Dual-circuit pressure control valve for hydraulic brake systems

Dual-circuit pressure control valves are known in which the control pistons are acted upon by a common control force. A preload distributor constructed as beams of balance to compensate for differences in the control behavior of the two control valves due to manufacturing tolerances have not provided the desired compensation. Therefore, according to the present invention an improved compensation arrangement is provided in the form of a compensating piston which is subjected to the outlet pressures of the two brake circuits.

BACKGROUND OF THE INVENTION 
The present invention relates to a dual-circuit pressure control valve for 
hydraulic brake systems having two control valves arranged in a housing in 
a parallel side by side relationship each assigned a different one of the 
two brake circuits, each of the two control valves having a separate 
control piston located between an associated inlet chamber and an 
associated outlet chamber, and the control pistons are subjected to a 
common control force acting on the pistons through the intermediary of a 
preload distributor. 
In a known dual-circuit pressure control valve of the aforementioned type 
such as disclosed in German Pat. DE-OS No. 2,614,080, the two control 
pistons are arranged in a parallel side by side relationship. The preload 
distributor includes a semicircular disc made of elastic material which 
abuts with its entire circumferential surface on a supporting element with 
the same radius. The elastic disc has both end surfaces embraced by 
further wall components of the supporting element and which has a rigid 
beam at the diameter surface, against which beam the control pistons are 
adapted to bear. The supporting element is carried by a lever upon which a 
control force acts which is variable dependent upon the vehicle's axle 
load. With varying pressures prevailing at the outlet of the two pressure 
control valves, the system comprising the two control pistons and the 
distributor will be displaced in such a way that additional pressure fluid 
is fed to the brake circuit having the lower pressure until the balance in 
pressure is re-established. Thus, it is possible to compensate to a 
certain extent for discrepancies in tolerance occurring in the manufacture 
of the control valves. Moreover, if one brake circuit fails, the pressure 
in the still intact brake circuit is allowed to increase. It has to be 
taken into consideration, however, that the distributor is largely made of 
elastic material, in particular rubber, and is, therefore, subjected to 
substantial aging and wear phenomena. This is especially true due to the 
considerable fluctuations in temperature occurring in automotive vehicles, 
due to the ingress of dirt and the strong forces to be absorbed during 
each braking operation, which leads to a deformation or a grinding along 
the supporting surface. Besides, difficulties arise if the two control 
pistons do not act upon the distributor precisely symmetrically. 
SUMMARY OF THE INVENTION 
It is an object of the present invention to provide a dual-circuit pressure 
control valve of the type referred to hereinabove, in which all tolerances 
of the control valves are compensated for and in which the two outlet 
pressures of the pressure control valve are equal during the entire life 
span of the device. 
A feature of the present invention is the provision of a dual-circuit 
pressure control valve for hydraulic brake systems comprising two control 
valves disposed in a housing in a parallel side by side relationship, each 
of the two control valves controlling a different one of two brake 
circuits, having a control piston disposed between an inlet chamber and an 
outlet chamber and subjected to a common control force acting on each of 
the control pistons through a preload distributor; and means disposed in 
the housing disposed between the outlet chamber of one of the two control 
valves and the outlet chamber of the other of the two control valves to 
provide a hydraulic balancing of the outlet pressure in each of the outlet 
chambers. 
As a result of this measure, all manufacturing tolerances of the control 
valves and of the preload distributor are compensated for and the closure 
travel of the control valves is kept small. 
In a suitable improvement upon the subject matter of the present invention, 
the means for the hydraulic balancing is a piston with end surfaces of 
equal size, which piston is acted upon by the pressures in the outlet 
chambers and on account of whose movement the valve of the one brake 
circuit is adapted to be controlled. It is attained by this arrangement 
that the piston has to move only in the area of the closure travel of the 
valve. To keep the loss in volume of the one brake circuit at a minimum 
possible rate in the event of failure of the other brake circuit and to 
ensure an increase of the change-over point in the event of a circuit 
failure independently of the piston's diameter, it is suggested that the 
piston be displaceable within limits. This may be accomplished in a 
particularly simple way by providing the piston with a radial extension 
which is located between a shoulder and a stop ring fastened in the 
housing. Advantageously, the radial extension is formed by a collar 
arranged at the end surface of the piston defining the outlet chamber. 
