Fluid pressure control device

A ball valve supporter is fixedly attached to an internal wall of a bore formed in an end plug of a fluid pressure control device. The valve supporter lacks an aperture which penetrates the supporter and opens into a ball valve chamber from the supporter at a portion confronting the ball valve. The valve supporter is formed outside the portion confronting the ball valve with passage means opening into the ball valve chamber.

BACKGROUND OF THE INVENTION 
The present invention relates generally to a braking pressure control 
device which regulates the output fluid pressure for the wheel cylinders 
of the motor vehicle to increase at the same rate as the input fluid 
pressure from the master cylinder when the input fluid pressure is below a 
critical fluid pressure. The regulation provides an increase at a rate 
smaller than the rate of increase in the input fluid pressure when the 
input fluid pressure is above the critical fluid pressure. Further, the 
critical fluid pressure is varied in accordance with variation in the 
weight of the motor vehicle. More particularly, the invention relates to 
an improved braking pressure control device of this type including barrier 
means for preventing the flow of the input fluid pressure from striking 
the inertia responsive movable ball valve and exerting thereon a thrust or 
dynamic pressure which would move the valve to the valve seat to close the 
port thereof. The barrier means causes the input fluid pressure to flow 
through the surroundings of the valve to the valve seat. 
As is well known in the art, usual motor vehicle hydraulic braking systems 
are such that the brakes are applied to the front and rear wheels 
concurrently. In this instance, if an excess amount of braking force is 
applied to the front wheels, the front wheels are locked earlier than the 
rear wheels to make it impossible for the driver to handle the motor 
vehicle. On the contrary, if the rear wheels are braked excessively, they 
are locked prior to the front wheels to cause the rear portion of the 
motor vehicle to swing transversely to the longitudinal direction thereof. 
Accordingly, in order to assure and increase the safety and stability of 
the vehicle during the braking operation, it is necessary to effect the 
distribution of the braking forces to lock the front and rear wheels 
concurrently. 
When the vehicle is braked, the so-called nosedive phenomenon takes place 
in which the vehicle weight applied on the front wheels increases and the 
vehicle weight applied on the rear wheels decreases. Accordingly, it is 
necessary for concurrently locking the front and rear wheels to distribute 
to the front wheels a braking force greater than a braking force 
distributed to the rear wheels. It is also necessary that the distribution 
of the braking forces to the front and rear wheels is varied in accordance 
with variation in the vehicle weight. Thus, ideal characteristics of 
distribution of the braking forces to the front and rear wheels, when are 
illustrated on oblique coordinates having the axes of abscissa and 
ordinate indicating respectively thereon the ratios (deceleration rate 
ratios) Bf/W and Br/W of the braking forces Bf and Br on the front and 
rear wheels versus the vehicle weight W, are expressed by curves having 
tangents the angles of inclination of which are relatively large within a 
range of the origin to a certain value and are relatively small outside 
the range. Furthermore, the heavier the vehicle weight is, the higher the 
ideal characteristics curve is located on the coordinates. 
It is accordingly necessary for providing the distribution of the braking 
forces which is close to the ideal characteristics curve to feed to the 
rear wheel cylinders a fluid pressure increasing at a rate smaller than 
that of increase in a fluid pressure fed to the front wheel cylinders or 
at a rate of zero when the fluid pressure fed to the front wheel cylinders 
exceeds a predetermined or critical fluid pressure. As an expedient for 
solving the problem, a limiting valve, proportioning valve or G valve was 
disposed in a rear braking circuit leading to the rear wheel cylinders. 
The limiting valve generates an output fluid pressure increasing at a rate 
of zero when an input fluid pressure exceeds a critical fluid pressure. 
The proportioning valve generates an output fluid pressure increasing at a 
rate lower than that of increase in an input fluid pressure when the input 
fluid pressure exceeds a critical fluid pressure. The G valve generates an 
output fluid pressure increasing at a rate less than an input fluid 
pressure when a predetermined rate of deceleration is attained. However, 
the output fluid pressure generated by these valves merely carried out the 
distribution of the braking forces approximating to a single ideal 
characteristics curve which accordingly, corresponds to a predetermined 
vehicle weight and, when the vehicle weight is varied, effected a 
distribution of the braking forces which largely deviated from an ideal 
characteristics curve corresponding to the vehicle weight varied. 
