JP6592356B2 - Brake device actuator and brake device including the same - Google Patents

Description

  The present invention relates to an actuator for a brake device and a brake device including the same.

  Patent Document 1 discloses an actuator for a brake device in which a piston that supports a brake member that is pressed against a brake disk is configured by a base end side piston and a front end side piston. In the brake device actuator of Patent Document 1, the front end side piston is supported by a receiving hole formed on the front end side of the base end side piston via a spherical bearing, and is configured to be tiltable with respect to the base end side piston. The

Japanese Patent Laid-Open No. 10-267058

  In the actuator for a brake device of Patent Document 1, the tip side piston is configured to be tiltable with respect to the base side piston, so that even when tilting occurs between the brake and the brake disk, the brake is controlled. The friction surface of the wheel can be brought into contact with the brake disk in parallel.

  However, in the brake device actuator disclosed in Patent Document 1, the front-side piston is tilted with respect to the proximal-side piston by sliding the outer ring and the inner ring of the spherical bearing. For this reason, at the time of braking, due to the sliding resistance of the spherical bearing, energy loss occurs in the pressing of the restrictor against the brake disk.

  The present invention has been made in view of the above problems, and an object thereof is to provide an actuator for a brake device and a brake device that reduce energy loss during braking.

A first invention is an actuator for a brake device used in a brake device that brakes a rotating body by a frictional force generated in a frictional force generating unit, the piston being pressed against the frictional force generating unit to generate the frictional force, A first screw member to which the piston is attached; a second screw member screwed into the first screw member; an electric motor that rotates the second screw member to linearly move the first screw member; and a passage hole through which the piston passes. A first screw member and a housing that accommodates the second screw member , and the first screw member includes a main body portion formed with a screw portion that is screwed with the second screw member, and an outer diameter of the main body portion. An insertion shaft portion through which the piston is inserted with a smaller outer diameter, and a step portion formed between the main body portion and the insertion shaft portion, and the piston is between the step portion and the axial direction. Insert shaft with clearance in the axial direction With attached, attached to the insertion shaft portion with a radial clearance between the radial direction between the insertion shaft portion and the apertures, the relative movement in the axial direction and the radial direction of the first threaded member relative to the first screw member It is possible to be provided .

According to a second aspect of the present invention, the insertion shaft portion is provided with a retaining member that prevents the piston from coming off from the insertion shaft portion .

In the first and second inventions, since the piston is provided so as to be relatively movable in both the axial direction and the radial direction of the first screw member, the central axis of the first screw member does not slide with the first screw member. Can be tilted with respect to a plane perpendicular to. Thus, the piston can be tilted without sliding with the first screw member, and the piston can be brought into contact with the frictional force generating portion in parallel.

The third invention is further characterized by further comprising a dust seal that closes a gap between the inner peripheral surface of the passage hole and the outer peripheral surface of the piston.

According to the third aspect of the invention, it is possible to prevent the dust from entering the inside of the housing and prevent the movement of the first screw member from being hindered by the dust.

The fourth invention further includes a carrier that supports the housing so that the piston faces the frictional force generating part, and the frictional force generating part is provided with a rotating disk that rotates together with the rotating body and a non-rotatable rotating disk. A non-rotating disk that is displaceable toward the disk, and a disk support member that receives the pressing force of the piston that is provided on the opposite side of the piston across the rotating disk and the non-rotating disk and is applied to the rotating disk and the non-rotating disk. The axial clearance and radial clearance are between the disc support member and the carrier when the piston is pressed against the frictional force generating part from the non-braking state where the piston is separated from the frictional force generating part. The piston is set to a size that can swing with respect to the axis of the first screw member by the change in angle.

In the fourth aspect of the invention, the piston can be tilted so as to follow the inclination of the frictional force generating part in the braking state.

According to a fifth aspect of the present invention, there is provided a rotation restricting portion for restricting the rotation of the first screw member by locking the base end portion of the first screw member on the opposite side in the axial direction from the insertion shaft portion to which the piston is coupled. Is further provided.

In the fifth aspect of the invention, the rotation restriction of the piston and the first screw member and the guide of the movement in the axial direction are performed by the rotation restriction portion that locks the proximal end portion of the first screw member in the circumferential direction. Therefore, the radial gap between the piston and the passage hole can be an appropriate gap in order to follow the inclination of the frictional force generating portion.

A sixth aspect of the invention includes any one of the first to sixth brake device actuators and a frictional force generator that generates a frictional force that brakes the rotating body.

  ADVANTAGE OF THE INVENTION According to this invention, the energy loss at the time of braking in the actuator for brake devices and a brake device is reduced.

1 is a plan view of a brake device according to an embodiment of the present invention. FIG. 2 is a cross-sectional view taken along line A-O-A in FIG. It is an expanded sectional view showing the structure around the piston of the actuator for brake devices concerning the embodiment of the present invention. It is a top view which shows the guide stopper mechanism of the actuator for brake devices which concerns on embodiment of this invention. It is an expanded sectional view showing the state where the piston of the actuator for brakes concerning an embodiment of the present invention contacted the non-rotating disk. It is an expanded sectional view showing the structure around the piston of the actuator for brake devices concerning the 1st modification of the present invention. It is an expanded sectional view showing the structure around the piston of the actuator for brake devices concerning the 2nd modification of the present invention. It is sectional drawing of the actuator for brake devices which concerns on the 1st comparative example of this invention. It is an expanded sectional view showing the structure around the piston of the actuator for brake devices concerning the 2nd comparative example of the present invention.

  Hereinafter, a brake device and an actuator for a brake device according to an embodiment of the present invention will be described with reference to the drawings.

  With reference to FIGS. 1-7, the brake device and actuator for brake devices which concern on embodiment are demonstrated.

  Hereinafter, the aircraft brake device 101 and the brake device actuator 100 that brake the vehicle body by braking a wheel (not shown) mounted on a fixed wing aircraft will be described. Hereinafter, the brake device actuator 100 is simply referred to as an “actuator 100”.

