Electro-hydraulic actuator for brake

An electro-hydraulic actuator (1) for actuating a brake (2) with a hydraulic thrust unit comprises an electric motor (3), a converting mechanism (5) to convert a rotational motion of the motor (4) into a translational motion, a cylinder (8), and a piston (9) connected to the converting mechanism, wherein the converting mechanism (5) is configured so that, for a given angular speed of the drive shaft (4), the translation speed of the piston (9) decreases from a maximum value in a rear length (14) of the piston stroke to a minimum value in a front length (13) of the piston stroke.

It is the object of the present invention an electro-hydraulic actuator for a brake, in particular for a disc brake with a hydraulic cylinder-piston thrust unit for motor vehicles, motor cycles, and commercial and industrial vehicles.

From U.S. Pat. No. 6,623,087 and DE19527936, braking systems for motor vehicles of the BBW type (“Brake By Wire”) are known, in which a linear transducer connected to a brake pedal detects the stroke of the brake pedal and transmits an electrical signal indicative for the request of braking torque by the user to a control unit. The control unit processes the signals of the transducer and controls an electric motor of a hydraulic pump as a function of the required braking torque. The hydraulic pump actuated by the electric motor pressurizes and conveys a hydraulic fluid to the hydraulic pressure units of the vehicle brakes.

Compared to the conventional braking systems, in which the brake pedal directly acts on the hydraulic circuit, the advantages of the “Brake By Wire” systems is that they allow generating and controlling the braking system hydraulic pressure without the aid of the force applied through the brake pedal. Furthermore, the at least partial replacement of the hydraulic circuits by electric circuits allows saving hydraulic fluid, reducing weight, and reducing the environmental impact of the braking system. Finally, the management of the braking system by an electronic control unit based on electric signals representing the required braking torque, allows a more ergonomic designing of the brake pedal or the brake lever, and a response of the braking system that is more targeted and differentiated according to the road conditions and the vehicle conditions upon braking, in addition to the possibility of performing a regenerative braking action (with an at least partial recovery of the kinetic energy) and blending.

In spite of the several advantages of the “Brake By Wire” systems, under extreme braking conditions, for example, in the case of an abrupt efficiency loss of the brake (the so-called “fading”) following an overheating after prolonged and deep braking actions, the brake requires very high hydraulic pressures, which translate into an oversizing of the pump electric motor, which otherwise would risk to be burned. Such oversizing of the electric motor does not lead to any advantages in 99% of operative situations of the braking system, yet involving additional manufacturing and operative costs, as well as a high weight and high overall dimensions.

Therefore, the object of the present invention is to provide an electro-hydraulic actuator for a hydraulic brake, having such characteristics as to obviate the drawbacks set forth with reference to the prior art.

A particular object of the invention is to propose an electro-hydraulic actuator for a hydraulic brake, in which an electric motor of the electro-hydraulic actuator is dimensioned for the standard operative conditions of the brake, and in which the electro-hydraulic actuator is configured to be able to generate, under extreme conditions, by the same electric motor, an exceptionally high fluid pressure without the risk of overheating the motor.

These and other object are achieved by an electro-hydraulic actuator for actuating a brake having a hydraulic thrust unit, said actuator comprising:an electric motor with a drive shaft,a converting mechanism connected with the drive shaft and adapted to convert a rotational motion of the drive shaft into a translational motion of a translatable portion,a hydraulic pump connected to the converting mechanism and adapted to carry out, in response to the translational movement, an increase in a hydraulic liquid pressure, in which the hydraulic pump comprises:a cylinder and a piston received in the cylinder and constrained to the translatable portion so as to translate together with it with respect to the cylinder along a piston stroke extending from a rear end of stroke to a front end of stroke, said piston stroke comprising:a front length including the front end of stroke and having a length less than or equal to half the length of the piston stroke, anda rear length including the rear end of stroke and having a length less than or equal to half the length of the piston stroke,a pressure chamber defined by the cylinder and the piston and having a volume that varies as a function of the position of the piston from a maximum volume when the piston is at the rear end of stroke position to a minimum volume when the piston is at the front end of stroke position,a supply duct in communication with the pressure chamber, and adapted to be connected with the hydraulic thrust unit of the brake,
in which the converting mechanism is configured so that, for a given angular speed of the drive shaft, the translation speed of the translatable portion decreases from a maximum value in a position of the piston at the rear length to a minimum value in a position of the piston at the front length.