On the other hand, most motor vehicles are in recent years provided with a 
hydraulic braking system of the tandem type which comprises front and rear 
braking circuits leading respectively separately to the front and rear 
wheel cylinders. However, a braking pressure control valve which is 
disposed in the rear braking circuit generated the same output fluid 
pressure as in the event of no failure of the fluid pressure in the front 
braking circuit in the event of the failure of the fluid pressure in the 
front braking circuit and as a result caused the deficiency of the braking 
force. 
Thus, the applicants have proposed a braking pressure control device 
comprising a fluid pressure regulating valve serving as a proportioning 
valve or limiting valve, biasing means which urges the fluid pressure 
regulating valve and to which the master cylinder fluid pressure is fed to 
control the force of the biasing means, and an inertia responsive movable 
ball valve responsive to a predetermined rate of deceleration. The ball 
valve closes the path of flow of the fluid pressure to the biasing means 
to maintain the fluid pressure having been fed to the biasing means at a 
predetermined value. Consequently, the critical fluid pressure is varied 
in accordance with variation in the vehicle weight to generate the output 
fluid pressure which provides the distributions of the braking forces 
which approximate to the ideal characteristics curves corresponding to the 
vehicle weight varied. The fluid pressure regulating valve is also biased 
by the fluid pressure in the front braking circuit in a direction opposite 
to the biasing direction by the biasing means. Thus, in the event of the 
failure of the fluid pressure in the front braking circuit, the critical 
fluid pressure is sufficiently increased to generate the output fluid 
pressure high to compensate the deficiency of the braking force. 
However, the conventional braking pressure control device has had a 
drawback in that support means for supporting the ball valve is formed 
through its central end portion with an aperture for passing the master 
cylinder fluid pressure to the biasing means. Therefore, the flow of the 
fluid pressure strikes the ball valve to exert a thrust thereon. The 
thrust moves the ball valve into a position to close an inlet port to the 
biasing means to hinder the braking pressure control device from 
exhibiting its desired function completely. 
SUMMARY OF THE INVENTION 
It is, therefore, an object of the invention to provide an improved braking 
pressure control device in which the flow of the master cylinder fluid 
pressure is prevented from striking the ball valve. Consequently, a thrust 
is not produced and does not move the ball valve into a position to close 
an inlet port to the biasing means. Thus, the braking pressure control 
device exhibits its desired function completely. The control device 
includes barrier support means for supporting the ball valve. The barrier 
support means lacks an aperture penetrating its central end portion to 
thereby prevent the master cylinder fluid pressure from passing 
therethrough. Flow of the fluid pressure is diverted to the periphery of 
the barrier means through passage means formed between the periphery of 
the barrier means and the ball valve. The fluid pressure diverted by the 
barrier means and through the passage means passes on to the biasing means 
.

DESCRIPTION OF SPECIFIC EMBODIMENTS 
Referring to FIG. 1 of the drawings, the ideal characteristics curves 
a.sub.1 and a.sub.2 as per the introduction of the specification of the 
distribution of the braking forces to the front and rear wheels are 
illustrated on oblique coordinates having the axes of abscissa and 
ordinate indicating respectively thereon the ratios (deceleration rate 
ratios) Bf/W and Br/W of the braking forces Bf and Br on the front and 
rear wheels versus the vehicle weight W. The curves a.sub.1 and a.sub.2 
are the ideal characteristics curves at the time when the weight of the 
vehicle is W.sub.1 (no load) and W.sub.2 (the vehicle carries load), 
respectively. Generally, the heavier the vehicle weight is, the higher or 
the more the ideal characteristics curve is positioned or extends steeply 
from the origin O in the graph of FIG. 1. 
As is apparent from the graph, the angle of inclination of a tangent of 
each of the curves a.sub.1 and a.sub.2 is relatively large within a range 
of the origin O to a certain value and is relatively small outside the 
range. In the graph of FIG. 1, there is also illustrated the 
characteristics lines b.sub.1 and b.sub.2 of the distribution of the 
braking forces to the front and rear wheels which distribution is provided 
to approximate respectively to the ideal characteristics curves a.sub.1 
and a.sub.2 by a motor vehicle hydraulic braking system incorporating 
therein a braking pressure control device according to the invention. 