  The brake device 101 brakes the wheel using the frictional force generated by the frictional force generator 102 as a braking force. The brake device 101 is driven when a brake pedal (not shown) mounted on the aircraft is operated by an operator.

  As shown in FIGS. 1 and 2, the brake device 101 controls the frictional force generating unit 102 by pressing the piston 10 against the frictional force generating unit 102 and the frictional force generating unit 102 that generates a frictional force that brakes the wheel. And four actuators 100 that generate frictional force as power. The actuators 100 are driven in synchronism with each other by a controller (not shown), and press the piston 10 against the frictional force generator 102, respectively.

  Aircraft wheels are configured by attaching tires (not shown) to the outer periphery of wheels (not shown). The wheel is attached to an axle (not shown) via a bearing (not shown). That is, the axle does not rotate even if the wheel rotates.

  As shown in FIG. 2, the frictional force generator 102 includes a plurality of rotating disks 103 that rotate together with the wheels, a non-rotating disk 104 that is non-rotatably attached to the axle via a torque tube (not shown), and the actuator 100. And a backing plate 105 as a disk support member for supporting the pressing force of the piston 10 applied to the rotating disk 103 and the non-rotating disk 104. When the piston 10 is pressed against the non-rotating disc 104 by the actuator 100, the frictional force generating unit 102 abuts the rotating disc 103 and the non-rotating disc 104 to generate a frictional force as a braking force.

  A plurality of rotating disks 103 are provided at a predetermined interval. A non-rotating disk 104 is provided on both surfaces of the rotating disk 103. Therefore, a plurality of non-rotating disks 104 are also provided at a predetermined interval. The frictional force generator 102 is arranged so that two rotating disks 103 are sandwiched between three non-rotating disks 104. The number of each of the non-rotating disc 104 and the rotating disc 103 is not limited to this, and is arbitrarily set according to the required braking force.

  The rotating disk 103 is formed in a disk shape. The outer periphery of the rotating disk 103 is supported by the inner periphery of the wheel, and rotates as a unit with the wheel. The rotating disk 103 is supported so as to be displaceable in the axle direction (up and down direction in FIG. 2) with respect to the wheel, that is, slidable.

  The non-rotating disk 104 is formed in a disk shape. The inner periphery of the non-rotating disk 104 is supported on the axle via a torque tube or the like. The non-rotating disc 104 does not rotate even if the wheel rotates. The non-rotating disc 104 is supported to be displaceable in the axle direction, that is, slidable.

  The backing plate 105 is provided on the opposite side of the piston 10 with the rotating disk 103 and the non-rotating disk 104 interposed therebetween. The backing plate 105 abuts on the non-rotating disc 104 located farthest from the piston 10 and supports the rotating disc 103 and the non-rotating disc 104. That is, the backing plate 105 supports the rotating disk 103 and the non-rotating disk 104 from the side opposite to the piston 10 and receives the pressing force by the actuator 100. Thereby, the inclination of the rotating disk 103 and the non-rotating disk 104 due to the pressing force of the actuator 100 is suppressed. The backing plate 105 is integrally formed with a torque tube to which the non-rotating disk 104 is attached.

  The actuator 100 includes a piston 10 that is pressed against the non-rotating disk 104 of the frictional force generator 102, a housing 20 having a bottom housing 21 in which a passage hole 22A through which the piston 10 passes is formed, and a first to which the piston 10 is coupled. A ball screw 31 as a screw member, a ball nut 35 as a second screw member screwed to the ball screw 31, an electric motor 25 as a motor for rotating the ball nut 35 to linearly move the ball screw 31, and electric A speed reduction unit 40 that decelerates the rotation of the motor 25 and transmits it to the ball nut 35, a guide stopper mechanism 60 as a rotation regulation unit that regulates the rotation of the ball screw 31, and a carrier 50 attached to the housing 20.

  In the actuator 100, the ball screw 31 and the ball nut 35 constitute a linear motion mechanism 30 that moves the piston 10 forward and backward with respect to the non-rotating disk 104. That is, the linear motion mechanism 30 is a ball screw mechanism that converts rotation transmitted from the electric motor 25 into linear motion of the piston 10.

  The piston 10 is provided to face the non-rotating disk 104 so as to be able to advance and retreat toward the non-rotating disk 104. The piston 10 is provided with a predetermined gap between the piston 10 and the non-rotating disk 104 in a non-braking state (the state shown in FIG. 2) in which the brake device 101 does not generate a braking force. That is, the piston 10 faces the non-rotating disk 104 with a predetermined gap.

  Further, the piston 10 has a circular cross section with a cross section perpendicular to the central axis. The circular cross-sectional shape is a substantially perfect circle and means that the width of the piston 10 across the central axis is substantially uniform in the circumferential direction.

  The housing 20 accommodates the linear motion mechanism 30 and the speed reduction unit 40. The housing 20 is decelerated between the bottom housing 21 attached to the carrier 50, the first cover member 23 that houses the linear motion mechanism 30 between the bottom housing 21, and the electric motor 25 and the bottom housing 21. And a second cover member 24 that accommodates the portion 40.

  The bottom housing 21 has a boss 22 that is inserted into a mounting hole 50 </ b> A formed in the carrier 50. A passage hole 22 </ b> A through which the piston 10 passes is formed in the boss portion 22.

  The passage hole 22A is formed in a shape corresponding to the cross-sectional shape of the piston 10. That is, like the piston 10, the passage hole 22A is formed in a circular cross-sectional shape.