To enable the piston to directly control the valve without insertion of 
transmitting members, the piston and the stepped piston are arranged in 
series on a common axis. 
For increasing the change-over point of the intact brake circuit to the 
double pressure valve upon failure of the other brake circuit, the preload 
distributor is favorably guided in the direction of the axes of the 
stepped pistons. To vary the control force acting on the stepped pistons 
dependent upon the vehicle's axle load, it is advantageous that the 
preload distributor is a lever adapted to swivel around a transverse axis 
of the stepped pistons. Preferably, the stepped piston includes a 
clearance relative to the preload distributor. It is thereby obtained that 
a valve closes to begin with and that a further pressure increase in this 
circuit is effected by controlling the valve through the piston. To be 
able to determine the clearance exactly, it is suitable to provide a means 
for adjustment of the clearance. A spring may be arranged between the 
stepped piston and the preload distributor which will load the stepped 
piston when depressurized back to its end position close to the outlet 
chamber. The same effect as that of the clearance between stepped piston 
and preload distributor may be achieved by providing the control valves 
with different closure travels. 
A dual-circuit pressure control valve, in which the control pistons are of 
smaller diameter, is advantageously constructed in such a manner that the 
control valves each include one slidable sleeve being sealed at its outer 
periphery and forming a valve seat at the one end surface, that the shanks 
of the control pistons penetrate each of the sleeves with clearance and 
have a valve plate at its end portion, that the one sleeve is synchronized 
with the compensating piston and that a stop is provided for fixing the 
rest position of the other sleeve. 
The control piston has a comparatively small diameter in this construction, 
since it does not have to incorporate any valve components in its inside. 
Due to the smaller cross section, it is possible to operate the device 
with a lower amount of control force than heretofore. Therefore, a less 
powerful and-with regard to the manufacturing tolerances-less exact spring 
may be employed, or a smaller lever transmission is sufficient. Both 
solutions result in a reduced space requirement. The diameter to be sealed 
is comparatively small, the friction forces which have to be overcome upon 
a displacement are correspondingly insignificant. Thereby a very precise 
operation of the control piston is achieved. The sleeve bearing the valve 
seat is freely accessible at its periphery. It is, therefore, not 
difficult to couple the one sleeve with the compensating piston. Valve 
closure springs are not required. This avoids in addition the occurrence 
of the reactive effect of such a spring on the compensating piston. 
Advantageously, the compensating piston is of a larger outside diameter 
than the sleeve. It is thereby accomplished that even slightest 
differences in the outlet pressure cause a sufficiently great force to 
displace the associated sleeve. 
Favorably, the compensating piston and the associated sleeve are coaxially 
arranged in tandem and are rigidly connected to each other by means of a 
bridge extending over the valve plate cooperating therewith. With the aid 
of the bridge, the sleeve may be loaded axially from the compensating 
piston despite the existence of the valve plate. 
For the resetting of the sleeve, the control piston advantageously includes 
a shoulder which is able to act upon an end surface of the sleeve remote 
from the valve seat. Since the control pistons are reset by the control 
force and entrain the sleeve, no separate return spring is required for 
the sleeve. The advantage is that the compensating piston is able to 
displace the sleeve without having to overcome a spring force, thereby 
rendering possible a still more precise pressure balance. 
For fixing the rest position of the other sleeve, a stop formed in the 
housing may cooperate with a step at the outer periphery of the sleeve. 
This results in a particularly simple construction. 
In a dual-circuit pressure control valve with unsymmetrical construction, 
slight differences in the control behavior of the valves cannot be avoided 
completely, and, moreover, the control valve assigned to the compensating 
piston has to close first. It is, therefore, particularly expedient to 
have the compensating piston act on a distribution device which will 
influence both control valves at the same time, but in an opposite sense 
in the event of an actuation. 
