JP2000240689A - Rolling bearing clutch - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a rolling bearing clutch which can improve durability to on-off operation of the clutch and improve reliability in high torque transmission by preventing inner and outer rings and a roller from directly coming into contact with each other in turning on the clutch and transmitting torque. SOLUTION: Traction oil is interposed in a contact part between a roller 4 seen from the aixial direction of the roller 4 and raceway surfaces 5, 6 of an inner ring 2 and an outer ring 3, and the roller 4 is made approach the raceway surfaces 5, 6 of the inner ring 2 and the outer ring 3 to generate normal load Fc in traction oil. Shearing strain is applied to traction oil, and tangential force F1 produced by shearing stress proportional to the strain is taken as tractive force to transmit torque between the roller 4 and the inner ring 2 and between the roller 4 and the outer ring 3.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a rolling bearing clutch which performs an on / off operation by moving the raceways of an inner ring and an outer ring toward or away from each other.

[0002]

2. Description of the Related Art As a rolling bearing clutch, for example, one having a configuration as shown in FIG. 9 is known. That is, as shown in FIG.
0, an outer ring 101, and a number of rollers 104 interposed between the outer circumference of the inner ring 100 and the inner circumference of the outer ring 101.
As shown in FIG. 9B, the roller 104
They are arranged at a predetermined twist angle β with respect to the central axis x of 0,101. The raceway surface 102 on the outer periphery of the inner ring 100 is a tapered surface having a truncated cone shape that tapers, and the raceway surface 103 on the inner periphery of the outer ring 101 is also formed in a truncated cone shape corresponding to the raceway surface 102 of the inner ring 100. In addition, since the track surfaces 102 and 103 need to be in line contact with the roller 104, the track surfaces 102 and 103 become a single-plane hyperboloid which is a trajectory obtained by rotating the roller 104 about the central axis x. Inner and outer rings 100, 1
The track surfaces 102 and 103 of 01 are inclined at a predetermined socket angle φ with respect to the central axis x.

The roller 104 contacts the raceway surfaces 102 and 103 of the inner and outer rings 100 and 101 at a predetermined twist angle β, and rotates with the rotation of the inner and outer rings 100 and 101.
Will spirally roll between the raceway surfaces 102 and 103. As shown in FIG. 9B, when the inner ring 100 is rotated in a direction of being screwed into the outer ring 101 via the roller 104, the inner ring 100 and the outer ring 1
01 acts in the axial direction, the roller component 104 between the inner ring 100 and the outer ring 101 is locked, and torque transmission becomes possible.

[0004] On the other hand, when rotated in the reverse direction, FIG.
As shown in FIG.
A rolling component force F2 is generated in a direction in which the rollers 00 and 101 are separated from each other, and the rollers 104 roll and rotate in a free state.
Utilizing such characteristics, it is used as a one-way clutch that transmits torque in one direction and rotates freely in the other direction. The roller 104 is connected to the inner and outer rings 100, 1
01, the inner ring 100 and the outer ring 101
A preload spring 105 for applying a light preload therebetween is provided.

[0005] The rolling bearing clutch having the above configuration is turned on and off.
When used as an off-clutch, a structure shown in FIG. 10 is known. FIG. 10A is a diagram showing the entire structure of the on / off clutch, and FIG. 10B is an explanatory diagram of its operation. As shown in FIG. 10A, the outer ring 101 of the rolling bearing clutch is housed as a piston in a cylinder 107, and fluid pressure such as air pressure is supplied from a port 108 to provide a raceway between the roller 104 and the inner and outer rings 100, 101. 10
The outer ring 101 is moved in a direction in which the outer ring 101 and the outer ring 100 do not come into contact with each other to enable rotation in both forward and reverse directions. To be locked in one direction. A ball spline 109 is interposed between the cylinder 107 and the outer ring 101. Reference numeral 110 denotes a retainer.

[0006]

However, in the above-mentioned conventional rolling bearing clutch, when the clutch is turned on and torque is transmitted, the metal is not interposed between the inner ring and the roller and between the outer ring and the roller without an oil film. However, there is a great difficulty in the durability of the on / off operation of the clutch, and reliability in high torque transmission is lacking. The present invention has been made in view of the above circumstances, and there is no direct metal-to-metal contact between the inner and outer rings and the roller when the clutch is turned on and torque is transmitted. It is an object of the present invention to provide a rolling bearing clutch that can improve the reliability in high torque transmission and can improve the reliability.