Referring to FIG. 2 of the drawings, there is shown a motor vehicle 
hydraulic braking system incorporating therein a braking pressure control 
device or valve according to the invention. The hydraulic braking system, 
generally designated by the reference numeral 10, includes a master 
cylinder 12 operated from a brake pedal 14. First and second hydraulic 
fluid circuits 16 and 18 lead from the master cylinder 12. The first fluid 
line 16 is connected to front wheel cylinders 20 cooperating with brakes 
(not shown) of front wheels 22 of a motor vehicle. The second fluid line 
18 is connected to the control device, generally designated 24, which is 
connected through a fluid line 26 to rear wheel cylinders 28 cooperating 
with brakes (not shown) of rear wheels 30 and 32 of the vehicle. The first 
and second fluid lines 16 and 18 are further connected to the control 
device 24 through branch lines 34 and 36, respectively. The control valve 
24 is mounted on the body (not shown) of the vehicle to have its axis 38 
inclined at an angle of .theta. from the horizontal plane 40 so that the 
forward end portion of the control valve 24 is positioned above the 
rearward end portion thereof. 
Referring to FIG. 3 of the drawings, the detailed construction of the 
braking pressure control valve 24 according to the invention is shown. The 
control valve 24 comprises a housing 42 having a first cavity 44 and inlet 
and outlet ports 46 and 48 which are formed in its front portion 43. The 
inlet and outlet ports 46 and 48 are connected respectively to the second 
fluid line 18 and the fluid line 26. An annular sealing member 50 such as 
a lip type seal is fixedly attached to a wall defining the cavity 44 and 
divides the same into first and second chambers 52 and 54 into which the 
inlet and outlet ports 46 and 48 open, respectively. The annular sealing 
member 50 has formed therethrough an aperture 56. A plunger 58 extends 
through the aperture 56 and is axially movable in the first and second 
chambers 52 and 54. The aperture 56 provides an annular clearance between 
the annular sealing member 50 and the plunger 58 to provide fluid 
communication between the first and second chambers 52 and 54. A plug 
member 60 is firmly fitted in a bore 61 formed in the forward end portion 
62 of the housing 42 and closes the forward end portion 62. The plug 
member 60 has formed therein an inlet port 64 connected to the branch line 
34 of the first fluid circuit 16, and a bore 65 communicating with the 
port 64. The plunger 58 has a stem portion 66 located in the first chamber 
52, an annular projection 68 having a cross sectional area of A.sub.1, and 
forward and rearward end portions 70 and 72 having cross sectional areas 
of A.sub.2 and A.sub.3, respectively, both of which are smaller than 
A.sub.1. The annular projection 68 is located in the second chamber 54 and 
is engageable with the annular sealing member 50 to obstruct fluid 
communication between the first and second chambers 52 and 54. The forward 
end portion 70 is connected to the annular projection 68 and is slidably 
supported in an aperture 74 formed through the forward end wall 76 of the 
cavity 44 and extends into the bore 65 of the closure member 60 through 
the aperture 74. The bore 65 is sealed from the second chamber 54 by a 
seal member 78. The rearward end portion 72 is connected to the stem 
portion 66 and is slidably supported in an aperture 80 formed through a 
rearward end wall 82 of the cavity 44. The rearward end portion 72 is 
formed therein with a blind bore 84 in which a push rod 86 is received. 
The housing 42 further has a second cavity 88 formed in its mid portion 89, 
and two opposite bores 90 and 92 formed in opposite end walls 94 and 96 of 
the cavity 88 and both opening into the cavity 88. Two pistons 98 and 100 
are slidably fitted in the bores 90 and 92, respectively. The push rod 86 
extends from the bore 84 of the plunger 58 into the bore 90 and engages 
the plunger seat 98. A spring seat 102 is slidably fitted in the cavity 88 
and is in abutting engagement with the end wall 96 or the piston 100. An 
inner compression spring 104 is located between the piston 98 and the 
spring seat 102 to urge these two members in opposite directions. An outer 
compression spring 106 is located between the end wall 94 and the spring 
seat 102 to urge the latter against the end wall 96 or the piston 100. The 
piston 100 has a cross sectional area of A.sub.4. The bore 90 is sealed 
from the first chamber 52 by a seal member 107. A chamber 108 is defined 
in the bore 92 between the piston 100 and an end wall 110 of the bore 92. 