This distribution device results in both control pistons and both control 
valves operating completely equally in status. There is no need for a lost 
motion for the one control piston. If the control valves include locking 
springs, the effect of these springs will be equalized by each other. The 
desired like pressure in both outlet chambers is achieved quickly in each 
case by the opposing influence on the two control valves. In addition to 
this, the components are accommodated symmetrically in the housing so that 
a simpler and more compact construction is achieved. 
According to a preferred embodiment, the distribution device comprises a 
distribution element which is slidably arranged transversely to the axes 
of the two control pistons and each includes an inclined surface to 
additionally govern a control valve with the inclined surfaces being 
oppositely sloped relative to each other. This leads to a very simple 
construction of the distribution element and permits a space-saving 
construction in comparison with tiltable distribution elements or the 
like. 
In particular, the compensating piston and the distribution element may be 
integrally formed. This combination saves component parts. Since the 
compensating piston is situated transversely to the control pistons, a 
housing is obtained which has a short length in direction of the control 
pistons' axes. 
When using a cylindrical compensating piston, the compensation piston is 
provided with conical inclined surfaces close to its end portions. 
It is, moreover, favorable to have a transmission piston seated on the 
inclined surface, which transmission piston is guided on the same axis 
with the associated control piston in the housing and acts on the control 
valve. This transmission piston ensures that the valve closure member is 
not loaded in the transverse direction on account of friction between the 
inclined surface and a valve tappet. 
In another development of the present structure, an auxiliary inclined 
surface being oppositely sloped can join the outer ends of each inclined 
surface. In particular, the inclined surface and the auxiliary inclined 
surface can form a double cone. The auxiliary inclined surface enables, 
upon failure of the one circuit, the control valve of the other circuit to 
be urged compulsorily to the open position so that in this circuit the 
outlet pressure follows the inlet pressure even in the case of higher 
values. In this arrangement, the remaining stroke of the stepped piston is 
desired to be only somewhat greater than the closure travel of the valve. 
Besides, a differential pressure indicating device can be provided, whose 
actuating element engages in a groove in the piston portion between the 
inclined surfaces. The groove requires including the seals on both sides 
of a specific axial length of the compensating piston. Since the 
compensating piston is no longer permitted to be arranged in axial 
prolongation of the one control piston but transversely thereto, it is not 
difficult to construct the compensating piston with this specific length.

DESCRIPTION OF THE PREFERRED EMBODIMENTS 
In FIG. 1, a housing 1 accommodates two parallel, stepped bores 2 and 2' in 
which the control valves are disposed. Since the two control valves are 
substantially alike, the valves will not be described separately in the 
following; however, the valve components will be assigned the reference 
numerals of both devices. Slight differences will be pointed at where need 
be. 
A sleeve 3,3' is arranged in the enlarged bore portion of bore 2,2' and 
sealed relative to housing 1 by means of a sealing ring 4,4'. Sleeve 3,3' 
abuts a first shoulder 5,5' of housing 1 and is secured against axial 
displacement by means of a ring 6,6' fastened in the housing 1. Located in 
each bore 2,2' is a stepped piston 7,7' which is guided by its larger 
diameter portion in bore 2,2' and with its smaller diameter portion in 
sleeve 3,3' and which is sealed by seals 8,8' and 9,9'. 
An inlet chamber 10,10' is bounded by an annular surface between the steps 
of stepped piston 7,7' and an outlet chamber 11,11' by the end surface of 
the larger diameter portion of stepped piston 7,7'. A radial bore 12,12' 
and a coaxial fluid passageway 13,13' in the inside of stepped piston 7,7' 
connects inlet chamber 10,10' to outlet chamber 11,11'. Disposed in fluid 
passageway 13,13' is a valve closure member 14,14' being loaded against a 
valve seat 16,16' by a spring 15,15'. Valve closure member 14,14' 
accommodates a tappet 17,17' projecting from stepped piston 7,7'. 
Arranged in an extension of bore 2 tapered in its diameter is a piston 18 
which is sealed relative to housing 1 by means of two seals 19 and 20. 
Piston 18 defines with its end surfaces, being of equal size, outlet 
chamber 11, on the one hand, and a pressure chamber 21, on the other hand. 