  As shown in FIG. 3, a radial gap T1 is formed between the outer peripheral surface of the piston 10 and the inner peripheral surface of the passage hole 22A. A radial gap T1 between the piston 10 and the passage hole 22A is closed by the dust seal 15. By providing the dust seal 15 on the outer periphery of the piston 10, the piston 10 is positioned in the radial direction. The dust seal 15 is formed in an annular shape according to the cross-sectional shape of the piston 10, and is formed of a member having high heat resistance. Thus, since the dust seal of the actuator 100 is annular, a general-purpose product can be used, and the cost is reduced. The dust seal 15 is configured by an elastic member so as not to hinder the later-described tilting (relative movement in the radial direction) of the piston 10.

  The dust seal 15 is a process in which the piston 10 moves from the retracted position (the state shown in FIG. 2) that is the farthest from the non-rotating disk 104 to the braking position (for example, the state shown in FIG. 5) pressed against the non-rotating disk 104. In FIG. 4, the piston 10 is provided at a position that always faces the outer peripheral surface of the piston 10. Specifically, the dust seal 15 is provided as close as possible to the non-rotating disk 104 in a state where the piston 10 is in the retracted position. As a result, the dust seal 15 is prevented from falling off due to the movement of the piston 10, and dust can be reliably prevented from entering the housing 20 through the space between the piston 10 and the passage hole 22A. Therefore, malfunction of the linear motion mechanism 30 due to dust is prevented.

  The first cover member 23 is attached to the bottom housing 21 and accommodates the linear motion mechanism 30 together with the bottom housing 21. The second cover member 24 is attached to the bottom housing 21 and accommodates the speed reducing portion 40 together with the bottom housing 21. The first cover member 23 and the second cover member 24 are attached to the bottom housing 21 so as not to contact each other.

  The electric motor 25 is provided such that the rotating shaft 26 is inclined from the moving direction of the piston 10 by the linear motion mechanism 30 (vertical direction in FIG. 2), and is attached to the second cover member 24 of the housing 20. Specifically, the electric motor 25 is provided so that the rotating shaft 26 is orthogonal to the moving direction of the piston 10, in other words, the rotating shaft 26 extends in the left-right direction in FIG.

  A rotation sensor (not shown) that detects the rotation speed of the rotation shaft 26 is attached to the electric motor 25. The rotation sensor may be built in the electric motor 25 or may be externally attached. The detection result of the rotation sensor is output to the controller.

  The ball screw 31 includes a main body portion 32 on which a male screw portion 32A is formed, an insertion shaft portion 33 which is an end portion to which the piston 10 is attached, and a base end portion 34 provided on the opposite side in the axial direction from the insertion shaft portion 33. Have.

  As shown in FIG. 3, the insertion shaft portion 33 has an outer diameter smaller than the outer diameter of the main body portion 32. Thereby, an annular step portion 33 </ b> A is formed between the insertion shaft portion 33 and the main body portion 32. The insertion shaft portion 33 is inserted into the shaft portion insertion hole 11 formed in the piston 10. The outer diameter of the insertion shaft portion 33 is formed smaller than the inner diameter of the shaft portion insertion hole 11. Therefore, a radial gap T <b> 2 is formed between the outer peripheral surface of the insertion shaft portion 33 and the inner peripheral surface of the shaft portion insertion hole 11. A radial clearance T2 between the insertion shaft portion 33 of the ball screw 31 and the shaft insertion hole 11 of the piston 10 is set to be approximately the same size as the radial clearance T1 between the piston 10 and the passage hole 22A. . The radial gaps T1 and T2 do not have to be substantially the same size.

  A C-ring (snap ring) 37 as a retaining member that prevents the piston 10 from coming off from the insertion shaft portion 33 is provided at the tip of the insertion shaft portion 33. The C ring 37 is accommodated in a ring recess 12 formed on an end surface of the piston 10 that is pressed against the non-rotating disk 104 (hereinafter referred to as “pressing surface 10 </ b> A”). The C-ring 37 has an axial length (distance) L1 from the step portion 33A of the ball screw 31 greater than an axial length L2 between the end surface of the piston 10 facing the step portion 33A and the bottom surface of the ring recess 12. Is also provided in the insertion shaft portion 33 so as to be larger. For this reason, an axial clearance T3 is formed between the piston 10, the step portion 33A, and the C ring 37.

  As described above, the piston 10 forms the radial gaps T1 and T2 between the passage hole 22A and the insertion shaft part 33, and the axial gap T3 between the step part 33A and the C ring 37. Form. Therefore, the piston 10 can move relative to the ball screw 31 both in the axial direction and in the radial direction. By being provided so as to be relatively movable in both the axial direction and the radial direction, the piston 10 can be tilted with respect to a plane perpendicular to the axis of the ball screw 31.

  As shown in FIG. 2, the ball nut 35 has a female screw portion 35 </ b> A that engages with a male screw portion 32 </ b> A of the ball screw 31, and is supported by the bottom housing 21 of the housing 20 via a radial bearing 46 described later.

  When the rotation of the electric motor 25 is transmitted and the ball nut 35 rotates, the ball screw 31 moves linearly in the axial direction together with the piston 10 according to the rotation. Thus, when the electric motor 25 rotates, the piston 10 advances and retreats with respect to the non-rotating disc 104 through the passage hole 22A by the linear motion mechanism 30. The position of the ball screw 31 is detected by a position sensor (not shown) and output to the controller.

  The speed reduction unit 40 is configured by a combination of a plurality of gear mechanisms, and decelerates the rotation of the rotating shaft 26 of the electric motor 25 and converts it into rotation of the ball nut 35 of the linear motion mechanism 30. The speed reduction unit 40 transmits the rotation to the output gear 43 from the bevel gear mechanism 41 to which the rotation of the rotating shaft 26 of the electric motor 25 is transmitted, the output gear 43 connected to the outer periphery of the ball nut 35, and the bevel gear mechanism 41. And a transmission unit 44 for transmitting.

  The bevel gear mechanism 41 includes a first bevel gear 41 </ b> A connected to the rotation shaft 26 of the electric motor 25, and a second bevel gear 41 </ b> B that meshes with the first bevel gear 41 </ b> A. In the bevel gear mechanism 41, the rotation axis of the first bevel gear 41A that is the input side gear and the rotation axis of the second bevel gear 41B that is the output side gear are orthogonal to each other.

