilot 
servo valve 60, very short correction times for the positioning of the 
main control piston 6 can thus be achieved. 
The range of axial movement of main control piston 6 is bounded by axially 
opposed first and second end positions, which are defined mechanically by 
an annular stop surface 16 on a shoulder formed on the portion of main 
control piston in control chamber 26 and by an end stop surface 17 at the 
second end of the main control piston, respectively. When first control 
chamber 26 is not pressurized, main control piston 6 is urged downwardly 
by a return spring 24, which is disposed, for example, in 
pressure-equalizing chamber 25, until stop surface 16 abuts against a 
counter-surface 16' formed in the internal surface of control sleeve 5. In 
the first end position, the piston is disposed so that first control edge 
28' closes first hydraulic connection 28 while second control edge 29' 
opens second hydraulic connection 29, so that working port A is 
disconnected from pressure port P and connected to tank port T. Further, 
fourth control edge 31' opens fourth hydraulic connection 31 and closes 
the third hydraulic connection 30, so that working port B is connected 
(via auxiliary connecting chamber 22) to pump port P and is disconnected 
from tank port T. Therefore, in this position, working port A is 
depressurized while working port B is pressurized. 
In many applications, such as presses and injection-molding machines, the 
cylinder controlled by the servo valve must operate fail-safe. That is, if 
a safety cut-out occurs or if the drive electronics fail or develop a 
fault, the controlled cylinder must not move. To achieve this result, both 
working ports A and B must either be depressurized (coupled to tank port 
T) or shut off. This has not previously been possible for servo valves 
with zero-overlapping control edges, but is achieved in the second 
embodiment of the present invention, illustrated in FIGS. 2A and 2B. 
In the second valve embodiment, main control piston 6 has a first coaxial 
auxiliary piston collar 32, which has first auxiliary control edge 32'. 
When main control piston 6 is in its first end position, first auxiliary 
control edge 32' closes fourth hydraulic connection 31 (from auxiliary 
connecting chamber 22 to second working port B), while third and fourth 
control edges 30' and 31' (on second coaxial piston collar 9) 
simultaneously open third hydraulic connection 30 (between tank port T and 
working port B)--working port B is thus depressurized. First working port 
A is similarly depressurized since second control edge 29' (on first 
coaxial piston collar 8) opens second hydraulic connection 29. Both 
working ports, and therefore both working chambers in the energy-consuming 
unit, are depressurized. 
The alternative fail-safe mode (in which both working ports A and B are 
shut off when main control piston is in its first end position, thereby 
locking the driven cylinder in place even when external loads are imposed 
on it) is achieved by a third valve embodiment, illustrated in FIGS. 3A 
and 3B. In this embodiment, main control piston 6 has a second coaxial 
auxiliary piston collar 33, which has second auxiliary control edge 33'. 
Further, as compared to the second embodiment, first auxiliary control 
edge 32' is moved closer to fourth control edge 31'. Therefore, when main 
control piston 6 is in its first end position, first auxiliary control 
edge 32' closes (as in the second embodiment) fourth hydraulic connection 
31 (between the auxiliary connecting chamber 22 and the working port B) 
and fourth control edge 31' closes third hydraulic connection 30 (between 
tank port T and working port B). Simultaneously, first control edge 28' 
closes first hydraulic connection 28 (between pump port P and working port 
A), and second auxiliary control edge 33' closes second hydraulic 
connection 29 (between tank port T and working port A). Thus, both working 
ports A and B are shut off, both to the tank side and to the pressure 
side. 
In a fourth valve embodiment illustrated in FIG. 4, main control piston 6 
is configured to produce an over-compensating hydrostatic force on the 
piston that acts in concert with the bias force of return spring 24 to 
urge the piston toward its first end position. As shown in FIG. 4, this is 
achieved by increasing the diameter of pressure-equalizing protrusion 21, 
so that pressure-equalizing surface 13 is greater in area than piston 
end-surface 12. 