The housing 42 further has a third cavity 112 and a bore 114 which are 
formed in its rearward end portion 116, and a bore 118 formed in an end 
wall 120 of the cavity 112. The bore 114 communicates with cavity 112 and 
has an open end. An inertia responsive movable ball member 122 is 
rotatably or rollably fitted in the cavity 112. A valve seat member 124 is 
firmly fitted in the bore 118 and has formed therethrough an aperture 126 
opening into the cavity 112 and communicating with the chamber 108 through 
a passage 128. The ball member 122 constitutes an inertia responsive valve 
129 together with the seat member 124 and is seated on the valve seat 124 
in response to a predetermined rate of deceleration or inertia force to 
obstruct fluid communication between the chamber 108 and the inlet port 
136. The cavity 112 has a groove 130 formed around the ball member 122. An 
end plug member 132 is threaded in the bore 114 to close the open end of 
bore 114. Plug member 132 includes a bore 134 opening into cavity 112 and 
an inlet port 136 communicating with the bore 134 through an orifice 137 
and connected to the branch 36 of the second fluid circuit 18. A ball 
support member or valve supporter 138 is press fitted in the bore 134 or 
is fixedly attached to an internal wall of bore 134, unmovably relative to 
end plug 132. Valve supporter 138 is formed in its circumferential surface 
139 or outside a portion confronting ball valve 122 with an axial groove 
or grooves 140 as shown in FIG. 4 of the drawings which communicates with 
the inlet port 136. The ball support member 138 is not formed at the 
portion confronting ball valve 122 with an aperture passing from its outer 
end 142 to its inner end 144. However, groove 140 provides fluid 
communication between the groove 130 and the inlet port 136 to cause the 
pressurized hydraulic fluid from the inlet port 136 to flow into the 
aperture 126 through the grooves 130 and 140. Thus, flow of the hydraulic 
fluid is prevented from striking the ball member 122 and from exerting 
thereon a thrust which moves the ball member 122 into a position to close 
the valve seat member 124. It is preferable that the ball support member 
138 is formed in the outer end 142 with radial grooves (not shown) leading 
to the grooves 140 for preventing or minimizing pressure loss of the 
hydraulic fluid from the inlet port 136. 
The braking pressure control device 24 thus far described is operated as 
follows: 
When the brake pedal 14 is depressed, the master cylinder 12 delivers a 
hydraulic fluid pressure Pm into the first and second fluid lines 16 and 
18. The fluid pressure Pm in the first fluid line 16 is fed into the front 
wheel cylinders 20 and through the inlet port 64 into the bore 65 of the 
pressure control valve 24. The fluid pressure Pm in the second fluid line 
18 is fed into the first chamber 52 of the pressure control valve 24 
through the inlet port 46 and is then delivered into the second chamber 54 
through the aperture 56 of the annular sealing member 50 as an outlet 
hydraulic fluid pressure Pr which has been modulated or unmodulated. The 
outlet fluid pressure Pr in the second chamber 54 is fed into the rear 
wheel cylinders 28 through the outlet port 48. The fluid pressure Pm in 
the second fluid circuit 18 is also fed into the chamber 108 of the 
pressure control valve 24 through the inlet port 136, the grooves 140 and 
130, and the aperture 126 of the seat member 124. 
