GB2111150A - Control valve for vehicle brakes - Google Patents

Abstract

A dual circuit brake control valve has inlets (2,3) connected to respective outlets (4,5) via respective metering valves (6,8). Each metering valve comprises telescopically arranged outer (12), intermediate (14) and inner (13) pistons have respective effective cross-sectional areas of A1, A2 and A3. The clearance space between the confronting ends of the inner pistons (13) is determined by a deceleration responsive member, e.g. a weight (38) rotatably mounted above its centre of gravity and having a helical surface complementary to a helical surface (27) on one inner piston (13). The net effective area of the metering valves (6, 8), which determines the cut-in pressure and input-versus-output characteristics, is A1-A3 or A1-A2 depending on the clearance space (determined by rotation of weight (38)) between the inner pistons at the moment when the pressure subsisting the brake system is sufficient to move the inner (13) and intermediate (14) pistons. Other embodiments show different deceleration responsive members, viz. a ball-and-ramp mechanism (figs 4, 5 not shown), or a pendulum (figs 6, 7 not shown). <IMAGE>

Description

SPECIFICATION Control valve assembly This invention relates to a control valve assembly, and more particularly to a control valve assembly for the hydraulic brake system of a motor vehicle.

It is known to provide control valve assemblies in hydraulic brake circuits in order to vary the hydraulic pressure applied to one or more wheels of the vehicle in relation to the hydraulic pressure applied to other wheels. For example, a control valve may be provided to ensure that the brake pressure supplied to rear brakes of the vehicle is less than the brake pressure supplied to the front brakes of the vehicle under certain operating conditions.

According to the present invention there is provided a control valve assembly to be located between a source of hydraulic operating fluid and a wheel brake of a motor vehicle hydraulic brake circuit, the control valve assembly comprising: an inlet; an outlet; a valve for controlling communication between the inlet and the outlet, the valve being operated by a piston which is movable for the purpose of operating the valve by hydraulic fluid pressure in the brake circuit acting on a net effective area of cross-section of the piston; and means responsive to deceleration of the vehicle for varying said net effective area of cross-section of the piston.

In the preferred embodiments of the invention a pair of metering valves associated with separate brake circuits are mounted in a common housing.

The deceleration responsive means is mounted in the housing between the metering valves and comprises a weight which is pivotally mounted for rotation about an axis located above its centre of gravity, or a ball.

The above and further features and advantages of the invention will become clear from the following description of preferred embodiments thereof, given by way of example only, reference being had by the accompanying drawings wherein: Figure 1 is a cross-sectional view of a control valve assembly; Figure 2 is a perspective view of the deceleration responsive weight of the valve of Figure 1; Figure 3 illustrates operating characteristics of the valve of Figure 1; Figure 4 shows a modification of the valve assembly of Figure 1; Figure 5 illustrates operation of the embodiment of Figure 4; Figure 6 shows a further embodiment of the invention; and Figure 7 is a view in the direction of the arrow VII of Figure 6.

Referring firstly to Figures 1 and 2 the valve assembly 1 is for use in a dual circuit hydraulic brake system of a motor vehicle. The control valve assembly 1 includes two inlets 2, 3 for receiving hydraulic fluid from a suitable source, for example a tandem master cylinder, and two outlets 4, 5 for communication with respective rear wheel brakes.

Communication between the inlet 2 and the outlet 4 is controlled by a metering valve 6 located within a first housing part 7, whilst communication between the inlet 3 and outlet 5 is controlled by metering valve 8 located within a second housing part 9.

The housing parts 7 and 9 are substantially identical and are joined together at a common interface 10 to form the housing of the control valve assembly.

It should be noted that the control valve assembly 1 does not provide any fluid communication between the metering valve 6 on the one hand and the metering valve 8 on the other hand. The space 11 within the housing between the metering valves 6, 8 is vented to atmosphere.

