JP2002213576A - Electric power steering device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To increase the power transmission efficiency of a gear reduction mechanism for transmitting the auxiliary torque of a motor to a rack and pinion mechanism. SOLUTION: In this electric power steering device 30, the auxiliary torque corresponding to steering torque is generated by the motor 82 and transmitted to the rack and pinion mechanism 32 of a steering system via the gear reduction mechanism 110, and a steering wheel is steered by the rack and pinion mechanism. The gear reduction mechanism is a combination structure of a worm 112 and a worm wheel 113. A porous film is formed by electroless plating on one or both of the tooth flanks of the worm and the tooth flanks of the worm wheel, the porous film is impregnated with a low-friction agent, therefore one or both of the tooth flanks of the worm and the tooth flanks of the worm wheel are coated with a film layer made of the low-friction agent.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

The present invention relates to an improvement of an electric power steering device.

[0002]

2. Description of the Related Art In recent years, electric power steering devices have been frequently used in order to reduce the steering force of a steering wheel to provide a comfortable steering feeling. This type of electric power steering apparatus generates an auxiliary torque corresponding to a steering torque by an electric motor and transmits the auxiliary torque to a rack and pinion mechanism of a steering system. 2. Description of the Related Art A power steering device (hereinafter referred to as "conventional technology") is known.

In the above prior art, as shown in FIG. 4 of the publication, an auxiliary torque corresponding to a steering torque is generated by a motor 11 (numbers quoted in the publication; the same applies hereinafter). This auxiliary torque is transmitted to the small bevel gear 7b and the large bevel gear 7
Through a, the power is transmitted to a rack and pinion mechanism composed of a combination of a pinion 2a and a rack shaft 5, and the steered wheels are steered by the rack and pinion mechanism.

[0004]

In the above prior art, the auxiliary torque is transmitted to a rack-and-pinion mechanism via a gear-type reduction mechanism comprising a combination of a small bevel gear 7b and a large bevel gear 7a. There is room for improvement in increasing the power transmission efficiency of such a gear type speed reduction mechanism.

An object of the present invention is to provide a technique capable of increasing the power transmission efficiency of a gear type speed reduction mechanism for transmitting an auxiliary torque to a rack and pinion mechanism of a steering system.

[0006]

In order to achieve the above object, a first aspect of the present invention is to generate an auxiliary torque corresponding to a steering torque by an electric motor, and to generate the auxiliary torque via a gear type reduction mechanism.
It is transmitted to the steering rack and pinion mechanism,
In the electric power steering apparatus in which the steered wheels are steered by the rack and pinion mechanism, the gear type reduction mechanism has a combination structure of a worm and a worm wheel, and any one of the worm tooth surface and the worm wheel tooth surface is used. A porous film is formed on one or both surfaces by an electroless plating method, and this porous film is impregnated with a low friction agent to reduce one or both of the worm tooth surface and the worm wheel tooth surface. It is characterized by being coated with a coating layer made of a friction agent.

[0007] The coefficient of friction of the sliding surface between the worm tooth surface and the worm wheel tooth surface can be reduced with a low friction agent. Therefore, the power transmission efficiency of the gear type speed reduction mechanism having the combined structure of the worm and the worm wheel can be improved.

[0008]

Embodiments of the present invention will be described below with reference to the accompanying drawings. The drawings should be viewed in the direction of reference numerals. FIG. 1 is a schematic diagram of a vehicle steering system according to the present invention. The vehicle steering system 10 is an electric power steering system including an electric power steering device 30 in a steering system 22 from a steering handle 11 of a vehicle to steered wheels (wheels) 21 and 21.

More specifically, a vehicle steering system 1
Reference numeral 0 denotes an input shaft 31 that connects the input shaft 31 of the electric power steering device 30 to the steering handle 11 via the steering shaft 12 and the universal joints 13, 13.
1, a rack and pinion mechanism 32 is connected, and left and right steering wheels 21 and 21 are connected to the rack and pinion mechanism 32 via left and right tie rods 37 and 37.
The rack and pinion mechanism 32 includes a pinion 33 provided on the input shaft 31 and a rack shaft 35 provided with a rack 34 for meshing with the pinion 33. The electric power steering device 30 includes a steering torque sensor 70.

[0010] Such a vehicle steering system 1
According to 0, the driver steers the steering handle 11, and steers the left and right steering wheels 21 and 21 via the input shaft 31, the rack and pinion mechanism 32 and the left and right tie rods 37 and 37 by the steering torque. Can be.

Further, the steering torque of the steering system 22 generated by the steering handle 11 is detected by a steering torque sensor 70, and control means 81 is provided based on the detected signal.
A control signal is generated by the motor 82. An auxiliary torque corresponding to the steering torque is generated by the electric motor 82 based on the control signal, and the auxiliary torque is transmitted to the steering system 22 via the torque limiter 90, the gear type reduction mechanism 110, and the input shaft 31. The power is transmitted to the rack and pinion mechanism 32, and the left and right steered wheels 21 and 21 can be steered by the rack and pinion mechanism 32 and the left and right tie rods 37. Therefore,
Steering wheels 21 and 21 can be steered by a composite torque obtained by adding the assisting torque of electric motor 82 to the steering torque of the driver.