  The output gear 43 is interposed between the first cover member 23 and the bottom housing 21 via a thrust bearing 45 provided between the first cover member 23 and a radial bearing 46 provided between the bottom housing 21. It is a spur gear supported by. As the output gear 43 rotates, the ball nut 35 rotates.

  The thrust bearing 45 supports a reaction force generated by pressing the piston 10 against the non-rotating disk 104 acting on the ball nut 35 via the output gear 43. Thereby, the reaction force by the pressing of the piston 10 acts, and the rotation of the output gear 43 is not hindered, and the output gear 43 can be smoothly rotated.

  The radial bearing 46 is provided in the bottom housing 21 while being sandwiched in the axial direction between the output gear 43 and the bottom housing 21. The radial bearing 46 is provided on the outer periphery of the ball nut 35 and rotatably supports the ball nut 35. When the ball nut 35 is supported by the radial bearing 46, the ball nut 35 and the ball screw 31 screwed to the ball nut 35 are positioned in the radial direction.

  The transmission unit 44 transmits the rotation of the second bevel gear 41 </ b> B to the output gear 43. The transmission unit 44 includes a plurality of gear mechanisms. The transmission part 44 should just be what transmits the rotation of the 2nd bevel gear 41B to the output gear 43, and can be set as arbitrary structures.

  The speed reduction unit 40 has a bevel gear mechanism 41 configured by first and second bevel gears 41A and 41B whose axes are orthogonal to each other, and the axis of rotation of the electric motor 25 is parallel to the moving direction of the piston 10. Convert to rotation around. Therefore, even if the rotating shaft 26 of the electric motor 25 is orthogonal to the moving direction of the piston 10, the rotation of the rotating shaft 26 is converted into the rotation of the ball nut 35 of the linear motion mechanism 30 to move the piston 10 linearly. Can do. The reduction ratio of the reduction unit 40 is appropriately set according to the specifications of the electric motor 25 and the required braking force.

  The guide stopper mechanism 60 restricts the ball screw 31 from rotating together with the ball nut 35 by locking the base end portion 34 of the ball screw 31 in the circumferential direction. Specifically, as shown in FIGS. 2 and 4, the guide stopper mechanism 60 includes a guide plate 61 attached to the base end portion 34 of the ball screw 31 and received in the receiving recess 23 </ b> A, and the central axis of the ball screw 31. And two guide pins 65 that are attached to the housing 20 and locked in the circumferential direction.

  As shown in FIG. 4, the guide plate 61 is a disk-shaped member, and is provided in the housing recess 23A with a radial gap between the guide plate 61 and the inner periphery of the housing recess 23A. The guide plate 161 has an insertion hole 62 into which the base end portion 34 is inserted, a fixing hole 63 that penetrates the guide plate 61 in the radial direction via the insertion hole 62, and a guide pin 65 that is inserted into the guide pin 65 and around the guide plate 65. A pair of guide holes 64 that can be locked in the direction are formed.

  The base end portion 34 is formed with a base end fixing hole 34A having substantially the same size as the fixing hole 63 of the guide plate 61 and penetrating in the radial direction. The guide plate 61 and the base end portion 34 are connected to each other by a fixing pin 66 that is press-fitted through the fixing hole 63 of the guide plate 61 and the base end fixing hole 34A of the base end portion 34 that are aligned in a straight line.

  In a state where the guide plate 61 and the base end portion 34 are connected, the end surface 34B of the base end portion 34 is on the piston 10 side (lower side in FIG. 2) with respect to the end surface 61A of the guide plate 61 facing the bottom of the housing recess 23A. Located in. As a result, when the ball screw 31 moves upward in FIG. 2 so as to separate the piston 10 from the non-rotating disk 104, the end surface 34B of the base end 34 does not contact the bottom of the housing recess 23A, and the guide plate 61 end face 61A contacts.

  The guide plate 61 can have a large cross-sectional area as long as space permits, and can increase the pressure receiving area that comes into contact with the bottom of the housing recess 23A. For this reason, it is possible to easily prevent deformation of the bottom of the housing recess 23A by bringing the end surface 61A of the guide plate 61 into contact with the bottom of the housing recess 23A without contacting the end surface 34B of the base end portion 34. it can.

  As shown in FIG. 4, the pair of guide hole portions 64 are notches formed by cutting out from the outer periphery of the guide plate 61 in an arc shape inward in the radial direction, and penetrate the guide plate 61 in the thickness direction. The pair of guide hole portions 64 are formed at positions facing each other across the central axis of the ball screw 31.

  The two guide pins 65 are respectively inserted into the pair of guide hole portions 64 with a gap in the radial direction. The guide pin 65 is a rod-shaped member having a circular cross section, and is press-fitted and fixed to the bottom of the housing recess 23A.

  When the ball screw 31 tries to rotate together with the guide plate 61, the guide hole 64 and the guide pin 65 of the guide plate 61 are locked in the circumferential direction, so that the rotation of the ball screw 31 is restricted. Further, when the guide plate 61 moves in the axial direction together with the ball screw 31, the guide hole portion 64 of the guide plate 61 can slide with respect to the guide pin 65. Therefore, the rotation of the ball screw 31 is restricted when the guide hole 64 of the guide plate 61 connected to the base end portion 34 is locked with the guide pin 65 in the circumferential direction, and the guide hole 64 is guided by the guide hole 64. By sliding in the axial direction with respect to the pin 65, it is possible to move in the braking direction while being guided in the axial direction.

  As shown in FIG. 2, the outer diameter of the guide plate 61 is formed larger than the inner diameter of the ball nut 35. When the ball screw 31 is stroked in the braking direction by a predetermined amount, the guide plate 61 comes into contact with the end face of the ball nut 35, and the stroke of the ball screw 31 in the further braking direction is restricted. This prevents the ball screw 31 from coming off in the axial direction. That is, the guide plate 61 also functions as a retainer that prevents the ball screw 31 from coming off in the axial direction.

