DE1243526B - Brake pressure control system for the brakes of a vehicle wheel, in particular a chassis wheel of an aircraft - Google Patents

Description

Brake pressure control system for the brakes of a vehicle wheel, in particular of a landing gear wheel of an aircraft The invention relates to a brake pressure control system for the brakes of a vehicle wheel, in particular a chassis wheel of an aircraft, which can be actuated as a function of a variable fluid pressure, with a Breins valve operated by the pilot and a slip signal actuated control valve and with a pressure control valve device, the inlet side Via the valve operated by the pilot with a pressure fluid source and on the output side is connected to the brake and has a, relief line and a sub Includes spring-loaded valve and a valve actuator that controls of the output pressure in accordance with an on the valve actuator acting manipulated variable that. via a throttle device fed from the inlet side is produced.

The control system according to the invention is particularly advantageous in brake systems for intercepting and compensating for slipping processes in which a brake is partially relieved or completely released depending on the beginning of a sliding process of a braked wheel. The invention provides particular advantages in connection with brake control systems for the chassis wheels of an aircraft. For this reason, the invention is described in connection with the braking of the landing wheels of an aircraft.

There are known Breinsregelsysteine for intercepting and regulating slipping processes, in which the wheel brakes are automatically triggered depending on an abnormal braking or deceleration of the wheel rotation. Such a delay in the speed of rotation of the wheel indicates an incipient slip: process. Furthermore, systems are known which, initially as a function of a first signal indicating a slip process, only relieve the brake, while the brake is only fully released when this initial relief of the brakes is not sufficient to prevent the incipient slip process Bring control. A major difficulty with such previously proposed systems is that after the wheel has recovered from its slipping state, the brake pressure is usually immediately restored to its original value, with the risk that a new slip process will be initiated immediately.

It is known to have a tensionable elastic in brake control systems Provide component having hydraulic control device, by the one of Pressure fluctuations in the brake valve and in the control valve controllable control force to the Valve actuator can be applied. It has also become known, throttle devices provide that directly and solely the rate of change of the brake pressure limit.

In a known brake control valve has two separate valve arrangements are required to regulate the braking pressure. These valves are connected in parallel between a clamp that is under braking pressure and the return line leading to the pump. When each valve opens, fluid flows back to the pump, reducing the brake pressure. The valves mentioned are operated completely independently of one another and have different functions. The actual brake pressure depends on which valve opens when the brake pressure is lower.

A valve is direct and immediate from that used by the pilot pressure exerted by a hand-operated valve. The second valve on the other hand has the tasks of slip and speed control. On the other hand, at the invention of the control pressure from the command pressure measured by the pilot can be derived, which has the advantage that all with the pressure design related functions can be combined in a single valve.

The object on which the invention is based is initially to modulate the control pressure applied by the driver of a vehicle to the brake pressure regulator, independently of the level of the brake pressure, in such a way that the braking effect only increases in accordance with the control pressure that can be changed at your predetermined limited speed. This reduces the risk that the braking pressure will be increased rapidly beyond a value at which an optimal braking effect would otherwise result, whereby a slip process could then be initiated. This modulation effect should be achieved without causing a noticeable delay in the supply of a sufficient amount of brake fluid to the brake to initiate the braking process.

It is also possible to use the brakes after a recovery to gradually reassemble from a slippery state. Such a gradual reinvestment the pressure gives the wheel and the landing gear time to return to normal Return to operating behavior, while at the same time counteracting a slip process Regelsystein can return to its state of equilibrium. During one of these Gradual reapplication of the brakes is a repetition of the previous one Slipping unlikely, so that the average braking efficiency is significant is increased. If the driver or pilot continues to apply excessive pressure applies the brakes, which increases the risk of periodic wheel slipping results, then by gradually reapplying the brakes according to the present Invention, the frequency of such slipping processes greatly reduced. This will also the risk of a resonance deformation of the landing gear of the aircraft to a large extent eliminated.

According to the invention, the aforementioned object is achieved by that the pressure control valve in no position connects the input line with the relief line connects and has a valve actuator which can be acted upon on the white side, wherein the throttling point is placed in the connecting line between the regulating connectors and only one of these is directly connected to the inlet line of the valve.

Further features within the scope of the invention are set out in the subclaims marked.

The invention and its advantages will be better understood from the following description of exemplary embodiments in conjunction with the drawings. In the drawings, F i g. 1 shows a schematic representation of a brake control system actuated with a pressure medium using the invention, FIG . 2 shows an axial sectional view through a pressure modulation device, FIG. 3 shows an axial sectional view through a brake control device which also includes a pressure modulation device according to FIG . 2 in combination with a valve construction for regulating a slip process, F i g. 4 shows an axial, partial sectional view through a modified embodiment, FIG. 5 shows an axial partial sectional view through a further modified embodiment, and FIG. 6 is an axial partial sectional view through another modified embodiment.

F i g. 1 shows in schematic form a brake control system. A braked vehicle wheel 20 is attached to the vehicle frame 22 with the aid of a support 24 which, for example, can represent the landing gear of an aircraft. A brake for the wheel 20 is shown schematically at 26 and is connected to the wheel 20 by a mechanical operative connection 27. The brake 26 is regulated and actuated by a liquid pressure medium. Air pressure turbine systems will be described later. The brake itself may be of known type and normally applies a brake torque to the wheel via brake line 30 corresponding to the fluid pressure.

The pressure fluid for actuating the brake 26 is supplied to the brake from a pressure source of a suitable type, which is shown schematically as a hydraulic pump 32 #. The pump 32 supplies the pressure fluid via a pressure line 36 and receives the pressure fluid coming at low pressure via the return line 34. The pressure from the line 36 is fed to the brake through a suitable valve, which is shown here as a conventional brake valve 40 with a brake lever 42. The brake valve 40 supplies the pressure fluid to a supply line 44 at a pressure which corresponds to the force exerted on the brake lever 42 by the pilot. If the force on the lever 42 exerted by the pilot removed, then the pressure in the supply line 44 is by switching off of the hydraulic fluid via the line 45 to the low-pressure circuit reduces return line 34th

A measuring device 50 responds to an incipient slip process of the wheel 20. It is driven in accordance with the speed of rotation of the wheel via a mechanical operative connection, which is indicated by the dashed line 51 . The measuring device 50 responds to a delay in the rotational movement of the wheel and supplies one or more electrical signals via lines 52 when the delay in the rotational movement of the wheel assumes an extraordinarily high value. The signals arriving via lines 52 - , indicating slipping, are fed to a slip control valve 54, which is electrically operated. The valve 54 is located between the inflow line 44 and the brake line 30 and normally allows a substantially unimpeded flow between the two lines. Depending on a slip signal on the lines 52 , the actuating device 55, which may contain an electromagnet which is mechanically connected to the valve 54, causes the valve to shut off the brake line 30 from the inflow line 44 and thereby prevent a further increase in the brake pressure. The valve 54 can also bring about a pressure relief of the brake line 30 via the line 56 after the low-pressure return line 34, whereby the brake is then triggered.

The valve device 54 relieves the brake pressure in response to a first, a loading start of forming slip operation indicative signal only partially and only then releases the brake completely, if this slip signal continues, when it is more or when it is delivered by the measuring means 50, an additional signal .

