Hydraulic brake system with slip control

The present invention relates to a hydraulic brake system with slip control, including a braking pressure generator which is connected hydraulically to at least one wheel brake by way of a main pressure line, a return line connected to the wheel brake and in communication with a pressure fluid collecting device, an auxiliary-pressure pump having an auxiliary-pressure line which is connected to the braking pressure generator, and at least one pressure modulation valve acting in the main pressure line and in the return line to close or keep open the pressure fluid passage in the main pressure line and the return line. An orifice is arranged in the main pressure line between the pressure modulation valve and the wheel brake, and an unimpeded hydraulic passage between the main pressure line and the wheel brake is provided in a first switch position, and the orifice limits the pressure fluid passage in the main pressure line leading to the wheel brake in another switch position.

BACKGROUND OF THE INVENTION 
The present invention relates to a hydraulic brake system with slip 
control. 
The discontinuous pressure fluid control in slip-controlled brake systems 
by way of digitally operated inlet and outlet valves produces an 
undesirable emission of sound due to the pulse-like pressure variation. 
Therefore, various arrangements to reduce noises during anti-lock/traction 
slip control operations have been disclosed. In this respect, reference is 
made to international patent application WO-A-90/12713, as an example. 
This application discloses the arrangement of pressure pulsation dampers 
inside the main pressure lines, and for that matter, in the vicinity of 
the braking pressure generator, or in the by-pass line of the 
electromagnetic inlet and outlet valves. The pressure pulsation dampers 
used are configured as vibration-damping elastomeric energy-accumulating 
elements which, by way of their defined expansion in volume, flow length 
and restricting properties, shall prevent the transmission of sounds 
produced by valve switching frequencies to the mass-loaded and, thus, 
vibrating brake system. 
A relatively soft pedal feel and an increased pedal travel are undesirable 
consequences of the additional pressure fluid volume input caused by the 
given pressure pulsation dampers during braking operations. 
European patent application No. 0 317 305 discloses a solenoid valve which 
is appropriate for use in anti-lock hydraulic brake systems. The solenoid 
valve includes a magnetic core accommodating a coil which is confined by a 
magnetic armature on one side and by a diaphragm member on the other side. 
The diaphragm member has a passage closable by a valve needle. The 
diaphragm member is axially movably arranged between the magnetic core and 
a housing cover containing the pressure fluid inlet. Thus, in the switch 
position of the valve needle closing the supply channel in the diaphragm 
member, a pressure differential is operative on both sides of the 
diaphragm member and causes a displacement of the valve needle relative to 
the magnetic armature. This fact preloads a compression spring compressed 
between the valve needle and the magnetic armature. The preloading force 
of the compression spring, caused by the difference in pressure on the 
diaphragm member, causes rapid release of the magnetic armature from the 
magnetic core when the electromagnetic energization is interrupted. The 
result is short valve opening times. A spring restoring force acting on 
the valve needle in the opening sense assists the rapid opening of the 
supply channel in the diaphragm member so that pressure fluid, after 
having passed the open passage across the valve needle, propagates to an 
annular slot which is formed between the hollow cylindrical inside wall of 
the magnetic core and the outside wall of a cylindrical part which guides 
the valve needle. The fluid emanating from the valve inlet is conducted to 
a pressure fluid connection leading to the pressure fluid consumer by way 
of the supply channel in the diaphragm member and the subsequent annular 
slot. 
UK patent application No. 2 252 140 discloses a valve assembly for 
anti-lock control including an inlet valve and an outlet valve. The inlet 
valve accommodates an annular piston adapted to be actuated in response to 
hydraulic pressure differences. The piston is configured as a stepped 
piston which is pressure-balanced, subjected to the hydraulic pressure of 
the master cylinder. A compression spring interposed between the valve 
carrier of the electromagnetically operable inlet valve and the annular 
piston keeps the annular piston in an unrestricted open position, with the 
result that pressure fluid is unimpeded to flow from the master cylinder 
to the wheel brake. When the inlet valve is energized electromagnetically, 
the valve closure member of the inlet valve adopts its closed position, 
and the electromagnetically energized outlet valve discharges the wheel 
braking pressure in the direction of a supply reservoir. This reduces the 
level of wheel pressure which is conducted from a branch line of the 
outlet valve to a piston step on the annular piston. Thus, the annular 
piston is no more pressure-balanced and, with its sealing seat, closes the 
originally unimpeded pressure fluid passage in the inlet valve from the 
master cylinder to the wheel brake in opposition to the effect of the 
compression spring. A new pressure increase in the wheel brakes occurs by 
pulsed operation of the magnetic coils of the inlet valve so that pressure 
fluid propagates in a restricted fashion to the wheel brake by way of the 
orifice bore in the annular piston. To this end, the outlet valve is in an 
operating position in which the connection to the supply reservoir is 
closed. Upon deactivation of the master cylinder, the annular piston 
adopts again a position in which the pressure fluid connection to the 
wheel brake is unrestrictedly open because the compression spring 
positions the annular piston on its stop. 