Pressure chamber 21 communicates with outlet chamber 11' via a pressure 
fluid channel 22. The end portion of piston 18 close to outlet chamber 11 
is radially enlarged in the form of a collar 23, with collar 23 being 
disposed between a second shoulder 24 of housing 1 and a stop ring 25 
fastened in housing 1. Piston 18 is displaceable within limits, with its 
respective end position being defined by shoulder 24 and stop ring 25. 
Tappet 17 of valve closure member 14 projecting from stepped piston 7 bears 
against collar 23, and tappet 17' of valve closure member 14' projecting 
from stepped piston 7' bears against the bottom of bore 2'. Because of 
this, both control valves are opened in the inactivated position of the 
device. 
Secured to housing 1 by means of a pivot 26 is a lever 27 which is adapted 
to swivel around the longitudinal axis of pivot 26. Lever 27 bears against 
one end of one of stepped pistons 7,7' in the inactivated position of the 
device. At its point lying opposite the end of piston 7, lever 27 includes 
a threaded portion 30 in which an adjusting screw 31 is accommodated. In 
the embodiment shown, the piston end of stepped piston 7 has a clearance 
relative to adjusting screw 31, which is smaller than the valve closure 
travel and whose significance will be described below when the operation 
of the device is described. 
A longitudinal cross section taken along the line II--II in FIG. 1 is shown 
in FIG. 2. The reference numerals of the individual elements correspond to 
those in FIG. 1. From FIG. 2, the moving ability of lever 27 can be 
clearly seen, and that a force designated F acts on lever 27. Force F is 
variable and serves as a control force of valve pistons 7,7'. In addition, 
FIG. 2 shows the pressure fluid ports leading to the master cylinder Hz 
and the wheel cylinder Rz. 
The mode of operation of the braking pressure control unit illustrated in 
FIGS. 1 and 2 will be first described assuming that both brake circuits 
are operable. In the inactivated position of the device, the movable parts 
will be positioned as illustrated with the exception that stepped piston 7 
abuts adjusting screw 31 and clearance a is situated between collar 23 and 
housing shoulder 24. Closure members 14,14' of the control valves are 
opened in this position. 
When the brake is actuated, the pressure fluid in both brake circuits will 
first of all be allowed to flow unhindered from the master cylinder Hz to 
the wheel cylinders Rz. Acting in each case on the stepped pistons 7,7', 
due to different sized pressurized surfaces, is a differential of force 
which causes a movement of stepped pistons 7,7' in opposition to lever 27 
and control force F acting thereupon. 
Since the closure travel of the valve 14,16 is reduced by the amount of the 
clearance a, it will be valve closure member 14 that closes fluid 
passageway 13 first by abutting the valve seat 16 after the remaining 
closure travel has been overcome. The other valve 14',16', the closure 
travel of which is not reduced by the amount of the clearance a, will be 
still open in the event of the same displacement travel of both stepped 
pistons 7,7', and the pressure build-up will thus be continued 
undiminished in the associated brake circuit. Due to the slight difference 
in pressure occurring in outlet chambers 11 and 11', a resultant force 
will act on piston 18 and displace piston 18 in the direction of stepped 
piston 7. Thereby, collar 23 acts upon tappet 17 and lifts valve closure 
member 14 from its valve seat 16. As a result, additional pressure fluid 
is supplied to outlet chamber 11 until the same pressure prevails there as 
in outlet chamber 11'. This compensation procedure will be repeated, if 
necessary, until stepped piston 7', too, has overcome the closure travel 
of valve 14',16'. 
A further pressure increase on the inlet side results in a reduced pressure 
increase on the outlet side, with the pressures in the outlet chambers 
11,11' being always of equal amount due to the compensating effect of 
piston 18. The pressure level, at which the reducing effect of the control 
valves occurs, is dependent on the magnitude of control force F which may 
be fixedly set or may be variable. 
When the pressure is decreased on the inlet side, valve closure members 
14,14' will be lifted from valve seats 16,16' on account of the pressure 
still prevailing on the outlet side causing the pressure being decreased 
there, too. Control force F causes stepped pistons 7 and 7' to move back 
to their inactive positions, with stepped piston 7' moving in abutment 
with the end surface of housing 1 and the clearance a being maintained 
between the housing shoulder 24 and collar 23. 