[0007]

In order to achieve the above-mentioned object, the present invention comprises an inner ring, an outer ring, and a number of rollers interposed between raceway surfaces of an inner ring outer periphery and an outer ring inner periphery. The rollers are inclined at a predetermined torsion angle (β) with respect to the center axis of the inner and outer rings, and have a hyperboloid shape which is a rotation locus when the rollers of the inner ring and the outer ring are rotated with respect to the center axis. In a rolling bearing clutch configured to lock when the inner and outer rings are relatively rotated in one direction and freely rotate in the opposite direction, the orbits of the roller, the outer ring, and the inner ring as viewed from the axial direction of the roller. The traction oil is interposed in the contact area with the surface, and the roller approaches the raceway surface of the outer ring and the inner ring to generate a normal load (Fc) on the traction oil, giving shear strain to the traction oil and proportional to the strain. It is characterized in that to allow torque transmission between the tangential force (Ft) roller and the inner ring and roller and the outer ring as tractive force caused by Rusendan stress.

According to the present invention, traction oil having a characteristic of behaving like a solid under high pressure is interposed in a contact portion between the roller and the raceway surfaces of the outer ring and the inner ring as viewed from the axial direction of the roller, and the roller is provided at the contact portion. By applying high pressure (1 to 3 GPa) to the traction oil by bringing the traction oil close to the raceway surfaces of the outer ring and the inner ring, the traction oil generates a normal load (pressing force) Fc on the traction oil, A small shear strain is applied to the oil, and torque transmission, that is, traction drive between the roller and the inner ring and between the roller and the outer ring is made possible by using a tangential force Ft generated by a shear stress proportional to the strain as a traction force. Therefore, when the clutch is turned on, the traction oil is always interposed between the contact surfaces between the inner and outer wheels and the rollers while the clutch is on and transmitting the torque, so that the clutch is repeatedly turned on and off in durability. Can be dramatically improved, and the reliability in high torque transmission can be improved.

[0009]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS A rolling bearing clutch according to the present invention will be described below based on the illustrated embodiment. FIG.
It is a sectional view showing the rolling bearing clutch concerning a 1st embodiment of the present invention. The rolling bearing clutch 1 includes an inner ring 2 and
The vehicle includes an outer race 3, a number of cylindrical rollers 4 interposed between the outer periphery of the inner race 2 and the inner periphery of the outer race 3, and a retainer 7 for holding the rollers 4 at predetermined positions. The inner ring 2 is directly connected to the output shaft 8. The outer ring 3 is connected to an outer ring 10 via a ball spline mechanism 9, and the outer ring 3 is relatively movable in the axial direction with respect to the outer ring 10 and integrally rotates in the rotation direction. A guide flange 11 is fixed to the outer ring 3. An input shaft 20 is fixed to the outer ring 10.

As shown in FIG. 1, the outer ring 10 is connected to the input shaft 2.
0, when the inner ring 2 is directly connected to the output shaft 8, when the input shaft 20 is rotated in the direction shown in the figure, the roller 4 moves in a direction sandwiched between the outer ring 3 and the inner ring 2;
, The power is transmitted to the output shaft 8, but when rotated in the opposite direction, the roller 4 moves in a direction away from the input shaft 2, and
Zero power is not transmitted to the output shaft 8. On the other hand, when the outer ring 10 is connected to the output shaft 8 and the inner ring 2 is connected to the input shaft 20, power is transmitted when rotated in the direction shown in FIG. 1, and is not transmitted when rotated in the opposite direction.

A preload spring 12 is interposed between the facing surfaces of the outer ring 3 and the outer ring 10, and the outer ring 3 is constantly pushed rightward by the urging force of the preload spring 12. A radial bearing 13 is interposed between the outer ring 10 and the inner ring 2.

Next, the relationship between the inner ring 2, the outer ring 3, and the rollers 4 will be described in detail. FIG. 2 shows inner ring 2, outer ring 3 and roller 4
2 (a) is a schematic cross-sectional view showing the inner ring 2, the outer ring 3 and the roller 4, and FIG. 2 (b) and FIG.
(C) is a figure which shows the relationship between a roller and a track surface. As shown in FIGS. 2A and 2B, the roller 4 is disposed at a predetermined twist angle β with respect to the center axis x of the inner and outer rings 2 and 3. The raceway surface 5 on the outer periphery of the inner ring 2 is a tapered surface having a truncated cone shape that tapers, and the raceway surface 6 on the inner periphery of the outer ring 3 is also formed into a truncated cone shape corresponding to the raceway surface 5 of the inner ring 2.