When the inlet fluid pressure Pm is less than a critical fluid pressure Ps, 
the outlet fluid pressure Pr in the second chamber 54 is equal to the 
inlet fluid pressure Pm, that is, 
EQU Pr = Pm 
In this condition, the fluid pressure Pm in the bore 65 exerts on the 
forward end portion 70 of the plunger 58 a force Pm .times. A.sub.2 which 
urges the plunger 58 rearwardly. When the inlet fluid pressure Pm is 
increased to the critical fluid pressure Ps, the force Ps .times. A.sub.2 
exceeds the force F.sub.1 of the inner spring 104 to move the plunger 58 
into a closed position in which the annular projection 68 engages or is 
pressed against the annular sealing member 50 to obstruct fluid 
communication between the first and second chambers 52 and 54. At this 
time, the following relation is defined: 
EQU Ps .times. A.sub.2 = F.sub.1 
accordingly, the critical fluid pressure Ps is expressed as 
EQU Ps = F.sub.1 /A.sub.2 Eq. 1 
In this instance, since the displacement of the plunger 58 is extremely 
small, an increase in the force of the spring 104 is so little as to be 
neglected. 
When the input fluid pressure Pm is subsequently further increased, the 
fluid pressure Pm in the first chamber 52 exerts on the annular projection 
68 a force which urges the plunger 58 into an open position to unseat the 
annular projection 68 from the annular sealing member 50. When the annular 
projection 68 is unseated from the annular sealing member 50, the fluid 
pressure Pm in the first chamber 52 is allowed to flow into the second 
chamber 54 to cause an increase in the output fluid pressure Pr. At this 
time, i.e., when Pm .gtoreq. Ps, the following equilibrium equation is 
established: 
EQU PmA.sub.2 +Pr(A.sub.1 -A.sub.2)=Pm(A.sub.1 -A.sub.3)+F.sub.1 Eq. 2 
Accordingly, the output fluid pressure Pr is expressed as 
##EQU1## 
The output fluid pressure Pr delivered from the outlet port 48 under the 
control of the braking pressure control valve 24 is in this manner given 
by either of the Equations 1 and 3 in accordance with the valve of the 
input fluid pressure Pm. Thus, when the input fluid pressure Pm increases 
from zero, the output fluid pressure Pr increases at the same rate as the 
input fluid pressure Pm until the input fluid pressure Pm reaches the 
critical fluid pressure Ps. When the input fluid pressure Pm increases 
above the critical fluid pressure Ps, the output fluid pressure Pr 
increases at the rate of m [wherein m = (A.sub.1 -A.sub.3 
-A.sub.2)/(A.sub.1 -A.sub.2)]which is smaller than the rate of increase in 
the input fluid pressure Pm. 
On the other hand, when the braking force B on the wheels increases with 
increases in the input fluid pressure Pm, the ratio of the rate a of 
deceleration versus a gravitational acceleration g also increases. This 
deceleration rate ratio a/g is equal to the ratio of the braking force B 
versus the overall weight W of the motor vehicle as follows: 
EQU a/g = B/W Eq. 4 
The braking force B is proportional to the input fluid pressure Pm as 
follows: 
EQU B = CPm (wherein C is a constant) Eq. 5 
When the deceleration rate ratio a/g reaches a predetermined value of 
(a/g).sub..theta. which is a function f(.theta.) of the angle .theta. of 
inclination of the pressure control valve 24, the ball member 122 of the 
inertia responsive valve 129 rolls forwardly in response to the force of 
inertia to seat on the valve seat member 124 to close the aperture 126 to 
obstruct fluid communication between the chamber 108 and the inlet port 
136. Thus, even if the inlet fluid pressure Pm subsequently increases, the 
fluid pressure in the chamber 108 is maintained at a fluid pressure Pg 
which is the input fluid pressure Pm at the moment when the aperture 126 
of the seat member 124 has been closed by the ball member 122. The fluid 
pressure Pg is expressed from the Eqs. 4 and 5 and the Eq. 6 
[(a/g).sub..theta. =f(.theta.)]as 
##EQU2## 
At this time, from the condition of equilibrium of the piston 100 and the 
Eq. 7 the following equation is obtained: 
##EQU3## 
where F.sub.2 is the force of the outer spring 106. 