FIGS. 2A and 2B are principle diagrams of a steering torque sensor according to the present invention. The steering torque sensor 70 is
This is a magnetostrictive torque sensor that, when a torque acts on an input shaft 31 having a magnetostrictive characteristic such as a steel material, a magnetostrictive effect generated according to the torque is electromagnetically detected by an electric coil. Such a magnetostrictive torque sensor is a known sensor as disclosed in Japanese Patent Application Laid-Open No. 6-221940, “Magnetostrictive Torque Sensor”. Hereinafter, the outline of the steering torque sensor 70 will be described.

A steering torque sensor 70 shown in FIG. 1A has an exciting coil 71 formed in a substantially figure eight shape and an exciting coil 7.
A detection coil 72 having a size substantially the same as that of 1 and formed in a figure of eight is superimposed substantially concentrically and substantially orthogonal to each other,
These excitation / detection coils 71 and 72 are arranged as a set of magnetic heads 73 near the outer peripheral surface of the input shaft 31. That is, facing the outer peripheral surface of the input shaft 31,
A substantially figure-shaped excitation coil 71 was arranged, and a substantially figure-shaped detection coil 72 was superimposed on the excitation coil 71 while changing the phase by 90 °. In this case, the exciting coil 71
Are arranged substantially parallel to the outer periphery of the input shaft 31 or substantially parallel to the axial direction. 74 is an excitation voltage supply source, and 75 is an output voltage amplifier.

When a high-frequency AC voltage (excitation voltage) of about 20 to 100 kHz is supplied from the excitation voltage supply source 74 to the excitation coil 71, the detection coil 72 is supplied to the detection coil 72 in response to the magnetostrictive effect of the input shaft 31 based on torque. Thus, an AC voltage (output voltage) having the same frequency as the excitation voltage can be obtained. The output voltage has the same phase or the opposite phase as the excitation voltage depending on the direction of the torque acting on the input shaft 31. The amplitude of the output voltage at this time is proportional to the magnitude of the torque. Therefore, if the output voltage is synchronously rectified with reference to the phase of the excitation voltage, the magnitude and direction of the torque can be detected.

The output voltage is amplified by the output voltage amplifier 75 and is sent to the control means 81 as a detection signal of the steering torque sensor 70. If the magnetizing force of the input shaft 31 is small, the number of turns of the exciting coil 71 and the detecting coil 72 may be increased, and the exciting / detecting coils 71 and 72 may be alternately arranged one by one.

The steering torque sensor 70 shown in FIG. 2B has a magnetic head 73 comprising excitation / detection coils 71 and 72 connected to two.
A set is prepared, and these two sets of magnetic heads 73, 73 are arranged near the outer peripheral surface of the input shaft 31 and at symmetrical positions with respect to the axis of the input shaft 31. And the output voltage amplifier 75
By amplifying the difference between the output voltages from the detection coils 72, 72, it is possible to obtain a steering torque signal that does not change much with changes in the environmental temperature.

The steering torque sensor 7 described in (a) or (b) above
By adopting 0, there is no need to divide the input shaft 31 in the longitudinal direction and connect these divided shafts with a torsion bar as in the case of detecting the steering torque in the conventional electric power steering device. . Therefore, the input shaft 31 can have a simple configuration, and the input shaft 31 can be set sufficiently long. Moreover, FIG.
In the case of processing the pinion 33 provided on the input shaft 31 shown in (1), it is easy to set the input shaft 31 on a processing machine, and the processing accuracy can be further improved. As the processing accuracy increases, the engagement accuracy between the pinion 33 and the rack 34 also increases. As a result, the rack and pinion mechanism 3
2 can improve the power transmission efficiency.

FIG. 3 is an overall configuration diagram of the electric power steering apparatus according to the present invention, and is a sectional view of a left end portion and a right end portion. In this figure, the rack shaft 35 of the electric power steering device 30 is moved in the vehicle width direction (the left-right direction in the figure).
In the housing 41 extending in the axial direction. The rack shaft 35 is a shaft in which ball joints 36, 36 are screwed to both ends in the longitudinal direction projecting from the housing 41, and left and right tie rods 37, 37 are connected to the ball joints 36, 36. The housing 41 is provided with brackets 42 for mounting on a vehicle body (not shown) and stoppers 43 on both ends in the longitudinal direction.

When the rack shaft 35 slides to the right by a predetermined amount, the contact end surface (rack end) 38 of the left ball joint 36 contacts the stopper 43. When the rack shaft 35 slides a predetermined amount to the left, the right ball joint 36
The contact end face (rack end) 38 of the contact member 43 contacts the stopper 43. By regulating the amount of movement of the rack shaft 35 in this manner, the maximum steering angle of the left and right steered wheels 21 and 21 (see FIG. 1) can be limited. That is, when the rack shaft 35 moves to the end of movement, the left and right steered wheels 21 and 21 are moved.
Steering angle is maximized. In the figure, reference numerals 44, 44 denote boots for dust sealing.