Each metering valve 6, 8 comprises an outer piston 12, and an inner piston 13, and, located between the outer and inner pistons, a sleeve-like intermediate piston 14. Seals 15 provide sliding sealing contact between the outer pistons 12 and stepped bores 1 6 of the housing parts 7, 9. Seals 1 7 provide sliding sealing contact between the outer pistons 12 and the intermediate pistons 14, and between the intermediate pistons 1 4 and inner pistons 13. Radial passages 18 formed in each outer piston 1 2 connect the zone between associated seais 1 5 to the stepped bore 1 9 of the outer piston 12.An axial passage 20 extends from the stepped bore 1 9 of each outer piston and is connected by radial passages 21 to a respective outlet 4 or 5. A seat 22 formed at the juncture of the stepped bore 1 9 and axial passage 20 is engageable by a valve member 23 as hereinafter described to isolate the stepped bore 1 9 from the axial passage 20.

The pistons 1 3 terminate at the ends thereof adjacent the centre of the valve 1 in enlarged heads which define shoulders 24 for engagement by pistons 14. The head 25 of the right hand piston as viewed in Figure 1 is in the form of a flange, whilst the head 26 of the left hand piston as viewed in Figure 1 is formed with a helical cam surface 27 described in more detail hereinafter.

Each metering valve 6, 8 includes a spring 28 which biases the valve member 23 away from inner piston 13 and towards seat 22, and a spring 29 which biases the intermediate piston 14 towards the centre of the valve. A spring 30 biases the outer pistons 1 2 away from each other.

In the rest condition of the valve illustrated in Figure 1 the outer pistons 12 are biased by spring 30 against respective end walls 31 of the housing parts 7, 9; the valve members 23 are biased by spring 28 into engagement with the end walls 31 and are spaced from their associated seats 22; and intermediate pistons 14 are biased by springs 29 into engagement with the shoulders 24 of the inner pistons 13.

A pin 32 engages a slot 33 in a spring seat 34 associated with left hand outer pistons 1 2 in order to prevent rotation of that piston within the housing 7, and a pin 35 secured to left hend outer piston 12 engages a slot 36 in head 26 to prevent rotation of the left hand inner piston 1 3 relative to the left hand outer piston 1 2.

Circlips 37 are located in grooves formed in outer pistons 1 2 and, in the rest condition of the valve illustrated, are separated from respective intermediate pistons 14 by a small gap.

Pivotally mounted between the inner pistons 1 3 for rotation about the longitudinal axis of the inner pistons 13 is a weight 38, which is shown in more detail in Figure 2. The weight 38 is pivotally mounted by means of two pins 39 which are biased apart by a spring 40 and which seat in respective bores formed in the inner pistons 1 3.

The centre of gravity of the weight 38 is offset from its axis of rotation so that in the absence of any acceleration or deceleration the weight hangs in the position shown in the drawings. The orientation of the control valve assembly is such that deceleration of the vehicle will cause the weight to rotate in the direction of the arrow A (Figure 2).

Referring now to Figure 2, the weight 38 includes an end face 41 which is located in a radial plane relative to the axis of rotation of the weight, and a helical cam face 42. The cam face 42 extends for a full revolution of the weight 38, the opposite ends of the cam face being connected by a flat wall 43 which extends parallel to the axis of rotation of the weight. The cam face 42 is complementary to the cam face 27 of the head 26. The cam face 27 extends through less than a full revolution in order to permit rotation of the weight 38 in the direction of the arrow A of Figure 2.

The spring 40 is sufficiently strong to ensure that in the rest conditions of the control valve illustrated in Figure 1 the end face 41 is spaced by a small gap from the adjacent face of the head 25 and the cam face 42 is spaced by a small gap from the adjacent cam face 27.

In use, hydraulic actuating fluid is supplied from a suitable source, for example a tandem master cylinder, to the inlet 2 and 3. When the brakes are initially applied the various components of the control valve assembly are in the positions illustrated in Figure 1 and hydraulic fluid flows freely from each inlet 2, 3 via respective radial passages 18, stepped bores 19, axial passages 20 and radial passages 21 to respective outlets 4, 5.