If one of the brake circuits fails due to a defect, stepped piston 7,7' of 
the still intact brake circuit will have to overcome the entire control 
force F prior to the reducing effect of the valve taking place. The 
changeover pressure of the valve will be increased to double the value. 
For providing the volume input of the still intact brake circuit to be 
increased only slightly upon failure of a circuit, the displacement travel 
of piston 18 is bounded by shoulder 24 of housing 1 and stop ring 25. 
The embodiment shown in FIG. 3 distinguishes from the embodiment of FIG. 1 
merely in that piston 7 includes a step 28 close to its end portion 
projecting from housing 1 and a spring 29 having one end bearing against 
step 28 and its other end acting on lever 27. An adjusting screw, as is 
illustrated in FIGS. 1 and 2, can also be included in this arrangement. In 
principle, spring 29 could be inserted in a different place. However, in 
the illustrated device it affords the advantage of having no detrimental 
effects (for example, in the form of a pressure difference in the outlet 
chambers 11 and 11'). 
The braking pressure control unit in accordance with FIG. 3 corresponds in 
its mode of operation with the preceding description referring to FIGS. 1 
and 2. However, when the device is depressurized, spring 29 always causes 
stepped piston 7 to move to its end position closed to outlet chamber 11. 
In the embodiment of FIG. 4, two parallel stepped bores 52 and 52' are 
provided in a housing 51. Located in bore 52 is a control valve 53 and a 
compensating piston 54 and located in bore 52' is a control valve 53'. 
Control valve 53 includes a sleeve 55 forming a valve seat 56 at one end 
surface thereof. Sleeve 55 is sealed at its outside by means of a seal 57 
which is held between a guide ring 58 and a prop ring 59. A control piston 
60 includes a valve plate 61, a shank portion 62 of smaller diameter 
penetrating sleeve 55 and a shaft portion 63 of greater diameter leading 
outwardly and surrounded by a seal 64. Seal 64 is arranged in an inset 65 
which is held between prop ring 59 and a circlip 66. Valve plate 61 
includes an elastic closure element 67 which is held by a sheet metal top 
68 forming a guide flange at the same time. Control valve 53 separates an 
inlet chamber 69 communicating with a port Hz of a tandem master cylinder 
from an outlet chamber 70 connected to at least one wheel cylinder Rz. A 
step 71 between shank portions 62 and 63 cooperates with the end surface 
72 of sleeve 55 in order to entrain sleeve 55 to the left-hand position. 
Notches 73 permit fluid flow in this position. 
The compensating piston 54 is provided with two seals 74 and 75. A bridge 
76 connects compensating piston 54 to sleeve 55. Bridge 76 is permeable to 
liquid. For example, bridge 76 may be welded annularly with the end 
surface of compensating piston 54 and may engage with resilient ribs in an 
annular groove of sleeve 55. Compensating piston 54 is located between 
outlet chamber 70 and a pressure chamber 77, the latter being connected 
via a bore 78 with outlet chamber 70' of control valve 53'. The movement 
of the compensating piston is limited to the right by a circlip 79 and to 
the left by an end surface 80 of housing 51. 
Control valve 53' is of the same construction as control valve 53. 
Therefore, like parts are assigned like reference numerals, however, 
marked with an apostrophe. In this arrangement, the inlet chamber 69' is 
connected with a second port Hz' of a tandem master cylinder. The outlet 
chamber 70' leads to at least one other wheel cylinder Rz'. In addition, 
provided in the guide ring 58' is a stop 81 which cooperates with a step 
82 at sleeve 55'. Acting on the end surfaces of both control pistons 60 
and 60' leading out of housing 51 is a lever 83 which is adapted to swivel 
around pivots 84 and, located outside the drawing plane, to be loaded by a 
control force F. Force F may be constant or responsive to the vehicle 
load. 