FIG. 4 is a sectional view taken along line 4-4 of FIG. 3, and shows a longitudinal sectional structure of the electric power steering apparatus 30. The electric power steering device 30 houses an input shaft 31, a rack and pinion mechanism 32, a steering torque sensor 70, a torque limiter 90 (see FIG. 1), and a gear type reduction mechanism 110 in a housing 41, and opens an upper opening of the housing 41. It is closed by the lid 45. Steering torque sensor 7
Numeral 0 is attached to the housing 41 or the lid 45.

The housing 41 is set vertically by rotatably supporting the lower end portion and the longitudinal center portion of the input shaft 31 via upper and lower two bearings 51, 52, and a rack guide 60 is provided. Is provided. 53 is a lid mounting bolt, and 54 is a retaining ring.

The input shaft 31 is a pinion shaft integrally formed with a pinion 33 at a lower portion, a screw portion 55 at a lower end portion, and an upper end portion projecting outward from a lid 45. The rack 34 is formed integrally with a rack shaft 35. By screwing the nut 56 into the screw portion 55, the movement of the input shaft 31 in the longitudinal direction (axial direction) can be restricted. 57 is a cap nut, 58 is an oil seal,
59 is a spacer.

The rack guide 60 comprises a guide portion 61 which contacts the rack shaft 35 from the side opposite to the rack 34, and an adjustment bolt 63 which presses the guide portion 61 via a compression spring 62. According to such a rack guide 60, the compression spring 6 is adjusted by the adjustment bolt 63 screwed into the housing 41.
By pressing the guide part 61 with an appropriate pressing force via the
The guide section 61 applies a preload to the rack 34 to
Can be pressed against the pinion 33. 64 is a contact member for sliding the back surface of the rack shaft 35, and 65 is a lock nut.

FIG. 5 is a sectional view taken along the line 5-5 in FIG. 4, and shows the relationship among the input shaft 31, the electric motor 82, the torque limiter 90, and the gear type reduction mechanism 110. The electric motor 82 has an output shaft 83
Is mounted sideways on the housing 41, and the housing 4
1, the output shaft 83 is extended.

The gear type reduction mechanism 110 is a torque transmission means for transmitting the auxiliary torque generated by the electric motor 82 to the input shaft 31 and has a combined structure of a drive gear and a driven gear.
It consists of a worm gear mechanism. Specifically, the gear-type reduction mechanism 110 is configured such that the output shaft 83 of the electric motor 82 is connected to the torque limiter 9.
0, a worm (drive gear) 112 formed on the transmission shaft 111, and a worm 112
And a worm wheel (driven gear) 113 coupled to the input shaft 31. The auxiliary torque of the electric motor 82 can be transmitted to the rack and pinion mechanism 32 (see FIG. 1) via the input shaft 31.

The present invention is characterized in that one or both of the tooth surface of the worm 112 and the tooth surface of the worm wheel 113 are coated with a coating layer made of a low friction agent by performing a surface treatment. I do. In order to cover with a film layer made of a low friction agent, for example, coating with a low friction agent or impregnation with a low friction agent is performed. By reducing the friction coefficient of the sliding surface between the tooth surface of the worm 112 and the tooth surface of the worm wheel 113 to a predetermined value with a low friction agent, power transmission efficiency can be increased. As the low friction agent, for example, polytetrafluethylene (abbreviation: PT
FE (registered trademark; Teflon). Fluororesins have a very low coefficient of friction.

As the surface treatment method using the low friction agent, there are the following first surface treatment method and second surface treatment method. (1) In the first surface treatment method, the material of the worm 112 and the worm wheel 113 is changed to carbon steel (JIS-G-
4051) etc., and worm 112 in a predetermined processing solution.
Electroless nickel and PTFE are co-deposited on the tooth surface of the worm wheel and the tooth surface of the worm wheel 113, and the PTFE is uniformly contained in the coating up to a volume ratio of 10 to 30%. ) Is a treatment method in which a coating of electroless nickel and PTFE is firmly adhered to the tooth surface. The thickness of the coating is 5 to 20 μm. This first surface treatment method is known as "Nif Grip (registered trademark)" of ULVAC TECHNO, LTD.

(2) The second surface treatment method uses the worm 11
The worm wheel 113 and the worm wheel 113 are made of steel such as carbon steel for machine structure (JIS-G-4051), and the tooth surface of the worm 112 and the tooth surface of the worm wheel 113 are coated with nickel and phosphorus by electroless plating. A porous coating is formed, the porous coating is impregnated with PTFE, followed by a heat treatment (about 40
(Firing at 0 ° C.) to firmly adhere the coating to the tooth surface. The coating adhered to the tooth surface is obtained by impregnating a porous coating of nickel / phosphorous precipitated in the form of particles with PTFE, and has a thickness of 5 to 20 μm. This second surface treatment method is known as "Nidax (registered trademark)" of ULVAC TECHNO, LTD.