This reduces the motor torque and thus the electric power supply required to generate a high fluid pressure in the front length of the piston stroke, thus obviating the risk of overheating the electric motor in extreme braking situations. At the same time, when the piston is in the rear length of the piston stroke (situation of standard operative braking actions), the actuator allows an operation of the electric motor in a motor torque capacity and absorbed power supply suitable for the dimensioning of the motor.

Furthermore, by virtue of the fact that the minimum value of the transmission and conversion ratio (or, in other words: the minimum value of the quotient between the linear translation speed of the translatable portion and the angular speed of the drive shaft) is reached when the piston is in the front length of the piston stroke, it is possible to generate very high fluid pressures under high fluid shift conditions outside the cylinder.

In accordance with an aspect of the invention, the front length of the piston stroke has a length less than one third of the length of the piston stroke, preferably less than one fourth of the length of the piston stroke. Still more preferably, the minimum value of the transmission and conversion ratio (or, in other words: the minimum value of the quotient between the linear translation speed of the translatable portion and the angular speed of the drive shaft) is reached when the piston is at the front end of stroke.

In accordance with a further aspect of the invention, the maximum value of the transmission and conversion ratio (or, in other words: the maximum value of the quotient between the linear translation speed of the translatable portion and the angular speed of the drive shaft) is reached when the piston is at the rear end of stroke.

In accordance with a further aspect of the invention, the converting mechanism comprises a crank and connecting rod mechanism for a conversion of the rotational motion of the drive shaft into the translational motion of the translatable portion. In the crank and connecting rod mechanisms, the translation speed of the translatable portion is a non-linear function of the angular position of the crank and also of the stroke of the translatable portion. Considering now that the fluid pressure is approximately a linear function of the piston stroke, and by combining such linear function of the pressure with the non-linear function of the translation speed of the piston, a non-linear ratio between the motor torque and the stroke of the piston is obtained, in which, in a prevailing portion of the rear length (in the first half) of the piston stroke, the motor torque increases in an approximately proportional (linear) manner as the piston advances, and in the front length (in the second half) of the piston stroke, the motor torque increases in a sub-proportional (or decreasing) manner, or it decreases as the piston advances. Just this characteristic allows an optimal operation of the electric motor and a precise control of the fluid pressure under normal braking conditions (with the piston at the rear length of the piston stroke) and a very high pressurization of the fluid under exceptional braking conditions (with the piston at the front length of the piston stroke).

In accordance with a further aspect of the invention, the converting mechanism comprises a cam mechanism for a conversion of the rotational motion of the drive shaft into the translational motion of the translatable portion. In such a case, the cam surface has a shape and a radial distance to its rotation fulcrum such as to obtain the above-mentioned relationship between the translation speed of the piston and the angular speed of the drive shaft and, preferably, a non-linear ratio between the motor torque and the piston stroke, in which, in a prevailing portion of the rear length (in the first half) of the piston stroke, the motor torque increases in an approximately proportional (linear) manner as the piston advances, and in the front length (in the second half) of the piston stroke, the motor torque increases in a sub-proportional (or decreasing) manner, or it decreases as the piston advances.

In accordance with a further aspect of the invention, the converting mechanism comprises a rotor coupled with a stator by the engagement between volving members connected to one of them, and a helicoidal track formed in the other one, in which the angle of the coil of the helicoidal track varies along the length thereof, so as to obtain the above-mentioned relationship between the translation speed of the piston and the angular speed of the drive shaft and, preferably, a non-linear ratio between the motor torque and the piston stroke, in which, in a prevailing portion of the rear length (in the first half) of the piston stroke, the motor torque increases in an approximately proportional (linear) manner as the piston advances, and in the front length (in the second half) of the piston stroke, the motor torque increases in a sub-proportional (or decreasing) manner, or it decreases as the piston advances.