The chosen channel arrangement and pressure fluid ducts to provide a 
connection between the master cylinder, the inlet and outlet valve and the 
wheel brake necessitates a structural series arrangement of the individual 
elements in the inlet valve. Also, the selected switching function of the 
annular piston in response to the amount of pressure conducted from the 
wheel brake by way of the outlet valve necessitates the chosen 
arrangement. 
SUMMARY OF THE INVENTION 
Therefore, an object of the present invention is to maintain a simple basic 
structure of the customary brake systems with slightest possible 
modifications and to provide a solution to effectively minimize noises, 
which are responsive to the valve operating positions, and the propagation 
of noise. 
Thus, the present invention is based on the idea to reduce the pressure 
pulses of different intensities, which are caused by the switching 
frequencies of the pressure modulation valves, in response to the 
difference in pressure between the wheel brake and the braking pressure 
generator, by way of a switchable orifice or throttle member to the end of 
varying the opening cross-section between the inlet valve, and, if 
necessary, also the outlet valve, and the wheel brake. 
Further features, advantages and possible applications of the present 
invention can be seen in the following description of several embodiments.

DETAILED DESCRIPTION OF THE INVENTION 
FIG. 1 shows schematically a hydraulic diagram of the brake system of the 
present invention. A braking pressure generator 5 is connected to a wheel 
brake 3 by way of a main pressure line 2. An electromagnetic, normally 
open pressure modulation valve 1 acting as an inlet valve and an orifice 4 
which is connected downstream of the pressure modulation valve 1 and is 
inoperative in its basic position are included in the main pressure line 
2. The switchable orifice 4 is represented as a two-way/two-position 
directional control valve. A first control pressure port 6 of the valve is 
acted upon by the pressure of the braking pressure generator 5, and an 
oppositely acting second control pressure port 6' of the 
two-way/two-position directional control valve is acted upon by the 
pressure in the wheel brake. The control pressure port 6' can also be 
connected to the main pressure line 2 between the pressure modulation 
valve 1 and the orifice 4 (variant A) as an alternative of the connection 
(as shown) of the second control pressure port 6' downstream of, i.e. 
after, the switchable orifice 4 (variant B), when viewed in the drawing. 
Advantageously, an invariable orifice 11 is arranged downstream of, i.e. 
after, the control pressure port 6' to prevent an excessively great 
difference in pressure upstream and downstream of the orifice 4 from 
becoming effective in the slip-free normal braking mode upon quick 
application of the brake. Undesirable throttling effects are thereby 
avoided. In case of need, an invariable orifice 25 can also be inserted 
into the main pressure line 2 upstream of, i.e. before, the inlet valve. 
Exactly as the above-mentioned invariable orifice 11, this last mentioned 
invariable orifice is calibrated related to the vehicle and, typically, 
can be sized somewhat larger in its cross-section than the invariable 
orifice 11. This prevents a premature undesirable switch-over and, thus, 
coming into effect of the switchable orifice 4 in all possible operating 
conditions. A compression spring 7, represented on the 
two-way/two-position directional control valve, ensures an initially 
unthrottled pressure fluid passage to the wheel brake 3 in the brake 
release position and the normal position. Further, a return line 16 is 
connected to the wheel brake 3 to extend to the suction side of an 
auxiliary-pressure pump 18 by way of an electromagnetically operated 
pressure modulation valve 1 which is closed in its basic position and acts 
as an outlet valve. To reduce the noise of the electromagnetically 
operable outlet valve, a switchable orifice 4 having corresponding control 
lines acted upon by the pressure in front of and behind the outlet valve 
may be arranged, exactly as has been described in detail with respect to 
the inlet valve. Because by far smaller pressure variations in front of 
and behind the outlet valve will occur in the pressure control mode, 
however, the noise damping measure, proposed with respect to the inlet 
valve, regarding an arrangement of an orifice connected downstream of the 
wheel brake 3 on the outlet side is not absolutely necessary. 