In normal operation, the arrangement illustrated in FIG. 4 operates as 
follows: 
With pressure fluid subjected to increasing pressure being fed to the two 
inlet chambers 69,69', the pressure in the outlet chambers 70,70' rises in 
the same manner. When the inlet pressure in each brake circuit multiplied 
by the surface Q2 exceeds half the control force F, both control valves 
53,53' will close. Thereupon, a rise of the outlet pressure will occur, 
said rise being reduced with reference to the rise of the inlet pressure 
according to the relation 
##EQU1## 
In the event of control valve 53 closing prior to control valve 53', a 
higher pressure will develop in outlet chamber 70' and thus in pressure 
chamber 77 causing compensating piston 54 with sleeve 55 to be displaced 
to the right in the drawing so that control valve 53 opens anew until a 
pressure balance prevails in both outlet chambers 70,70'. In the event of 
control valve 53' closing first, compensating piston 54 and sleeve 55 will 
be moved to the left in the drawing so that control valve 53 closes as 
well. If the inlet pressure is decreased, control pistons 60,60' will 
return to the drawn initial position under the influence of the control 
force F. In doing so, pistons 60,60' entrain sleeves 55,55' via the step 
71,71'. The backstroke is terminated when step 82 of sleeve 55' abuts stop 
81. The inactivated position of the system is defined this way. 
When one brake circuit fails, the associated control piston remains in its 
rest position. The other control piston will then be loaded with the full 
control force F resulting in the change-over pressure of the valve being 
increased to double the value, which is desired in such cases. 
In FIG. 5, two like control valves are accommodated in a housing 101, one 
of which will be described. The other control valve is assigned like 
reference numerals, however, marked with an apostrophe. 
A stepped bore 102 is provided in its enlarged bore portion with a sleeve 
103 which is sealed relative to housing 101 by means of a sealing ring 
104. Sleeve 103 bears against a first shoulder 105 of housing 101 and is 
located against axial displacement by means of a ring 106 which is 
fastened in housing 101. A stepped piston 107 is guided with its larger 
diameter portion in bore 102 and with its smaller diameter portion in 
sleeve 103 and sealed by means of seals 108 and 109. 
An inlet chamber 110 is defined by an annular surface between the larger 
and the smaller diameter portions of stepped piston 107 and an outlet 
chamber 111 is defined by the end surface of the larger diameter portion 
of stepped piston 107. Inlet chamber 110 and outlet chamber 111 
communicate with each other via a radial bore 112 and a coaxial pressure 
fluid passageway 113 on the inside of stepped piston 107. Situated in 
pressure fluid passageway 113 is a valve closure member 114, which is 
loaded by a spring 115 against a valve seat 116. Valve closure member 114 
includes a tappet 117 extending from stepped piston 107. 
A compensating piston 118 which is sealed relative to housing 101 by means 
of two seals 119 and 120 is arranged in a transverse bore 121, bore 121 
being closed by a plug 122 which is sealed relative to housing 101 by 
means of a seal 123 and which is fastened by means of a circlip 124. 
Compensating piston 118 is integrally formed with a distribution element 
which includes in this case two cones 125 and 126 whose inclined surfaces 
127 and 128 are oppositely sloped with respect to each other. Seated on 
cones 125 and 126 are transmission pistons 129 and 130 which are guided in 
bores 131 and 132 by means of ribs. Via the channels remaining between the 
ribs, outlet chamber 111 is connected to the pressure chamber 133 on the 
one side of compensating piston 118, and outlet chamber 111' is connected 
to the pressure chamber 134 on the other side of compensating piston 118. 
Pressure chambers 133 and 134, on their part, communicate with ports 135 
and 136, to which ports lines leading to wheel cylinders Rz and Rz' are 
connectible. The inlet chambers 110,110' are connected with two outlets Hz 
and HZ' of a tandem master cylinder. 
A preloading force F acts on a lever 137 which is pivotable around two 
pivots 138 and 139 which are displaced from the drawing plane. Force F can 
be constant or vary load-responsively. 
