Hydraulic pressure control apparatus for controlling a pressure of a working fluid in a brake cylinder

Abstract: An apparatus for controlling a working fluid pressure in a brake cylinder comprises: a low-pressure source (31); a solenoid-operated pressure control valve (58, 62) between low-pressure source and brake cylinder including a solenoid-operated seating valve (130) having a valve seat (134), a valve member (136) movable to be seated on and unseated from said valve seat, and a device (132) for generating an electromagnetic drive force (Fs) such that said force acts on said valve member in a first direction for moving said valve member to be seated on said valve seat, said seating valve being constructed such that the pressure difference force (Fp) between the fluid pressures in said low-pressure source (31) and said brake cylinder (50, 52) acts on said valve member in a second direction opposite to said first direction; and a controller for controlling the energy applied to said electromagnetic force generating device, depending upon the fluid pressure in said brake cylinder, when said valve member of said seating valve is held seated on said valve seat.

Description: BACKGROUND OF THE INVENTION Field of the invention The present invention relates to a hydraulic pressure control apparatus for controlling a fluid pressure in a hydraulically operated brake cylinder. Discussion of the Related Art JP-A-10-322803 (published on December 4, 1998, before the present invention was made) discloses such a hydraulic pressure control apparatus using an electrically operated hydraulic pressure source as a high-pressure source. This hydraulic pressure control apparatus includes (1) the above-indicated high-pressure source which includes a pump, an electric motor for operating the pump, and an accumulator for storing a working fluid pressurized by the pump, for supplying the pressurized fluid to a brake cylinder, (2) a low-pressure source for storing the fluid discharged from the brake cylinder, (3) a solenoid-operated pressure increase control valve disposed between the high-pressure source and the brake cylinder, and electrically controlled for controlling the pressure of the fluid in the brake cylinder, to a level depending upon an electric energy applied thereto, (4) a solenoid-operated pressure reduction control valve disposed between the low-pressure-source and the brake cylinder, and electrically controlled for controlling the pressure of the fluid in the brake cylinder, to a level depending upon an electric energy applied thereto, and (5) a control device for controlling the pressure increase control valve and the pressure reduction control valve to regulate the pressure of the fluid in the brake cylinder, while controlling the electric motor to hold the pressure of the fluid in the accumulator at a predetermined level. When the fluid pressure in the brake cylinder is increased under the control of this hydraulic pressure control apparatus, the electric energy to be applied to the solenoid-operated pressure increase control valve is controlled by the hydraulic pressure control apparatus, with the solenoid-operated pressure reduction control valve held in its closed state. The pressure of the fluid to be supplied from the accumulator to the brake cylinder through the pressure increase control valve is controlled according to the controlled electric energy applied to the pressure increase control valve. When the fluid pressure in the brake cylinder is reduced under the control of the hydraulic pressure control apparatus, the electric energy to be applied to the pressure reduction control valve is controlled by the apparatus with the pressure increase control valve held in its closed state. The pressure of the fluid in the brake cylinder is reduced to a level corresponding to the controlled electric energy. Thus, the pressure of the fluid in the brake cylinder is controlled by controlling the electric energies to be applied to the pressure increase control valve and the pressure reduction control valve. The assignee of the present invention proposed a hydraulic pressure control apparatus as disclosed in Japanese Patent Application No. 10-102476 which has not been laid open at the time the present invention was made). The hydraulic pressure control apparatus disclosed in this Application includes a normally open, solenoid-operated shut-off valve which is disposed between the brake cylinder and a reservoir and which is placed in its open state when an electric energy is not applied thereto and in its closed state when the electric energy is applied thereto. In a conventional hydraulic pressure, a solenoid-operated shut-off valve disposed between the brake cylinder and the reservoir is a normally closed valve which is placed in its closed state when no electric energy is applied thereto, in order to prevent delayed initiation of a braking effect provided by the brake cylinder. After each braking operation of the brake cylinder by releasing of a brake operating member, for instance, this normally closed shut-off valve is held in the open state for a predetermined time which is estimated to be required for substantially zeroing the pressure of the fluid in the brake cylinder. Then, the electric energy is completely removed from the shut-off valve, so that the shut-off valve is returned to the original closed state. However, the brake cylinder may have a residual fluid pressure even after the expiration of the above-indicated predetermined time, that is, even after the pressurized fluid has been discharged for the predetermined time. This residual fluid pressure may cause a so-called "brake drag", namely, some degree of braking even after the braking operation in question. This brake drag can be avoided where the shut-off valve is a normally open valve which is returned to the open state by removal of an electric energy after each braking operation. On the other hand, however, the normally open shut-off valve must be quickly brought to the closed state when it is required to increase the fluid pressure in the brake cylinder. To this end, the normally open shut-off valve disclosed in the above-identified Patent Application is energized with a maximum electric energy by the pressure control apparatus, when the fluid pressure in the brake cylinder is increased. Accordingly, this normally open shut-off valve requires a comparatively large amount of consumption of the electric energy, and suffers from a relatively large operating noise, a relatively short life, and other problems. Thereto WO 97/23372 A and GB-A-2 320 748 disclose a hydraulic pressure control apparatus according to the preamble of claim 1 in which electric energy to be applied to the solenoid-operated pressure control valve is controlled in relation to a pressure difference such that the electromagnetic drive force of said valve is larger than said pressure difference by a constant offset. Object of the present invention is to improve a hydraulic pressure control apparatus according to the preamble of claim 1 in that the electromagnetic drive force is controlled more effectively. This object is solved according to claim 1 in that said electromagnetic drive force is controlled to be larger than a pressure difference force by at least a predetermined margin value which is determined by at least a maximum velocity of increase in the fluid pressure at said solenoid-operated pressure control device. The solenoid-operated pressure control valve in the hydraulic pressure control apparatus according to the invention is a normally open valve which is placed in the open state when no electric energy is applied thereto. When the electromagnetic drive force generated by application of an electric energy to the pressure control valve is larger than the pressure difference force, the pressure control valve is brought to the closed state with its valve member being seated on its valve seat. The pressure difference force is generated by the pressure difference between the fluid pressure in the low-pressure source such as a reservoir and the fluid pressure in the brake cylinder. Since the fluid pressure in the low-pressure source can be considered to be approximately equal to the atmospheric pressure, the pressure difference force corresponds to the brake cylinder pressure. To hold the valve member seated on the valve seat, the application of the maximum electric energy to the pressure control valve is not essential. That is, the valve member can be held seated on the valve seat, as long as the electromagnetic drive force generated by application of the electric energy is larger than the pressure difference force. Thus, the pressure control valve may be held in the closed state by controlling the electric energy depending upon the brake cylinder pressure. In this arrangement, the required amount of consumption of the electric energy by the pressure control valve is smaller than in the arrangement wherein the maximum electric energy is applied to the pressure control valve for holding it in the closed state. The electric energy applied to the solenoid-operated pressure control valve may be changed with a change in the brake cylinder pressure, or may be kept constant at a predetermined value, for instance, at the value upon initiation of an anti-lock braking pressure control operation of controlling the fluid pressure in the brake cylinder for braking a wheel of an automotive vehicle in an anti-lock fashion. Even when the electric energy is kept constant at a suitable value determined by the wheel brake cylinder pressure at a given point of time during a control of the brake cylinder pressure, the required amount of electric energy can be made smaller than when the maximum electric energy is kept applied to the pressure control valve. Where the electric energy is controlled to change with a change in the brake cylinder pressure, the electric energy may be continuously changed or in steps. The solenoid-operated pressure control valve may include a biasing device for biasing the valve member in the above-indicated second direction for moving the valve member away from the valve seat. To hold the pressure control valve in the closed state in this case, the electromagnetic drive force is required to be larger than a sum of the pressure difference force and a biasing force of the biasing device. However, the biasing force can be ignored, since it is used for merely holding the valve member away from the valve seat and can therefore be considerably small. In a hydraulic pressure control apparatus according to the present invention the electric energy control means controls the electric energy to be applied to the electromagnetic force generating device such that the electromagnetic drive force is larger than the pressure difference force. When the electromagnetic drive force is larger than the pressure difference force, the valve member can be held seated on the valve seat with high stability. In a hydraulic pressure control apparatus according to the present invention the electric energy control means controls the electric energy to be applied to the electromagnetic force generating device such that the electromagnetic drive force is larger than the pressure difference force by at least a predetermined margin value. When the electromagnetic drive force Fs is larger than the pressure difference force Fp by at least the predetermined margin value Fr (where Fs > Fp + Fr), the valve member can be held seated on the valve seat with high stability, while assuring significantly reduction of the required amount of consumption of the electric energy. For instance, the margin value Fr may be determined on the basis of the maximum velocity of increase in the fluid pressure and a force required for forcing the valve member onto the valve seat, for example, as described below in detail. The margin value Fr may be either constant or variable. For instance, the margin value Fr may be a variable which is changed in proportion to the pressure difference force Fp. This arrangement is reasonable since the absolute values of detection errors and control errors of the pressure difference force and electromagnetic drive force increase as the pressure difference force and the electromagnetic drive force increase. In a hydraulic pressure control apparatus according to the present invention the margin value is determined by at least a maximum velocity of increase in the fluid pressure at the solenoid-operated pressure control device. When the fluid pressure presently controlled by the solenoid-operated pressure control valve is Pn, the fluid pressure Pn+1 a given time ΔT after the present time will not exceed a sum of the fluid pressure Pn and a maximum pressure increase amount ΔP = α·ΔT (an amount by which the fluid pressure is expected to be increased at the maximum pressure increasing velocity α). Namely, Pn+1 ≤ Pn + α·ΔT. Therefore, the valve member of the electromagnetic force generating device can be held seated on the valve seat by controlling or changing the electric energy to be applied to the electromagnetic force generating device with a cycle time equal to ΔT such that the electromagnetic drive force Fs is equal to or larger than the pressure difference force Fp(Pn + α·ΔT) corresponding to the fluid pressure (Pn + α·ΔT). in this case, the margin value Fr is equal to or larger than the pressure difference force Fp(ΔP) corresponding to the maximum pressure increase amount ΔP = α·ΔT). The maximum pressure increasing velocity α is determined by the maximum fluid pressure of a hydraulic pressure source, a flow resistance of the fluid through a fluid passage connecting the hydraulic pressure source and the brake cylinder, and a maximum cross sectional area of flow of the fluid through a solenoid-operated pressure increase control valve if provided in the fluid passage. The fluid flow resistance and the maximum cross sectional area of flow of the fluid through the pressure increase control valve are determined by the structural characteristics of the fluid passage and the pressure increase control valve. Where the hydraulic pressure source includes a pump, the maximum fluid pressure of the hydraulic pressure source is a maximum delivery pressure of the pump. Thus, the maximum pressure increasing velocity a of the solenoid-operated pressure control valve can be calculated based on the above-indicated factors. Further, the temperature of the working fluid may be taken into account in determining the maximum pressure increasing velocity. As the temperature of the fluid rises, the viscosity of the fluid is lowered, and the pressure increasing velocity is increased. The maximum pressure increasing velocity may be obtained by experimentation. Where the hydraulic pressure source includes a master cylinder operated by a brake operating member to generate a pressurized fluid whose pressure corresponds to an operating force acting on the brake operating member, on the other hand, the maximum pressure increasing velocity of the solenoid-operated pressure control valve is proportional to the velocity of increase in the fluid pressure in the master cylinder when the brake operating member is operated at an expected highest speed. As described above, the maximum pressure increasing velocity can be the maximum velocity at which the fluid pressure can