The forces F.sub.1 and F.sub.2 of the inner and outer springs 104 and 106 
are expressed respectively as the sums of the amounts f.sub.1 and f.sub.2 
of preset or initial loads of the springs 104 and 106 and the products of 
the amounts of deflection or shrinkage of the springs 104 and 106 by a 
compressive force from the piston 100 and the spring constants K.sub.1 and 
K.sub.2 of the springs 104 and 106. In this instance, since the amounts of 
deflection of the springs 104 and 106 are equal to each other, the 
following equation is obtained: 
##EQU4## 
From the Eqs. 8 and 9, the force F.sub.1 of the spring 104 is obtained as 
##EQU5## 
Substitution of the Eq. 10 into the Eqs. 1 and 3 results in 
##EQU6## 
When Pm .gtoreq. Ps 
##EQU7## 
It is apparent from the Eq. 11 that by selecting the variables in the Eq. 
11 in a manner to make the value of (f.sub.2 -f.sub.1 .multidot.K.sub.2 
/K.sub.1) positive, a relation between the critical fluid pressure Ps and 
the overall vehicle weight W as shown in FIG. 6 of the drawings is 
obtained in which relation the critical fluid pressure Ps increases at a 
rate greater than the rate of increase in the vehicle weight W when the 
vehicle weight increases. As a result, it is possible to make the 
characteristics of fluid pressure or braking force distribution close to 
the ideal characteristics curves a.sub.1, a.sub.2 of FIG. 1 in accordance 
with increases in the vehicle weight W. 
In the event of the failure of the input fluid pressure Pm in the first 
fluid circuit 16, since PmA.sub.2 = 0 in the Eq. 2, the following equation 
is obtained: 
EQU Pr(A.sub.1 -A.sub.2)=Pm(A.sub.1 -A.sub.3)+F.sub.1 
accordingly, the output fluid pressure Pr is obtained as 
##EQU8## 
In this instance, between the braking force B on the wheels and the input 
fluid pressure Pm the following relation is provided: B = C'Pm where C' &lt; 
C. Hence, the force F.sub.1 ' of the spring 104 is expressed as 
##EQU9## 
When the input fluid pressure Pm is at a critical fluid pressure Ps', the 
following equation is defined: 
EQU Ps' (A.sub.3 -A.sub.2) = F.sub.1 ' 
accordingly, the critical fluid pressure Ps' is obtained as 
##EQU10## 
where Ps' &gt; Ps. Accordingly, it is apparent that the critical fluid 
pressure Ps' is increased to a considerably high value which provides a 
braking force so great as to compensate the failure of the input fluid 
pressure Pm in the first fluid circuit 16. 
Support member 138 serves as the barrier means and prevents the flow of the 
fluid pressure Pm from the inlet port 136 from striking the rear surface 
146 of the ball member 122 and from exerting on the ball member 122 a 
thrust which moves the same toward the valve seat 124. Thus, ball member 
122 does not impede the admission of the fluid pressure Pm into the fluid 
chamber 108 to provide the forces of the springs 104 and 106. These forces 
are controlled by the fluid pressure Pm through the piston 100 to a 
desired or predetermined value and the ball member 122 moves to the valve 
seat 124 in response to a desired or predetermined deceleration rate 
accurately to have the control device 24 perform its desired function or 
operation accurately. 
It will be appreciated that the invention provides a braking pressure 
control device comprising support and baffle means for diverting the flow 
of the fluid pressure thereto to the periphery or edge thereof. The 
support and baffle means prevents the flow of the fluid pressure from 
striking the rear end of a ball valve. Thus, a thrust does not produce a 
bad influence upon the admission of the fluid pressure into a piston 
chamber nor hinder the ball valve moving to close an inlet port to the 
piston chamber in response to a predetermined deceleration rate. The 
control device also includes passage means for passing the diverted flow 
of the fluid pressure to the inlet port between the circumference or 
circumferential surroundings of the support and baffle means and the ball 
valve consequently the braking pressure control device exhibits its 
desired function completely or best. 
The braking pressure control device has support and baffle means press 
fitted in a closure member for the rearward end of the device body to be 
integral with the closure member. Therefore its furnishing to the device 
body is easy as compared with a conventional ball support means held 
between a ball valve body and a closure member. The support and barrier 
means can be easily exchanged for a new one without removing the whole of 
the control device from the vehicle body or at a state in which the 
control device remains mounted on the vehicle body. 
Although the invention has been described as being applied to a braking 
fluid pressure control device comprising a proportioning valve, the 
invention can be applied to a braking fluid pressure control device 
comprising a limiting valve in place of a proportioning valve.