The transmission shaft 111 is arranged concentrically with the output shaft 83, and is connected to the housing 4 via two bearings 114 and 115.
1 is a shaft rotatably supported. Housing 41
The first bearing 114 is located near the output shaft 83 so as not to move in the axial direction, and the second bearing 115 located far from the output shaft 83 is fitted so as to move in the axial direction.
Further, the end surface of the outer ring of the second bearing 115 is
6 is pushed toward the output shaft 83 by the adjustment bolt 117. By applying a preload to the first and second bearings 114 and 115 with the pressing force of the adjusting bolt 117 and the thin disk-shaped leaf spring 116, the transmission shaft 111 is adjusted so that there is no play in the axial direction. You can play back. In addition, by adjusting the axial displacement of the worm 112, the mesh between the worm 112 and the worm wheel 113 can be adjusted so that there is no backlash while maintaining appropriate friction. Further, the transmission shaft 111 is moved by the elastic force of the leaf spring 116.
Can absorb thermal expansion in the axial direction. Reference numeral 118 denotes a lock nut, and 119 denotes a retaining ring.

The present invention is characterized in that the tooth surface of the worm 112 and the tooth surface of the worm wheel 113 are engaged with each other without backlash. As a configuration for eliminating backlash, for example, the following (1) to
It consists of the combination of (4). (1) The worm 112 is made of a metal product, and its tooth surface is covered with a coating layer made of the above low friction agent. (2) The worm wheel 113 is made of a resin product. (3) the distance X from the center O ₁ of the worm 112 to the center O ₂ of the worm wheel 113 is set to a predetermined theoretical value (reference value). (4) The reference pitch circle diameter d1 of the worm 112 or the reference pitch circle diameter d2 of the worm wheel 113 is set to be slightly larger than a predetermined theoretical value (reference value).

When the gear type speed reduction mechanism 110 is assembled, the tooth surface of the worm 112 and the tooth surface of the worm wheel 113 are engaged with each other by applying a preload by an amount corresponding to the larger reference pitch circle diameter d1 or d2. Become. As a result, there is no backlash (meshing gap) between the tooth surface of the worm 112 and the tooth surface of the worm wheel 113, so that there is no play in the meshing. Because there is no rattling,
The impact torque due to the inertia of the electric motor 82
Does not act on the tooth surface of the worm wheel 113. Therefore, the durability of the gear type reduction mechanism 110 is
Even higher.

However, since there is no backlash, the meshing resistance (friction) between the tooth surface of the worm 112 and the tooth surface of the worm wheel 113 increases. In order to solve this problem, the tooth surface of the worm 112 was covered with a coating layer made of the above low friction agent. As a result, the friction coefficient of the sliding surface between the tooth surface of the worm 112 and the tooth surface of the worm wheel 113 can be reduced with the low friction agent. Therefore, even though the gears 112 and 113 are adjusted so that there is no play between the worm 112 and the worm wheel 113, the friction between the tooth surfaces of the gears 112 and 113 can be appropriately maintained, and the power transmission efficiency can be increased.

FIG. 6 is a sectional view of the torque limiter according to the present invention. The present invention relates to an electric motor 82 and a gear type reduction mechanism 11.
0, and a torque limiter 90 is interposed therebetween. The torque limiter 90 is connected to the output shaft 8 of the electric motor 82.
3, the inner member 91 serrated to the transmission shaft 1
11 is a torque limiting mechanism fitted to a cylindrical outer member 93 serrated and coupled to the outer member 93.

The inner member 91 is formed on the outer peripheral surface 92 by the transmission shaft 1.
11 is a male member having a tapered shape toward the front end.
The outer member 93 is a female member having an inner peripheral surface 94 tapered forward so that the outer peripheral surface 92 of the inner member 91 fits. The torque limiter 90 is assembled by fitting the tapered outer peripheral surface 92 to the tapered inner peripheral surface 94 and retaining the rear end surface 95 of the inner member 91 with the retaining ring 97 while resiliently projecting with the disc spring 96. be able to. 101 is a spacer, 10
2 is a washer and 103 is a disc spring.

The outer peripheral surface 92 is pressed against the inner peripheral surface 94 by the resilient force of the disc spring 96 to apply a preload to the outer peripheral surface 9.
2 can be connected to the inner peripheral surface 94 with a predetermined frictional force. Since such a torque limiter 90 is used,
When a large torque exceeding a predetermined frictional force acts, the outer peripheral surface 92 and the inner peripheral surface 94 slip. As a result,
The auxiliary torque transmitted from the electric motor 82 to the gear-type reduction mechanism 110 can be limited, that is, the overtorque can be cut. Therefore, no excessive torque is generated in the electric motor 82, and no excessive torque is transmitted to the load side.