It will be appreciated that hydraulic pressure subsisting in the stepped bores 1 9 acts on the inner pistons 13 and intermediate pistons 14 over the entire area A2 of the larger portion of the stepped bore 1 9. Since the space 11 is vented to atmosphere, hydraulic pressure acting over the area A2 is resisted only by the spring 40 and at a predetermined pressure (hereinafter referred to as the setting pressure) Pe the pistons 1 3 and with them the associated intermediate pistons 14 will start to move towards the centre of the valve.

If we now consider the vehicle in a heavily laden condition, the pressure Pe subsisting in the braking system when the pistons 1 3 and 14 begin to move will be insufficient to produce a significant deceleration of the vehicle, and accordingly the weight 38 will be at or very near the rest position illustrated in Figure 1. Thus, the gap between the head 25 and end face 41 on the one hand and the gap between the confronting cam faces 27, 42 on the other hand will rapidly be ciosed and the weight 38 will be trapped in position. This will occur before the gaps between the intermediate pistons 14 and associated circlips 37 have been closed.Thus, after the setting pressure Pe has been reached the inner piston 1 3 and intermediate piston 14 of each metering valve will act as a single fixed rigid piston having an effective cross-sectional area of A2, and hydraulic pressure will act on each outer piston 12 over a net effective area of A1 - A2 where A, is the overall cross-sectional area of the piston 12. As hydraulic pressure continues to rise it will eventually reach a value at which the hydraulic force produced on the pistons 1 2 is sufficient to overcome the spring 30 and the pistons 1 2 will start to move towards each other.The pressure at which this occurs is referred to as the cut-in pressure and, in the fully laden condition, is for the purposes of this description P, and is determined by the area A1 - A2 and the preload of spring 30.

Before the inner ends of the pistons 1 2 abut, the respective valve seats 22 will engage the associated valve members 23 to isolate inlets 2, 3 for associated outlets 4, 5.

A further rise in inlet pressure will cause the outer pistons 1 2 to move away from each other thereby moving the valve members 23 out of engagement with the associated seats 22 to produce an increase in outlet pressure. As will be well understood by those skilled in the art, the outlet pressure will rise at a lower rate than the inlet pressure by the ratio: A1 -A2 A1 Referring to Figure 3, where outlet pressure Pout is plotted against inlet pressure Pin the above described operation of the control valve assembly wherein the vehicle is heavy laden is represented by the line OAB.

If we now consider the operation of the control valve when the vehicle is only lightly laden, setting pressure Pe will produce a significant deceleration of the vehicle and this will cause the weight 38 to rotate about the axis of the pistons 1 3 in the direction of the arrow A of Figure 2 to increase the clearance between the confronting cam faces 27, 42. As they result, when the pistons 13 and 14 begin to move at the setting pressure the intermediate pistons 14 will engage circiips 37 before the inner pistons engage the faces 41,42 of the weight 38. Thus, at pressures above P0 it is only the inner pistons 1 3 having an effective area A3 which act as rigid fixed pistons. The intermediate pistons 1 4 move with the outer pistons 10 and thus the effective area over which hydraulic pressure acts on the outer pistons 10 is A1 - A3 i.e. is a larger area than in the heavily laden condition referred to above. Metering operations of the metering valve 6, 8 under increasing pressure conditions will be generally as described above except that the cut-in pressure (Pc in this case) will be then determined by the area Aj - A3 and will thus be less than the cut-in pressure Pa relating to the heavily laden condition.

Further, at pressures above Pc the outlet pressure will be reduced relative to the inlet pressure by the ratio: A -A3 3 A1 and accordingly a given rise inlet pressure will produce a smaller rise in outlet pressure than in the case of fully laden conditions as described above.

The characteristics of the control valve assembly 1 in the lightly laden conditions are shown by the line OCD of Figure 3.