Further, the return line 16 includes a low-pressure accumulator 19 for the 
intermediate storage of excessive pressure fluid. The pressure side of the 
auxiliary-pressure pump 18 is connected to the braking pressure generator 
5 or, in turn, the main pressure line 2, by way of an auxiliary-pressure 
line 20. The valve switching position according to the present invention 
renders it possible to reduce the valve switching noises otherwise caused 
in the pressure modulation valve 1 (inlet valve) as soon as the orifice 4 
in the main pressure line 2 becomes operative during the switching 
operation of the inlet valve and in response to a change-over point 
defined by construction. It is provided, by way of defined design criteria 
of the brake system, to activate the orifice function, for example, with 
an increasing pressure difference on the pressure modulation valve 1 of 
more than 15 bar. 
Based on the general hydraulic circuit design of the brake system of the 
present invention, suitable valve constructions will be described in the 
following by way of FIGS. 2 to 5. 
FIG. 2 shows, in a cross-section taken through a valve accommodating member 
9, two structurally parallel arranged pressure modulation valves 1. The 
pressure modulation valve 1 shown in the left of the drawing functions as 
an inlet valve, and the pressure modulation valve 1 shown in the right of 
the drawing functions as a normally closed outlet valve. The two pressure 
modulation valves 1 are hydraulically interconnected by way of a pressure 
fluid channel, and a blind-end bore is arranged transversely to the 
pressure fluid channel to accommodate a slide piston 8. Preferably on its 
peripheral surface, the slide piston 8 has a circumferential throttling 
notch which takes the function of the switchable orifice 4. In the 
drawing, the slide piston 8 abuts with its first frontal end on a stop 10 
inserted in the blind-end bore. Stop 10 accommodates a compression spring 
7 which is supported on the frontal end of the slide piston 8. Opposite to 
the first frontal end, the further frontal end of the slide piston 8 is 
alternatively acted upon by the pressure of the main pressure line 2, as a 
control pressure, and, if necessary, by the pump pressure which propagates 
also in the direction of the main pressure line 2 through the pump 
pressure line 12. Thus, the channel portion arranged between the 
invariable orifice 25 and the slide piston 8 functions as the control 
pressure port 6. The invariable orifice 25 calibrates the pressure fluid 
flow rate to the slide piston 8 so that an undesirable premature switching 
of the orifice 4 due to a quick brake application is prevented. The valve 
closure member 17 of the inlet valve, which is open in the initial 
position, permits an unimpeded pressure fluid supply in the direction of 
the wheel brake 3 by way of the normally closed pressure modulation valve 
1 (outlet valve), because the compression spring 7 initially keeps the 
connecting channel between both pressure modulation valves 1 
unrestrictedly open in the basic position of the slide piston 8. The 
position of the slide piston 8, as shown in the drawing, will be achieved 
only by a considerable rise in pressure in the main pressure line 2 or in 
the pump pressure line 12. When the slide piston 8 is immersed in the 
blind-end bore, hydraulic attenuation of the slide piston 8 is caused. 
Normally, the difference in pressure on the two frontal ends of the slide 
piston 8 is at least 15 bar in the orifice switching position. In addition 
to the constructive design of the present invention as shown, a non-return 
valve opening in the direction of the main pressure line 2 may be inserted 
upstream of the invariable orifice 25, with the result that pressurization 
for control of the slide piston 8 on its bottom frontal end is exclusively 
caused by the pressure in the pump pressure line 12. 
FIG. 3 shows another embodiment of a switchable orifice in the form of a 
member which acts as a throttle needle 13 that is also immersed into a 
slide piston 8. Slide piston 8 is modified compared to FIG. 2. Regarding 
the pressure fluid ports, the design corresponds largely to the basic 
design of the valve accommodating member 9 known from FIG. 2. Thus, the 
main pressure line 2 and the pump pressure line 12 in FIG. 3 are also 
connected downstream of the slide piston 8 in order to effect a 
travel-responsive continuous restriction of the pressure fluid, which 
flows from the pressure modulation valve 1 (inlet valve) to the slide 
piston 8, the restriction being responsive to the pressurized frontal end 
remote from the stop 10. Thus, responsive to the stroke of the slide 
piston 8, a defined quantity of pressure fluid is conducted to the second 
pressure modulation valve 1 (outlet valve) by way of transverse and 
longitudinal bores in the slide piston 8. In the brake release and 
slip-free normal braking position, the compression spring 7 arranged above 
the slide piston 8 keeps the slide piston 8 in its bottom position so that 
the throttle needle 13 attached to the stop 10 is immersed only slightly 
into the slide piston 8 and is passed by fluid in an initially 
unrestricted manner by way of the piston's transverse and longitudinal 
bores. As mentioned in FIG. 2, the throttling position of the slide piston 
8 is caused either by a defined increase in pressure in the main pressure 
line 2 or by the pressure in the pump pressure line 12. Thus, when the 
pump pressure is utilized, exclusively the magnitude of the pump pressure 
and the spring force of the compression spring 7 would be most important 
for the respective position of the slide piston 8. 