A pressure modulating valve is switched on in the pressure line between the slip control valve 54 and the brake 26 . Such a pressure modulating valve is shown in FIG . 1 shown at 60 . The pressure is fed to this valve from the outlet side of the valve 54 via the input line 62 , and the valve 60 in turn supplies a pressure fluid with modulated or changed pressure via the brake line 30 to the brake. Furthermore, a connection from the modulating valve 60 via the return line 64 to the low-pressure return line 34 is provided. While valve 60 is shown and described as a separate unit from slip control valve 54, the two valves can be constructed as a single valve device.

One embodiment of a pressure modulating valve 60 is shown in FIG. 2 shown schematically in axial section. The valve consists of a valve housing 61, the longitudinal axis of which is denoted by 63. In the construction shown here, the axis 63 is perpendicular and is also mentioned in the description for the sake of simplicity, although this alignment is not necessary. Housing 61 includes upper and lower main sections 66 and 67, respectively, between which suitable sealing means 68 are inserted. The upper section 66 is closed by a threaded cap 69 and the lower section 67 is closed by a pipe fitting 65 to which the brake line 30 is connected. The housing 61 encloses a series of coaxially arranged chambers. The lowermost chamber 70 is referred to as the outlet chamber and is directly connected to the brake line 30 . Immediately above the outlet chamber 70 is the inlet chamber 72, which is connected to the inlet line 62 via a transverse bore 74. A relief chamber 76 is arranged above the inlet chamber 72 and is connected to the relief line 64 via an inclined transverse bore 78. An axial passage between the relief chamber 76 and the inlet chamber 72 snugly fits the cylindrical. Part 82 of a valve body 80 which can be reciprocated in the axial direction and for which this passage serves as a guide. The valve body 80 extends down through the inlet ducts 72 and from there through an axial bore into the outlet chamber 70. A radial flange 71 in this bore serves as a valve seat for the lower end of the valve body 80 , which widens downwards, and thus forms the pressure control valve 73. A coil spring 83 urges the valve body 80 slightly upwards. This movement closes the pressure control valve 73 and thus blocks the connection between the inlet chamber 72 and the outlet chamber 70.

Suitable sealing means 84 prevent communication between the inlet chamber 72 and the relief ducts 76 along the outer surface of the valve body 80 . An axially through bore 85 in the valve body 80 establishes a connection between the outlet chamber 70 and the relief chamber 76 as soon as the upper end of the valve body is not covered. The valve body is normally covered at its upper end and its axial position is determined in part by the construction described below.

The relatively large, cylindrical spring chamber 90 and the cylinder-shaped regulating chambers 92, 93 are located above the chambers 76 within the housing 61. The chambers 90 and 92 are separated from one another by a transverse wall 94 which has an axially extending drilling 96 inside -the regulating piston 98 is inserted in a sliding manner. Suitable sealing means 99 are provided between the piston 98 and its cylinder wall 96 . The valve actuator 100 is slidably engaged with the cylindrical side wall 102 of the chamber 92 via seal means 1103. The piston 98 and valve actuator 100 are constrained to move axially by any suitable coupling mechanism capable of providing any desired ratio between their two movements. The pistons in their embodiment shown. contain a control component 104 integrally connected therewith and can therefore be regarded as such that they are connected to one another via a coupling mechanism with a drive ratio of 1: 1. The effective work surface. of the valve actuator 100 is much larger than that of the control piston 98, as shown here in the drawing. The ratio of these areas can, for example, be on the order of 10: 1 .

The spring chamber 90 contains a resiliently resilient mechanism which can be tensioned under the controlling effect of the inlet pressure prevailing in the line 62. For example, a compressible helical spring 112 is provided for this purpose, although other elastic bodies, for example a compressible amount of gas enclosed between two axially movable pistons or another mechanical spring arrangement designed as a tension spring or compression spring can be provided. The control piston 98 extends down through the bore 96 into the spring canister 90 , where it rests against the upper surface of the spring plate 110 . The coaxial helical spring 112 is arranged between the spring plate 110 and the component 120 used for actuating the valve and for damping. The component 120 consists of an upper piston-like section 122, which at the lower end of the spring chamber 90 is slidingly engaged with the cylindrical side wall 124 via sealing means 125 , as well as the lower section 126 acting as a valve tappet, which extends through with the interposition of sealing means 129 the axial bore 128 in the housing 61 extends into the relief chamber 76. The lower end surface 127 of the valve stem 126 is in engagement with the upper end of a tubular valve body 80 and normally closes the axial bore of this valve body.

A downward movement of the valve stem 126 opens the pressure regulating valve 73, so that a flow connection between the outlet chamber 70 and the inlet chamber 72 is established. An upward movement of the valve stem 126 allows an upward movement of the valve body 80 under the action of the spring 83 , whereby the pressure control valve 73 is closed and this flow connection is interrupted. With a further upward movement of the valve tappet 126 , its lower surface 127 lifts off the valve body 80 and thus establishes a flow connection between the outlet chamber 70 and the relief chamber 76 via the axial bore 85 of the valve body 80 . These movements of the valve stem 126 are caused by fluid pressures. which act on the various parts, as well as by the spring force exerted on the component 120 by the spring 112. The magnitude of this force changes with the position of the control piston 98 and increases with a downward movement to tension the spring.

In the present embodiment, the pressure corresponds to the Federkammer90 substantially the pressure of the return line 64, normally a connection is made to this line by a housing mounted Durchlaß'132. Pressure equalization on opposite working surfaces of the damping piston 122 is achieved through the bore 134 made in the piston. This bore is chosen so large that rapid and immediate movements of the component 120 are possible, but still sufficiently small that it effectively supports the friction and inertia of the piston in the damping of vibrational movements of the valve tappet 126 , which causes a rattling of the valve body 80 is prevented when both valves are slightly open. The control pressure for actuating the control piston 98 is supplied from the inlet line 74 to the control chamber 92 via a passage 136 . Via this passage the liquid comes under the pressure of the inlet line 62 immediately after the lower control chamber 92 below the valve actuating member 100.

The chamber section 93 above the valve actuating member 100 is closed off with suitable connecting channels which allow a limited flow of liquid between the top and bottom of the piston, i. H. between the upper chamber 93 and the lower chamber 92. Such a channel can be seen in the piston itself, as shown at 140. The throttle point in the form of the bore 140 is adjustable during the assembly of the device in such a way that a predetermined narrowing of the flow cross-section results between the top and bottom of the valve actuating member 100 , as is provided for a specific pressure modulating valve.

Regarding the operation of the device, it should be noted that when the valves 73 and 130 are both fully closed or at least substantially closed, the pressure in the inlet chamber 72 exerts virtually no axial pressure on the valve body 80. The pressure exerted on the valve body 80 by the pressure medium in the outlet chamber 70 is practically zero, so that its position is actually only determined by the spring 83 and the position of the valve tappet 126 . The spring 83 holds the valve body 80 in engagement with the valve stem 126, whereby the valve 130 is closed until an upward movement of the valve body 80 is prevented by the closure of the pressure regulating valve 73 . The force of the spring 80 is normally small compared to the hydrostatic forces to be described, but is sufficient to overcome the weight and friction of the valve body 80 .