Here, the raceway surfaces 5, 6 of the inner and outer races 2, 3 satisfy the function only when the roller 4 comes into line contact, as shown in FIGS. 3 (a) and 3 (b). Is the orbit of the outer and inner circumferences of the roller 4 when is revolved around the central axis x, and has a hyperboloidal shape.

The track surfaces 5, 6 of the inner and outer rings 2, 3 are shown in FIG.
As shown in (c), it is inclined at a predetermined socket angle φ with respect to the central axis x. Figure 2 shows the socket angle φ.
As shown in (b), it is the angle between the tangent a and the center axis x at the roller contact point P (x, y) having a hyperbolic track cross section cut along a plane passing through the center axis x of the inner ring 2 and the outer ring 3. .

The rollers 4 are held at predetermined intervals by a holder 7. The torsion angle β of the roller 4 is not held by the cage 7, but the raceway surfaces 5, 6
Is automatically maintained by the hyperboloid shape of. Cage 7
Is to keep the rollers 4 from falling apart when the inner and outer rings 2 and 3 are disassembled.

Here, the bite angle (ψ) will be described with reference to FIG. FIG. 4A is a diagram illustrating a tapered screw. As shown, the thread 34 of the taper screw 33
Is spirally wound at a predetermined lead angle β with respect to the imaginary cylindrical surface 35, and its outer periphery is
Is located on the outer periphery of the truncated cone 36 whose diameter increases upward at a predetermined socket angle φ. Therefore, the outer diameter of the thread 34 has a wedge shape that gradually increases at a predetermined angle θ as it goes upward with respect to the virtual cylindrical surface 35. This angle θ is defined as a rolling wedge angle.

The rolling wedge angle θ is obtained by calculating the angle of increase or decrease of the screw diameter due to the torsion of the male screw and the female screw when the taper screw is used. The sum of the wedge angles is the actual wedge angle. This wedge angle is defined as the bite angle ψ.

In the case of the rolling bearing clutch 1, the inner race 2
Since the roller 4 interposed between the inner ring 2 and the outer ring 3 has a torsion angle β, when the inner ring 2 and the outer ring 3 are relatively rotated via the roller 4, as if the screw is present, as if by the rolling direction of the roller 4. The inner ring 2 is screwed into the outer ring 3, and the roller 4 bites between the raceways 5, 6 of the inner ring 2 and the outer ring 3, and has exactly the same relationship as the tapered screw 33.

FIG. 4B is a diagram in which the inner and outer rings 2 and 3 of the rolling bearing clutch 1 are spirally cut in the rolling direction of the roller 4. As shown, the raceway surface 5 on the inner ring 2 side
Is inclined so as to gradually increase in diameter with a rolling wedge angle θi in the rolling direction of the roller 4, and the raceway surface 6 on the outer ring 3 side gradually decreases in diameter with a rolling wedge angle θo. It is inclined.

FIG. 4C is a schematic view of the roller 4 interposed between the raceway surfaces 5 and 6 as viewed from the axial direction. If the inner ring 2 and the outer ring 3 are relatively moved in the direction in which the raceway surfaces 5 and 6 become narrower, the roller 4 bites. The biting state of the roller 4 is determined by the rolling wedge angles θi and θi of the inner ring 2 and the outer ring 3.
Since both of o work, they are synthesized and the bite angle ψ
To be evaluated. Here, the raceway surfaces 5 of the inner ring 2 and the outer ring 3
The precondition that the roller 4 rolls into and locks between the rollers 6 and locks is that the contact portions between the roller 4 and the raceway surfaces 5 and 6 do not slip.

However, in the present invention, the roller 4 moves in the rolling and biting direction between the raceway surfaces 5 and 6 of the inner race 2 and the outer race 3, but without being locked, the roller 4 and the raceway surfaces 5 and 6 are not locked.
, On the condition that a slight slip occurs at the contact portion with traction oil. For this purpose, in the present invention, the traction oil is interposed in the contact portions between the raceway surfaces 5 and 6 of the inner race 2 and the outer race 3 and the roller 4 so that the traction drive transmits torque between the inner race 2 and the outer race 3. I have to. Here, traction drive is a drive method that uses traction oil that behaves like a solid under high pressure, applies a slight shear strain to this, and transmits the tangential force with a shear stress proportional to the strain. It refers to a driving method different from large friction transmission. The lubrication state is different from boundary lubrication and fluid lubrication in EHL (hydrodynamic lubrication). Although a slight slip occurs on the surface transmitting the tangential force, the slip ratio is a very small amount, so it is called creep and is distinguished from slip (gross slip) by boundary lubrication or fluid lubrication. In the embodiment of the present invention, as a method of interposing traction oil in the contact portion between the raceway surfaces 5 and 6 of the inner and outer rings 2 and 3 and the roller 4, the entire inner and outer rings 2 and 3 and the roller 4 are immersed in the traction oil. I am trying to do it.