  As described above, the guide stopper mechanism 60 functions as a stopper of the ball screw 31 in both the braking direction and the release direction and a guide for guiding the linear motion of the ball screw 31 in addition to the function of restricting the rotation of the ball screw 31. As a single function.

  The carrier 50 is attached to the bottom housing 21 via bolts (not shown). As shown in FIG. 1, the carrier 50 is configured so that the electric motor 25, the linear motion mechanism 30, and the speed reduction unit 40 do not directly face the non-rotating disk 104 of the frictional force generation unit 102 except for the mounting hole 50 </ b> A. The size is formed so as to cover all of these. Thereby, since the frictional heat generated in the frictional force generating unit 102 is blocked by the carrier 50, the influence of the frictional heat on the electric motor 25, the linear motion mechanism 30, and the speed reducing unit 40 is reduced. Therefore, durability of the electric motor 25, the linear motion mechanism 30, and the speed reduction part 40 is improved.

  The carrier 50 is formed by a member having a lower thermal conductivity than the housing 20. Thereby, the frictional heat generated in the frictional force generating unit 102 is effectively cut off, and the durability of the electric motor 25 and the speed reduction unit can be improved. The carrier 50 is not limited to a member having a lower thermal conductivity than the housing 20, and may be formed of any material.

  Next, the operation of the actuator 100 and the brake device 101 will be described.

  Hereinafter, the direction in which the piston 10 moves toward the non-rotating disk 104 (downward direction in FIG. 2) is referred to as “braking direction”, and the direction in which the piston 10 is separated from the non-rotating disk 104 (upward direction in FIG. ". The rotation direction of the electric motor 25 that moves the piston 10 in the braking direction is referred to as “forward direction”, and the rotation direction of the electric motor 25 that moves the piston 10 in the release direction is referred to as “reverse direction”.

  In the non-braking state in which the amount of operation of the brake pedal is zero, as shown in FIG. 2, the piston 10 is separated from the non-rotating disc 104 with a predetermined gap (retracted position).

  When the brake pedal is depressed by the operator from the non-braking state, the controller outputs a command signal to the electric motor 25 so as to rotate in the forward direction according to the operation amount of the brake pedal. Specifically, the controller instructs the electric motor 25 to exert a desired braking force based on the detection results of the position sensor that detects the position of the ball screw 31 and the rotation sensor that detects the rotation of the electric motor 25. Output a signal.

  The rotation of the electric motor 25 in the forward direction is decelerated by the bevel gear mechanism 41, the transmission unit 44, and the output gear 43 of the reduction unit 40 and transmitted to the ball nut 35 of the linear motion mechanism 30. When the rotation of the electric motor 25 is transmitted and the ball nut 35 rotates, the ball screw 31 moves in the braking direction together with the piston 10.

  At this time, the rotation of the ball screw 31 is restricted by the guide hole 64 of the guide plate 61 and the guide pin 65 being locked in the circumferential direction, and the guide hole 64 is pivoted with respect to the guide pin 65. By sliding in the direction, it moves in the braking direction while being guided in the axial direction. Thereby, the ball screw 31 can be stably linearly moved.

  When the ball screw 31 linearly moves in the braking direction, the pressing surface 10A of the piston 10 is pressed toward the non-rotating disc 104, and the non-rotating disc 104 and the rotating disc 103 come into contact with each other (braking position). In this way, the rotating disk 103 and the non-rotating disk 104 are pressed by the piston and the backing plate 105, and the frictional force generating unit 102 generates a frictional force, thereby exerting a desired braking force. In addition, since the non-rotating disk 104 is provided so as not to rotate, no frictional force is generated between the piston 10 and the non-rotating disk 104.

  Here, in order to facilitate understanding of the present invention, actuators 200 and 300 according to comparative examples will be described with reference to FIGS. 8 and 9. In the actuator 200 according to the first comparative example shown in FIG. 8, the piston 110 is not relatively movable in both the axial direction and the radial direction with respect to the ball screw 31, that is, is fixedly attached to the ball screw 31. Further, in the actuator 300 according to the second comparative example shown in FIG. 9, the ball screw 31 and the piston 210 are connected via the spherical bearing 230, and the outer ring 230A and the inner ring 230B of the spherical bearing 230 slide, The piston 210 tilts with respect to the ball screw 31.

  The pressing force that presses the piston 10 against the non-rotating disk 104 by the actuators 100, 200, and 300 is supported by the backing plate 105. For this reason, the inclination of the rotating disk 103 and the non-rotating disk 104 is suppressed by the pressing force.

  However, when a large pressing force is applied by the actuators 100, 200, and 300, the tilt cannot be suppressed, and the rotating disk 103 and the non-rotating disk 104 may tilt as shown in FIG.

  When such an inclination occurs in the actuator 200, the non-rotating disk 104 and the pressing surface 110A of the piston 110 do not come into surface contact, and only a part of the pressing surface 110A on the radially inner side (left side in FIG. 8) of the wheel. Touch. Thereby, the pressure distribution acting on the non-rotating disk 104 from the piston 110 becomes non-uniform, and the contact between the rotating disk 103 and the non-rotating disk 104 becomes non-uniform. For this reason, uneven wear occurs between the rotating disk 103 and the non-rotating disk 104. Further, since the piston 110 contacts the non-rotating disc 104 only inside the wheel in the radial direction, the length of the moment arm that is the radial length from the wheel axle to the contact portion between the piston 110 and the non-rotating disc 104 is small. Shorter. As a result, the force (moment) acting on the rotating disk 103 is reduced and the braking force is insufficient. Further, in the piston 110 as well, only a part of the pressing surface 110 </ b> A contacts the non-rotating disk 104, so that a bending moment is generated in the ball screw 31 by the reaction force. For this reason, the life of the ball screw 31 is reduced, and a high load is also applied to the electric motor 25 for moving the ball screw 31.