With reference to the Figures, an electro-hydraulic actuator is generally indicated with the reference1. The actuator1is provided for actuating a brake2with a hydraulic thrust unit and comprises an electric motor3with a drive shaft4, a converting mechanism5connected with the drive shaft4and adapted to convert a rotational motion of the drive shaft4into a translational motion of a translatable portion6, as well as a hydraulic pump7connected to the converting mechanism5and adapted to generate, in response to the translational movement, an increase in pressure of a hydraulic liquid.

The hydraulic pump7comprises a cylinder8and a piston9received in the cylinder8and constrained to the translatable portion6so as to translate together with it with respect to the cylinder8along a piston stroke10extending from a rear end of stroke11to a front end of stroke12. The piston stroke10comprises a front length13including the front end of stroke12and having a length less than or equal to half the length of the piston stroke10, and a rear length14including the rear end of stroke11and having a length less than or equal to half the length of the piston stroke10.

The hydraulic pump7further comprises a pressure chamber15defined by the cylinder8and the piston9and having a volume that varies as a function of the position of the piston9from a maximum volume when the piston9is at the rear end of stroke position11to a minimum volume when the piston9is at the front end of stroke position12. The pressure chamber15is in communication with a supply duct16, which is preferably formed in the cylinder8(but it could also be formed in the piston9) and adapted to be connected with the hydraulic thrust unit of the brake2.

According to an aspect of the invention, the converting mechanism5is configured so that, for a given angular speed of the drive shaft4, the translation speed of the translatable portion6(i.e., of the piston9) decreases from a maximum value in a position of the piston9in the rear length14to a minimum value in a position of the piston9in the front length13.

This reduces the motor torque and thus the power supply required to generate a high fluid pressure in the front length13of the piston stroke10, thus obviating the risk of overheating the electric motor3under extreme braking situations. At the same time, when the piston9is in the rear length14of the piston stroke10(situation of standard operative braking actions), the actuator1allows an operation of the electric motor3in a motor torque capacity and absorbed power supply that is suitable for the dimensioning of the same motor.

Furthermore, by virtue of the fact that the minimum value of the transmission and conversion ratio VLIN,9/VANG,4(or, in other words: the minimum value of the quotient between the linear translation speed of the translatable portion6(or piston9) and the angular speed of the drive shaft4) is reached when the piston is in the front length13of the piston stroke10, it is possible to generate very high fluid pressures when the fluid itself has already undergone a high shift. In this manner, a rapid approach and engagement of the pads against the brake disc under normal braking conditions (rear length of the piston stroke) is obtained, and very high pressures can be generated when the pads have already engaged the brake disc and do not require further substantial shifts.

Furthermore, the converting mechanism5is configured so as to ensure also a controlled backward movement of the piston9from the front end of stroke (12) position to the rear end of stroke (11) position, as a response of a retro-movement of the motor3. Such reversibility of the motion is indispensable for the braking systems at issue.

According to an embodiment, the front length13of the piston stroke10has a length less than one third, preferably less than one fourth, of the length of the piston stroke10. Still more preferably, the minimum value of the transmission and conversion ratio VLIN,9/VANG,4(or, in other words: the minimum value of the quotient between the linear translation speed of the translatable portion6(piston9) and the angular speed of the drive shaft4) is reached when the piston9is at the front end of stroke12.

The maximum value of the transmission and conversion ratio VLIN,9/VANG,4(or, in other words: the maximum value of the quotient between the linear translation speed of the translatable portion6(piston9) and the angular speed of the drive shaft4) is preferably reached when the piston is at the rear end of stroke11.

According to a preferred embodiment, the converting mechanism5comprises a crank17and connecting rod18mechanism for a conversion of the rotational motion of the drive shaft4into the translational motion of the translatable portion6(piston9). In the crank and connecting rod mechanisms, the translation speed of the translatable portion6VLIN,9is a non-linear function of the angular position of the crank and also of the stroke of the translatable portion6. Considering now that the fluid pressure is approximately a linear function of the stroke of the piston9(i.e., the position of the piston9along the piston stroke10) and by combining such linear function of the pressure of the fluid with the non-linear function of the translation speed of the piston, a non-linear ratio between the motor torque and the stroke of the piston9is obtained, in which, in a prevailing portion of the rear length14(in the first half) of the piston stroke10, the motor torque increases in an approximately proportional (linear) manner as the piston advances9, and in the front length13(in the second half) of the piston stroke10, the motor torque increases in a sub-proportional (or decreasing) manner, or it decreases as the piston advances9, as indicated by the continuous curve in the diagram inFIG. 10. Just this characteristic of the converting mechanism5allows an optimal operation of the electric motor3and a precise control of the fluid pressure under normal braking conditions (with the piston9in the rear length14of the piston stroke10) and a very high pressurization of the fluid under exceptional braking conditions (with the piston9in the front length13of the piston stroke10).