The outlet pressure in chamber 70 acts on bottom surface # 127 of valve stem 126 and attempts to raise it. This effect is canceled by the downward force of the spring 112. The actual valve positions are thus given by the equilibrium of the forces acting on the valve tappet 126. With a downward movement of the valve tappet, the valve 130 remains closed and the pressure regulating valve 73 is opened. An upward movement of the valve lifter closes the pressure regulating valve 73 and then, if it is continued, also opens the valve 130. Since the spring 83 does not contribute anything to this last part of the movement, the pressure required on the valve lifter surface 127 is thus to control the valve 130 to open, usually slightly greater than the pressure required to close valve 73.

The downward force exerted on the valve stem 126 by the spring 112 increases the tension of the spring, which in the present case results from a compression of the spring. Since the axial movement of the valve stem, which is necessary to actuate the valves 73 and 130 , is normally very small, the spring force depends essentially entirely on the axial position of the upper end of the spring, which is determined by the action of the control piston 98 and the valve actuator 100 is determined.

Under equilibrium conditions, when the brakes are applied, the compression of the spring 112 corresponds to the downwardly directed fluid pressure acting on the control piston 98. The working area of piston 98 can be viewed as an area of the entire surface area of valve actuator 100 that is the same cross-section as piston 98. Under equilibrium conditions, the pressure applied to this area is equal to the inlet pressure exerted from inlet conduit 62 via passage 136 and the restriction 140 is transmitted. The diameter of the control piston 98 is in practice approximately equal to the effective diameter of the working surface 127 of the valve stem 126 which is exposed to the fluid pressure of the outlet chamber 70 when the valve 130 is closed. In the present case, this working area essentially also corresponds to the area of the opening on the valve element 71. Using the mutual relationships described here, the modulating valve under, equal weight conditions, has the effect that the brake pressure in the fuel line is essentially equal to that in line 62 the prevailing inlet pressure.

When the inlet pressure increases, whereby the pressure in the control chambers 92 and 93 also increases, the control piston 98 is pushed down and tries to compress the spring 112. The speed at which such a compression can take place, however, is limited by the liquid flowing through the throttle point 140. Therefore, the brake pressure normally increases more slowly than the inlet pressure. The amount of lag of the brake pressure behind the inlet pressure can be viewed as the result of two effects, each of which can be used on its own to achieve such a beneficial effect.

First of all is the work surface. of the control piston 98 effectively has an area of the upper surface of the control member 104 which is the same size as the cross section of the piston 98. Since the working surface is exposed to the pressure in the upper chamber 93 , the downward pressure exerted on the piston 98 depends on the pressure in the chamber 93 instead of the pressure in the lower control chamber 92 . When the regulating component 104 moves downwards , the operating pressure can only be maintained and increased as the pressure fluid flowing through the throttle point 140 fills the increasing volume of the upper chamber 93. The flow volume required for this to flow through the bore 140 is equal to the working area of the valve actuating member 100 multiplied by -, d6m path of the piston. This volume can be freely selected and changed within wide limits of the construction. For example , you can increase the flow volume per unit of force. thereby increasing the spring force in that the - or enlarged surface of the valve actuator 100 has a softer. Spring used. On the other hand, any pressure difference in a fluid flowing upward through the bore 140 causes a force that urges the valve actuator 100 upward. The effective working area of the valve actuator 100 is its total area less than the working area described. of the control piston 98. Since this effective working area of the valve actuating member 100 is considerably larger than the area of the control piston 98, even a slight excess pressure below the valve actuating member 100 can completely compensate for the force exerted by the control piston 98. In fact, this force is also counteracted by the upward force of the spring 112 supported on the spring plate 110. Thus, the available pressure difference, which has to press the pressure fluid upwards, is approximately equal to the pressure which is necessary to exert an upwardly directed force on the valve actuating member 100 which is the same as that. The downward force developed by the piston 98 is less than the force of the spring 112. This available pressure difference is a small fraction of the total inlet pressure prevailing in the lower canister 91. Tn ,. Idle state of the system when the brake fluid pressure across the modulating valve 60 is practically preferably as far biased equal to the pressure of the return line 64. The spring 112 is that they 98 holds the control piston at the upper end of its stroke, so that the pressure control valve 130 is closed and the Pressure regulating valve 73 is open, the spring simultaneously opposing an upward movement of valve tappet 126 with a defined, predetermined force. When brake pressure is then applied to the brake by the pilot's brake valve 40, the increasing inlet pressure is transmitted unhindered to the outlet chamber 70 and to the brake itself, whereby a slight braking effect is immediately initiated. The resulting pressure in the outlet chamber 70 exerts an upward force on the plunger 126 , which is soon sufficient to overcome the bias of the spring 112 and compress this spring, and which tries to close the pressure regulating valve 73, whereby a further rapid growth of the brake pressure is prevented. The initially rapid increase in brake pressure is particularly desirable in the case of brakes that require a considerable amount of fluid to flow in at the beginning of a braking process in order to apply the brakes.

The increasing inlet pressure is simultaneously transmitted to the lower control chamber 92 via the passage 136 and to the upper control chamber 93 via the bore 140. If the pressure above the valve actuator 100 becomes greater, then the resulting downward force acting on the control piston 98 actually does not cause any downward movement of the control component 104 until this force becomes the upward force of the pretensioned spring 112 overcomes. During this initial phase with increasing pressure in the chamber 93, the bore 140 causes only a negligible delay in the rate at which the pressure increases, since only a very small volume of liquid is required to increase the pressure under static conditions. Thus, the downward movement of the regulating member 104 is initiated practically at the same time that the inlet pressure exceeds a certain critical value which corresponds to the selected preload of the spring 112.

If the inlet pressure rises above this critical value, then the control piston 98 moves down at a limited speed and thereby tensions the spring 112 even further. During this phase of the working cycle, the pressure in the outlet chamber 70 is not directly determined by the pressure in the inlet chamber 72 , but is modulated by the pressure regulating valve 73 in accordance with the gradually increasing force exerted by the spring 112 on the valve stem 126. Because of the resilience of the spring, even a small increase in spring force, and hence in outlet pressure, requires a substantial downward movement of the top of the spring. The piston 98 therefore moves progressively downwards, each position of the piston corresponding to a certain equilibrium value of the starting pressure.

At every point of the downward movement of the control piston, the forces acting on it are practically in equilibrium. This approximate relationship can be expressed by the equation F = a (P - D) - AD , where F is the spring force of spring 112 at the instantaneous compression, a is the size of the working surface of control piston 98, A is the equal areas of the upper and lower lower working surfaces of the valve operating member 100, P the pressure below the valve operating member 100, which is substantially equal to the inlet pressure ' and D the pressure difference between the top and the bottom of the valve operating member 100 represents. Equation (1) expresses the fact that the force of the spring 112 at the momentary compression is equal to the downward force exerted by the pressure on the working surface of the control piston 98 minus the upward force that is exerted due to the pressure difference between the upper and lower working surface of the modulating piston. Equation (2) results from equation (1). The expression represents the pressure which would have to act on the working surface of the control piston 98 in order to compensate for the force F of the spring 112. Therefore , the expression in parentheses in equation (2) can be viewed as the difference between the actual inlet pressure P and the pressure to which the current spring position corresponds. This pressure difference is essentially equal to the pressure difference which is across the pressure control valve 73 , but which is the pressure difference between the inlet pressure B and the outlet pressure which is applied to the brake in the current operating state of the modulating valve.