  In order to prevent the above problems, as shown in FIG. 9, it is conceivable to connect the piston 210 and the ball screw 31 by a spherical bearing 230 and attach the piston 210 to the ball screw 31 so as to be tiltable.

  However, in this case, since the spherical surfaces slide and the piston 210 tilts, energy loss occurs in the pressing of the piston 210 by the actuator 200 due to the sliding resistance. Further, when galling occurs in the sliding portion, the piston 210 cannot be tilted. In order to prevent such galling and reduce sliding resistance, it is conceivable to lubricate the contact portion. However, since high frictional heat is generated between the rotating disk 103 and the non-rotating disk 104, a lubricant having high heat resistance must be used, which increases costs.

  On the other hand, in the actuator 100 of the present embodiment, the piston 10 can move relative to the ball screw 31 in the axial direction and the radial direction, and therefore slides with the ball screw 31 as shown in FIG. It can tilt without. As a result, the piston 10 follows the inclination of the non-rotating disc 104, and the entire pressing surface 10A of the piston 10 can be uniformly pressed against the non-rotating disc 104. Accordingly, it is possible to prevent uneven wear of the rotating disk 103 and the non-rotating disk 104, insufficient braking force, and a decrease in durability of the ball screw 31 and the electric motor 25, respectively.

  Further, since the piston 10 does not slide with the ball screw 31 when tilted, energy loss in pressing the piston 10 can be reduced. Further, since no lubricant is required and there is no possibility of galling, the cost can be reduced and the piston 10 can be tilted reliably.

  The tilt angle of the piston 10 is a change in angle between the frictional force generator 102 and the actuator 100 when changing from a non-braking state to a braking state, specifically, an angle change between the backing plate 105 and the carrier 50. It is desirable that the angle is equal to or greater than the corresponding angle. That is, the radial gaps T1, T2 between the piston 10 and the insertion shaft part 33 and the passage hole 22A, and the axial gap T3 between the piston 10 and the step part 33A of the ball screw 31 and the C ring 37 are: It is desirable that the piston 10 be set to tilt by an angle equal to or greater than the angle change between the backing plate 105 and the carrier 50. According to this, the piston 10 can be tilted so as to follow the tilt of the non-rotating disk 104 in the braking state.

  When the amount of operation of the brake pedal by the operator decreases, the controller outputs a command signal to the electric motor 25 so as to rotate in the reverse direction.

  The rotation of the electric motor 25 in the reverse direction is decelerated by the decelerating unit 40 and transmitted to the ball nut 35 as in the case of rotating in the forward direction. Thereby, the ball nut 35 rotates and the ball screw 31 moves in the release direction together with the piston 10.

  When the ball screw 31 moves in the release direction, the ball screw 31 moves in the axial direction while the rotation of the ball screw 31 is restricted by the guide hole portion 64 and the guide pin 65 of the guide plate 61 as in the case of moving in the braking direction. Guided.

  When the ball screw 31 moves in the releasing direction by a predetermined stroke, the piston 10 is separated from the non-rotating disk 104 and the braking of the brake device 101 is released. When the amount of operation of the brake pedal by the operator becomes zero, as shown in FIG. 2, braking of the brake device 101 is completely released and the piston 10 is separated from the non-rotating disc 104 so as to leave a predetermined gap.

  As described above, in the actuator 100, the ball screw 31 linearly moves while being guided by the guide plate 61 and the guide pin 65 in the axial direction while being restricted in rotation, and the piston 10 follows the inclination of the non-rotating disk 104. Tilt like so. Therefore, the actuator 100 can exert a stable braking force by bringing the pressing surface 10A of the piston 10 into surface contact with the non-rotating disk 104.

  In order to guide the piston 10 and the ball screw 31 in the axial direction and linearly move while restricting the rotation, the piston 10 is not formed in a circular cross section as in the present embodiment. Can be considered. That is, it is conceivable that the piston 10 and the passage hole 22A are locked in the circumferential direction, and the movement of the piston 10 in the axial direction is guided by the passage hole 22A. However, in order to restrict the rotation and guide the movement in the axial direction, it is desirable that the radial gap between the piston 10 and the passage hole 22A is small, while in order to tilt the piston 10, the piston 10 and the passage hole. It is desirable that the radial gap T1 between 22A is larger. Therefore, in this case, it is difficult to set an appropriate radial clearance between the piston 10 and the passage hole 22A.

  On the other hand, in the actuator 100, the restriction of the rotation of the piston 10 and the ball screw 31 and the guide of the movement in the axial direction are performed by the guide plate 61 and the guide pin 65 provided at the proximal end portion 34 of the ball screw 31. Therefore, the radial gap T1 between the piston 10 and the passage hole 22A can be an appropriate gap to follow the inclination of the non-rotating disk 104. Therefore, the actuator 100 can exhibit a stable braking force.

  Next, a modification of the above embodiment will be described with reference to FIGS.

  In the said embodiment, it is comprised as a clearance gap between the piston 10 and the ball screw 31, and another member is not provided. Instead, if the piston 10 is relatively movable in the axial direction of the ball screw 31 and can be tilted with respect to the axis of the ball screw 31, an elastic member or the like is provided between the piston 10 and the ball screw 31. It may be provided.