Since at least some of the above-mentioned advantages of the so-configured converting mechanism5can be obtained by alternative kinematics, some of which will be described by way of exemplary, non-limiting example only herein below, an aspect of the invention also relates to the general technical concept whereby, in a prevailing portion of the rear length14(in the first half) of the piston stroke10, the motor torque increases in an approximately proportional (linear) manner as the piston advances9, and in the front length13(in the second half) of the piston stroke10, the motor torque increases in a sub-proportional (or decreasing) manner, or it decreases as the piston advances9in the direction of the front end of stroke12. The concept of a “sub-proportional” trend of mathematical functions is sometimes also referred to by the term “sub-linear”.

In accordance with an embodiment, the actuator comprises a housing19receiving the converting mechanism5and supporting or receiving the electric motor3, the cylinder8, as well as a reservoir20for the pressurized fluid.

The converting mechanism5may comprise an (optional) reduction unit22to demultiply the rotational motion of the drive shaft4and a converting unit30to convert the rotational motion into a translational motion. Furthermore, first transmission means21of the rotational motion of the drive shaft4to the reduction unit (preferably epicycloidal) and second transmission means23of the rotational motion of the reduction unit to the converting unit30can be provided.

According to an embodiment, the first transmission means21comprise a connecting portion formed on the end of the drive shaft4and having an outer toothing meshing with satellite toothed wheels24of a first train of satellites of the epicycloidal reduction unit22, so that the drive shaft4forms the central pinion of a first reduction stage of the epicycloidal reduction unit22. The epicycloidal reduction unit22preferably comprises two reduction stages, the first reduction stage of which includes the above-mentioned central pinion formed by the end of the drive shaft4, an outer crown25with internal toothing locked in rotation and the above-mentioned first train of satellite toothed wheels24meshing with both the drive shaft4and the crown25. The first satellite toothed wheels24are supported by a first satellite bearing plate26which, in turn, comprises a toothed central portion27composing a central pinion (planet) of a second reduction stage.

Such second reduction stage comprises, in addition to the second central pinion (planet) formed by the toothed central portion27of the first satellite bearing plate26, the same toothed crown25and a second train of satellite toothed wheels28meshing with both the second central pinion27and the crown25.

The second satellite toothed wheels28are supported by a second satellite bearing plate29implementing the connection with the converting unit30.

The toothed crown25can be manufactured as a distinct part from the housing19and subsequently inserted therein. This allows optimizing the thicknesses and the materials of the toothed crown25and the housing19independently from one another to reduce the weight and manufacturing cost of the actuator1. In particular, the housing19could be manufactured in a material different from that of the toothed crown, for example, in a plastic material or by die casting in an aluminium/magnesium alloy.

By way of non-limiting example, the housing19can be in plastic material injection moulded over the toothed crown25in metallic material, for example, steel, reducing the tolerances and assembling costs of the actuator1.

The electric motor3may comprise a front flange31connectable to the housing19, for example, by means of screws insertable in holes obtained in the front flange31and that can be screwed in two internally threaded holes of the housing19. The front flange31may comprise centering means, for example, engaging seats or surfaces, adapted to engage the toothed crown25so as to position and center it with respect to the drive shaft4. In this manner, an expensive precision mechanical machining of a plurality of inner surfaces of the housing19is obviated. The front flange31of the electric motor3could form also one or more protuberances to lock in rotation the toothed crown25. Alternatively, the means to lock the toothed crown25in rotation (e.g., of the recess-tooth, or the recess-key type) are formed and act between the toothed crown25and the housing19.