The pressure difference B across the control bore 140 is not equal to the pressure drop across the pressure control valve 73, but only a small fraction of this value, this fraction being approximately the same is as indicated in equation (2). So if A is much larger than a, then that fraction is approximate Therefore, the available differential pressure forcing the pressure fluid through the throttle point 140 is only a small fraction of the pressure difference. between inlet pressure and outlet pressure. The value of this fraction is determined by the structural details of the device and can be relatively small, e.g. B. one tenth can be chosen. As a result, the implementation of an effective control is made much easier, without it being necessary to make the throttle point 140 hardly manufacturable small. For a given value of a, which normally corresponds to the opening cross-section of the pressure control valve 73 , a progressive increase in the area A of the modulating piston simplifies the control not only because of the described reduction in the differential pressure above the throttle point 140, but also because it increases the volume of liquid that must pass through the throttle point 140 when the piston moves by a unit of length.

In summary, when the pilot applied brake pressure increases rapidly, the modulating valve causes the actual pressure applied to the brake to increase proportionally up to a predetermined critical value which is determined primarily by the bias of the spring 112. The pressure then rises at a reduced speed, the value of which is slightly determined by such constructive variables as the spring constant of the spring 112, the ratio of the piston areas and the dimensions of the restriction 140 can be determined. If the pilot then maintains a constant pressure on the actuation lever of the brake valve 40, then the modulating valve 60 gradually increases the actual brake pressure until the pressure gradually equals the pressure set by the pilot. In the steady state, the pressure difference with respect to the valve actuator 100 disappears and the spring 112 is compressed to such an extent that F = aP (3) . If the inlet pressure applied to the valve 60 is now reduced, either by releasing the brake lever 42 by the pilot or by the action of the slip control valve 54, this pressure reduction is immediately passed on via the line 136 to the control chambers 92, 93 , whereby a pressure difference directed in this way is transmitted the valve actuator 100 is established that the piston is pushed down. In particular, when the pressure drop occurs rapidly, the existing equilibrium is disturbed and the pressure control valve 73 is opened so that the brake pressure can escape immediately and immediately from the brake line 30 to the input line 62. This reduction in pressure in the outlet chamber 70 reduces the upward force exerted on the valve tappet 126 so that the brake release is cumulative. Thus, the brake pressure decreases more directly proportionally with the decreasing inlet pressure, without any noticeable delay occurring.

If such a brake release is due to the action of the slip control valve 54, an incipient slip process is normally intercepted after a small fraction of a second, whereupon the valve 54 again applies the full brake pressure set by the pilot to the inlet line 62 . The modulating valve 60 immediately transmits this increasing pressure to the brake line 30, which, however, only increases up to a critical value described by the pretensioning of the spring 112. Under the currently assumed conditions, the tension of the spring 112 is not limited by the above-mentioned preload, but corresponds to the existing position of the control piston 98. During the few milliseconds in which the pressure is normally removed to absorb a slip, this piston moves under the force the spring 112 upwards by a distance which is limited by the speed at which the pressure fluid can escape through the throttle point 140. The speed of this flow is determined by the size of the bore and by the pressure difference on both sides of this bore, as has already been described for an increase in pressure. During the pressure release, the pressure below the valve actuating member 100 can be assumed to be zero, and the pressure above this piston is approximately given by the spring force F which acts on the total area A + a on the upper side of the piston. But with this the differential pressure D 'is approximately the same Shortly before the pressure was released, the spring force was approximately equal to a - P from equation (3), so that the pressure difference immediately after the pressure was released is approximately given by Like Equation (2), this expression includes a factor that is approximately equal to the ratio and usually much less than 1 . One sees therefore immediately that during a release process of the brakes to intercept a slipping process, the control piston 98 moves only a small distance upwards and thus only causes a correspondingly small reduction in the tension of the spring 112. If the pressure is thus supplied to the valve 60 again, then this pressure is immediately passed on to the brake up to a critical value which is generally only slightly less than the value at which the beginning of the slip process has occurred. Above this critical value, the speed of the pressure increase is modulated by the valve in the manner already described. Thus there is a reasonable time delay before the brake pressure again reaches its full value at which the wheel previously started to slip. During this time, the Brein system and the wheel itself return to their stable state, which reduces the risk of a new slip process being initiated. Even if there is another slipping afterwards, an effective braking effect has been achieved for a noticeable time. The throttle point 40 can, if desired, be replaced by two separate bores which are provided with corresponding, oppositely directed check valves, so that one bore has a limited flow upwards through and around the piston 100 and the other bore a similar flow from controls the current flowing downwards. This arrangement has the advantage that the pressure modulation caused by the piston during the pressure increase can be determined by selecting a suitable cross-section of the bore without influencing the pressure modulation caused by the piston during the pressure decrease. In this way, however, this last-mentioned effect can also be set independently of the regulation when the pressure increases by suitable selection of the dimensions of the bore which controls the downward flow.

With the construction described up to this point, if the pressure coming from the source decreases very slowly, it can be achieved that this pressure becomes smaller than the brake pressure actually present, without the pressure control valve 73 being opened as a result. In this case, the brake pressure would not decrease in direct proportion to the decrease in the applied pressure, but would only decrease due to the escape of pressure fluid through the valve 130 after the line 64 called the relief or return line, corresponding to the upward movement of the control component 104. While this mode of operation may be desirable for certain purposes , it is usually preferable to use special means to ensure that the pressure control valve 73 opens when the pressure drops. This can be achieved, for example, by a design which, under equilibrium conditions, slightly overcompensates for the upward force exerted on the surface 127 by the braking pressure.

For this purpose, for example, a downward force can be applied to the component 120 by hydraulic or mechanical means directly in relation to the housing 61 or via a force transmitted via the spring 112. For example, for this purpose the working area of the control piston 98 can be made somewhat larger than the effective area of the working area 127, so that when the brake pressure and inlet pressure are equal, the component 120 is moved downwards, whereby the pressure control valve 73 opens. The difference between the two work surfaces need only be about 5 to 1011 / o for this purpose. However, this arrangement may have the disadvantage that the force actually required to open the pressure control valve 73 under equilibrium conditions also increases with increasing pressure. It may therefore be difficult to achieve a completely trouble-free operation at low pressure without applying simultaneously a greater overcompensation that is desirable at higher pressures.