be increased in a braking system including the present hydraulic pressure control apparatus. However, the maximum pressure increasing velocity can be the maximum velocity at which the fluid pressure can be increased at the present moment of control or in the present control cycle, and which can be obtained, for instance, by the cross sectional area of fluid flow through the pressure increase control valve and the fluid pressure in the hydraulic pressure source at the present moment of control. The maximum pressure increase amount ΔP determined by the maximum pressure increasing velocity a of the pressure control valve may vary with the control cycle time. If the maximum pressure increase amount ΔP is excessively large, the advantage of reducing the required amount of consumption of the electric energy is reduced. If the amount ΔP is excessively small, the valve member cannot be held seated on the valve seat with high stability. In view of this, the upper limit of the maximum pressure increase amount ΔP is desirably about 2%, 1% or 0.5% of the fluid pressure in the hydraulic pressure source, and the lower limit is desirably about 0.1%, 0.2% or 0.3% of the fluid pressure in the hydraulic pressure source. In a hydraulic pressure control apparatus according to a preferred embodiment of the present invention the margin value is determined by at least a force required for forcing the valve member onto the valve seat. The valve member of the electromagnetic force generating device is seated on the valve member when the electromagnetic drive force is just equal to the pressure difference force. In this state, it is theoretically true that the working fluid does not flow through the solenoid-operated pressure control valve. Actually, however, a minute local gap is generally present between the valve member and the valve seat due to a dimensional or configurational error of the contact surfaces of the valve member and the valve seat. As a result, the pressure control valve generally suffers from some amount of leakage of the working fluid therefrom. The amount of such a minute gap may be reduced, and the amount of the fluid leakage may be reduced, by causing elastic deformation of the valve member and the valve seat by forcing the valve member onto the valve seat. Accordingly, the amount of the fluid leakage can be reduced to within a practically tolerable range in which the accuracy of control of the fluid pressure in the brake cylinder is satisfactory, if the electromagnetic drive force Fs is larger than the pressure difference force Fp by a margin value which is equal to a force Fa required to cause elastic deformation of the valve member and the valve seat which is sufficient to reduce the fluid leakage amount to within the tolerable range. The required force Fa is determined depending upon the roundness values and other dimensional and configurational accuracies, surface smoothness (roughness), modulus of elasticity of the materials, and other factors of the valve member and the valve seat. Relationships between the operating stroke S and the fluid leak amount ε and electromagnetic drive force Fs (force Fa) can be obtained by calculation, or by experimentation, as shown in the graph of Fig. 25. According to these known relationships, it is possible to obtain the force Fa = Fs* - Fso required to force the valve member onto the valve seat for reducing the fluid leak amount to a tolerable value ε*. The tolerable value ε* may be a predetermined constant, or a variable which varies with the target value of the fluid pressure in the brake cylinder. This force Fa is an example of the force required for forcing the valve member onto the valve seat. In a further preferred embodiment of the present invention the margin value is determined by at least an amount of leak of the working fluid through a gap between contact surfaces of the valve member and the valve seat. As described above, the amount of leak of the working fluid through the gap between the contact surfaces of the valve member and the valve seat is reduced as the force for forcing the valve member onto the valve seat is increased. In controlling the fluid pressure in the brake cylinder, it is desirable to minimize the amount of leak of the fluid. However, the fluid leak amount need not be zeroed, provided the fluid leak amount is reduced to a tolerable value. As this tolerable value is increased, the force required for forcing the valve member onto the valve seat is reduced, and the required amount of consumption, of the electric energy is accordingly reduced. In a hydraulic pressure control apparatus according to a preferred embodiment of the present invention, the electric energy control controls the electric energy to be applied to the electromagnetic forge generating device, further depending upon a temperature of the working fluid flowing through the solenoid-operated seating valve. In the hydraulic pressure control apparatus according to the above embodiment wherein the temperature of the working fluid flowing through the solenoid-operated seating valve is taken into account in controlling the electric energy to be applied to the electromagnetic force generating device, the accuracy of control of the fluid pressure in the brake cylinder is less likely to be influenced by a change in the fluid temperature, than in an apparatus wherein the Fluid temperature is not taken into account. As described in detail with respect to the presently preferred embodiments of this invention, the electric energy in the form of an electric current i is determined according to an equation (8) (which will be described), on the basis of a density ρ, a compression coefficient (volume change coefficient) β and a flow rate coefficient C of the fluid, a pressure difference ΔP across the seating valve, and a rate of change dP/dt of the target value of the fluid pressure in the brake cylinder. The pressure difference ΔP is a difference between the fluid pressure in the brake cylinder and the fluid pressure in the low-pressure source (reservoir). Since the fluid pressure in the reservoir is almost equal to the atmospheric pressure, the pressure difference ΔP can be considered to be substantially equal to the fluid pressure in the brake cylinder. The density ρ and compression coefficient β of the working fluid are determined by the temperature of the fluid. Therefore, the electric current i determined according to the equation (8) reflects the temperature of the fluid. In a preferred embodiment the controller further includes slowly energy increasing means for slowly increasing the electric energy to be applied to said electromagnetic force generating device, when the valve member of the seating valve is seated on the valve seat. When the valve member is required to be seated onto the valve seat, the electric energy to be applied to the electromagnetic force generating device may be relatively rapidly increased so that the electromagnetic drive force is rapidly increased to a value sufficient for seating the valve member on the valve seat with high stability. In the hydraulic pressure control apparatus according to the above embodiment, the electric energy is slowly increased so that the velocity at which the valve member is seated onto the valve seat is made low enough to reduce an impact and a noise of seating of the valve member against the valve seat. The present arrangement has a further advantage of increasing the durability of the seating valve of the solenoid-operated pressure control valve. Further, the required amount of consumption of the electric energy by the pressure control valve is reduced. The slowly energy increasing means which provides the advantages indicated above may be considered to be means for reducing the velocity of seating of the valve member onto the valve seat, means for reducing the noise of seating of the valve member onto the valve seat, means for reducing the impact of seating of the valve member against the valve seat, and means for assuring slow or soft seating of the valve member onto the valve seat. In the solenoid-operated pressure control valve included in the present hydraulic pressure control apparatus, the pressure difference force acts on the valve member in the direction for moving the valve member away from the valve seat, while the electromagnetic drive force generated by the electromagnetic force generating device acts on the valve member in the direction for moving the valve member toward the valve seat. To cause the valve member to be seated on the valve seat, the electromagnetic drive force must be larger than the pressure difference force. However, an excessively large amount of the electromagnetic drive force causes problems such as an excessively high velocity of seating of the valve member onto the valve seat. These problems may be solved by controlling the electric energy so that the electromagnetic drive force is only slightly larger than the pressure difference force. However, the control of the electric energy to accurately control the electromagnetic drive force as described above is difficult. If the electromagnetic drive force is smaller than the pressure difference force, the valve member cannot be seated on the valve seat. According to the above embodiment, the electric energy is slowly increased to slowly increase the electromagnetic drive force. This arrangement permits a higher degree of stability of seating of the valve onto the valve seat, while avoiding an excessively high velocity of seating of the valve member on the valve seat. Where the pressure difference force is zero, too, the slow increase of the electromagnetic drive force is effective to avoid the excessively high seating velocity of the valve member. The slowly energy increasing means provided in the above embodiment is activated when the valve member which is spaced apart from the valve member is required to be seated onto the valve seat. When it is required to activate the brake cylinder for providing a braking effect while the fluid pressure in the brake cylinder is not being controlled, (that is, while the solenoid-pressure pressure reduction control valve is in the open state, for instance) this pressure reduction control valve must be controlled to control the fluid pressure in the brake cylinder. For example, the slowly energy increasing means is activated when the brake cylinder pressure is increased while the brake cylinder pressure is not being controlled, that is, while the brake cylinder is not in operation, or when the brake cylinder pressure is increased after termination of reduction of the brake cylinder pressure, or when the brake cylinder pressure is held constant after termination of reduction of the brake cylinder pressure. The slowly energy increasing means may be considered to include means for slowing increasing the electric energy when the brake cylinder is activated to initiate a braking operation, means for slowing increasing the electric energy immediately after initiation of an operation to increase the brake cylinder pressure, and means for slowing increasing the electric energy immediately after termination of an operation to reduce the brake cylinder pressure. The slowly energy increasing means may be adapted to slowly increase the electric energy in a predetermined fixed pattern, or in a pattern which varies depending upon the fluid pressure in the brake cylinder, for instance, in a pattern described below with respect to the following preferred embodiments of this invention. The rate of increase of the electric energy by the slowly energy increasing means may be held constant throughout the operation of this means, or may be changed as needed. The pattern in which the electric energy is slowly increased immediately after initiation of an operation to increase the brake cylinder pressure may be identical with or different from the pattern in which the electric energy is slowly increased immediately after termination of an operation to reduce the brake cylinder pressure. in the latter case, the rate of increase of the electric energy is desirably made higher when the electric energy is slowly increased immediately after initiation of the operation to increase the brake cylinder pressure. In a further preferred embodiment the slowly energy increasing means increases the electric energy to be applied to the electromagnetic force generating device, at a rate corresponding to a maximum velocity of increase in the fluid pressure in the brake cylinder. The maximum velocity of increase of the fluid pressure in the brake cylinder is strictly a maximum velocity of increase of the fluid pressure at the solenoid-operated pressure control valve. By increasing the electric energy at a rate corresponding to the maximum velocity of increase of the fluid pressure in the brake cylinder, the valve member can slowly seated onto the valve seat while avoiding a delay in the operation to increase the brake cylinder pressure. The electric energy may be increased at a rate corresponding to a sum of the maximum pressure increase velocity and a suitable margin value. In another further preferred embodiment the slowly energy increasing means increases the electric energy to be applied to said electromagnetic force generating device, on the basis of the fluid pressure in said brake cylinder and a deceleration coefficient. The deceleration coefficient γ is a coefficient for reducing the velocity of seating of the valve member onto the valve seat. For instance, the deceleration coefficient γ may be a coefficient such as (1 - 1/2 t ), which increases from "0" to "1" as the time passes. As described below with respect to the presently preferred embodiments of the invention, the electric energy applied to the electromagnetic force generating device is controlled so that the generated electromagnetic drive force Fs is equal to Fs' which is a product of the electromagnetic drive force generated in the apparatus according to any one of the above embodiments and the deceleration coefficient γ or a reciprocal (1/γ) of this deceleration coefficient γ. Namely, Fs' = Fs · (1 - 1/2 t ), or Fs' = Fs/(1-1/2 t ). Alternatively, the electric energy is controlled so that the generated electromagnetic drive force Fs is equal to Fs' which is a product of the pressure difference force Fp(Pref) corresponding to the target brake cylinder pressure Pref and the deceleration coefficient γ or a reciprocal (1/γ) of this deceleration coefficient γ. Namely, Fs' = Fp (Pref)·(1 - 1/2 t ), or Fs' = Fp(Pref)/(1 - 1/2 t ). Where the generated electromagnetic drive force Fs' is equal to the product of the value Fs and the reciprocal (1/γ) of the deceleration coefficient γ, the force Fs' is relatively large with respect to the force Fs or pressure difference force Fp(Pref) immediately after initiation of the control of the electric energy, the force Fs' approaches the force Fs or Fp(Pref) as the time passes. The time t may be expressed by the following equation: t = T c + ΔTn, In the above, equation, "T c ", "ΔT" and "n" respectively represent the control initiating time, the control cycle time and the number of control cycles. A preferred embodiment further comprises pressure increasing means for increasing the fluid pressure in the brake cylinder when a predetermined condition for effecting an increase in the fluid pressure