The second satellite bearing plate29comprises a shaft portion34for the transmission of the rotational motion in output from the epicycloidal reduction unit (and precisely of the rotational motion of the second satellite bearing plate29) to the converting unit30. The shaft portion34is connected to the housing19and centered by a first volving radial bearing32(ball-bearing, reducer side) and a second plain radial bearing32′ (converting unit side) received in corresponding seats33,33′ of the housing19.

The shaft portion34engages the crank17of the crank17-connecting rod18mechanism, so as to be able to apply a torque to the crank17and to make it rotate (FIGS. 3, 4, 5).

The crank17itself is connected to the housing19, preferably by a roller bearing37(for example, a roll- or needle-bearing), so as to be able to rotate around a crank fulcrum or axis38transversal, preferably perpendicular, to the translation direction39of the piston9. The connecting rod18has a preferably elongate shape, with a first end40hinged to the crank17in an eccentric point (at a radius R, seeFIGS. 6 and 9) with respect to the crank fulcrum38, and a second end forming the translatable portion6and permanently engaging a rear portion41of the piston9.

The shaft portion34can form or support a toothed wheel or pinion35meshing with a toothed sector36of a crank37. When, as illustrated in the Figures, the toothed sector36has a radius5greater than that of the pinion35, the rotation of the shaft portion34not only rotates the crank17around the crank fulcrum38, but also demultiplies the rotational speed thereof.

As it can be noticed inFIGS. 5 and 6, with the piston9at the rear end of stroke11position (FIG. 5), the rotational angle of the crank17Theta (θ) defined as the angle between the translation axis39of the piston9and the plane42defined by the crank axis38and the hinge axis of the first end40of the connecting rod18, is maximum (and preferably ranging from 75° to 110°, still more preferably from 80° to 100°, and still more preferably from 84° to) 92°. With the piston at the position of front end of stroke12(FIG. 6), such rotational angle of the crank17Theta (θ) is minimum (and preferably ranging from 3° to 35°, still more preferably from 15° to 30°, and still more preferably da 22° a 28°).

In order to obtain that in a prevailing portion of the rear length14(in the first half) of the piston stroke10the motor torque increase in an approximately proportional (linear) manner with the advancement stroke of the piston9, and in the front length13(in the second half) of the piston stroke10, the motor torque increases in a sub-linear (or decreasing) manner, or it decreases as the piston advances9in direction of the front end of stroke12, and to be able to set the desired trends of the motor torque in the approximately linear phase and in the sub-linear phase, it can be advantageous to arrange the crank axis38at a perpendicular distance from the translation axis39of the piston9. Such perpendicular distance is preferably less than the distance between the crank axis38and the hinge axis of the first end40of the connecting rod18(radius R inFIG. 5). Furthermore, such perpendicular distance preferably ranges from 0.4L to 0.6L, where L is the distance between the first and the second ends of the connecting rod18(length connecting rod L).

One or more return elastic springs57connected between the crank17and the housing19ensure an automatic return of the crank17and of the piston9to the rear end of stroke position11(position ofFIG. 5) in the case of a failure or operative stop of the electric motor.

The cylinder8may comprise a connecting flange43connectable to the housing19, for example, by means of screws insertable in holes obtained in the connecting flange43and that can be screwed in two internally threaded holes of the housing19(FIGS. 2, 5). Alternatively, the cylinder8can be directly formed by the housing19, or manufactured separately and subsequently inserted in a cylinder seat of the housing19.

The piston9is slidably received in the cylinder10by the interposition of a first gasket in the proximity of a front portion44thereof and of a second gasket in the proximity of the rear portion41thereof. The rear portion41of the piston9has a rounded cavity receiving in pressing contact and rotatably the second end (translatable portion6) of the connecting rod18. Such pressing contact between the rear portion41of the piston9and the connecting rod18is ensured by a return spring45arranged with elastic pre-load in the pressure chamber15between a bottom wall of the cylinder8and the front portion44of the piston, so as to bias the piston9permanently elastically towards the position of rear end of stroke11.