It will thus be appreciated that the provision of the deceleration responsive weight 38 not only alters the cut-in pressure of the control valve assembly 1 in dependence upon vehicle loading, but also provides a different meter ratio in the laden and unladen conditions. If necessary, the confronting surfaces of the weight 38 and the heads 25, 26 can be roughened to increase friction and prevent rotation of the weight 38 due to force applied to it by inner pistons 1 3.

If there is a fault in one brake circuit, for example the brake circuit associated with inlet 3, application of the brakes will result in a greater rise in pressure in the one stepped bore 19, in this case the left hand stepped bore 19, than in the other stepped bore 1 9. Under these conditions at a pressure below Pe the pistons 13 and 14 associated with the higher pressure brake circuits will begin to move towards the centre of the valve 1 because of the lack of balancing pressure in the other brake circuit. The intermediate piston 40 will soon come into engagement with the associated circlip 37, but the inner piston 13 will continue to move to take up the clearance between a head 44 of a stem 45 secured to the piston 13 and a flange 46 of cage 47 secured to the valve member 23.

Once the clearance between the head 44 and the flange 46 has been taken up further movement of the piston 13 will draw the valve member 23 away from the end wall 31 and thus prevent the valve member 23 from being engaged by seat 22.

Thus, a fault in either brake circuit resulting in reduced pressure in that brake circuit will automatically disable the metering valve of the other brake circuit and permit full inlet pressure to be transmitted to the wheel brake of the other circuit.

Referring now to Figures 4 and 5 a modified embodiment of the valve of Figure 1 is shown. In this modification the portion of the valve illustrated in Figure 4 replaces that portion of the valve of Figure 1 located between the planes P and P2 of Figure 1.

The pistons 1 3A of Figure 4 each include an outwardly directed flange 50, and are normally biased apart by a spring 51 so that the flanges 50 engage respective circlips 52 located in grooves 53 in the outer pistons 12A. The intermediate pistons 14 are biased against the circlips 52 by springs 29. The axial extent of the grooves 53 is greater than the axial length of the circlips 52 so that the circlips are permitted limited axial movement relative to the outer piston 12A. In the normal rest configuration of the valve the spring 51 is sufficiently strong to maintain the circlips 52 in the illustrated positions, i.e. at the axially outward end of their respective grooves 53.

The head 54 of the right-hand piston (as viewed in Figure 4) defines a conical recess 55 in which is located a ball 56. In use, the valve of Figure 4 is orientated such that the arrow F of Figure 4 points towards the front of the vehicle, the arrow R points towards the rear of the vehicle, and the arrow U points upwardly. Thus, when the valve of Figure 4 is orientated in its use position the recess 55 faces upwardly and the ball 56 will rest under the influence of gravity at the base of the recess 55.

The left-hand piston (as viewed in Figure 4) includes an extension 57 which can engage the ball 56 to hold the ball 56 in the recess 55 es described in more detail hereinafter.

Operation of the valve of Figure 4 will now be described in more detail.

Considering firstly the vehicle in a heavily laden condition, the setting pressure Pa at which the pistons 1 3A, 14 begin to move will be insufficient to produce a significant deceleration of the vehicle. Thus, the pistons 1 3A will move towards each other and the extension 57 will engage the ball 56 to prevent the ball leaving the base of the recess.This will occur before the circlips 52 have moved through the full available extent of the grooves 53, and accordingly at pressures above Pa the pistons 1 3A and 14 act as a single rigid piston of effective area A2, and hydraulic pressure will act on each outer piston 1 2A over an effective area A1 - A2. Operation of the valve will thus be as described above with reference to Figures 1 and 3.

Considering now the lightly laden case, setting pressure Pa will be sufficient to produce a significant deceleration of the vehicle, and this deceleration of the vehicle will be sufficient to run the ball 56 to the periphery of conical recess 55.