in the brake cylinder is satisfied, and wherein the slowly energy increasing means includes preliminary energy applying means for applying the electric energy to the electromagnetic force generating device before the condition is satisfied upon detection of a sign indicative of a high possibility that said condition will be satisfied. The slow increase of the electric energy applied to the electromagnetic force generating device may be initiated when the predetermined condition for effecting an increase in the brake cylinder pressure is satisfied. in this case, however, there is a risk that some amount of the working fluid is discharged from the brake cylinder to the low-pressure source through the solenoid-operated pressure control valve (seating valve) which has not been fully closed. Accordingly, the increase in the brake cylinder pressure may be delayed. According to the above mode (34) of this invention, the application of the slowly increased electric energy to the electromagnetic force generating device of the pressure control valve is initiated before the above-indicated condition is satisfied, more precisely, when a sign indicative of a high possibility that the condition will be satisfied is detected. This arrangement reduces or eliminates the delayed increase of the brake cylinder pressure, improving a response of the pressure control valve to a command for initiating the increase of the brake cylinder pressure. Further, the present arrangement permits a slower increase of the electric energy than when the application of the electric energy is initiated when the predetermined condition is satisfied. Accordingly, the velocity at which the valve member is seated onto the valve seat can be lowered, so that the operating noise of the pressure control valve is reduced, and the durability of the valve is improved. The sign indicative of a high possibility that the predetermined condition for effecting an increase in the brake cylinder pressure is satisfied in the near future is detected by a suitable detecting means. Where the predetermined condition is an operation of a brake operating member, this detecting means may be adapted to detect releasing of the brake operating member. Where the predetermined condition is satisfied when a predetermined condition for initiating a traction control or vehicle stability brake control is satisfied, the detecting means may be adapted to detect a predetermined condition which is satisfied before the condition for initiating the traction control or vehicle stability brake control is satisfied. The detecting means may be adapted to detect a sign indicative of a high possibility that the pressure control mode will be changed from the pressure reduction mode to the pressure hold or increase mode during a normal braking pressure control, an anti-lock braking pressure control, a traction control or a vehicle stability brake control. Means for effecting the normal braking pressure control may be adapted to control the fluid pressure in the brake cylinder so that a braking effect provided by the brake cylinder corresponds to an amount of operation of the brake operating member. Where the hydraulic pressure control apparatus is used for an automotive vehicle, the brake cylinder pressure may be controlled so that the deceleration of the vehicle obtained by activation of the brake cylinder meets the vehicle operator's intention as represented by the operating amount of the brake operating member. The preliminary energy applying means need not be adapted to complete the seating of the valve member onto the valve seat before the predetermined condition for effecting the brake cylinder pressure increase is satisfied. In other words, the preliminary energy applying means may be adapted to cause at least a movement of the valve member toward the valve seat, so that the valve member can be seated onto the valve seat at a relatively low velocity in a short time after the predetermined condition is satisfied. In this case, too, the delay in the increase of the brake cylinder pressure can be reduced. Where the application of the electric energy to the electromagnetic force generating device is initiated while the brake cylinder pressure is almost equal to the atmospheric pressure, the pressure difference force is zero, so that the valve member can be moved and seated on the valve seat by a relatively small electromagnetic drive force, even when a biasing force acts on the valve member in the direction opposite to the direction in which the electromagnetic drive force acts on the valve member. It will be understood from the above explanation that the preliminary energy applying means functions as means for reducing a delayed increase of the brake cylinder pressure. In a yet preferred embodiment the electric energy control means includes energy application terminating means for termination the application of the electric energy to the electromagnetic force generating device by the preliminary energy applying means when a predetermined time has passed after initiation of the application without satisfaction of the predetermined condition. Where the application of the electric energy is initiated while the brake cylinder is not activated, it is desirable to terminate the application of the electric energy when the predetermined time has passed without satisfaction of the predetermined condition, that is, without initiation of an operation to increase the brake cylinder pressure. Continued application of the electric energy for a long time results in the solenoid-operated pressure control valve (e.g., pressure reduction control valve) being held in its closed state for the long time. Although the continued closure of the pressure control valve per se does not cause a problem, the continued application of the electric energy to the electromagnetic force generating device causes a rise of the temperature of the components of the pressure control valve and the temperature of the working fluid, resulting in an increase in the fluid pressure in the brake cylinder and a risk of causing a brake drag, and further resulting in unnecessary consumption of the electric energy. In this respect, it is desirable to terminate the application of the electric energy for a long time without an initiation of an operation to increase the brake cylinder pressure. The energy application terminating means meets this desire. The predetermined time indicated above may be considered a time during which the electric energy is applied to the electromagnetic force generating device by the preliminary energy applying means. Preferrably the pressure increasing means includes a hydraulic pressure source, and a solenoid-operated pressure increase control valve disposed between the hydraulic pressure source and the brake cylinder, and wherein the controller further includes means for controlling an electric energy to be applied to the solenoid-operated pressure increase control valve, to increase the fluid pressure in the brake cylinder. Yet preferrably the slowly energy increasing means includes means for initiating the application of the electric energy to the electromagnetic force generating device when a predetermined condition for terminating an operation to reduce the fluid pressure in the brake cylinder. By initiating the application of the electric energy to the electromagnetic force generating device when the predetermined condition for terminating an operation to reduce the brake cylinder pressure is satisfied, it is possible to reduce a delay in a subsequent operation to increase the brake cylinder pressure. The predetermined condition for terminating an operation to reduce the brake cylinder pressure may be satisfied during a normal braking pressure control operation, an anti-lock braking pressure control operation, a traction control operation or a vehicle stability brake control operation, or when a vehicle accelerating member is released. Where the pressure control valve is a normally open pressure reduction control valve, this control valve is normally placed in a state for reducing the fluid pressure in the brake cylinder. Therefore, the predetermined condition for terminating an operation to reduce the brake cylinder pressure may be a condition in which it is no longer necessary to hold the pressure control valve in the open state, namely,a condition in which there is a high possibility that the pressure control valve will be switched to the closed state in the near future. BRIEF DESCRIPTION OF THE DRAWINGS The above and optional objects, features, advantages and technical and industrial significance of this invention will be further clarified by reading the following detailed description of presently preferred embodiments of the invention, when considered in connection with the accompanying drawings, in which: Fig. 1 is a schematic view illustrating a braking system including a hydraulic pressure control apparatus according to an embodiment of this invention; Fig. 2 is an elevational view in cross section of a front wheel linear valve device included in the braking system of Fig. 1; Fig. 3 is an elevational view in cross section of a rear wheel linear valve device included in the braking system of Fig. 1; 4 Fig. 4A is a view schematically illustrating forces acting on a pressure increase control valve and a pressure reduction control valve which are included in the front wheel linear valve device of Fig. 2; Fig. 4B is a view schematically illustrating forces acting on a pressure reduction control valve included in the rear wheel linear valve device of Fig. 3; Fig. 5 is a view showing an example of control modes established by the linear valve devices of the hydraulic pressure control apparatus of Fig. 1; Fig. 6 is a graph indicating relationships between operating stroke and fluid leak amount and electromagnetic force of the control valve of the linear valve device; Fig. 7 is a graph showing an example of control of electric current applied to the pressure reduction control valve, for slow seating of a valve member onto a valve seat, immediately after termination of a pressure reducing operation in the hydraulic pressure control apparatus of Fig. 1; Fig. 8 is a graph showing an example of preliminary current application to the pressure reduction control valve by the hydraulic pressure control apparatus; Fig. 9 is a flow chart illustrating a linear valve device control routine executed according to a program stored in ROM of the hydraulic pressure control apparatus; Fig. 10 is a flow chart illustrating a linear valve device control in step S140 of the flow chart of Fig. 9; Fig. 11 is a view showing an example of control of the linear valve device by the hydraulic pressure control apparatus of Fig. 1; Fig. 12 is a flow chart illustrating a linear valve device control routine executed according to a program stored in ROM of a hydraulic pressure control apparatus according to yet another embodiment of this invention; Fig. 13 is a graph showing an example of control of the electric current applied to the pressure reduction control valve immediately after initiation of a pressure increasing operation in the hydraulic pressure control apparatus of Fig. 1; Fig. 14 is a graph showing examples of control of the electric current immediately after initiation of the pressure increasing operation in still further embodiments of the invention; Fig. 15 is a view showing other examples of control of the electric current according to still further embodiments of the invention; Fig. 16 is a view showing a data map representing a compensation coefficient for controlling the electric current applied to the pressure reduction control valve, according to a yet further embodiment of the invention, which is stored in ROM of the hydraulic pressure control apparatus; and Fig. 17 is a schematic view illustrating a braking system including a hydraulic pressure control apparatus, according a yet further embodiment of the invention. DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS Referring first to Fig. 1, there is shown a hydraulically operated braking system for an automotive vehicle, which includes a brake operating member in the form of a brake pedal 10, and a master cylinder 12. The master cylinder 12 is of a tandem type having two pressurizing chambers, one of which is connected through a fluid passage 14 to a brake cylinder 18 for a front left wheel 16 of the vehicle, and the other of which is connected through a fluid passage 20 to a brake cylinder 24 for a front right wheel 22 of the vehicle. Thus, the two front wheel brake cylinders 18, 24 are connected to the respective two pressurizing chambers of the master cylinder 12. Solenoid-operated master cylinder cut-off valves 26, 28 are connected to the respective two fluid passages 14, 20. Each of these master cylinder cut-off valves 26, 28 is switchable between a closed state and an open state by energization and de-energization of its solenoid coil. When the pressure of a working fluid in the corresponding wheel brake cylinder 18, 24 is controlled, the master cylinder cut-off valve 26, 28 is held in its closed state with an electric current being applied to the solenoid coil. When the fluid pressure is not controlled, the master cylinder cut-off valve 26, 28 is held in its open state without an electric current being applied to the solenoid coil. Thus, each cut-off valve 26, 28 is a normally open valve, which is brought to the open state in the event of occurrence of any electrical abnormality. The present braking system further includes a pump device 30, which includes a low-pressure source in the form of a master reservoir 31, two pumps 32, 34, and two check valves 35, 36. As shown in Fig. 1, the two pumps 32, 34 are disposed in parallel connection with each other. The pump 32 is a high-pressure small-capacity gear pump (hereinafter referred to as "high-pressure pump" where appropriate), while the pump 34 is a low-pressure large-capacity gear pump (hereinafter referred to as "low-pressure pump" where appropriate). The high-pressure pump 32 has a relatively higher maximum delivery pressure and a relatively small maximum delivery amount per unit time, while the low-pressure pump 34 has a relatively lower maximum delivery pressure and a relatively larger maximum delivery amount per unit time. The high-pressure pump 32 is operated by a high-pressure pump motor 38, while the low-pressure pump 34 is operated by a low-pressure pump motor 40. The check valve 35 is provided to prevent the delivery pressure of the low-pressure pump 34 from acting on the high-pressure pump 32, while the check valve 36 is provided to prevent the delivery pressure of the high-pressure pump 34 from acting on the low-pressure pump 34. Further, a pressure relief valve (not shown) is connected in parallel with each of the high-pressure and low-pressure pumps 32, 34, for limiting the delivery pressure of the pump 32, 34 to the predetermined maximum level. The pressure and flow rate of the working fluid delivered from each pump 32, 34 (delivery pressure and rate of the pump 32, 34) can be adjusted by controlling the corresponding pump motor 38, 40. Both of the high-pressure and low-pressure pumps 32, 34 need not be gear pumps. That is, at least one of these pumps 32, 34 may be a plunger pump. If a solenoid-operated shut-off valve is