Further, the inner member 91 is connected to the outer member 9.
3 is fitted with a taper, the assembly accuracy of both is high,
Alignment is also easy. Accordingly, the assembly accuracy of the transmission shaft 111 with respect to the output shaft 83 is high, and the alignment is easy. In addition, the motor 82 and the gear type reduction mechanism 1
10 and the torque limiter 90 is interposed between
The torque limiter 90 can be small, and the torque limiter 9 can be used.
0 can be reduced in size and cost. Since the small torque limiter 90 requires a small space for arrangement, it can be easily housed in the housing 41.

FIGS. 7A to 7D are configuration diagrams of a rack and pinion mechanism according to the present invention. In addition, in order to facilitate understanding, the pinion 3
3 is arranged and represented. L1 is the center line of the pinion, L2 is the center line of the rack axis, and L3 is a line perpendicular to the tooth surface of the rack. 7 and 8, the case where the center line L1 of the pinion is orthogonal to the center line L2 of the rack axis is described.
explain.

(A) shows the rack and pinion mechanism 32
The pinion 33 and the rack 34 are "helical gears" (helical gears). That is, the pinion 3
Numeral 3 is a helical pinion, and rack 34 is a helical rack. For example, the “helical gear” forming the pinion 33 is, as shown in (b), a tooth line 3 which is an intersection line between the peripheral surface of the cylinder 33a serving as a reference pitch surface and the tooth surface 33b.
3c is a cylindrical gear having a helical winding having a predetermined twist angle θ. The “torsion angle θ” is an acute angle θ formed between the helical winding 33d and the bus 33e of the cylinder 33a considering the helical winding 33d.

FIG. 3C is a partially enlarged perspective view of the "helical gear" forming the rack 34. The torsion angle of the "helical gear" forming the pinion 33 is different from that of the "helical gear". Indicates that they are the same. The present invention is characterized in that the torsion angle θ of the helical gear forming the pinion 33 and the rack 34 is set so as not to exceed the friction angle of the helical gear. The reason will be described later.

(D) shows the pinion 33 and the rack 34
2 is an enlarged cross-sectional view of the tooth profile of the “helical gear”, which shows that the tooth profile of the helical gear is an arc tooth profile. For details on the arc gear, see "[New Gears and Their Applications] Arc Gear" (Mechanical Design, Vol. 26, No. 3, March 1982, pp. 47-51, published by Nikkan Kogyo Shimbun). Are known from the literature. Hereinafter, the outline of the arc tooth profile will be described.

An arcuate gear is defined as one of a pair of gears.
At least one of the gear tooth flank surfaces is referred to as a reference pitch line Pi.
A gear having an arcuate tooth shape formed on an arc surface having a substantially center on the top, and at least a tooth root surface of the other gear formed on an arc surface having a center on a reference pitch line Pi,
Also called W / N gear. The arc-shaped gears include a symmetrical arc-shaped gear and an asymmetrical arc-shaped gear. Here, the root surface is a portion of the tooth surface between the root surface and the reference pitch line Pi, and the apical surface is the tooth tip surface and the reference pitch line P
i is the portion of the tooth surface between the i.

In the pinion 33, the symmetrical arc tooth shape means that, as shown in (d), the end face 33f is formed in an arc face and the root face 33g is also formed in an arc face, that is, the end face is formed. The surface 33f and the surface 33g of the tooth root are formed in an arc surface substantially point-symmetrical with respect to the reference pitch line Pi, and examples thereof include Novikov gear type 3 and shinmark gear. r is the radius of the arc surface. The symmetrical arc tooth profile of the rack 34 is also the same as the symmetric arc tooth profile of the pinion 33, and the tip end surface 34a and the root surface 34b are substantially point-symmetric with respect to the reference pitch line Pi. It is formed on a circular arc surface of the shape.

On the other hand, an asymmetrical arc tooth profile means that, among a set of gears, one gear tooth is formed only by a tooth tip arc substantially centered on a reference pitch line Pi, and the other gear tooth is formed. Is an arc tooth profile formed only by a root arc that is substantially centered on the reference pitch line Pi, and includes, for example, Novikov Gears 1 and 2 and Sirkirk Gear. In the present invention,
To make the helical gear tooth profile a symmetrical arc tooth profile,
More preferred.

The front tooth profile of the involute tooth profile is a mesh between the convex tooth surfaces. In contrast, the present invention
The tooth profile of the helical gear is an arc tooth profile. The front tooth profile of the arc tooth profile is an engagement between the concave tooth surface and the convex tooth surface. Since the relative radius of curvature in the tooth ridge direction is large, when a load is applied, the contact line becomes a region having a large area. In general,
The arc tooth profile has a surface fatigue strength of 6 to 7 times, a bending strength of 1.5 to 1.6 times, and a bending fatigue strength of 1.5 to 1.6 times as compared with the involute gear.