The in F i g. The embodiment described in FIG. 2 provides a substantially constant force which, under equilibrium conditions, tends to open the pressure control valve 73. This force is generated by a C spring 145 which is applied between the top surface of the damping piston 122 and the inner shoulder 146 at the base of the upper housing portion 66. The spring 145 exerts a downward force on the component 120 which is sufficient to exceed the minor force of the spring 83 by an amount which is about a small fraction, about one twentieth to one tenth of that biasing force corresponds to the force exerted by the control spring 112. This force is substantially constant, since the axial movement of the component 120 is small. When the outlet pressure will always be controlled by the control action of the pressure regulating valve 73, then causes the spring 145, that the regulated The value of the outlet pressure is higher than it would otherwise be, namely by that of the spring force of the spring 14 5 corresponding amount. The direct effect of this difference on the braking effect is negligible for most practical cases. After a step-like increase in the inlet pressure, however, the increasing outlet pressure, for example, reaches the level of the inlet pressure before the differential pressure over the modulating piston has dropped completely to zero. As this differential pressure decreases further, with a further downward movement of the control part 104, an additional, downwardly directed force is exerted on the valve body 80 , as a result of which the pressure control valve 73 is inevitably opened. Then, when the inlet pressure gradually decreases, the pressure fluid can freely escape from the outlet chamber 70 to the inlet chamber, so that the fluid pressure acting on the surface 127 decreases in accordance with the fluid pressure, whereby the pressure regulating valve 73 is reliably kept open.

It is of great advantage to provide additional devices in the construction of the modulating valve which, in the event of slipping, effect a brake control, as is the case in connection with the valve 54 in FIG. 1 . A combination of modulating and slip control valve is shown in one embodiment in FIG . 3 and is denoted by 60a. If the valve 60a in place of the modulator valve 60 in the one shown in FIG. 1 is employed, the slip control valve 54, its control mechanism 55 and the flow channel 56 can be practically omitted. The lines 44 and 62 are then connected directly to one another. In addition, the electrical lines 52 via which the slip signals arrive from the measuring devices 50 , as in FIG. 3 , can be replaced by lines 154 and 164 which are directly connected to valve 60a. The in F i g. 3 not only compactly combines the slip control and the pressure modulation, as already described, but also uses the pressure modulator to, in addition to the precisely measured pressure increase already described, also a precisely measured pressure decrease under certain operating conditions to be described achieve.

The essential parts of the valve 60 a in F i g. 3 are the same as those already in connection with the valve 60 in FIG. 2 and are therefore given the same reference numerals. The two constructions shown differ mainly in that in FIG. 3 control valves are provided in certain flow channels, whereby the previously described mode of operation of the modulating valve can be modified as a function of control signals, as shown by a slip measuring device 50 in FIG. 1 can be delivered. In the embodiment shown here, two such slip control valves are used. One is in the flow channel from the channel 136 - F i g. 2, the other is inserted into a passage which corresponds to the passage 132 in FIG. 2 corresponds. These two slip control valves practically control two different levels of slip control as a function of corresponding signals from the slip measuring device 50, which correspond to various successive slip states with increasing degrees of difficulty. The first stage of regulation is controlled by valve 150 shown in FIG. 3 is shown in its normal position and is displaceable to the left depending on the excitation of the electromagnet 152. A spring 151 holds the valve ball in contact with the actuating pin 156, which is passed tightly through the housing wall. A spring 155 is in engagement with the armature 157 of the electromagnet and overcomes the relative force of the spring 151 when the electromagnet is not energized. The second stage of slip control is controlled by valve 160 , which is shown in FIG. 3 is also shown in its rest position. It is held in this position by the spring 161 and can move to the left depending on the excitation of the electromagnet 162.

In the rest position of the valve 150 , the connecting lines 136a and 136b, or also called passages, are directly connected to one another and thus connect the regulating chamber 92 to the inlet line 94, as in FIG . 2. When the electromagnet 152 is energized, the valve 50 blocks the passage 136b and connects the passage 136a to the return chamber 76 via a transverse passage 154. This effect occurs in response to a first slip signal arriving on lines 154 from measuring device 50 ( FIG. 1), i.e. practically at a point in time when a noticeable pressure is applied to the brakes the brake pressure applied by the pilot via his brake control valve 40 has already moved the control piston 98 down a noticeable distance from its rest position and thus has caused a noticeable brake pressure to be delivered through the modulating valve to the brake line 30. The actuation of the valve 150 reduces the pressure in the control chamber suddenly transferred 92 from the inlet pressure of the line 62 to the relatively low return pressure. a larger portion of this Druckverwinderung is substantially directly to the upper control chamber 93 above. of Ventübetätigungsglieds 100, since a change in pressure in the enclosed chamber only a negligible flow through the Requires throttle point 140. The on the work surface of the Regelk Downward force acting on piston 98 is therefore practically canceled, while the upward force exerted by compressed spring 112 is practically maintained. Therefore, a pressure difference D ' approximately corresponding to equation (4) is built up across the valve actuating member 100 . Accordingly, the piston 98 moves upward as rapidly as is possible due to the pressure fluid flowing through the throttle point 140 as a function of the pressure difference. This movement largely corresponds to the upward movement of the control piston, as already in connection with FIG. 2, which occurs in response to an abrupt decrease in inlet pressure. ] [In the present case, however, the inlet rack remains practically constant or even increases depending on the pilot's brake actuation. Thus, the immediate release of the brake pressure described above does not occur. Instead, the brake pressure is gradually reduced at a limited rate determined by the upward movement of the control piston 98 . While this movement causes a relief of the spring 112, whereby the downward force exerted on the valve tappet 26 is reduced, brake fluid can escape from the brake line 30 via the valve 130 to the return chamber 76 and the return line 64, whereby the brake pressure corresponds directly to the gradual decrease in the tension F of the spring 112 decreases.

This feature is that the precisely measured decrease of the braking pressure by the same mechanism is achieved which previously supplied the precisely measured increase in the brake pressure, is a great advantage. In both cases, the accuracy and reliability of the control are considerably better than in the case of a control which can be achieved with a device in which a bore with a narrowed cross section is connected directly in series with the brake. This is particularly true for brake systems in which the movement of a small volume of pressure fluid from the brake and after the brake causes a relatively large change in the brake pressure. The pressure modulation of the brake pressure works in the valve with a relatively large volume of fluid and with a relatively small pressure difference above the throttle point 140.

If the gradual decrease in brake pressure achieved by actuating the valve 150 is sufficient to intercept an incipient slip process, which is often the case, the signal coming from the measuring device 50 stops and the electromagnet 152 drops, so that the valve 150 returns to its rest position. The inlet pressure prevailing in the inlet line 62 is therefore immediately fed back to the lower control chamber 92. As a result, the piston 98 is gradually moved under the modulating control action of the piston 100 in such a direction that the brake pressure is brought back to a value which corresponds to the inlet pressure. If the inlet pressure has remained constant, for example, then this actually means a gradual increase in the pressure, and a noticeable time interval is required before the brake pressure can again reach the value at which the slip process was previously initiated.