provided in place of the check valve 36, the fluid pressure in the wheel brake cylinders 18, 24 may be controlled by operating the low-pressure pump 34 in the reverse direction. A pump passage 42 is connected to the delivery side of the pump device 30. The pump passage 42 is split into four branch passages 44 which are respectively connected at their ends to the above-indicated brake cylinders 18, 24 for the front wheels 16, 22, and brake cylinders 50, 52 for rear left and right wheels 46, 48. Thus, the four wheel brake cylinders 18, 24, 50, 52 are connected to the pump device 30 such that the wheel brake cylinders 18, 24, 50, 52 are connected in parallel with each other. The braking system includes a solenoid-operated pressure increase control valve 179 in each of the four branch passages 44 which connects the pump device 30 to the respective wheel brake cylinders 18, 24, 50, 52. Each solenoid-operated pressure increase control valve 179 is a normally closed valve, which is placed in the closed state when an electric current is not applied thereto. The pressure increase control valves 179 and the pressure reduction control valves 58, 62 are linear solenoid valves. The pressure increase control valves 179 and the pressure reduction control valves 58 for the front wheel brake cylinders 18, 24 constitute a front wheel linear valve device 184, while the pressure increase control valves 179 and the pressure reduction control valves 62 for the rear wheel brake cylinders 50, 52 constitute a rear wheel linear valve device 186, as generally indicated in Fig. 1. If the pressure reduction control valves 62 for the rear wheel brake cylinders 50, 52 were normally closed valves, these valves 62 would have to be held open with an electric current applied thereto at the end of braking actions of the rear wheel brake cylinders 50, 52, for a predetermined period of time necessary for the fluid to be discharged from the rear wheel brake cylinders. In this case, the pressure reduction control valves 62 are returned to the original closed state after the predetermined period of time has passed. However, this arrangement may cause some amount of the pressurized fluid left in the rear wheel brake cylinders 50, 52, causing a so-called "brake drag". In the present embodiment, the normally open pressure reduction control valves 62 are returned to their original open state with their coils being de-energized at the end of their braking actions , permitting the pressurized fluid to be completely discharged from the rear wheel brake cylinders 50, 52, thereby avoiding the otherwise possible brake drag. The master cylinder cut-off valves 26, 28 are also normally open valves, so that the front wheel brake cylinders 18, 24 are held in communication with the master cylinder 12, for fail-safe purpose, while the coils of those cut-off valves 26, 28 are in the de-energized state (while the front wheel brake cylinders 26, 28 are not activated, or while the braking system has an electrical abnormality) . The normally open master cylinder cut-off valves 26, 28 permit the pressurized fluid to be discharged from the front wheel brake cylinders 18, 24 and returned to the master cylinder 12, without flowing through the pressure reduction control valves 58. Therefore, the pressure reduction control valves 58 need not be normally open valves, to prevent the brake drag in connection with the front wheel brake cylinders 18, 24. If an electrical abnormality of the braking system takes place during an operation of the braking system, the master cylinder cut-off valves 26, 28 are returned to the original open state, while the pressure increase control valves 179 are returned to the closed state. As a result, the front wheel brake cylinders 18, 24 are disconnected from the pump device 30, and are connected to the master cylinder 12, so that the front wheel brake cylinders 13, 24 can be activated by the pressurized fluid delivered from the master cylinder 12. If the pressure reduction control valves 58 for the front wheel brake cylinders 18, 24 were normally open valves, these control valves 58 could not be returned to the original normally open state in the event of occurrence of an electrical abnormality, and the pressurized fluid would be discharged from the front wheel brake cylinders 18, 24 and returned to the master reservoir 31 through the open pressure reduction control valves 58. In the present embodiment, the normally closed pressure reduction control valves 58 permit activation of the front wheel brake cylinders 18, 24 even in the event of an electrical abnormality. Thus, the normally closed pressure reduction control valves 58 for the front wheel brake cylinders 13, 24 and the normally open pressure reduction control valves 62 for the rear wheel brake cylinders 50, 52 permit activation of the wheel brake cylinders 18, 24, 50, 52 without any brake drag, even in the event of occurrence of an electrical abnormality in the braking system. Referring Figs. 2 and 3, there are schematically shown the front wheel linear valve device 184 and the rear . wheel linear valve device 186. In the front wheel linear valve device 184 of Fig. 2, the pressure reduction control valve 58 and the pressure increase control valve 179 have the following construction. Each of the solenoid-operated pressure reduction control valves for the front wheel brake cylinders 18, 24 includes a seating valve 82 and an electromagnetic force generating device 84. The seating valve 32 includes a valve member 90, a valve seat 92, an electromagnetically movable member 94 movable with the valve member 90, and an elastic member in the form of a spring 96 which serves as a biasing device for biasing the electromagnetically movable member 94 in a direction for moving the valve member 90 to be seated on the valve seat 92. The electromagnetic force generating device 34 includes a coil 100, a holder member 102 made of a resin for holding the coil 100, a first magnetic path forming body 104 and a second magnetic path forming body 106. By energizing the coil 100, there is produced a magnetic field. A substantive portion of the magnetic flux in the magnetic field passes through the first magnetic path forming body 104, the electromagnetically movable member 94, and an air gap between the movable member 94 and the second magnetic path forming body 106. A magnetic force which acts between the electromagnetically movable member 94 and the second magnetic path forming body 106 can be changed by controlling the electric current to be applied to the coil 100. This magnetic force increases with an increase in the amount of electric current to be applied to the coil 100. Since the electric current amount and the magnetic force generated has a known relationship, the force for biasing the electromagnetically movable member4 94 can be changed by continuously changing the electric current to be applied to the coil 100, according to the known relationship. This force (which will be referred to as "electromagnetic drive force Fs") acts on the electromagnetically movable member 94 in the direction away from the valve member 90, that is, in the direction for moving the valve member 90 away from the valve seat 32. This direction in which the electromagnetic drive force Fs acts on the movable member 94 is opposite to a direction in which a biasing force Fk of the spring 96 acts on the movable member 94. As indicated above, the electromagnetic drive force Fs and the biasing force Fk of the spring 96 act on the valve member 90 through the electromagnetically movable member 94 in the pressure reduction control valve 58. In addition, a force Fp based on a difference between fluid pressures in an input port 103 and an output port 108 of the seating valve 82 acts on the valve member 90 in the direction for moving the valve member 90 away from the valve seat 92, that is, in the direction in which the electromagnetic drive force Fs acts on the valve member 90. when a sum of the electromagnetic drive force Fs and the force Fp (hereinafter referred to as "pressure difference force") is larger than the biasing force Fk of the spring 96 (where Fs + Fp > Fk), the pressure reduction control valve 58 (seating valve 82) is held in the open state with the valve member 90 being spaced apart from the valve seat 92. When the sum of the electromagnetic drive force Fs and the pressure difference force Fp is smaller than the biasing force Fk (where Fs + Fp < Fk), the pressure reduction control valve 58 is held in the closed state with the valve member 90 seated on the valve seat 92. Since the biasing force Fk of the spring 96 is determined to be larger than the pressure difference force Fp corresponding to the maximum delivery pressure of the pump device 30, the pressure reduction control valve 58 is held in the closed state when the electromagnetic drive force Fs is zero. That is, the pressure reduction control valve 58 is a normally closed valve, as indicated above. In the present braking system, the electric current to be applied to the coil 100 of the pressure reduction control valve 58 is continuously controlled to regulate the fluid pressure in the corresponding front wheel brake cylinder 18, 24. The difference between the fluid pressures in the input and output ports 108, 109 of the pressure reduction control valve 58 corresponds to a difference between the fluid pressure in the wheel brake cylinder 18, 24 and the fluid pressure in the master reservoir 31. Since the fluid pressure in the master reservoir 31 can be considered to be equal to the atmospheric pressure, the pressure difference force Fp corresponds to the fluid pressure in the wheel brake cylinder 18, 24. The sum of the pressure difference force Fp and the electromagnetic drive force Fs can be controlled by controlling the electromagnetic drive force Fs. Therefore, the fluid pressure in the wheel brake cylinder 18, 24 can be controlled by controlling the electromagnetic drive force Fs. The fluid pressure in the wheel brake cylinder 18, 24 can be reduced by increasing the electromagnetic drive force Fs. Described more precisely, the electric current to be applied to the coil 100 is controlled so that the actual fluid pressure in the wheel brake cylinder 18, 24 changes toward a determined target value. The seating valve 82 of the pressure increase control valve 179 has a port 180 connected to the front wheel brake cylinders 18, 24 and the pressure reduction control valve 58, and a port 182 which is closed by seating of the valve member 90 on the valve seat 92. The port 182 is connected to the pump device 30. In the rear wheel linear valve device 186 of Fig. 3, the pressure reduction control valve 62 has the following construction. Each of the solenoid-operated pressure reduction control valves for the rear wheel brake cylinders 50, 52 includes a seating valve 130 and an electromagnetic force generating device 132, like the pressure reduction control valves 58. The seating valve 130 includes a valve seat 134, a valve member 136 movable to be seated on and unseated from the valve seat 134, an elastic member in the form of a spring 138 serving as a biasing device for biasing the valve member 136 in a direction away from the valve seat 134, a drive member 140 for moving the valve member 136, and an electrically movable member 142. The drive member 140, valve member 136 and electromagnetically movable member 142 are movable as a unit. The electromagnetic force generating device 132 includes a coil 144, a holder member 145 for holding the coil 144, a first magnetic path forming body 146, and a second magnetic path forming body 148 fixed to the housing of the control valve 62. When the coil 144 is energized, a magnetic field is produced with a magnetic flux passing through the first magnetic path forming body 146, the electromagnetically movable member 142, and an air gap between the movable member 142 and the second magnetic path forming body 148. As a result, an electromagnetic drive force Fs acts on the electromagnetically movable member 142 in a direction for moving the movable member 142 toward the second magnetic path forming body 143, so that the movable member 142 is moved to move the valve member 136 toward the valve seat 134. When the electromagnetic drive force Fs is zeroed, the valve member 136 is moved away from the valve seat 134 by a biasing force Fk of the spring 138. The pressure reduction control valve 62 is held in its open state when the electromagnetic drive force Fs is smaller than a sum of the pressure difference force Fp and the biasing force Fk (where Fs < Fp + Fk). Since the biasing force Fk is very small, it can be ignored. When the electromagnetic drive force Fs is zero, the control valve 62 is held in the open state. That is, the control valve 62 is a normally open valve. When the electromagnetic drive force Fs is larger than the sum of the pressure difference force Fp and the biasing force Fk (where Fs > Fp + Fk), the control valve 62 is held in the closed state. The fluid pressure in the wheel brake cylinder 50, 52 can be reduced by reducing the electromagnetic drive force Fs. In the present embodiment, the electric current to be applied to the coil 144 is controlled so that the actual fluid pressure in the wheel brake cylinder 50, 52 changes toward a determined target value. The pressure increase control valve 179 of the valve device 186 is identical in construction with the pressure increase control valve 179 of the front wheel valve device 184. As shown in Fig. 4A, the valve member 90 of the pressure increase control valve 179 receives a biasing force Fk of the spring 96, a pressure difference force Fp based on a pressure difference across the seating valve 82, and an electromagnetic drive force Fs. The pressure difference across the seating valve 82 is detected as a difference between the output or delivery pressure of the pump device 30 and the fluid pressure in the wheel brake cylinder 18, 24, 50, 52. When a sum of the pressure difference force Fp and the electromagnetic force Fs has become larger than the biasing force Fk of the spring 96, the valve member 90 is moved away from the valve seat 92. When the electromagnetic force Fs is zero, the valve member 90 is moved away from the valve seat 92 when the pressure difference force Pp has become larger than the biasing force Fk. The pressure difference across the seating valve 82 necessary to open the pressure increase control valve 178 when the electromagnetic force Fs is zero is larger than 18MPa (about 134kg/cm 2 ) which is the maximum delivery pressure of the pump 30. Therefore, the pressure increase control valve 178 will not be opened without energizing the solenoid coil 100. in the pressure reduction control valve 58, too, the valve member 90 receives the biasing force Fk, pressure difference force Fp and electromagnetic drive force Fs, as also indicated Fig. 23A. The opening pressure difference of this pressure reduction control valve 58 is also set to be larger than 18MPa, so that the valve 58 will not be opened when the electromagnetic drive force Fs is zero, even if the fluid pressure in the front wheel brake cylinder 18, 24 is increased to the maximum delivery pressure of the pump device 30. That is, the fluid will not be discharged from the front wheel brake cylinder 18, 24 to the master reservoir 31, without energizing