By forming the pinion 33 and the rack 34 as the helical gears having the above-mentioned arc-tooth shape, the strength of these gears can be further increased. For example, the following effects are exhibited. When the left and right steered wheels are steered to the maximum steering angle, that is, when the rack shaft 35 moves to the end of movement in FIG.
Hits the stopper 43 and the right ball joint 36
Hits the stopper 43, the rack 34 (see FIG. 1) stops immediately. At this time, an extremely large torque acts on the pinion 33 (see FIG. 1) and the rack 34 as compared with the normal steering. Even in such a case, the pinion 33 and the rack 34 with increased strength can sufficiently receive a large torque.

FIGS. 8A to 8C are conceptual diagrams of the rack and pinion mechanism according to the present invention. (A) shows that the rack 34 was moved leftward in the figure by rotating the pinion 33. When the left and right steered wheels are steered to the right to the maximum steering angle, that is, when the rack shaft 35 moves to the end of movement, the right ball joint 36 hits the stopper 43, and the rack 34 immediately stops. Since the torque at this time is an impact torque, it is a maximum torque much larger than that during normal steering.

(B) shows the toothed surface of the pinion 33 on the rack 3
FIG. 5 is a schematic diagram showing a state where a tooth surface of No. 4 is pushed to the left in the figure. (C) is a diagram schematically showing (b), wherein a right triangle with the tooth surface of the rack 34 as the hypotenuse D is represented by △.
Expressed as ABC. In (b) and (c), the inclination angle of the hypotenuse D is the same as the torsion angle θ of the helical gear. The force for pushing the hypotenuse D with the teeth E of the pinion 33 is W ₀ , and this force W ₀ corresponds to the force acting on the pitch circle of the pinion 33 in the circumferential direction (the rotational force of the pinion 33). Thus, the force W ₀ acts in a direction perpendicular to the line AB.

The rack 34 stopped at the moving end is
When the tooth E is further pushed by the pinion 33, the tooth E slides on the hypotenuse D in the direction of the point A and tends to move. The direct pressure acting between the hypotenuse D and the tooth E, that is, the direct pressure acting between the tooth surface of the pinion 33 and the tooth surface of the rack 34 (force acting at right angles to the tooth surface) W ₁ is: W ₁ = W ₀ × cos θ (1) The force acting in parallel to the hypotenuse D (the force acting in parallel to the tooth surface of the rack 34) W ₂ is: W ₂ = W ₀ × sin θ (2) )

It is necessary to support the tooth E with a force P ₀ parallel to the line AB so that the tooth E does not slide on the hypotenuse D in the direction of the point A due to the force W ₂ . This supporting force P ₀ is pinion 3
3 and the direction of the force P ₀ is orthogonal to the direction of the force W ₀ . Component force of the force P ₀ support is a perpendicular force component P ₁ and the oblique component force P ₂ Doo parallel to D on the hypotenuse D, it can be expressed by the following equation. P ₁ = P ₀ × sin θ (3) P ₂ = P ₀ × cos θ (4)

Since the sum of the components of the force in the direction perpendicular to the hypotenuse D, that is, the composite direct pressure R is the sum of the direct pressure W ₁ and the component force P ₁ , this can be expressed by the following equation: R = W ₁ + P ₁ (5) When the maximum friction force between the tooth surface of the pinion 33 and the tooth surface of the rack 34 is F, the magnitude of the maximum friction force F is proportional to the combined direct pressure R. . This is expressed by the following equation: F = μ × R (6) where μ is the friction between the tooth surface of the pinion 33 as the helical gear and the tooth surface of the rack 34. It is a coefficient. If the friction angle of the helical gear corresponding to the friction coefficient μ is ρ, μ = tanρ (7) The tooth E slides on the hypotenuse D in the direction of the point A by the force W ₂ . since you try to move, the friction force F acts in the direction opposite to the direction of the force W _2.

The three forces F, W ₂ , and P ₂ parallel to the hypotenuse D have the following relationship. P ₂ = W ₂ −F (8) By substituting equations (1) to (6) into equation (8), P ₀ × cos θ = W ₀ × sin θ-μ × R = W ₀ × sinθ−μ (W ₁ + P ₁ ) = W ₀ × sin θ−μ (W ₀ × cos θ + P ₀ × sin θ) = W ₀ × sin θ−μ × W ₀ × cos θ−μ × P ₀ × sin θ (9) When formula (9) is rearranged, P ₀ (cos θ + μ × sin θ) = W ₀ (sin θ−μ × cos θ) (10) P ₀ = W ₀ (sin θ−μ × cos θ) / (cos θ + μ × sin θ) (11) By substituting equation (7) into equation (11),

[0052]

(Equation 1)

As is apparent from the above equation (12), θ =
For ρ, P ₀ = 0. When θ <ρ, P ₀ <0. Accordingly, in a state in which the rack 34 is stopped, also act large torque to the pinion 33, a thrust to the pinion 33, i.e., the force P ₀ does not work. This is the reason that the torsion angle θ of the helical gear is set in a range not exceeding the friction angle ρ of the helical gear (0 ° <θ ≦ ρ).

On the other hand, when the rack 34 is not stopped at the left or right end of movement and is in a normal state, the pinion 33
In the case of driving the rack 34 to the left or right, the force pushing the hypotenuse D teeth E of the pinion 33 is much smaller force than the force W _0. Thrust corresponding to this small force acts on the pinion 33. Thus, the thrust acting on the pinion 33 can be suppressed to an extremely small value.