Thus, as the pistons 1 3A move towards each other, the extension 57 must engage the base of the recess 55 before the pistons 1 3A act as a solid piston. Before this can occur, the circlips 52 will have travelled the full axial extent of the grooves 53 and will accordingly limit the travel of the intermediate pistons 1 4. Thus, at pressures above Pa the configuration of the valve components will be as shown in Figure 5, that is to say the pistons 1 3A will be in contact to form a single rigid piston of effective area A3, whilst the intermediate pistons 14 are constrained to move with the outer pistons 12A.Thus, the net effective area of the outer pistons is A1 - A3, and operation of the valve at pressures above setting pressure is as described above with reference to Figures 1 and 3.

The valve of Figures 4 and 5 has several advantages as compared with that of Figures 1 and 2. Firstly, the relatively complicated helical cam arrangement of Figures 1 and 2 has been replaced by the relatively simple ball and recess arrangement of Figures 4 and 5. This tends to reduce the cost of manufacture and makes the valve more reliable. The device is also effective to reduce rear braking in the presence of high lateral acceleration such as might occur during high speed cornering. If the brakes are applied during cornering, the acceleration to which the vehicle is subjected is not aligned with the fore-and-aft direction of the vehicle, and accordingly if a valve is used which is sensitive only to fore-and-aft acceleration, the valve may not respond correctly to the deceleration present when the vehicle is unladen and setting pressure is reached.This could result in the valve providing a high cut-in pressure even though the vehicle is relatively lightly laden. This would have an adverse effect on braking characteristics, especially during cornering. It will be appreciated, however, that the valve of Figures 4 and 5 is not highly sensitive to the direction of acceleration to which it is subjected, and will be effective to provide the desired low cut-in pressure even though the acceleration of the vehicle is not aligned with the fore-and-aft direction.

Figures 6 and 7 show a further embodiment of the invention. In this embodiment the deceleration responsive means 60 comprises a pendulum 61 located within a casing 62. The top edge portion 63 of the pendulum 61 is located within a slot 64 in the casing wall to form a pivot 65 for the pendulum 61, and the weight of pendulum is supported by a spring 66 which in turn is supported by a bracket 67 provided by a portion of the casing.

A plunger 68 having one end located in a recess provided on the pendulum extends through a seal 69 into the interior of the valve housing part 9B where it is linked to a cranked lever 70 by a cross-pin 71. In the rest condition the components are in the positions illustrated with the lever 70 held flat against the end wall 31 by the valve member 23. Movement of the plunger 68 to the right as viewed in Figure 6 causes the lever 70 to pivot about the bend thereof and move the valve member 23 towards the centre of the valve assembly.

The valve assembly includes a pair of inner pistons 13B, a pair of outer pistons 1 2B and a pair of intermediate pistons 14 having respective overall areas A3, A1 and A2. The intermediate pistons 14 are normally spring biased into engagement with retainers 72 secured to the outer pistons 1 2B. An abutment sleeve 73 is slidably mounted on the inner pistons and is a clearance fit within the retainers 72 so that the opposite ends of the abutment sleeve 73 may be engaged by the intermediate pistons as described in more detail hereinafter.

In operation, regardless of vehicle loading, increasing pressure at the inlets 2, 3 initially moves the inner pistons 1 3B into engagement with each other, and thereafter continues to rise until each outer piston 1 2B and its associated intermediate piston 1 4 begins to move. At this point, if the vehicle is lightly laden a significant deceleration will be present and this will bias the pendulum in the direction of the front of the vehicle (arrow F of Figure 6) with sufficient force to maintain the plunger 68, and with it the lever 70, in the illustrated position.The outer and intermediate pistons will move as a single piston having an effect area of A1 - A3 and the valve members 23 will engage their respective seats at a relatively low pressure Pc. Thereafter fluid will be metered as described above with reference to Figures 1 and 3.