the coil 100 of the pressure reduction control valve 58. In the pressure reduction control valve 62 of the rear wheel linear valve device 186 of Fig. 3, the valve member 140 receives a biasing force Fk of the spring 138, a pressure difference force Fp based on the pressure difference across the seating valve 130, and an electromagnetic force Fs, as shown in Fig. 4B. However, the direction in which the biasing force Fk and the electromagnetic drive force Fs act on the valve member 130 are opposite to the directions in which the biasing force Fk and the electromagnetic drive force Fs act on the valve member 90 of the pressure reduction control valve 58 of the front wheel linear valve device 184 of Fig. 2. When the electromagnetic force Fs is smaller than a sum of the biasing force Fk and the pressure difference force Fp, the pressure reduction control valve 62 is held open. When the electromagnetic drive force Fs has become larger than the sum of the biasing force Fk and the pressure difference force Fp, the valve 62 is closed. Since the biasing force Fk of the spring 138 is considerably small, it may be ignored. Since the pressure difference force Fp acts on the valve member 136 in the direction for moving the valve member 136 away from the valve seat 134, the valve member 136 is held in the open state when the electromagnetic drive force Fs is zero, even if the biasing force Fk is considerably small. An operation of the present braking system will be described. When the brake pedal 10 is operated, the control device 160 executes a normal braking pressure control routine, wherein the pump device 30 is operated, and the master cylinder cut-off valves 26, 23 are brought to the closed state, and the front and rear wheel linear valve devices 184, 186 are controlled so that the fluid pressure pressurized by the pump device 30 is controlled by the linear valve devices 184, 186 and applied to the wheel brake cylinders 18, 24, 50, 52. In the normal braking pressure control routine, a target value of the fluid pressure in the wheel brake cylinders 18, 24, 50, 52 is obtained depending upon a depression force acting on the brake pedal 10, and the actual value of the fluid pressure in the wheel brake cylinders is controlled so as to reduce a difference between the target and actual values. Namely, the fluid pressure in the wheel brake cylinders 19, 24, 50, 52 is controlled by the linear valve devices 184, 186 so that the braking effect provided by the wheel brake cylinders coincides with a target value as represented by the operating amount of the brake pedal 10. The linear valve devices 184, 186 are controlled according to a linear valve device control routine which will be described. The linear valve device control routine is formulated to select the pressure increase mode when a difference ΔP between the target wheel brake cylinder pressure obtained in the normal braking pressure control routine and the actual wheel brake cylinder pressure is larger than a predetermined threshold EPS. That is, the pressure increase mode is selected when the target wheel brake cylinder pressure is higher than the actual wheel brake cylinder pressure by more than the threshold EPS. When the pressure difference ΔP is smaller than a predetermined negative threshold -EPS (when the pressure difference ΔP is larger than the absolute value of the threshold -EPS), the pressure reduction mode is selected. In the other case, the pressure hold mode is selected. When the pressure control mode is changed from the pressure reduction mode to the pressure hold mode, the valve member 136 of the seating valve 130 of the pressure reduction control valve 62 is slowly seated on the valve seat 134. When the pressure control mode is changed from the pressure increase mode to the pressure hold mode, the valve member 90 of the seating valve 82 of the pressure increase control valve 179 is slowly seated on the valve seat 92. The pressure control modes will be described in detail. When the amount of slipping of at least one wheel of the vehicle has become excessive with respect to the friction coefficient of the road surface during brake application to the vehicle, that is, a predetermined anti-lock braking pressure control initiating condition is satisfied, an anti-lock braking pressure control is initiated. In the anti-lock braking pressure control, the master cylinder cut-off valves 26, 28 are switched to the closed state, and the front and rear wheel linear valve devices 184, 186 are controlled to control the fluid pressures in the wheel brake cylinders 13, 24, 50, 52 independently of each other, so that the amounts of slipping of the wheels are held within an optimum range. As in the normal braking pressure control, the target fluid pressure in each wheel brake cylinder is determined in the anti-lock braking pressure control. When the amounts of slipping of the rear drive wheels 46, 48 have become excessive with respect to the road surface friction coefficient, that is, a predetermined traction control initiating condition is satisfied, a traction control of the rear drive wheels 46, 48 is initiated. Described more specifically, the traction control initiating condition is satisfied when the rotating speed of the rear drive wheels 46, 48 has become higher than a traction control initiating upper limit VTB, which is higher by a predetermined amount than the vehicle running speed which is estimated on the basis of the rotating speed of the front driven wheels 16, 22. In the traction control, the fluid pressure in the rear drive wheels 50, 52 is controlled in the same manner by the rear wheel linear valve devices 186. Since the traction control is effected to the rear wheel brake cylinders 50, 52, the master cylinder cut-off valves 26, 28 provided for the front wheel brake cylinders 18, 24 are held in the open state. Accordingly, when the brake pedal 10 is depressed during the traction control, the fluid pressurized by the master cylinder 12 is immediately supplied to the front wheel brake cylinders 18, 24, so that the front wheel brake cylinders are activated without a delay. In the traction control, the target fluid pressure in the rear wheel brake cylinders 50, 52 is obtained. While the present sixth embodiment is adapted such that the delivery pressure of the pump device 30 is kept at the same value irrespective of whether the braking system is activated or not, the delivery pressure when the braking system is not activated may be made lower than when the braking system is activated. When the amount, of spinning or drift-out motion of the vehicle during turning has exceeded a predetermined threshold SVs or SVd, the control device 160 initiates a spinning control routine or a drift-out control routine. The amount of spinning or drift-out motion of the vehicle is obtained on the basis of the output signals of the steering angle sensor 170 and the yaw rate sensor 172. The spinning control routine or drift-out control routine is formulated to control the front or rear wheel linear valve device 184, 186 so that the appropriate wheel or wheels is/are braked so as to give the turning vehicle a yaw moment which is effective to reduce the amount of spinning or drift-out motion. In the spinning or drift-out control routine, the target fluid pressure of the appropriate wheel brake cylinder or cylinders is determined to brake the appropriate wheel or wheels so as to give the desired yaw moment to the vehicle. In the event of occurrence of an electrical abnormality, the master cylinder cut-off valves 26, 28 are returned to the open state, the pressure increase control valves 179 and the pressure reduction control valves 58 are returned to the closed state, while the pressure reduction control valves 62 are returned to the open states. As a result, the front wheel brake cylinders 18, 24 are disconnected from the pump device 30 and are connected to the master cylinder 12. With the pressure reduction control valves 58 placed in the closed state, the front wheel brake cylinders 18, 24 can be actuated by the pressurized fluid delivered from the master cylinder 12. On the other hand, the rear wheel brake cylinders 50, 52 are disconnected from both the pump device 30 and the front wheel brake cylinders 18, 24, and are connected to the master reservoir 31. Since the pressure increase control valves 179 provided between the front wheel brake cylinders 18, 24 and the rear wheel brake cylinders 50, 52 are placed in the closed state, the fluid in the front wheel brake cylinders 18, 24 is prevented from being discharged into the rear wheel brake cylinders 50, 52. The pressure increase control valves 179 will not be opened even when the fluid pressure at the portion 182 on the side of the pump device 30 is higher than the fluid pressure at the port 180 on the side of the wheel brake cylinders 18, 24, 50, 52. When the pressure increase mode is selected, the electric current (one form of the electric energy) to be applied to the solenoid coil 100 of the pressure increase control valve 179 is controlled so as to zero the pressure difference ΔP, namely, so that the actual wheel brake cylinder pressure approaches the target value, as indicated in Fig. 24. As described below, the electric current to be applied to the coil 100 is controlled by both a feedback control so as to zero the pressure difference ΔP, and a feed-forward control while taking account of the temperature of the working fluid and other parameters. In the pressure increase mode, the pressure reduction control valve 58 is held in the closed state with the solenoid coil 100 being in the de-energized state. On the other hand, the electric current to be applied to the solenoid coil 144 of the pressure reduction control valve 62 is controlled such that the electromagnetic drive force Fs generated is larger than the pressure difference force Fp by a predetermined margin value Fr. Although the maximum electric current may be applied to the coil 144 as in the conventional braking system, this arrangement requires a comparatively large amount of consumption of the electric energy by the pressure reduction control valve 62. In the present embodiment, the amount of electric current to be applied to the coil 144 is determined by the pressure difference ΔP (pressure difference force Fp), that is, by the fluid pressure in the rear wheel brake cylinders 50, 52, so that the electric energy consumption is minimized. Further, the electric current is determined so that the electromagnetic drive force Fs is larger than the pressure difference force Fp by the predetermined margin value Fr, the valve member 90 is stably held seated on the valve seat 92. The electric current is controlled by both a feedback control and a feed-forward control, as in the control of the electric current to be applied to the coil 100 of the pressure increase control valve 179. The margin value Fr indicated above is determined on the basis of a maximum pressure increasing velocity α of the pressure increase control valve 179 and a force Fa that should be applied to the valve member 136 to be seated on the valve seat 134. The maximum pressure increasing velocity α is a maximum amount of pressure increase by the pressure increase control wale 54 per unit time, which is determined by the specific constructional characteristics of the braking system, such as the maximum delivery pressure of the pump device 30, the resistance of fluid flow through the fluid passage 42, and the resistance of fluid flow through an orifice in the pressure increase control valve 179. The linear valve device control routine is executed with a predetermined cycle time ΔT. A fluid pressure Pn+1 controlled by the pressure increase control valve 179 in'the next control cycle will not be higher than (Pn + ΔPi), wherein Pn is the fluid pressure in the present control cycle, and ΔPi is an amount of increase (α·ΔT) of the pressure from the present value Pn at the maximum pressure increasing velocity α. Therefore, the valve member 136 can be held seated on the valve seat 134 by controlling the electric current to be applied to the coil 144 so that the electromagnetic drive force Fs is equal to or larger than the pressure difference force Fp corresponding to the fluid pressure (Pn + α·ΔT). The amount of leak of the fluid is determined by the operating stroke S of the valve member 136, as schematically indicated in Fig. 25. The operating stroke S is determined by the force by which the valve member 136 is forced against the valve seat 134. When the operating stroke S is equal to a nominal stroke S0 at which the valve member 136 is just seated on the valve seat 134, the fluid discharge from the rear wheel brake cylinder 50, 52 to the master cylinder- 31 is theoretically inhibited. However, there may exist a certain gap between the valve member 136 and the valve seat 134 due to an error in the manufacture of the seating valve 82, and the fluid may leak through this gap by an amount amount ε0. The amount of the gap is reduced by further forcing the valve member 136 against the valve seat 134 for elastic deformation thereof, whereby the fluid leak amount is reduced. To reduce the fluid leak amount to a tolerable value ε*, the operating stroke S should be increased from the nominal value S0 to S*, to increase the electromagnetic drive force Fs from Fs0 to Fs*. That is, the margin value Fr is equal to a difference (Fs* - Fs0) between the electromagnetic force Fs* corresponding to the increased operating stroke S* and the electromagnetic force Fs0 corresponding to the nominal operating stroke S0. Accordingly, the electric current to be applied to the solenoid coil 144 is controlled to produce the electromagnetic force Fs which is equal to a sum of the pressure difference force Fp and the margin value Fr = Fs* - Fs. With the fluid leak amount reduced to the tolerable value ε*, the accuracy of control of the fluid pressure in the pressure reduction control valve 62 can be improved. The required amount of consumption of the electric energy according to this arrangement is smaller than in the case where the fluid leak amount is zeroed. when the pressure reduction mode is selected, the pressure increase control valve 179 is held in the closed state with its coil 200 being in the de-energized state, while the electric currents to be applied to the pressure reduction control valves 58, 62 are controlled so as to zero the pressure difference ΔP, as indicated in Fig. 5. When the pressure hold mode is selected, the pressure increase control valve 179 and the pressure reduction control valve 58 are held in the closed state with their coils 100 being in the de-energized state. However, the electric current to be applied to the coil 144 of the pressure reduction control valve 62 is controlled as in the pressure increase mode, that is, so that the electromagnetic force Fs generated is equal to the pressure difference force Fp and the margin value Fr. The pressure difference force Fp corresponds to the fluid pressure in the rear wheel brake cylinders 50, 52 when the pressure hold mode is selected. when the pressure control mode is changed from the pressure reduction mode to the pressure hold mode, the electromagnetic drive force Fs generated by the pressure reduction control valve 62 is increased by an amount Fss which is necessary for seating of the valve