The mode of operation of the dual-circuit pressure control valve 
illustrated in FIG. 5 will first be described assuming that both brake 
circuits are operable. The two stepped pistons 107,107' bear with their 
end surface against the left-hand step of bores 102,102' in the rest 
position such that valves 114, 116 and 114', 116' are open. When the 
pressure increases at the inlet Hz,Hz', the pressure at the outlet Rz,Rz' 
will follow in the same way, since the two control valves 114,116 and 
114',116' are open. With increasing pressure, the stepped pistons 107,107' 
will move to the right in the drawing, until the aforementioned valves 
finally close. In case one control valve 114,116 closes earlier than the 
other control valve 114',116', the pressure in the pressure chamber 134 
will rise higher than the one in the pressure chamber 133, and the 
compensating piston 118 moves downwards. As a result, transmission piston 
130 and, thus, tappet 117 will be shifted to the right in the drawing by 
means of inclined surface 128 so that control valve 114,116 opens again. 
At the same time, valve closure member 114' and transmission piston 131 
will be displaced to the left in the drawing by locking spring 115' 
because inclined surface 127 offers the respective space therefor. Control 
valve 114',116' will be, therefore, moved in the closing direction. Due to 
this oppositely directed movement, both control valves will be closed 
approximately simultaneously and, for this reason, at the same pressure. 
With the inlet pressure continuing to rise, the outlet pressure follows 
along a characteristic curve with a reduced slope. If the control valve 
114',116' had closed first, compensating piston 118 would have been moved 
upwards so that control valve 114',116' would have opened again and the 
other control valve 114,116 would have been moved in the closing 
direction. 
When one brake circuit fails, compensating piston 118 will move to the stop 
provided by plug 122 or the stop provided by the oppositely disposed end 
surface of transverse bore 121, and the control valve of the operable 
brake circuit will function as usual. The valve will, however, close after 
a shorter travel of piston 107,107' and against an increased force because 
the preloading force F is no longer distributed onto the two pistons. In 
total, the changeover pressure of the intact valve will, therefore, be 
increased when a circuit fails. 
It is, however, frequently desired in the event of failure of the one brake 
circuit to keep the outlet pressure of the other brake circuit always at 
the same level as the inlet pressure. This may be achieved by the 
embodiment according to FIG. 6. 
In FIG. 6, like parts are assigned like reference numerals as in FIG. 5, 
increased by 100. In this structure, adjoining the two cones 225,226 with 
their inclined surfaces 227,228 are auxiliary cones 240,241 with 
oppositely sloped inclined surfaces 242,243, respectively. When the 
compensating piston 218 is urged into the one end position upon failure of 
the one brake circuit, these inclined surfaces cause the transmission 
piston 230,231 to be pressed completely to the right in the drawing of 
FIG. 5, and, by this means, to constantly keep open the pertinent control 
valve 114,116 or 114',116'. To the end that this is accomplished for sure, 
the maximum displacement travel of piston 107--as shown in FIG. 5--which 
always has to be greater than the closure travel of the valve 114,116, is 
smaller than the sum of the displacement travel of the transmission piston 
130 (in the case of piston 107', transmission piston 129 is meant) upon 
failure of a brake circuit plus the valve closure travel. Therefore, the 
wheel cylinders of the brake circuit in operation are supplied with the 
full inlet pressure over the entire range of pressure. 
In addition, compensating piston 218 has a groove 244 between inclined 
surfaces 227 and 228, in which a tripping pin 245 of a differential 
pressure indicating device 246 engages. This groove 244 is of such an 
axial length that the compensating function of compensating piston 218 is 
possible in the range of the usual pressure differentials of the inlet 
pressures without actuating the differential pressure indicating device 
246. After a predetermined travel, compensating piston 218 with its end 
surface 247 will be in operative connection via a spring plate 248 with a 
spring 249, or with a circlip 250 via a spring plate 251 with spring 249. 
As soon as the differential of pressure between the two brake systems 
acting on compensating piston 218 exceeds the preload of spring 249, 
tripping pin 245 is moved by groove 244 outwardly so that a switching 
operation occurs which provides an indication. 
The compensating piston is likewise able to act on a distribution device 
separated therefrom. The distribution device may, for instance, be 
composed of a lever which is tiltable around a point of rotation arranged 
between the two stepped pistons' axes. The inventive principle can be used 
in connection with control valves of different design as well. 
While we have described above the principles of our invention in connection 
with specific apparatus, it is to be clearly understood that this 
description is made only by way of example and not as a limitation to the 
scope of our invention as set forth in the objects thereof and in the 
accompanying claims.