Next, the thrust acting on the pinion 33 when the input shaft 31 is installed at an angle will be described with reference to FIG. FIGS. 9A and 9B are modified conceptual views of the rack and pinion mechanism according to the present invention. (A)
Indicates a case where a straight line orthogonal to the center line L2 of the rack shaft 35 is set as the reference line S, and the input shaft 31 is inclined by an inclination angle α from the reference line S to the left in the figure. (B) shows a case where the input shaft 31 is inclined by the inclination angle α from the reference line S to the right in the figure.

Here, it is assumed that the torsion angle of the rack 34 as a helical gear is β. The torsion angle of the rack 34 in (a) is β = θ + α, and the torsion angle of the rack 34 in (b) is β
= Θ-α. However, the twist angle θ of the pinion 33
Is constant regardless of the inclination angle α of the input shaft 31.
Since the twist angle θ of the pinion 33 is always constant, the thrust P ₀ acting on the pinion 33 is also constant, and the above (1)
2) It can be expressed by the following equation. As described above, the pinion 3
The torsion angle θ of No. 3 is set within a range not exceeding the friction angle ρ of the helical gear (0 ° <θ ≦ ρ). Therefore,
Regardless inclination angle α of the input shaft 31, in the state in which the rack 34 is stopped, also act large torque to the pinion 33, a thrust to the pinion 33, i.e., the force P ₀ does not work.

In summary, since the pinion 33 and the rack 34 are helical gears, relatively large torque can be transmitted as compared with the immediate gear (spur gear). As a result, the rack and pinion mechanism 32 can be reduced in size. Moreover, since the torsion angle θ of the pinion 33 as the helical gear is set so as not to exceed the friction angle ρ of the helical gear, the thrust acting on the pinion 33 is such that the rack 34 is moved to the left or right It is only a small force when it is not stopped and is in a normal state. Therefore, the input shaft 31 shown in FIG. 8A and the input shaft 31 shown in FIGS.
The thrust acting on the input shaft 31 of the input shaft 31 is small.
The thrust acting on the bearings 51 and 52 supporting the gear shaft and the gear type reduction mechanism 110 (see FIG. 4) connected to the input shaft 31 is also small. Therefore, despite the use of the helical gear, the input shaft 31, the bearings 51 and 52, and the gear-type reduction mechanism 11
It is not necessary to increase the strength of 0, and these members can be reduced in size and inexpensive.

FIG. 10 is a modified example (part 1) of the gear type speed reducing mechanism according to the present invention, and shows a modified example of the support structure of the transmission shaft 111 in the gear type speed reducing mechanism 110. The supporting structure of the modified example is such that the transmission shaft 111 is connected to the first and second bearings 114 and 11.
5 and the housing 41 via the eccentric sleeve 121.
It is characterized by being supported by. Specifically, the eccentric sleeve 121 is a tubular sleeve rotatably fitted into a hole of the housing 41, and first and second bearings 114 and 115 are fitted into holes inside the tubular sleeve, and these first and second bearings are fitted.・
The transmission shaft 111 is rotatably supported via second bearings 114 and 115. By pressing the eccentric sleeve 121 against the housing 41 in the axial direction with the ring-shaped bolt 122, the eccentric sleeve 121 can be fixed to the housing 41 by a frictional force.

FIG. 11 is a modified example (part 2) of the gear type reduction mechanism according to the present invention. The worm 112 and the worm wheel 113 in the cross section taken along line 11-11 of FIG.
And the eccentric sleeve 121. This figure, with respect to the center O ₃ of eccentric sleeve 121, indicating that the center O ₁ of the worm 112 (center O ₁ of the transmission shaft 111) are eccentric by the eccentricity δ to up or down. Since the transmission shaft 111 is supported at an eccentric position of the eccentric sleeve 121, when the eccentric sleeve 121 is rotated, the worm 1
Twelve centers O ₁ move so as to come into contact with and separate from the center O ₂ .
As a result, the distance X from the center O ₁ of the worm 112 to the center O ₂ of the worm wheel 113 is changed. Therefore, the backlash of the worm 112 with respect to the worm wheel 113 can be easily adjusted only by rotating the eccentric sleeve 121. Also in this modified example, it is preferable that the worm 112 is made of a metal product, the tooth surfaces of the worm 112 are coated with a coating layer made of the low friction agent, and the worm wheel 113 is made of a resin product.

Accordingly, the tooth surface of the worm 112 and the tooth surface of the worm wheel 113 are engaged with each other so that there is no backlash, and the tooth surface of the worm 112 and the tooth surface of the worm wheel 113 are preloaded. Can be engaged. Since there is no backlash, there is no backlash between the worm 112 and the worm wheel 113. Since there is no rattling of the electric motor 8
2 (see FIG. 10) does not affect the tooth surface of the worm wheel 113 from the tooth surface of the worm 112. Therefore, also in the modified example, the durability of the gear type reduction mechanism 110 is further enhanced.