member 136 on the valve seat 134. At the end of the pressure reduction mode, the valve member 136 is presumed to be spaced apart from the valve seat 134. Therefore, the electromagnetic drive force Fs must be increased by the amount Fss which is necessary to holding the valve member 136 seated on the valve seat 134. When the pressure increase mode is terminated, the valve member 90 of the seating valve 82 of the pressure increase valve 179 is slowly seated on the valve seat 92, for soft seating of the valve member 90 on the valve seat 92. That is, the pressure increase mode is slowly changed to the pressure hold mode, at a relatively low speed of seating of the valve member 90 on the valve seat 92. To this end, the electric current to be applied to the solenoid coil 10 is slowly reduced to zero, as indicated in Fig. 24, according to an equation i = i 0 β n , wherein "i 0 " represents an electric current applied upon termination of the pressure increase mode, and "β" represents a value smaller than "1", while "n" represents the number of control cycles to slowly reducing the electric current i. Thus, the speed at which the valve member 90 is seated on the valve seat 92 is reduced, making it possible to reduce the seating noise and the impact of the valve member 90 onto the valve seat 92, so that the durability of the pressure increase control valve 179 is increased. When the pressure reduction mode is terminated, the valve member 136 of the pressure reduction control valve 62 is slowly seated on the valve seat 134, for soft seating of the valve member 136 on the valve seat 134. That is, the pressure reduction mode is slowly changed to the pressure hold mode, at a relatively low speed of seating of the valve member 136 on the valve seat 134. To this end, the electric current to be applied to the solenoid coil 144 is slowly increased, as indicated in Fig. 5. The pressure reduction control valve 62 is adapted such that the pressure difference force Fp acts on the Valve member 136 in the direction for moving the valve member 136 away from the valve seat 146, as indicated in Fig. 4B. Accordingly, the electromagnetic drive force Fs generated must be larger than the pressure difference force Fp, so that the valve member 136 is seated onto the valve seat 134. If the electromagnetic drive force Fs is abruptly increased, the speed of seating of the valve member 136 onto the valve seat 134 is excessively large, causing an increased seating noise and an increased impact of the valve member 136 onto the valve seat 134, resulting in reduced durability of the seating valve 130. In the present embodiment, however, the electromagnetic drive force Fs is slowly increased so as to avoid an excessive speed of seating of the valve member 136 onto the valve seat 134, for thereby reducing the seating noise and improving the durability of the seating valve 130. In the present embodiment, the electric current is controlled so that the electromagnetic drive force Fs is slowly increased at a rate represented by (Fr + Fss)/n, as indicated in Fig. 7. "n" represents the number of control cycles which are repeated with the cycle time ΔT. For instance "n" is equal to "4". After the pressure reduction mode has been terminated, the pressure hold mode is established in most cases. Where the pressure reduction mode is changed to the pressure hold mode, the wheel brake cylinder pressure is held constant. To this end, the electric current to be applied to the solenoid coil 144 is controlled so that the electromagnetic drive force Fs is held equal to a sum of the force Fss necessary for seating of the valve seat 136 on the valve seat 134 and the margin value Fr. With this electromagnetic drive force Fs, the valve member 136 is held seated on the valve seat 134, since the wheel brake cylinder pressure is not increased. Thus, the electric current applied to the coil 144 is not abruptly increased, but is slowly increased, so that the valve seat 136 is slowly seated onto the valve seat 134 with a reduced impact therebetween. When a predetermined condition for initiating preliminary current application to the solenoid coil 144 of the pressure reduction control valve 62 is satisfied when the braking system is not in operation, the electric current is applied to the solenoid coil 144 before initiation of a fluid pressure increase in the rear wheel brake cylinder 50, 52, as indicated in Fig. 2, so that the valve member 136 is seated on the valve seat 134. The condition for initiating the preliminary current application is satisfied when there is a high possibility that the pressure increase mode will be selected in the near future. For instance, the condition for initiating the preliminary current application is satisfied when releasing of the accelerator pedal is detected by the throttle opening sensor 168, when the rotating speed of the rear drive wheels 46, 48 has exceeded a threshold VTB' which is lower than the threshold VTB for initiating the traction control described above, or when the amount of spinning or drift-out motion of the vehicle has exceeded a threshold SVs' or SVd' which is lower than the threshold SVs or SVd for initiating the spinning or drift-out control routine. In such a case, the electric current applied to the solenoid coil 144 is slowly increased at a predetermined low rate. Since the pressure difference force Fp is zero while the braking system is not' in operation, the valve seat 136 can be seated on the valve seat 134 with a relatively small value of the electromagnetic drive force Fs. This preliminary current application is effective to reduce a delay in the operation to increase the wheel brake cylinder pressure. In the present embodiment, the preliminary current application is effected for a predetermined time Tpre. If the pressure increase operation is not initiated within this time Tpre, the electric current applied to the coil 144 is zeroed to unseat the valve seat 136 from the valve seat 134. Since the wheel brake cylinder pressure is equal to the atmospheric pressure, the valve member 136 need not be seated on the valve seat 134. This arrangement is effective to prevent unnecessary consumption of the electric energy. In the preliminary current application, the electric current is increased in steps for "n" times (four times in this embodiment), so that the time Tpre is equal to the cycle time ΔT x n. The increase in the electric current in the preliminary current application is illustrated in Fig. 8. In the first control cycle, the electric current is increased to a predetermined value larger than zero. In the fourth and the following control cycles, the electric current is held constant at a predetermined level. The valve seat 136 can be seated on the valve seat 134 in a relatively short time by increasing the electric current to a relatively high value in the first control cycle. This arrangement reduces a delay in the increase in the wheel brake cylinder pressure even if the pressure increase mode is established immediately after the preliminary current application is initiated. Since the electric current is not abruptly increased, the speed of seating of the valve member 136 onto the valve seat 134 can be low enough reduce the seating impact. The preliminary current application may be effected to increase the electric current at a constant rate. in this case, the electric current may be controlled in the same manner when the pressure reduction mode is terminated. The operations to control the electric currents to be applied to the solenoid coils of the linear valve devices 184, 186 will be further explained. As described above, the electric currents are controlled by combination of the feedback and feed-forward controls. The feedback control of the electric currents is effected in a well known PID control fashion (proportional, integral and differential control), so as to zero the pressure difference ΔP between the target and actual values of the wheel brake cylinder pressure. The PID control may be replaced by a P control, an I control, a D control, a PI control or a PD control. The feed-forward control of the electric currents is effected by taking into account the temperature of the working fluid. As shown in Figs. 4A and 4B, the valve member 90, 136 of the pressure increase control valves 173 and pressure reduction control valves 58, 62 receive the biasing force Fk of the spring, electromagnetic drive force Fs and pressure difference force Fp. Where "x0", "x", "i" and "Δ P" respectively represent a displacement of the valve member relative to the valve seat in a steady state of the control valve 179, 58, 62, a displacement of the valve member from the steady state, an electric current applied to the solenoid coil 100, 144, and a pressure difference across the seating valve 82, 130, the biasing force Fk, electromagnetic drive force Fs and pressure difference force Fp are represented by the following equations, respectively: Fk = k0(x0 + x) Fs = ks1•i - ks2•x + α1 Fp = kp1•Δp - kp2•x + α2 wherein, k0, ks1, ks2, kp1, kp2, α1 and α2 are constants determined by the modulus of elasticity of the spring 96, 138 and other factors of the control valves 179, 58, 62. In the above equations, the terms (-ks2•x + α1) and (-kp2•x + α2) are necessary only when the control of the valves is non-linear, and may be zero when the control may be considered linear. In the pressure increase control valves 179 and the pressure reduction control valves 58 shown in Fig. 23A, the following equation (1) is satisfied for equilibrium of the forces Fk, Fs, Fp: Fk = Fs + Fp in the pressure reduction control valves 62, the following equation (2) is satisfied for equilibrium of the forces Fk, Fs and Fp: Fk + Fp = Fs Further, the orifice provided in each of those control valves 179, 58, 62 satisfies the following equation (3): Q = C•A•√(Δp/ρ) A change in the fluid pressure and a rate of change in the fluid volume have a relationship represented by the following equation (4): dV/V = β•dp In the above equation (3), "Q", "C", "A" and "ρ" respectively represent a flow rate of the working fluid, a coefficient of flow of the fluid, a cross sectional area of flow of the fluid through the control valve, and a density of the fluid. The cross sectional area A of flow of the fluid is proportional to the operating stroke x of the valve member 90, 136, and is represented by an equation A = ka•x. in the above equation (4), "β" represents a compression coefficient of the fluid, which is a reciprocal of the modulus of elasticity of volume, and "V" represents an initial value of the volume of the fluid, which is determined by the volume of the control valve. "dV" is a positive value when the volume is reduced. The above equation (4) may be converted into (1/Vβ)•dV/dt = dp/dt. Since "dv/dt" in this equation is equal to the flow rate Q of the fluid, the equation can be converted into the following equation (5): (1/Vβ)•Q = dp/dt From the above equations (3j and (5), the following equation (6) may be obtained: (1/Vβ)•C•ka•x•√(Δp/ρ) = dp/dt According to the above equations (1) and (2), the operating stroke x of the valve member 90, 136 can be expressed by the following equation (7) which includes the pressure difference Δp across the seating valve 82, 130, the electric current i and coefficients a1, a2 and a3: x = a1•Δp + a2•i + a3 For the pressure increase control valves 179 and the pressure reduction control valves 58, the coefficients a1, a2 and a3 are represented by the following equations: a1 = kp1/(k0 + ks2 + kp2) a2 = ks1/(k0 - ks2 + kp2) a3 = -(k0•x0 + α1 + α2)/(k0 + ks2 + kp2) For the pressure reduction control valves 62, the coefficients a1, a2 and a3 are represented by the following equations: a1 = kp1/(kp2 - k0 - ks2) a2 = -ks1/(kp2 - k0 - ks2) a3 = (k0•x0 + α2 - α1)/(kp2 - k0 - ks2) From the above equations (6) and (7), the electric current i is represented by the following equation (8): i = {1/(C•Vβ•ka)•(dP/dt)√(ρ/Δp) - a1•Δp - α 3}/a2 By inserting into the above equation (8) a target rate of change dP/dt of the target wheel brake cylinder pressure and a pressure difference Δp across the seating valve 82, 130 when the target wheel brake cylinder pressure P is expected to be established, the electric current i can be obtained. The density ρ and the compression coefficient β (volume change ratio) of the fluid vary with the temperature of the fluid. The electric current i obtained according to the above equation (8) wherein "ρ" and "β" are determined depending upon the fluid temperature is an optimum electric current reflecting the fluid temperature. In other words, the electric current i according to the above equation (8) permits optimum control of the control valves 179, 58, 62 without an adverse influence of the fluid temperature. The fluid temperature can be detected by a temperature sensor 174. The linear valve devices 184, 186 are controlled according to the linear valve device control routine illustrated in the flow chart of Fig. 9. The linear valve device control routine is initiated with step S131 to determine whether the target wheel brake cylinder pressure Pref is higher than the atmospheric pressure. Normally, a negative decision (NO) is obtained in step S131, and the control flow goes to step S132 to determine whether a PRELIMINARY CURRENT APPLICATION flag is set at "1". The target wheel brake cylinder pressure Pref is higher than the atmospheric pressure when any one of the normal braking pressure control routine, anti-lock braking pressure control routine, traction control routine and vehicle stability brake control routine is executed. The PRELIMINARY CURRENT APPLICATION flag is set to ''1" when the predetermined condition for initiating the preliminary current application is satisfied, and is reset to "0" when one of the above-indicated routines is initiated or when the predetermined time Tpre indicated above has passed. If a negative decision (NO) is obtained in step S132, the control flow goes to step S133 to determine whether the condition for initiating the preliminary current application is satisfied or not. As indicated above, this condition is satisfied when any sign indicating a high possibility that the pressure increase mode will be selected in the near future is detected. When an affirmative decision (YES) is obtained in step S133, the control flow goes to step S134 to set the PRELIMINARY CURRENT APPLICATION flat to "1", and then to step S135 to determine whether a content npre of a counter for counting the number of control cycles during the preliminary current application is larger than a predetermined value Nspre. If a negative decision (NO) is obtained, the control flow goes to step S136 in which the predetermined electric current is applied to the solenoid coil 144 of the pressure reduction control valves 62, and the content npre is incremented. The predetermined value Nspre corresponds to the predetermined time Tpre described above by reference to Fig. 8. Then, the control flow goes back to step S131. When step S132 is implemented again, an affirmative decision (YES) is obtained in step S132, and the control flow goes to step S135 skipping steps S133 and S134. Steps S132, S135 and S136 are repeatedly implemented until an affirmative decision (YES) is obtained in step S135. Thus, the electric current is applied to the coil 144 for the time Tpre. If the preliminary