If, in the alternative, the vehicle is heavily laden, the pressure subsisting when the pistons 1 2B, 14 begin to move will be insufficient to produce a significant deceleration of the vehicle, and thus the pendulum 61 will only apply a small loading to the plunger 68. Under these conditions the pressure at the outlet 5 acting over the area of the plunger 68 will apply a force to the plunger sufficient to overcome the loading of the pendulum and seal friction, and thus the plunger will move to the right as viewed in Figure 6, thereby moving the pendulum in the direction of the arrow R and moving the crank lever 70 to displace the valve member 23 towards the centre of the valve assembly. With the valve member so displaced, the intermediate pistons 14 will engage the ends of abutment sleeve 73 before both valve members 23 engage their respective seats.

Thereafter, the intermediate pistons 14 and abutment sleeve 73 together act as a single rigid member, and the net effective area of the piston becomes A1 - A2. Pressure thereafter must rise to a value Pa higher than P, before both valve members 23 engage their respective seats.

Metering thereafter occurs as described above.

Claims (12)

1. A control valve assembly to be located between a source of hydraulic operating fluid and a wheel brake of a motor vehicle hydraulic brake circuit, the control valve assembly comprising: an inlet; an outlet; a valve for controlling communication between the inlet and the outlet, the valve being operated by a piston which is movable for the purpose of operating the valve by hydraulic fluid pressure in the brake circuit acting on a net effective area of cross-section of the piston; and means responsive to deceleration of the vehicle for varying said net effective area of cross-section of the piston.

2. A control valve assembly according to claim 1 wherein the means responsive to deceleration of the vehicle comprises means operative when a predetermined pressure exists within the brake circuit to select a relatively large effective area of cross-section if the deceleration present is above a predetermined value and to select a relatively small effective area of crosssection if the deceleration present is below the predetermined value.

3. A control valve assembly according to claim 2 wherein the control valve assembly comprises telescopically arranged outer, intermediate, and inner piston portions having overall areas A1, A2 and A3 respectively, and wherein said means responsive to deceleration is operative at said predetermined pressure to fix the position of the inner piston portion if the deceleration is above said predetermined value whereby the net effective area of the piston is A1 - A3 and to fix the position of the inner and intermediate piston portions if the deceleration is below said predetermined value whereby the net effective area of the piston is A1 - A2.

4. A control valve assembly according to claim 3 wherein the means responsive to deceleration comprises a weight movable under the influence of vehicle deceleration and forming a movable abutment for the inner piston portion.

5. A control valve assembly according to claim 4 wherein the weight includes a helical surface which is movable, in response to sensed deceleration, relative to a complementary surface on the inner piston portion, to increase the clearance therebetween.

6. A control valve assembly according to claim 4 wherein the weight is a ball normally located in a recess, the ball being engageable by the inner piston portion if the sensed deceleration is below the predetermined value, and being movable out of the path of the inner piston if subjected to a deceleration above the predetermined value.

7. A control valve assembly according to any of claims 4 to 6 wherein two control valve assemblies are provided for controlling two brake circuits, the control valve assemblies being located on opposite side of a weight which is common to the two assemblies.

8. A control valve assembly according to claim 1 wherein said valve is operative above a cut-in pressure to provide a pressure at the outlet which is lower than the pressure at the inlet, and wherein the cut-in pressure is determined by the net effective area of cross-section of the piston.

9. A control valve assembly according to claim 8 wherein the relationship between the inlet pressure and the outlet pressure at pressures above the cut-in pressure is determined by the net effective area of cross-section of the piston.

1 0. A control valve assembly according to claim 8 or claim 9 wherein the means reponsive to deceleration of the vehicle is operative to determine the net effective area of cross-section of the piston in response to the deceleration of the vehicle produced by a predetermined brake pressure lower than the cut-in pressure.

11. A control valve assembly according to claim 1 wherein the valve includes a valve member and a valve seal which are moved relative to each other by the piston, wherein the means responsive to deceleration includes means for displacing the valve member to increase the travel of the seat required to close the valve, and wherein the net effective area of cross-section of the piston is determined by the piston travel necessary to close the valve.

12. A control valve assembly substantially as hereinbefore described with reference to and as shown in the accompanying drawings.