  Moreover, in the said embodiment, as shown in FIG. 3, between the penetration shaft part 33 and the main-body part 32 in the ball screw 31 is formed as an annular step part 33A perpendicular to the central axis of the ball screw 31. Instead, as shown in FIG. 6, a space between the insertion shaft portion 33 and the main body portion 32 may be formed as a spherical convex portion 133 </ b> A that protrudes toward the piston 10. In this case, it is desirable that a spherical concave portion 13 </ b> A capable of spherical contact with the spherical convex portion 133 </ b> A is formed in a portion of the piston 10 that faces the spherical convex portion 133 </ b> A. According to this, when the piston 10 is tilted, the spherical convex portion 133A of the ball screw 31 and the spherical concave portion 13A of the piston 10 come into spherical contact, and the piston 10 can tilt smoothly. On the other hand, as shown in FIG. 7, the space between the insertion shaft portion 33 and the main body portion 32 of the ball screw 31 is formed as a spherical concave portion 133B, and the portion of the piston 10 that faces the spherical concave portion 133B is formed into a spherical concave portion 133B. You may form as the spherical convex part 13B which protrudes toward. Also in this case, when the piston 10 is tilted, the spherical convex portion 13B of the piston 10 and the spherical concave portion 133B of the ball screw 31 are in spherical contact, and the piston 10 can be tilted smoothly.

  In the above embodiment, the guide stopper mechanism 60 has the two guide pins 65. In order to restrict the rotation of the guide plate 61, the guide stopper mechanism 60 desirably has two guide pins 65. However, the guide stopper mechanism 60 may have a single guide pin 65 or may have three or more guide pins 65. When three or more guide pins 65 are provided, it is desirable that the guide pins 65 are arranged at equal intervals in the circumferential direction.

  Moreover, in the said embodiment, the brake device 101 brakes the wheel of a fixed wing aircraft. Instead of this, the brake device 101 may brake a wheel of an automobile or a railway. Moreover, the brake device 101 may brake a rotor of a helicopter (rotary wing aircraft) as a rotating body.

  Moreover, in the said embodiment, the deceleration part 40 is comprised by the several gear mechanism. Instead of this, the speed reduction unit 40 may be any configuration as long as it transmits the rotation of the rotating shaft 26 of the electric motor 25 while reducing the rotation of the rotation shaft 26 to the rotation of the ball nut 35. For example, the speed reduction part 40 may have a chain mechanism or a belt mechanism.

  In the above embodiment, the rotating shaft 26 of the electric motor 25 is provided so as to be orthogonal to the moving direction of the piston 10. In order to save the space around the brake device 101, it is desirable that the moving direction of the rotating shaft 26 and the piston 10 be orthogonal, but the rotating shaft 26 may be inclined from the moving direction of the piston 10. In this case, the inclination angles of the rotation shafts of the first bevel gear 41A and the second bevel gear 41B may be determined according to the inclination angle between the rotation shaft 26 and the moving direction of the piston 10. If at least the rotating shaft 26 is inclined from the moving direction of the piston 10, the size of the actuator 100 is increased in the moving direction of the piston 10 compared to the case where the rotating shaft 26 is provided in parallel to the moving direction of the piston 10. Thus, the space around the brake device 101 can be saved. The rotating shaft 26 of the electric motor 25 may not be inclined from the moving direction of the piston 10, that is, provided in parallel with the moving direction.

  According to the above embodiment, there exist the following effects.

  In the actuator 100, the piston 10 is provided with clearances T 1, T 2, T 3 in both the axial direction and the radial direction of the ball screw 31, so that the piston 10 is perpendicular to the central axis of the ball screw 31 without sliding with the ball screw 31. Can be tilted with respect to any plane. Thus, since the piston 10 tilts without sliding with the ball screw 31, the piston 10 is prevented from generating energy loss in the pressing of the piston 10 by the actuator 100, and the piston 10 is connected to the non-rotating disk of the frictional force generating unit 102. 104 can be contacted in parallel.

  The piston 10 tilts without sliding with the ball screw 31 to follow the tilt of the non-rotating disk 104, so that the entire pressing surface 10A of the piston 10 can be pressed against the non-rotating disk 104. Accordingly, it is possible to prevent uneven wear of the rotating disk 103 and the non-rotating disk 104, insufficient braking force, and a decrease in durability of the ball screw 31 and the electric motor 25, respectively.

  Further, when the piston 10 is tilted, it does not slide with the ball screw 31, so that no lubricant is required and there is no possibility of galling. Therefore, the cost can be reduced and the piston 10 can be reliably tilted.

  In the actuator 100, the restriction of the rotation of the piston 10 and the ball screw 31 and the guide of the movement in the axial direction are performed by a guide plate 61 and a guide pin 65 provided at the proximal end portion 34 of the ball screw 31. Therefore, the radial gap T1 between the piston 10 and the passage hole 22A can be an appropriate gap to follow the inclination of the non-rotating disk 104. Therefore, the actuator 100 can exhibit a stable braking force.

  Hereinafter, the configuration, operation, and effect of the embodiment of the present invention will be described together.

  The brake device actuator 100 used in the brake device 101 that brakes the wheel by the frictional force generated in the frictional force generation unit 102 is attached to the piston 10 that is pressed against the frictional force generation unit 102 to generate the frictional force, and the piston 10 is attached. A ball nut 35 screwed into the ball screw 31, and an electric motor 25 that rotates the ball nut 35 to linearly move the ball screw 31. The ball screw 31 is attached so as to be relatively movable in the axial direction and the radial direction.

  The brake device actuator 100 further includes a housing 20 having a passage hole 22 </ b> A through which the piston 10 passes and accommodating the ball screw 31 and the ball nut 35, and the ball screw 31 is screwed into the ball nut 35. 32A is formed between the main body portion 32, the insertion shaft portion 33 having an outer diameter smaller than the outer diameter of the main body portion 32 and inserting the piston 10, and the main body portion 32 and the insertion shaft portion 33. The piston 10 is attached to the insertion shaft portion 33 with an axial clearance T3 between the step portion 33A and the axial direction of the step portion 33A, and is connected to the insertion shaft portion 33 and the passage hole 22A. It is attached to the insertion shaft portion 33 with radial gaps T1, T2 between the radial directions.

  Further, in the brake device actuator 100, the insertion shaft portion 33 is provided with a C ring 37 that prevents the piston 10 from coming off from the insertion shaft portion 33.