current application initiating condition is not satisfied, a negative decision (NO) is obtained in step S133, and the control flow goes to step S137 to reset the various counters and place the solenoid-operated valves 26, 26, 58, 62, 179 in the initial states, since the fluid pressures in any wheel brake cylinders 18, 24, 50, 52 are controlled, and the preliminary current application is not effected. A step similar to step S137 is implemented when the operation to control the wheel brake cylinder pressure is terminated. When the content npre of the counter has exceeded the predetermined value Nspre, an affirmative decision (YES) is obtained in step S134, and the control flow goes to step S138 in which the electric current applied to the coil 144 is zeroed, and the PRELIMINARY CURRENT APPLICATION flat is reset to ''0''. That is, the preliminary current application to the coil 144 is terminated since an operation to control the wheel brake cylinder pressure has not not initiated within the predetermined-time Tpre. When the target wheel brake cylinder pressure Pref is higher than the atmospheric pressure, an affirmative decision (YES) is obtained in step S131, and the control flow goes to step S139 to reset the PRELIMINARY CURRENT APPLICATION flat to "0", and then to step S140 in which the linear valve devices 184, 186 are controlled as illustrated in the flow chart of Fig. 10. Initially, step S142 is implemented to determine whether a rate of change ΔPref of the target wheel brake cylinder pressure Pref is higher than a predetermined value ΔPref'. If an affirmative decision (YES) is obtained in step S142, it means that the pressure increase mode is required. In this case, the control flow goes to step S143 to determine whether the pressure difference ΔP between the target value Pref and the actual value P is equal to or larger than the predetermined threshold EPS. If an affirmative decision (YES) is obtained in step S143, the control flow goes to step S144 in which the pressure increase mode is established to increase the fluid pressure in the wheel brake cylinder. When the pressure difference ΔP has been reduced to a value smaller than the threshold EPS, namely, when the pressure increase mode is terminated, a negative decision (NO) is obtained in step S143, and the control flow goes to step S145 to determine whether a content damp_count of a counter for slow seating of the valve member 90 of each pressure increase control valve 179 on the valve seat 92 is smaller than a predetermined value. If an affirmative decision (YES) is obtained in step S145, the control flow goes to step S146 in which the electric current applied to the coil 100 is reduced, and the content damp_count of the counter is incremented. Steps S145 and S146 are repeatedly implemented until the content damp_count has reached the predetermined value, that is, until a negative decision (NO) is obtained in step s145. Thus, the electric current applied to the coil 100 of the seating valve 82 of the pressure increase control valve 179 is slowly reduced for slow or soft seating of the valve member 90 on the valve seat 92. When the negative decision (NO) is obtained in step S145, the control flow goes to step S147 to establish the pressure hold mode and reset the content damp_count of the counter for the soft seating of the valve member 90. If a negative decision (NO) is obtained in step S142, it means that the pressure increase mode is not required. In this case, the control flow goes to step S148 to determine whether the rate of change ΔPref of the target wheel brake cylinder pressure Pref is smaller than a threshold -ΔPref'. If an affirmative decision (YES) is obtained in step S148, the control flow goes to step S149 to determine whether the pressure difference ΔP is equal to or smaller than a predetermined threshold -EPS, that is, whether the absolute value of the pressure difference ΔP is equal to or larger than the value EPS. If an affirmative decision (YES) is obtained in step S149, it means that the pressure reduction mode is required. In this case, the control flow goes to step S150 to establish the pressure reduction mode in which the pressure increase control valves 179 are held in the closed state with the coils 100 being in the de-energized state, and the electric currents to be applied to the coils 100, 144 of the pressure reduction control valves 5a, 62 are controlled so as to reduce the pressure difference ΔP, by the feedback and feed-forward controls indicated above. If a negative decision (NO) is obtained in step S149, the control flow goes to step S151 to determine whether the content damp_count of a counter for slow seating of the valve member 136 of each pressure reduction control valve 62 on the valve seat 134 is smaller than a predetermined value. If an affirmative decision (YES) is obtained in step S151, the control flow goes to step S152 in which the electric current applied to the coil 144 is increased, and the content damp_count is incremented. Steps S151 and S152 are repeatedly implemented until a negative decision (NO) is obtained in step S151. Thus, the electric current applied to the coil 144 of the seating valve 130 of the pressure reduction control valve 62 is slowly increased until the electromagnetic drive force Fs is increased to a sum of the required seating force Fss and the margin value Fr (described above by reference to Fig. 7), for slow or soft seating of the valve member 136 on the valve seat 134. When the negative decision (NO) is obtained in step S151, the control flow goes to step S153 to establish the pressure hold mode. When a negative decision (NO) is obtained in step S148, that is, when the rate of change ΔPref of the target wheel brake cylinder pressure Pref is in a predetermined range between ΔPref' and -ΔPref', the control flow goes to step S154 to establish the pressure hold mode. Thus, one of the pressure increase mode, pressure hold mode and pressure reduction mode is selected during an operation of the brake pedal 10, in not only the normal braking pressure control but also the anti-lock braking pressure control. By controlling the pressure increase control valves 179 and the pressure reduction control valves 62 as illustrated in the flow chart of Fig. 10, the electric currents applied to the coils 100 and 144 of these control valves 179, 62 are controlled as schematically indicated in the graph of Fig. 11. As shown in this graph, the amount of electric current applied to the coils 144 of the pressure reduction control valves 62 is made smaller than in the conventional control, thereby reducing the required amount of consumption of the electric energy by the control valves 62. Further, the electric current applied to the coils 144 is controlled so that the electromagnetic drive force Fs is larger than the pressure difference force Fp by the predetermined margin value Fr, whereby the valve member 136 can be held seated on the valve seat 134 with high stability. Further, the electric current applied to the coils 144 are slowly increased upon termination of the pressure reduction mode, so that the valve member 136 is slowly seated on the valve seat 134, for reduced impact of seating of the valve member 136 onto the valve seat 134, and accordingly increased durability of the pressure reduction control valves 62. In addition, the preliminary current application to the coils 144 of the control valves 62 before initiation of an operation in the pressure increase mode is effective to reduce a delay in the increase in the wheel brake cylinder pressures 50, 52 in the pressure increase mode. Further, the electric current applied to the coils 100 of the pressure increase control valves 179 is slowly reduced upon termination of the pressure increase mode, for soft or slow seating of the valve member 30 on the valve seat 92, for reducing impact of seating of the valve member 90 onto the valve seat 92 and accordingly increased durability of the pressure increase control valves 179. It will be understood from the foregoing description of the embodiment of this invention that a controller for controlling an electric energy to be supplied to the solenoid-operated pressure control valve 62 is constituted by the stroke sensor 71, pressure sensors 72, 74-78, control device 160, brake switch 166, throttle opening sensor 168, steering angle sensor 170 and yaw rate sensor 172. It will also be understood that the controller includes electric energy control means for controlling the electric energy to be supplied to the pressure control valve 62 depending upon the fluid pressure in the brake cylinder 50, 52. The electric energy control means is constituted by a portion of the control device 160 assigned to implement steps S144, S147, S153 and S154 of the flow chart of Fig. 10. It will further be understood that a portion of the control device 160 assigned to implement steps S134-S136 and S152 of the flow charts of Figs. 9 and 10 constitutes slowly energy increasing means for slowly increasing the electric energy to be supplied to the pressure control valve 62, for slow seating of the valve member 136 on the valve seat 134. It will also be understood that the slowly energy increasing means includes preliminary energy applying means for initiating application of the electric energy to the pressure control valve 62 before the condition for increasing the fluid pressure in the brake cylinder 50, 52 is satisfied. The preliminary energy applying means is constituted by a portion of the control device 160 assigned to implement steps S134-S136. It will further be understood that the pump device 30, pressure increase control valves 179 and a portion of the control device 160 assigned to implement steps S144 constitute pressure increasing means for increasing the fluid pressure in the brake cylinder 50, 52. In the embodiment of Figs. 9 and 10, the electric current to be applied to the pressure reduction control valves 62 is increased in different patterns in the preliminary current application control in steps S134-S136 and in the soft or slow seating current application control in step S152 after the pressure reduction mode. However, the electric current is increased in the same pattern in these preliminary current application control and soft seating current application control. The preliminary current application control is not essential, and may be replaced by a slow seating current application control effected immediately after the initiation of the pressure increase control in the pressure increase mode, so that the electric current applied to the coil 144 is slowly increased for a predetermined time after the initiation of the pressure increase control, as in a linear valve device control routine illustrated in the flow chart of Fig. 12 according to a further embodiment of the present invention. The linear valve device control routine of Fig. 12 is initiated with step S161 to determine whether the target wheel brake cylinder pressure Pref is higher than the atmospheric pressure. If an affirmative decision (YES) is obtained in step S161, it means that the wheel brake cylinder pressure is being controlled. In this case, the control flow goes to step S162 to determine whether the rate of change ΔPref is higher than the predetermined threshold ΔPref'. If a negative decision (NO) is obtained in step S161, it means that the wheel brake cylinder pressure is not being controlled. In this case, the control flow goes to step S166 in which the braking system is initialized, with various counters being reset and the solenoid-operated valves 26, 28, 58, 62, 179 being placed in the initial states. If an affirmative decision (YES) is obtained in step S162, the control flow goes to step S163 to determine whether the content damp_count of a counter for soft or slow seating immediately after initiation of the pressure increase control in the pressure increase mode is smaller than a predetermined threshold. If an affirmative decision (YES) is obtained in step S163, the control flow goes to step S164 to effect the slow seating current application control in which the electric current to be applied to the coil 144 of the pressure reduction control valves 62 is slowly increased. Steps S163 and S164 are repeatedly implemented until a negative decision. (NO) is obtained in step S163. That is, the electric current applied to the coil 144 is slowly increased until a predetermined time has passed after the pressure increase mode is required, that is, after the affirmative decision is obtained in step S162. When the negative decision is obtained in step S163, the control flow goes to step S165 and subsequent steps S167-S170 which are identical with steps S145-S150 of Fig. 10. Steps S171-S77 of Fig. 12 are identical with steps S143-S154 of Fig. 10. While the electric current to be applied to the coil 144 may be slowly increased in step S164 in the same pattern as described above with respect to the sixth embodiment by reference to the graph of Fig. 8, the control of the electric current in step S164 is effected in the present seventh embodiment such that the electromagnetic drive force Fs changes as indicated by solid line in the graph of Fig. 13. In this case, the margin value Fr is equal to the pressure difference force Fp based on the pressure difference ΔP, which is equal to α·ΔT, as described above, wherein "α" represents the maximum pressure increasing velocity, while the ΔT is the cycle time. Accordingly, the electromagnetic drive force Fs increases at a rate corresponding to the margin value Fr = Fp corresponding to the pressure difference ΔP. Namely, ΔFs (Δi) = Fp (ΔP)·n, wherein "n" represents the number of control cycles in which the electric current i is slowly increased up to a value i e . If the current value i e in the last control cycle is higher than a normal value i n according the normal current control in the pressure increase mode, the current i is not immediately reduced to the value i n , but is held at the value i e until the normal value in has increased to the value i e . Then, the current i is controlled according to the normal current control in the pressure increase mode. In the first control cycle, the electromagnetic drive force Fs is smaller than a sum of the pressure difference force Fp and the margin value Fr. In the second and subsequent control cycles, the electromagnetic force Fs is equal to or larger than the sum, as in the normal current control in the pressure increase mode, as indicated in Fig. 13. This arrangement assures stable seating of the valve member 136 on the valve seat 134, while avoiding an excessive amount of the electromagnetic drive force Fs upon seating of the valve member 136 onto the valve seat 134. In this case, the target wheel brake cylinder pressure Pref may be considered to be increased at the maximum pressure increasing velocity a. The control of the electric current in step S164 may be such that the electromagnetic drive force Fs increases according to the following equation (9), as indicated by solid line L in the graph of Fig. 14, according to an eighth embodiment of this invention. Fs(i) = Fp(Pref)/γ + Fr x γ In the above equation, "γ" represents a deceleration coefficient represented by 1 - 1/2 t . The deceleration coefficient γ approaches "1" as the time passes. The electric current controlled according to the