If the described action of the solenoid valve 150 of the first stage is insufficient to intercept an incipient slipping process, the measuring device 50 ( FIG. 1) delivers a second slip signal via line 1.64 which excites the solenoid 162 of the second stage and thus the valve 160 actuated. When the valve 160 is actuated, the line 132 a is separated from the return chamber 76 and instead is connected to the inlet chamber 72 via an inclined bore 166 . In this way, the full available pressure of the supply line 62 is supplied to the underside of the piston 120, the working area of which is large in comparison with the working area of the regulating piston 98. In the embodiment shown here, the part of the chamber located above the piston 120 is 90 is not connected to the lower part of the chamber through the piston, as in the embodiment according to FIG. 2, but is in constant contact with the return pressure. Such a connection can be made by a T-shaped flow channel 170 in FIG. 3 can be produced. With this arrangement, the operation results of the slip control valve of the second stage, a 'acting on the piston 120, so that upward force, which easily overcomes the oppositely directed force of the Feder112 - the valve plunger 126 moves nach'oben quickly and positively to the stop will. The valve body 80 follows this upward movement under the slight force of the spring 83 until this movement is terminated by the valve seat of the valve 73. The continued upward movement of the valve tappet 126 opens the valve 130, whereby the brake pressure drops rapidly, since a flow of brake fluid to the return line 64 is initiated.

This complete release of the brake results in a reliable interception of incipient slip areas. When the wheel has returned to its normal rotational speed, slip signals on lines 154 and 164 cease, causing the corresponding electromagnets to be triggered and the corresponding valves 150 and 160 to return to their rest position. The pressure differential across the piston 120 is thereby removed and the inlet pressure is again supplied to the lower control chamber 92. The remaining compression of the spring 112 immediately pushes the piston 120 downwards, closes the valve 130 and opens the pressure control valve 73, which remains open until the brake pressure in the line 30 has been increased again to its critical value, which is the Preload of the spring 112 corresponds.

Since the valve actuating member 100 has moved upwards under the force of the spring 112 during the entire slip control by the valves 150 and 160 , this critical braking pressure is noticeably lower than the pressure which triggered the beginning of the slip process. On the other hand, this pressure is still considerably higher than the value which corresponds to the full extension of the spring 112. Thus, the brake immediately receives a sufficiently high pressure so that the normal braking effect also begins immediately. This pressure is then gradually increased at a limited rate which, as already described, is limited by the action of the modulating piston 100. It can be seen that with increasing intensity of the beginning of the slip process and thus with increasing duration of the slip control or the control effected by the valves 150 and 160 , the brake pressure which is applied again immediately afterwards, in comparison to the brake pressure which triggered the slip process, correspondingly is greatly reduced.

Another embodiment is shown schematically in FIG. 4 shown. This construction is practically identical to that of FIG. 3, however, has some structural advantages, in particular a space-saving, shorter overall length in the axial direction. The lower section of the construction according to FIG. 4 with the valve body 80, the valves 150 and 160 and the directly assigned chambers and passages is identical in its arrangement and mode of operation to the corresponding part in FIG. 3 and therefore does not need to be described and shown again. The two AusführLmgsfoimen differ mainly in the place of the spring 112 and the construction of the - rule'bauteiles. Ih the embodiment according to _F i g. 4, the spring 112 is mounted above the partition 178 dividing the housing in the same chamber 92a as the control piston and the modulating piston. The piston 120 a serving for damping and braking is arranged in a chamber 90 a below the partition 78 , the axial length of which can be relatively small because the axial movement of this piston corresponds practically only to the short axial movement of the valve body 80.

The control component 104 a in FIG. 4 comprises a one-piece piston construction which can be moved in the axial direction in the cylindrical control chamber 92 a. The control component 104 a is directly in engagement with the upper end of the spring 112 and determines the tension of this spring. The fluid pressure is supplied to the lower control chamber 92a via the housing passage 3136c and the inclined bore 136d in the partition wall 178 . The passage 136 c is directly related to the chamber of the skid control valve 150 connected and receives the pressure under the control of this valve, as g in connection with the passage 136a in F i. 3 has been described. The lower control chamber 92 a below the piston 100 a and the upper section by section 93 a above the piston 100 a are interconnected by a mounted within the body of the piston narrowed bore 140a. The section of the regulating component 104a which is on the outside in the radial direction therefore corresponds to the modulating pin 100 in FIG . 3 and is denoted by 100 a. The effect of this piston is clear from the above nen Gegan P a description.

- The lower end of the control spring 112 is carried by a bracket 180 which is mounted on the valve member 182nd The valve member 182 is in axially aligned holes in the partition wall 178 and the partition wall in the housing between the comb ren 76 and 90 used a movable. The lower section of the valve component 182 forms the valve tappet 126a, which in terms of design and function corresponds to the valve tappet 126 in FIG. 3. The central section of this valve component carries the damping and slip control piston 120a firmly attached to it within the chamber 90a. The axial position of the valve component 182 controls the valves 73 and 130 and is essentially determined by the existing relationship between the downward force of the control spring 112 and the upward force acting on its underside 127 caused by the fluid pressure in the outlet chamber 70 certainly. The upper end of the valve component 182 slides in a cylindrical axial bore 184 in the control component 104 a and thus forms a piston that works in the cylinder 186 . The pressure in the cylinder is maintained equal to the substantially uniform pressure of the return chamber 76 , thereby exerting a uniform downward force on the upper end surface 183 of the valve member 182 and a uniform upward pressure on the upper end surface 99 of the cylinder. This uniform pressure is supplied, for example, by an axial bore 187 which extends over almost the entire length of the valve component 182 and ends in its transverse bore which leads to the return chamber 76 . A second transverse bore 189 provides the return pressure after the top Abschnit the chamber 90 a above the piston 120a, and thus corresponds to the passage operatively 170 g in F i. 3.

The upper end surface 99 of the cylinder 186 can be viewed in the present embodiment as the lower working surface of the control piston and is denoted by 98a. The upper working surface of this piston is then the middle section of the upper surface of the control component 104 a directly above the surface 99 and is the same area as this surface. Accordingly, the piston 98 a is subjected to a downward force which is equal to the pressure difference between the upper control chamber 93 a and the return pressure in the cylinder 186 . The resulting forces acting on the piston 98 a therefore correspond directly to the previously described operating mode of the control piston 98 in FIG. 3.

Another embodiment. is schematically in FIG. 5 shown. The embodiment. according to FIG. 5 uses the same spring arrangement as in FIG . 4 in combination with other structural features that offer certain advantages on the structural side and also result in a somewhat modified mode of operation of the device. The lower part of the construction according to FIG. 5 should be constructed in the same way as the corresponding parts according to FIG. 3 and is thus not shown and described.

In Fig. 5, the housing partition wall 178 separates a chamber 92 b, the control of the chamber 90 b is mounted axially movable in the one of the damping and skid control dierfender piston 120b. This piston is in one piece - formed with the valve component 182 a. The lower end of this valve component forms a valve tappet 126 which extends into the outlet chamber 76 and rests against the valve body 80 , as already described in connection with FIG. 3 described. The Mittelabächnitt of the valve component 182a extends through a sealed bearing guide in the housing partition wall 178 a and carries the lower end of the control spring 112 in the control kammer92b i using the Federbügels180a. The upper end of the valve component 182a is inserted in a sliding and sealed manner in an axial bore in the control component 104b, as was explained in a similar manner in connection with FIG. The piston and cylinder 186 a thereby formed in the regulating member 104b is supplied via an axial bore 187 with a pressure fluid. With this construction, the pressure fluid fluctuations in the chamber 92 b only via the control spring 112 transmitted as pressure on the valve member 182a. The movement of the valve component 182a upon actuation of the valves and also the much larger movement of the control piston 98b , as will be described below, do not result in any change in the effective total volume of the chamber 92b .