A further feature illustrated in the fourth embodiment is a clearing valve 
62, connected between four-way pilot servo valve 60 and first control 
chamber 26. The ports of pilot servo valve 60 are identified similarly to 
those of servo valve 3--its pilot pressure port (coupled to control 
pressure line X) is identified as P', its pilot tank port (coupled to 
unpressuized control line Y) is identified as T', its first pilot working 
port (coupled via line 56 and clearing valve 62 to first control chamber 
26) is identified as A', and its second pilot working port (coupled to 
second control chamber 27) is identified as B'. When clearing valve 62 is 
in its spring-biased position (i.e. the solenoid is not energized), first 
control chamber 26 (which contains actuating surface 14) is depressurized 
(relieved in the direction of the tank). Therefore, regardless of the 
position of pilot servo valve 60, main control piston 6 is urged into its 
first end position. Pilot servo valve 60 only becomes effective for 
positioning main control piston 6 when clearing valve 62 is energized into 
its second position, in which second pilot working port B' is coupled to 
first control chamber 26. 
In accordance with a fifth embodiment of the invention, illustrated in FIG. 
5, the cost of the fourth valve embodiment can be reduced by controlling 
the position of main control piston 6 with a simpler, three-way pilot 
valve 61. Pilot valve 61 has a pilot pump port P', a pilot tank port T', 
and a single pilot control port A'. Pilot pump port P' is pressurized via 
control pressure line X, while pilot tank port T' is connected to an 
unpressurized control line Y. Pilot control port A' coupled via clearing 
valve 62 to first control chamber 26. 
Since, in this embodiment as in the fourth embodiment, pressure-equalizing 
protrusion has an increased diameter, second actuating surface 15 in 
second control chamber 27 is smaller than first actuating surface 14 in 
first control chamber 26. Second control chamber 27 is constantly relieved 
in the direction of the tank via line 54 and unpressurized control line Y. 
When pilot valve 61 has not been triggered, main control piston 6 is in 
the workrest position. When pilot valve 61 is electrically triggered 
(i.e., its solenoid is energized), a hydraulic control force is generated 
by pressurization of the larger, first actuating surface 14, controlling 
the position of main control piston 6, as described above, with the 
electrohydraulic position-control feedback loop. To achieve this 
positioning, however, clearing valve 62 must be energized. When clearing 
valve 62 is relieved into its basic position, e.g. as a result of a fault, 
control chamber 26 is again depressurized so that, regardless of the 
position of pilot valve 61, main control piston 6 is urged into its first 
end position by the hydrostatic over-compensation and by return spring 24. 
If the additional safety feature of clearing valve 62 is not required, the 
valve can be omitted. Control chamber 26 is then connected directly to 
pilot control port A' of pilot valve 61. 
As an alternative to the electrical feedback loop illustrated in FIGS. 1 to 
5 and described above, which uses the electrical position transducer 63 
shown, a mechanical feedback loop may be used in accordance with a sixth 
embodiment, illustrated in FIG. 6. In this embodiment, a three-way piston 
slide valve 67 is disposed as an axial extension of main control piston 6. 
It has a pilot pump port P' (with associated line 58), a pilot tank port 
T' (with associated line 59), a pilot control port A' (with associated 
line 56) and a slide piston 68. Slide piston 68 is supported at one end on 
a spring plate 69 in pressure-equalizing chamber 25 and is connected at 
its second end to a proportional magnet 66. A measuring spring 65 is 
disposed in pressure-equalizing chamber 25 between spring plate 69 and 
main control piston 6. Slide piston 68 is axially bored through for 
hydrostatic pressure-equalization. Regardless of the position of slide 
piston 68, control port A' is constantly connected either to pilot pump 
port P' or to pilot tank port T'. Main control piston 6 has the same 
configuration as in the fourth and fifth embodiments, and thus has the 
same characteristics. 
The force build-up of proportional magnet 66 is proportional to an 
electrical control current, i.e. the desired value. The spring force of 
measuring spring 65 is proportional to the position of main control piston 
6, i.e. the actual value. The output control pressure of pilot slide valve 
67, which is acting on first actuating surface 14, is corrected in the 
event of differences between the desired and actual value until the 
electrically pre-determined position in the position-control feedback loop 
is achieved. 
A clearing valve 62 is coupled to pilot pump port P' (at one end of line 
58'). When clearing valve 62 is electrically relieved into its basic 
position, pilot pump port P' is depressurized by coupling it to 
unpressurized control line Y (via lines 59, 54). Regardless of the 
position of pilot slide valve 67, actuating surface 14 is therefore always 
depressurized, so that main control piston 6 is urged by the hydrostatic 
over-compensation and return spring 24 into its first end position. 
Again, if safety requirements are reduced, clearing valve 62 can be omitted 
without affecting the reliable basic functioning of servo valve 3.