above equation (9) is comparatively large immediately after the initiation of the pressure increase control, so that the valve member 136 is moved toward the valve seat 134 at a relatively high velocity and is seated on the valve seat 134. Subsequently, the electric current is controlled according to the normal current control in the pressure increase mode. The deceleration coefficient γ may be replaced by any other suitable coefficient. The electric current may be controlled according to the following equation (10): Fs(i) = Fp(Pref)/γ The electric current applied to the coil 144 may be controlled according to another equation similar to the above equations (9) and (10), provided the electric current is comparatively large in the first control cycle and approaches the normal value as the time passes, as in the above cases according to the above equations (9) and (10). Further, the electric current may be controlled according to a stored data map representative of a predetermined relationship between the electric current values and the control cycles. Further, the electric current may be controlled such that the electromagnetic drive force Fs increases according to the following equation (11), as indicated by one-dot chain line M in Fig. 14: Fs(i) = Fp(Pref) x γ In this case, too, the valve member 136 may be slowly seated on the valve seat 134, for reduced seating noise of the valve member 136 and increased durability of the pressure reduction control valves 62. The electric current applied to the coil 144 may be controlled as indicated by solid line L' in Fig. 15, so as to generate a relatively large electromagnetic drive force Fs in the first control cycle and subsequently hold the drive force Fs constant before the normal current control in the pressure control mode. Alternatively, the electric current may be controlled as indicated by two-dot chain line M' in Fig. 15, such that the electric current in the first control cycle is comparatively small, and such that the electric current in the second and subsequent control cycles are comparatively large. In the latter case, the pressure increase may be delayed, but the valve member 136 can be seated on the valve seat 134 at a sufficiently low velocity, and can be held seated with high stability. The tolerable amount of fluid leak used in determining the margin value Fr of the electromagnetic drive force Fs need not be a predetermined constant value ε*, but may be a variable which changes with the target wheel brake cylinder pressure Pref. It is also noted that the margin value Fr need not be determined on the basis of both the maximum pressure increasing velocity a and the force for forcing the valve member 135 onto the valve seat 134, but may be determined on the basis of one of those values. The maximum pressure increasing velocity α may also be a variable which varies with the delivery pressure of the pump device 30, for instance. Further, the electromagnetic drive force Fs need not be larger than the pressure difference force Fp by the margin value Fr, and may be equal to the pressure difference force Fp where the valve member 136 is seated on the valve seat 134 in a good condition. In this case, the electric energy consumption by the pressure reduction control valves 62 can be minimized. In view of a tendency that the amount of fluid leak through a gap between the valve member 136 and the valve seat 134 changes with the time, it is possible to automatically update the relationship between the operating stroke S of the valve member 136 and the fluid leak amount, as indicated in the graph of Fig. 6. The embodiments of Figs. 1-15 are adapted to control the electric currents applied to the coils 100, 144 of the control valves 58, 62, 179, while taking account of the temperature of the working fluid with which the density ρ and volume change ratio (compression coefficient) β. However, the electric currents may be adjusted by a compensation coefficient which is determined on the basis of the electric current i and the wheel brake cylinder pressure P, and according to a predetermined relationship among the compensation coefficient, electric current i and wheel brake cylinder pressure P, as shown in Fig. 16 by way of example. The predetermined relationship is represented by a data map stored in the ROM 154. In this case, the temperature sensor 174 is not necessary. For instance, the above-indicated relationship may be obtained as follows: The actual fluid pressure in the wheel brake cylinders when the electric current applied to the coil 100, 144 is determined according to the above-indicated equation (8) including the target wheel brake cylinder pressure P, etc. is detected. If the density p and compression coefficient β of the working fluid accurately correspond to the actual fluid temperature, the detected actual value and the target value of the wheel brake cylinder pressure are equal to each other. In other words, if the detected actual wheel brake cylinder pressure when the electric current corresponding to the target wheel brake cylinder pressure is applied to the coil 100, 144 is the same as the target wheel brake cylinder pressure, the electric current is not required to be compensated. In this case, the compensation coefficient is set at "1". It the actual and target wheel brake cylinder pressures are different from each other, the compensation coefficient is determined based on the difference. Thus, the predetermined relationship is obtained, and a data map representative of this relationship is stored in the ROM 154, so that the compensation coefficient is determined according to the data map, and on the basis of the wheel brake cylinder pressure and the electric current calculated according to the above equation (8). The electric current calculated according to the above equation (8) is adjusted by multiplying the calculated electric current by the determined compensation coefficient. The thus adjusted electric current reflects the changes in the density and compression coefficient of the working fluid with its temperature. In the embodiment of Fig. 16, it is not required to take into account the chronological change in the amount of the fluid leak from the closed control valves 58, 62, 179 when the electric current applied to the coil 100, 144 is determined. Accordingly, it is not required to update the relationship between the fluid leak amount and the operating stroke of the valve member 90, 136. The pressure increase control valves 179 need not be controlled, provided that the delivery pressure of the pump device 30 is controlled depending upon the amount of operation of the brake pedal 10, for example. The pressure increase control valves 179 are controlled only when the delivery pressure of the pump device 30 is excessively high, so that actual wheel brake cylinder pressure is close to the target value. in this case, the high-pressure and low-pressure pump motors 38, 40 are controlled so that the delivery pressure detected by the pump pressure sensor 74 corresponds to the depression force acting on the brake pedal 10. The pressure increase control valves 179 may be eliminated depending upon the application of the braking system. Further, the spring 133 of the pressure reduction control valves 62 is not essential. Where the spring 138 is eliminated, the pressure reduction control valves 62 may be placed in the closed state when the braking system is not in operation. However, this closed state is not positively maintained. The hydraulic circuit in the braking system is not limited to the details shown in Fig. 17. For instance, the braking system may use a pump device 200 including an accumulator 202 and a single pump 204, as shown in Fig. 17. The pump device 200 further includes a pump motor 206 for operating the pump 204 to deliver a pressurized fluid stored in the accumulator 202. The pump motor 206 is controlled so that the fluid pressure in the accumulator 202 is held within a predetermined optimum range. A pressure switch 208 is provided to detect a rise and a drop of the fluid pressure in the accumulator 202 above and below upper and lower limits of the optimum range. In the braking system of Fig. 17, each of the wheel brake cylinders 18, 24, 50, 52 is connected to both of the master cylinder 12 and the pump device 200. In a fluid passage connecting the pump device 200 and the front wheel brake cylinders 18, 22, there is provided a solenoid-operated shut-off valve 220. In a fluid passage connecting the pump device 200 and the rear wheel brake cylinders 50, 52, there is provided a solenoid-operated shut-off valve 222. These shut-off valves 220, 222 are normally closed valves. In a fluid passage connecting the master cylinder 12 and the front wheel brake cylinders 18, 22, there is provided a solenoid-operated shut-off valve 224. In a fluid passage connecting the master cylinder 12 and the rear wheel brake cylinders 50, 52, there is provided a solenoid-operated shut-off valve 226. These shut-off valves 224, 226 are normally open valves. In this braking system, all wheel brake cylinders 18, 22, 50, 52 are disconnected from the pump device 200 and are connected to the master cylinder 12, in the event of occurrence of an electrical abnormality of the braking system. in the fluid passage connecting the master cylinder 12 and the rear wheel brake cylinders 50, 52, there is also provided a pressure increasing valve 228 so that the fluid pressure generated by the master cylinder 12 is further increased by the pressure increasing valve 228, so that the thus increased fluid pressure is applied to the rear wheel brake cylinders 50, 52. in a fluid passage connecting the front wheel brake cylinders 13, 22, there is provided a solenoid-operated shut-off valve 230. In a fluid passage connecting the rear wheel brake cylinders 50, 52, there is provided a solenoid-operated shut-off valve 232. These shut-off valves 230, 232 are controlled to effect selective connection and disconnection of the two front wheel brake cylinders 18, 22 or two rear wheel brake cylinders 50, 52 to and from each other. When the fluid pressures in the two wheel brake cylinders 18, 22 (50, 52) are controlled independently of each other, the shut-off valve 230, 232 is placed in the closed state . When the fluid pressures in the two wheel brake cylinders are controlled to the same level, the shut-off valve 230, 232 is placed in the open state. In this latter case, only one of the linear valve devices 184, 186 is controlled to control the fluid pressures in the two wheel brake cylinders. In the present embodiment of Fig. 17, a leakage of the fluid from the shut-off valve 222 and/or pressure increase control valve 179 may cause the pressurized fluid to be supplied from the accumulator 202 to the rear wheel brake cylinder 50, 52, resulting in a so-called "brake drag", even while the brake pedal 10 is in the non-operated position. However, the pressure reduction control valves 62 are normally open valves, to permit the fluid to be discharged from the rear wheel brake cylinders 50, 52 to the master reservoir 31, for thereby preventing the "brake drag" which would otherwise take place due to the fluid leakage. In this respect, it is desirable that the pressure reduction control valves 58 for the front wheel brake cylinders 18, 22 are normally open valves. In the braking system of Fig. 17, the pump device 200 which is independent of the master cylinder 12 may be replaced by any other device such as a pressure booster or pressure increasing device. In this case, too, the fluid pressures in the rear wheel brake cylinders 50, 52 are controlled by controlling the pressure reduction control valves 62. It is also noted that the traction control and vehicle stability brake control are not essential.

Claims: A hydraulic pressure control apparatus for controlling a pressure of a working fluid in a brake cylinder, said apparatus comprising: a low-pressure reservoir (31) for storing the fluid discharged from said brake cylinder; a solenoid-operated pressure control valve (58, 62) disposed between said low-pressure reservoir and said brake cylinder; and a controller (160, 71, 72, 74-78, 166, 168, 170, 172) for controlling an electric energy applied to said solenoid-operated pressure control valve, wherein said solenoid-operated pressure control valve (62) includes a solenoid-operated seating valve (130) having a valve seat (134), a valve member (136) movable to be seated on and unseated from said valve seat, and an electromagnetic force generating device (132) for generating an electromagnetic drive force (Fs) corresponding to the electric energy applied thereto, such that the electromagnetic drive force acts on said valve member in a first direction for moving said valve member to be seated on said valve seat, said seating valve being constructed such that a pressure difference force (Fp) based on a difference between the fluid pressures in said low-pressure reservoir (31) and said brake cylinder (50, 52) acts on said valve member in a second direction opposite to said first direction, and wherein said controller further includes electric energy control means (160, S144, S147, 5152-S154) for controlling the electric energy to be applied to said electromagnetic force generating device, depending upon the fluid pressure in said brake cylinder, when said valve member of said seating valve is held seated on said valve seat, characterised in that said electric energy control means controls said electric energy to be applied to said electromagnetic force generating device (132) such that said electromagnetic drive force (Fs) is larger than said pressure difference force (Fp) by at least a predetermined margin value (Fr) which is determined by at least a maximum velocity (α) of increase in the fluid pressure at said solenoid-operated pressure control device (62). A hydraulic pressure control apparatus according to claim 1, wherein said margin value (Fr) is determined by a force (Fss) required for forcing said valve member (136) onto said valve seat (134) and said maximum velocity of increase in the fluid pressure at said solenoid-operated pressure control device. A hydraulic pressure control apparatus according to claim 1 or 2, wherein said margin value (Fr) is determined by an amount of leak of the working fluid through a gap between contact surfaces of said valve member (136) and said valve seat (134 ) and said maximum velocity of increase in the fluid pressure at said solenoid-operated pressure control device A hydraulic pressure control apparatus according to any one of claims 1-3, wherein said electric energy control means controls the electric energy to be applied to said electromagnetic force generating device (132), further depending upon a temperature of the working fluid flowing through said solenoid-operated seating valve (130).

Inventors: NAKAMURA KIYOHARU, OHKUBO MASAYASU, SAKAI AKIRA, SHIMURA AMANE, SOEJIMA SHINICHI, NAKAMURA, KIYOHARU, OHKUBO, MASAYASU, SAKAI, AKIRA, SHIMURA, AMANE, SOEJIMA, SHINICHI, Ohkubo, Masayasu, Toyota Jidosha Kabushiki Kaisha, Sakai, Akira, Toyota Jidosha Kabushiki Kaisha, Shimura, Amane, Toyota Jidosha Kabushiki Kaisha, Soejima, Shinichi, Toyota Jidosha Kabushiki Kaisha
Applicants: TOYOTA MOTOR CO LTD, TOYOTA JIDOSHA KABUSHIKI KAISHA
Application year: 1999
Main classification: B60T8
IPC codes: B60T8, B60T7