The control component 104 b in FIG. 5 comprises the valve actuating member 100 b, which is arranged so as to be movable to and fro in a cylindrical cam 92 b and acts directly on the upper end of the spring 112. The valve actuating member 100 b is provided with a narrowed bore 140 a that delimits the pressure fluid flow or the pressure equalization flow, as has already been shown in FIG. 4 was shown. However, the control piston works in a cylinder in which the fluid pressure is independent of the pressure in the modulating chamber 92 b . As shown, the control piston 98 b forms part of the control component 104 b, extends upwards from the valve actuating member 100 b and operates in a cylinder 210 which is formed by an axial housing bore. The fluid pressure is supplied to the cylinder 210 via a bore 136 e which is in communication with the chamber of the valve 150 , which is controlled in the manner already described. Thus, but the pressure in the cylinder 210 is generally equal to the inlet pressure, but dependence reduced by the operation of the valve 150 in the absence of a slip signal to the pressure of the return line.

An essential feature of the embodiment. is that the control chamber 92 b is completely isolated from the rest of the fluid system. In this case, the valve actuator 100 b the various control processes by modulating or changing the pressure, as described previously, perform. Any desired average value or equilibrium value of the fluid pressure can thereby prevail within the chamber 92b. This chamber can, for example, be filled with a liquid medium and completely sealed. Normally, however, it will be desirable to provide means to compensate for changes in temperature, thereby avoiding excessive changes in pressure. For most areas of application, this can be achieved, for example, by providing a flow channel which enables a relatively small movement of liquid between the chamber and a source of the same pressure. For example, a passage can be through the partition wall 178 a to the chamber 90 be b provided which lies on the pressure of the return line, wherein the passage is defined by a narrowed bore, the waste usually measurements in comparison with the control bore 140a are small. As in Fig. As shown in FIG. 5, a temperature compensation device is provided which does not require connection to other parts of the apparatus. For this purpose there is a laterally attached line 212 which is arranged on the side wall of the chamber 92 outside the area of the valve actuating member 100b and is provided with a narrowed bore 213 which limits the flow. The line 212 leads to a collecting container 214 of conventional design and of sufficient size to accommodate any changes in the volume of the liquid in the chamber 92b that may occur. The bore 213 is made sufficiently narrow that the fluid flow through this bore during the pressure modulation action of the device is too small to noticeably affect the degree of compression of the control spring 112. Even a very small bore can produce a sufficient flow, for example to compensate for an expansion of the liquid resulting from temperature changes.

An advantage of isolating the chamber 92b from the rest of the fluid system, as shown in FIG. 5 is that the fluid used as the operating medium for the valve actuator 106b may be different from the fluid used in the rest of the system. For example, in connection with a hydraulic brake system, it may be desirable for the modulating piston to operate in a liquid medium which has different properties than the pressure fluid. In particular, the pressure modulation can be given a piston exhaust measurement be increased by a given flow-restricting bore such that a liquid is used with greater viscosity.

In addition, the embodiment of FIG. 5 can be used to explain the usefulness of the device for improving the operation of pneumatic braking systems which are generally used for actuating the brakes of railroad trains, automobiles and aircraft. The embodiments described above are for use in systems that use pneumatic media, such as. B. use air is not suitable, since satisfactory operation of the valve actuator requires that it operates in a substantially incompressible medium. With the in F i g. 5 , the valve operating member can be in an incompressible medium, such as. B. one of the known hydraulic fluids work, while the rest of the system is pneumatic, e.g. B. is actuated by compressed air.

For example, the control system shown in FIG. 1 can be regarded as a typical pneumatically operating brake control system. The in Fig. 1 , the modulating valve 60 shown in FIG. 5 , the slip control valves 150 and 160 as shown in FIG. 2 are omitted. In addition , the construction according to FIG. 5 including valves 150 and 160 as in FIG. 3, both the modulating device 60 and the slip control devices 54 and 55 of FIG. 1 represent. In either case, of course, all of the return lines 34, 45, 56 and 64 are not necessary since the system's pressurized air can conveniently be vented to atmosphere and the pump 32 can compress air from the same source for delivery via the pressure line 36 .

Another advantage of the in FIG. 5 construction shown is that the relatively high supply pressure in the Leitung- 62 only be used in the cylinder 210 Wird- when the slip control "combined in the same device is housed, in order to raise nevertheless Kölben 120b. Since the other parts of the device are only exposed to relatively low pressures, the entire construction can be made correspondingly light, which saves weight in the finished product and thus also reduces manufacturing costs.

During the pressure modulating process of the valve device in FIG . In the embodiment shown in FIG. 5, the regulating component 104b is moved, controlled by the pressure of the printing block, in a similar manner to that which has already been explained in connection with the other embodiments. For example, during an increase in the brake fluid pressure supplied to control cylinder -210 via line 136 , the substantial balance between the spring force F and the fluid forces acting on the control member at any point in its downward movement can be expressed by the following equation: F # a - P - A - D, (5) where the first term on the right side of the equation downward force on the - represented b top side of the control piston 98, which is while the second term on the right side of the equation upward force, which is caused by the pressure difference D, which acts on the working surfaces of the valve actuator 100 b. From equation (5) one obtains If. is small, then the equations are different (5) and (6) do not differ significantly from equations (1) and (2), respectively. Equations (4) and (4a) discussed in connection with the upward movement of the control member can likewise be applied to the present embodiment. can be applied ' by replacing the expression A + a with A.

F i g. 6 shows a further embodiment which, like FIG. 5 is suitable for use in hydraulically and pneumatically operating brake systems. The parts in FIG. 6, which are not specifically described, are so constructed and operate in the same manner as FIG. 5.

As in the previously described embodiments, means are also provided here to allow a limited flow of liquid to pass from one side of the valve actuating member 110 to the other side thereof. In the construction in FIG. 6 , means, generally designated 140c, are provided for this purpose, which form part of the housing and comprise the flow channel 220 and the valve 222. This valve limits the flow through the flow channel 220 and can be adjusted from outside the housing by means of a screw 224. Such an easily accessible setting of the regulating bore can also be provided in the embodiments described above and has the advantage that the pressure modulation of the device corresponds to the various Operating conditions can be easily changed.

In Fig. 6 ends the valve component 182b , which otherwise corresponds in all parts to the valve component 182a in FIG . 5 corresponds - above the spring plate 180 b. The top surface 231 of this component is therefore exposed to the fluid pressure in the Kanirn he c 92 below the control component 104c. The resulting downward force is equal to the pressure prevailing in the chamber multiplied by the working surface of the valve component 182b at the point at which it passes through the housing partition 178b . This section of the valve component 182b therefore acts as a piston and is provided with the reference number 230 . The force acting on the working surface 231 of the piston 230 is transmitted directly to the valve tappet 126 and not just via a change in the preload of the control spring 112. The magnitude of this force can be changed within wide limits in the construction of the mechanism by suitable selection of the effective piston area. In Fig. 6 , this area is made slightly smaller than the working area of the control piston 98 c.