  In these configurations, the piston 10 is provided so as to be relatively movable in both the axial direction and the radial direction of the ball screw 31, so that it does not slide with the ball screw 31 and is perpendicular to the plane perpendicular to the central axis of the ball screw 31. Can be tilted. Thus, the piston 10 can be brought into parallel contact with the frictional force generating portion 102 by tilting the piston 10 without sliding with the ball screw 31. Therefore, the energy loss in pressing of the piston 10 during braking by the brake device actuator 100 is reduced.

  The brake device actuator 100 further includes a dust seal 15 that closes a gap between the inner peripheral surface of the passage hole 22A and the outer peripheral surface of the piston 10, and the piston 10 has a circular cross section.

  According to this configuration, the intrusion of dust into the housing 20 is prevented, and the movement of the ball screw 31 is prevented from being hindered by the dust.

  The brake device actuator 100 further includes a carrier 50 that supports the housing 20 so that the piston 10 faces the frictional force generation unit 102. The frictional force generation unit 102 includes a rotating disk 103 that rotates together with the wheels, an axle. And a non-rotating disc 104 provided so as to be displaceable toward the rotating disc 103, and a rotating disc 103 and a non-rotating disc provided on the opposite side of the piston 10 across the rotating disc 103 and the non-rotating disc 104. It has a backing plate 105 that receives the pressing force of the piston 10 applied to the disk 104, and the axial gap T3 and the radial gaps T1 and T2 are pistons from a non-braking state in which the piston 10 is separated from the frictional force generator 102. 10 is pressed against the frictional force generating part and is in a braking state Angle variation only the piston 10 between the backing plate 105 and the carrier 50 of the is set to a size that can swing with respect to the axis of the ball screw 31.

  In this configuration, the piston 10 can be tilted so as to follow the tilt of the non-rotating disk 104 of the frictional force generating unit 102 in the braking state. Therefore, a stable braking force can be exhibited by the brake device actuator 100.

  In addition, the brake device actuator 100 includes a guide stopper for restricting rotation of the ball screw 31 by locking a proximal end portion of the ball screw 31 on the opposite side in the axial direction to the insertion shaft portion to which the piston is coupled. A mechanism 60 is further provided.

  In this configuration, the rotation restriction of the piston 10 and the ball screw 31 and the guide of the movement in the axial direction are performed by the guide stopper mechanism 60 that locks the proximal end portion 34 of the ball screw 31 in the circumferential direction. Therefore, the radial gap T1 between the piston 10 and the passage hole 22A can be an appropriate gap to follow the inclination of the non-rotating disk 104. Therefore, the brake device actuator 100 can exhibit a stable braking force.

  The brake device 101 includes the brake device actuator 100 having any one of the above-described configurations, and a frictional force generator 102 attached to the wheel.

  The embodiment of the present invention has been described above. However, the above embodiment only shows a part of application examples of the present invention, and the technical scope of the present invention is limited to the specific configuration of the above embodiment. Absent.

  DESCRIPTION OF SYMBOLS 10 ... Piston, 15 ... Dust seal, 20 ... Housing, 22A ... Pass-through hole, 25 ... Electric motor, 31 ... Ball screw (1st screw member), 32 ... Main-body part, 33 ... Insertion shaft part, 33A ... Step part, 34 ... Base end part, 35 ... Ball nut (second screw part), 37 ... C ring (prevention member), 40 ... Deceleration part, 60 ... Guide stopper mechanism (rotation restricting part), 100 ... Actuator (actuator for brake device) ), 101... Brake device, 102... Friction force generator, 105... Backing plate (disc support member)

Claims (6)

An actuator for a brake device used in a brake device that brakes a rotating body with frictional force generated in a frictional force generating unit,
A piston that is pressed against the frictional force generating part to generate the frictional force;
A first screw member to which the piston is attached;
A second screw member threadably engaged with the first screw member;
A motor that rotates the second screw member to linearly move the first screw member;
A housing having a passage hole through which the piston passes, and housing the first screw member and the second screw member ;
The first screw member includes a main body portion formed with a screw portion to be screwed with the second screw member, an insertion shaft portion having an outer diameter smaller than an outer diameter of the main body portion, and inserting the piston. A step portion formed between the main body portion and the insertion shaft portion,
The piston is attached to the insertion shaft portion with an axial gap between the step portion and the radial direction between the insertion shaft portion and the passage hole. wherein attached to the insertion shaft portion, the first brake system actuator, characterized in that provided in a relatively movable in the axial direction and the radial direction before Symbol first screw member with respect to the screw member. 2. The brake device actuator according to claim 1 , wherein the insertion shaft portion is provided with a retaining member that prevents the piston from coming off from the insertion shaft portion . 3. The brake device actuator according to claim 1 , further comprising a dust seal that closes a gap between an inner peripheral surface of the passage hole and an outer peripheral surface of the piston . A carrier attached to the housing such that the piston faces the frictional force generating portion;
The frictional force generator is
A rotating disk that rotates with the rotating body;
A non-rotating disk provided non-rotatably and displaceable toward the rotating disk;
A disc support member that is provided on the opposite side of the piston across the rotating disc and the non-rotating disc and receives a pressing force of the piston applied to the rotating disc and the non-rotating disc;
The axial gap and the radial gap are the disc support member when the piston is pressed against the frictional force generating part from the non-braking state where the piston is separated from the frictional force generating part. according to any one of claims 1 to 3, the angle variation only the piston is characterized in that it is set to a size that can swing with respect to the axis of the first threaded member between said carrier Brake device actuator. A rotation restricting portion for restricting rotation of the first screw member by locking a proximal end portion of the first screw member on the opposite side to the insertion shaft portion to which the piston is connected in the circumferential direction; The actuator for brake devices as described in any one of Claim 1 to 4 characterized by the above-mentioned.   The brake device actuator according to any one of claims 1 to 5,
  A brake device comprising: the frictional force generating unit that generates a frictional force that brakes the rotating body.