The supply duct16is formed in a side wall of the cylinder10. Furthermore, a suction duct46is provided, which puts the pressure chamber15in communication with the reservoir20of the pressurized fluid. Such suction duct46,46′ can be formed in the cylinder8and/or in the piston9. In the embodiment illustrated in the Figures, a side surface of the piston9forms a (preferably circumferential) cavity47defining, together with the side wall of the cylinder8a (preferably annular) suction chamber. A first portion46of the suction duct extends from the reservoir20through the side wall of the cylinder8and opens to the cylinder side in the suction chamber, and a second portion46′ of the suction duct is provided with a valve48and extends from an opening in the front portion44of the piston9through the piston up to the cavity47. Such valve48allows a free communication between the pressure chamber15and the reservoir20with the piston at the rear end of stroke position, while preventing the fluid passage (in both directions) when the piston exceeds a given initial distance (idle stroke) from the rear end of stroke, thus allowing a pressurization of the pressure chamber15and a thrust of the fluid towards the thrust unit of the brake.

For the control of the electro-hydraulic actuator1:a control unit58to control the supply of the electric motor3,a drive member60, for example, a lever or a pedal,a drive sensor62, for example, a linear or rotating (potentiometric or magnetostrictive) transducer, connected to the drive member60and in signal communication with the control unit58, such drive sensor60being configured to generate a required braking torque signal as a function of a shift of the drive member60and to transmit the required braking torque signal to the control unit,a pressure sensor59connected to the pressure chamber15and in signal communication with the electronic control unit58, such pressure sensor59being configured to generate a pressure signal as a function of the fluid pressure in the pressure chamber15and to transmit the pressure signal to the control unit, may be providedin which the control unit58is configured to receive and process the pressure signal and the required braking torque signal and to control the power supply of the motor3as a function of the required braking torque and the detected fluid pressure.

The present invention also relates to the single brake2, which comprises the electro-hydraulic actuator1for the supply of the hydraulic thrust unit thereof.

The present invention further relates to a braking system56(FIG. 11) comprising a drive member60, a plurality of hydraulic brakes2with hydraulic thrust units (and optionally one or more electro-mechanical brakes61), in which each hydraulic brake2comprises its own electro-hydraulic actuator1connected to its own hydraulic thrust unit, and in which a control unit58is provided, which is configured to receive and process the required braking torque signals and the pressure signals of each electro-hydraulic actuator1and to control in an individual and targeted manner the power supply of each single motor3for each of the hydraulic brakes2as a function of the required braking torque and the detected fluid pressures.

This allows generating the braking power in an individual and targeted manner for each braked wheel and, when desired, also independently from the braking power required for the other wheels of the vehicle.

This results in the possibility of a targeted and individual optimization of the braking action of every single wheel, for example, under extreme conditions in which an anti-locking modulation (ABS) or curve stabilization of the vehicle (EPS) is required, also independently from the position of the brake pedal.

FIGS. 12 and 13illustrate two alternative embodiments of the invention.

In accordance with an alternative embodiment (FIG. 12) the converting mechanism5may comprise a cam mechanism, for example:a substantially circular cam48rotatably connected to the housing19around a cam fulcrum or axis49that is eccentric with respect to the center of a circumferential cam surface50thereof, said cam48being engaged by the reduction unit22or directly by the drive shaft4so as to be able to apply a torque to the cam48and rotate it,a translatable cam follower portion51connected to the piston9and in engagement with the cam surface50, in which the cam is shaped so as to obtain the characteristics of transmission ratio and the motor torque-piston stroke dependence as described before with reference to the first embodiment.

In accordance with a further alternative embodiment, the converting mechanism may comprise a rotor52coupled with a stator53by the engagement between volving members54connected to one of them and an helicoidal track55formed in the other one, in which the angle (or pitch) of the coil5of the helicoidal track55varies along the length thereof, so as to obtain the characteristics of transmission ratio and the motor torque-piston stroke dependence as described before with reference to the first embodiment.

In addition to the described advantages, the system according to the invention ensures the reversibility of the motion that is indispensable for the braking systems at issue, it is adapted to a control by means of sensors, and it is characterized by a law of numerically univocal motion and easily numerically simulable, and it reduces the power supply absorption compared to the prior art.

It shall be apparent that to the electro-hydraulic actuator, to the hydraulic brake, and to the braking system according to the present invention, those skilled in the art, in order to meet contingent, specific needs, will be able to make further modifications and variations, all of which are anyhow contained in the protection scope of the invention, as defined by the following claims.