Fig. 6 also shows a collection ebb, which is here generally provided with -dem Bezitigszeichen, 214 a and partially that in F i g. 5 corresponds to the collecting container 214 shown. For space-saving reasons of Sarnmel ebälter214a c is in the interior of the Regelbautelles 104 using a room housed, the g in the embodiment of F i. 5 was used in part for the cylinder 186 a. In this component, an axial chamber 233 is formed, which is accessible through an opening in the lower surface of the piston 100c. A piston 232 is axially slidable in the chamber 233 and is urged downward by a spring 234. The opening of the chamber 233 is almost completely blocked from the flow of liquid by a closure piece 236 which is inserted tightly and which has a very small bore 238 . This bore allows a considerably smaller flow than the control bore on valve 222 for a given pressure difference. The section of chamber 233 above piston 232 is filled with air and is kept at atmospheric pressure by a vent valve of a suitable type. As shown, a Duirchlaß 240 is provided through the side wall -the chamber 233 just below the sealing means 242 of the plunger 98 c. Furthermore, a passage is provided after the outer shift through the housing wall at 244 just above the lower seal 246. From Raumersparnisgränden 242, the top seal of an O-ring which is inserted c on the periphery of the piston 98, while the lower seal 246 is composed of an inserted into the housing wall O-ring The normal clearance of a few hundiertstel Millimeterii formed between the Wall of the piston 98 c and the housing and these O-rings, allows enough air for the present purposes between the passages 240 and 244 through. The comb he 92 c and also the part of the compensation channel 233 located below the piston 232 are precipitated with a suitable liquid medium: e-mem. When this fluid expands, for example with increasing temperature, the compensation piston 232 is moved upwards against the force of the spring 234, so that a substantially uniform pressure is maintained in the equilibrium state. Such B ewegung of the piston is also used to compensate for changes in the total volume of the Kanimem92c and 93b on both sides of the piston 100 c, resulting from axial movement of the control piston 98c and the piston 230th It can be seen that the change in volume due to the movement of the piston 230 is small, since it only covers a short distance. The change of volume of the chamber 92c due to the movement of the control piston has c 98 is a considerable value, but is much less than the movement accompanying this flow of liquid through the restricted bore, d. H. Throttle point 140 c.

In the embodiment according to FIG. 6 , the lower surface of the control piston 98 c is exposed to the liquid pressure in the chamber 92 c, the change of which corresponds essentially to the pressure difference or the pressure difference D. Thus the equilibrium of the forces corresponds to F i g. 6 equations (1) and (2) and not equations (5) and (6).

The c prevailing in the chamber 92 pressure also acts on the piston 230 a which forms a part of the valve component 182b. The constant portion of this force, which corresponds essentially to the constant tension of the compensation spring 234, has an influence on the operation of the valve which is essentially equivalent to an increase in the preload of the main regulating spring. This effect can easily be taken into account in the construction of the device.

The variable portion of the pressure acting on the piston 230 corresponds to the pressure difference D , in particular during the downward movement of the control component 104. The resulting force acting on the piston 230 is transmitted directly to the valve tappet 126 and thus acts directly on the valve body 80 and increases it thus the fluid pressure which is transmitted via the pressure control valve 73 after the brake. So that the immediate response of the piston 230 is clearly distinct from the previously described operation of the various forces acting on the control component 104c forces the kand influence the actual braking pressure only on the movement of the control spring 112 acting on the piston 230 force acts on the other hand parallel to the force exerted by the lower end of the spring 112 and is functionally equivalent to a change in the preload of this spring. However, this effective change in spring preload occurs without changing the actual spring tension.

In general, the described force acting on the piston 230 has the effect that the brake pressure actually applied is increased by a certain small amount while the brake pressure is being applied. This amount is the valve actuating member 100 c, a lag of the braking pressure after the pressure applied by the pilot to the inflow conduit 62 pressure causes the strength immediately and directly proportional to the differential pressure D, which can be regarded as a measure.

Thus, if the pressure applied by the pilot increases rapidly, the described force acting on piston 230 has the effect that the brake pressure actually applied follows this increase more precisely than would otherwise be the case. However, this effect does not change anything in the movement of the regulating component 104 c and thus clearly differs from any structural change that would cause this movement to take place more quickly.

The mode of operation described can be seen particularly during the first application of the brakes. In the previously described embodiments, the increasing pressure from the supply line is immediately supplied to the brake, but only up to a value which corresponds to the preload of the spring 112. Above this value, the brake pressure only increases to the extent that the control component moves downwards and the spring tension increases. In this embodiment, on the other hand, the increasing pressure of the supply line is fed directly to the brake up to a value which corresponds to the preload of the spring 112 plus a predetermined proportion of the applied pressure in excess of this preload value. During normal application of the brakes, this excess pressure is only very small, and the effect described causes practically only a slight change in the braking process. However, if the pilot suddenly applies full brake pressure, for example in an emergency, then the described effect of the differential pressure on the piston 230 can markedly accelerate the application of the brake pressure to the brake. The magnitude of this effect can easily be determined in the design of the device by suitable selection of the size of the working surface of the piston 230.

Claims (2)

Claims: 1. Brake pressure control system for the brakes of a vehicle wheel, in particular a chassis wheel of an aircraft, which can be actuated as a function of a variable fluid pressure, with a brake valve actuated by the pilot as well as with a control valve actuated by a slip signal and with a pressure control valve device, which on the input side via the valve operated by the pilot is connected to a pressure fluid source and on the output side with the brake and has a relief conduit and contains a valve under spring pressure and a valve actuator which is used to regulate the output pressure in accordance with a manipulated variable acting on the valve actuator which is controlled by one of the The throttle device charged on the input side is generated, characterized in that the pressure control valve (73) does not connect the input line (62) to the relief line (64) in any position and is acted upon on both sides ares valve actuator (100, 100a, 100b, 100e) , wherein the throttle point (140, 140a, 140c) in the connecting line between the control chambers (92, 93; 92 a, 93 a; 92 b; 92 c, 93 b) , and of these only one (92, 92 a, 93 b) is connected directly to the inlet line (62) of the valve (73) . 2. Control system according to claim 1, characterized in that the throttle device consists of cylindrical control chambers (92, 93) , the working surface of which is larger than the bore (96) into which the control piston (98) is slidably inserted, and the valve actuating member (100 ) is movably arranged in the cylindrical control chambers (92, 93) , the throttle point (140) establishing a connection between the cylindrical chambers (92, 93) separated from one another by the valve actuating member (100) . 3. Control system according to claim 1 and 2, characterized in that a valve (150) of the anti-slip device in the connecting line (136a, 136b) between the control chamber (92) and the input line (62) of the valve (73) is arranged ( Fig. 3), which responds to an abnormal decrease in wheel speed and thereby isolates the connection line from the input line and connects the connection line to a source of lower pressure. Documents taken into consideration: German Auslegeschrift No. 1061626; British Patent No. 704 925; USA. Patent Nos 2,698,205, 2,919 162. Contemplated prior patents: German patent no. 1156 218th.