The operation for adjusting the meshing of the gear type speed reduction mechanism 110 is as follows. (1) In FIG. 10 described above, the eccentric sleeve 121 is gradually rotated by a tool with the electric motor 82 and the ring-shaped bolt 122 removed. As a result, the center O of the transmission shaft 111 is
_{Since 1} moves, the backlash of the worm 112 with respect to the worm wheel 113 can be adjusted. (2) After the adjustment is completed, the ring-shaped bolt 122 is tightened, and the eccentric sleeve 121 is fixed to the housing 41 by a frictional force. (3) Torque limiter 90 installed in motor 82
Is inserted into the housing 41 and fitted to the transmission shaft 111. (4) The motor 82 is attached to the housing 41 with the bolts 123 to complete the operation. The bolt hole 124 of the electric motor 82 is a hole larger than a normal bolt hole diameter.
Therefore, the center of the output shaft 83 of the electric motor 82 is
Can easily be adjusted to the center O ₁ of

In the above embodiment, the torque limiter 90 is not limited to a friction type torque limiter. Further, the gear type reduction mechanism 110 is not limited to the worm gear mechanism, and may be, for example, a bevel gear mechanism or a spur gear mechanism.

[0063]

According to the present invention, the following effects are exhibited by the above configuration. The invention according to claim 1 is a gear type speed reduction mechanism having a combined structure of a worm and a worm wheel, wherein one or both of the worm tooth surface and the worm wheel tooth surface is made of a porous material by an electroless plating method. By forming a coating and impregnating the porous coating with a low friction agent, one or both of the worm tooth surface and the worm wheel tooth surface were coated with a coating layer made of a low friction agent. The coefficient of friction of the sliding surface between the worm tooth surface and the worm wheel tooth surface can be reduced with a low friction agent. Therefore, the power transmission efficiency of the gear type speed reduction mechanism having the combined structure of the worm and the worm wheel can be improved.

[Brief description of the drawings]

FIG. 1 is a schematic view of a vehicle steering system according to the present invention.

FIG. 2 is a principle diagram of a steering torque sensor according to the present invention.

FIG. 3 is an overall configuration diagram of an electric power steering device according to the present invention.

FIG. 4 is a sectional view taken along line 4-4 in FIG. 3;

FIG. 5 is a sectional view taken along line 5-5 of FIG. 4;

FIG. 6 is a sectional view of a torque limiter according to the present invention.

FIG. 7 is a configuration diagram of a rack and pinion mechanism according to the present invention.

FIG. 8 is a conceptual diagram of a rack and pinion mechanism according to the present invention.

FIG. 9 is a conceptual diagram of a modified rack and pinion mechanism according to the present invention.

FIG. 10 is a modified example of the gear type speed reduction mechanism according to the present invention (part 1).

FIG. 11 is a modified example of the gear type speed reduction mechanism according to the present invention (part 2).

[Explanation of symbols]

21: Steering wheel (wheel), 22: Steering system, 30:
Electric power steering device, 31: pinion shaft (input shaft), 32: rack and pinion mechanism, 82: electric motor, 110: gear type reduction mechanism, 112: worm, 11
3 ... Warm wheel.

Continued on the front page (51) Int.Cl. ⁷ Identification FI FI Theme Court II (Reference) F16H 55/06 F16H 55/06 (72) Inventor Katsuji Watanabe 1-4-1 Chuo, Wako-shi, Saitama Japan Honda Co., Ltd. Within the Technical Research Institute (72) Inventor Shigeru Yamawaki 1-4-1 Chuo, Wako-shi, Saitama Prefecture Inside Honda Research Institute (72) Inventor Atsuhiko Yoneda 1-4-1 Chuo, Wako-shi, Saitama Honda Co., Ltd. Inside the Technical Research Institute (72) Inventor Yasuhiro Terada 1-4-1 Chuo, Wako-shi, Saitama F-Term in Honda R & D Co., Ltd. (Reference) 3D033 CA04 CA22 3J009 DA20 EA06 EA19 EA23 EA32 EB08 FA08 3J030 AC10 BA03 BC03 CA10 4K022 AA02 AA48 BA14 BA16 DA01 EA04

Claims (1)

[Claims] 1. An electric motor generates an auxiliary torque according to a steering torque, and the auxiliary torque is transmitted through a gear type reduction mechanism.
It is transmitted to the steering rack and pinion mechanism,
In the electric power steering device configured to steer the steered wheels by the rack and pinion mechanism, the gear type reduction mechanism has a combination structure of a worm and a worm wheel, and the tooth surface of the worm and the tooth surface of the worm wheel are combined. A porous film is formed on one or both of the surfaces by an electroless plating method, and by impregnating the porous film with a low friction agent, one or both of the worm tooth surface and the worm wheel tooth surface are formed. An electric power steering apparatus characterized in that the power steering apparatus is coated with a coating layer made of a low friction agent.