FIG. 4 shows a particularly compact design of the circuit known from FIG. 1 
in a partial cross-section taken through a pressure modulation valve 1 
(inlet valve). With the valve closure member 17 of the pressure modulation 
valve 1 being switched to provide an unimpeded fluid passage in the basic 
position, pressure fluid can propagate from the main pressure line 2 in 
the direction of the wheel brake 3 by way of the orifice 4 in the valve 
carrier 14 and a closable channel 22. The special features of the present 
invention involve that the channel 22 is arranged in different planes of 
the valve carrier 14 so that, preferably, the first portion of the channel 
22 provides a connection to a hollow chamber 26 having an annular sleeve 
23 by way of the open valve closure member 17. The second portion of 
channel 22 connects the hollow chamber 26 to the main pressure line 2 
which leads to the wheel brake 3. The two portions of the channel 22 are 
offset in relation to each other, for example, by roughly 90 degrees, so 
that in the presence of only a slight difference in pressure upstream and 
downstream of the annular sleeve 23, a compression spring 7, in the 
capacity of a cup spring, initially keeps the hollow chamber 26 open (that 
interconnects the two portions of the channel 22), until the hydraulic 
pressure below the annular sleeve 23, which corresponds either to the 
pressure in the main pressure line 2 or the pressure in the pump pressure 
line 12, closes the two portions of the channels 22 in opposition of the 
compression spring 7. Thus, there is exclusively a pressure fluid 
connection between the pressure modulation valve 1 (inlet valve) and the 
wheel brake 3 by way of the orifice 4. 
Another design variant for the integration of the switchable orifice 4 in 
the pressure modulation valve 1 can be seen in FIG. 5. The orifice 4 is a 
component of an axially movable valve seat member 15 which is slidable in 
opposition to the effect of a compression spring 8 in the valve carrier 14 
of the pressure modulation valve 1. The valve seat member 15 has two 
sealing seats, a first sealing seat being provided by the interaction of 
the valve closure member 17 with a through-bore in the valve seat member 
15. In the direction of the valve closure member 17, the through-bore has 
a diameter which corresponds to the diameter of the sealing seat 21 remote 
from the valve closure member 17 so that the axially movable valve seat 
member 15 is hydraulically pressure-balanced in its basic position. The 
sealing seat 21 is arranged on the mouth of the main pressure line 2 or 
the pump pressure line 12 into the valve carrier 14 so that, with slight 
differences in pressure in front of and behind the valve seat member 15, 
the sealing seat 21 closes the hydraulic connection between the main 
pressure line 2 or the pump pressure line 12 and the hollow chamber 26 
arranged below the valve seat member 15. 
Thus, in the slip-free normal braking position, the parts described adopt 
their position as shown in the drawings, so that pressure fluid coming 
from the main pressure line 2 is not impeded to flow in the direction of 
the wheel brake 3 by way of the through-bore of the valve closure member 
17 that is in its open position. In the pressure maintain-constant phase, 
the valve closure member 17 closes the through-opening in the valve seat 
member 15 so that the valve seat member 15 remains in its position, as 
shown in the drawing In the pressure increase phase of a braking pressure 
control operation, the valve closure member 17, in turn, is in its open 
position. Starting from a defined difference in pressure on both frontal 
ends of the valve seat member 15, the valve seat member is displaced until 
it abuts on the valve closure member 17, and the sealing seat 21 
simultaneously opens the passage from the main pressure line 2 to the 
orifice 4 exclusively. Subsequently, for braking pressure reduction, 
release of the brake pedal causes an unimpeded supply of pressure fluid 
from the wheel brake 3 in the direction of the main pressure line 2 by way 
of the annular sleeve arranged on the valve seat member 15. Due to the 
front end of the valve seat member 15 remote from the valve closure member 
17 being relieved from pressure, the valve closure member with its sealing 
seat 21 abuts again on the valve carrier 14 under the effect of the 
compression spring 7.