Next, conditions for establishing a traction drive according to the present invention will be described. 5 and 6 are diagrams showing the relationship between the forces acting on the roller. FIG. 5 is a model (cross-sectional view passing through the center of the shaft) schematically showing the mechanical clutch. FIG. FIG. 3 is a diagram (a detailed cross-sectional view in the direction of the end face of the roller) assuming that the vehicle passes through an intersection of tangents formed at contact points of inner and outer rings.

(1) Traction coefficient (μ _t ) The traction coefficient (μ _t ) is defined as a normal force Fc at a contact portion between the raceway surfaces 5 and 6 of the inner ring 2 and the outer ring 3 and the roller 4.
And the ratio of the tangential force Ft. That is, the normal force of the wedge force Fw and the inner and outer rings contact portion acting on μ _t = Ft / Fc roller center Fw = Fcosinα = Fcisinα (∵α = (θo + θi) / 2) ··· Here, .theta.o, .theta.i is Rolling wedge angle determined from roller and hyperboloid. When the traction coefficient was μ _t,
Roller and power relationship of the inner and outer rings contact _{points, Fti = μ t Fc ··· '} Fto as Fco = Fci = Fc than _{Fti = μ t Fci ··· Fto =} μ t Fco ··· from Fig 6 = mu _t Fc ... 'At this time, attention is paid to the inner ring side. (A) Torque Ti by traction drive Input torque is Ti, turning radius of inner ring (at contact point)
The When ri Ti = ri · Fticosθi ··· = μ t · ri · Fccosθi ··· (∵ ') (b) Ftb a force parallel to the central axis acting on the torque Tiw roller center by wedges, rb roller radius Then, Tiw = Ftb (rbcosθi + ri) ... On the other hand, the relationship between Fw and Ftb is Fw = Ftbcos {(θo−θi) / 2}. {(Θo−θi) / 2} = [Fcsin ｛(θo + θi) / 2} (rbcosθi + ri)] / cos {(θo−θi) / 2}... Torque Ti by traction drive ≧ Torque Tiw by wedge Sometimes, a traction drive that transmits torque while causing slight slippage without causing slippage (gross slip) of the contact portion is established. Therefore, the limit traction coefficient μ _t is μ _t = {(1 / cos θi) + rb / ri} · sin {(θo + θi) / 2} / cos {(θo−θi) / 2}. _{If t} is larger than the value obtained by the equation, traction drive is established.

[0024] (2) and the traction curve traction curve, which is determined mainly by the shear stress characteristics of oil under high pressure, it is a relationship between the traction coefficient mu _t and Creep .DELTA.U / U (slip ratio).
The traction curve is the Hertz stress σ [Pa] acting on the contact part.
Is used as a parameter, for example, a curve as shown in FIG. 7 is drawn. Next, a method of obtaining a traction curve will be described. Dimensionless traction J ₄ traction force J elastic band
_4e and the traction force J _{4p in the} plastic region are expressed by the following equation. J ₄ = J _4e + (τ _0t / τ ₀ ) · J _4p ··· Here, J _4e = (3/2) · {S / (1 + S ² ) ² } J _4p = (3/8) · [( π / 2) -sin ^-1 ｛(1-
^{S 2) / (1 + S} 2)} + 2S (S 2 -1) / (1 + S
² ) ² ] S = (3/2) · (G · b / τ _ch · h) · (ΔU / U) G: bulk modulus, b: contact length, τ _c : critical shear stress, h: oil film Thickness (value at each torque) Here, G = αE ′, and α is the pressure viscosity coefficient [P
a ⁻¹ ], and E ′ = E / (1−ν ² ). Here, E is the Young's modulus [Pa], and ν is the Poisson's ratio of the material. Correction of critical shear stress by temperature τ _0t / τ ₀
From the equation of Jonson and Trevaarwerk, τ _0t / τ ₀ = 1 + T _f / (T _θ + D) ⁿ (n> 1) T _f : Temperature rise due to heat generation (obtained by repeated calculation) T _θ : Oil supply temperature D: Temperature constant _{and, J 4 = μ t / μ} max ··· μ max: the maximum traction coefficient, different formulas depending on the type of oil can be more calculated relation of the traction coefficient mu _t and creep .DELTA.U / U is. If this relationship is represented in a graph, a traction curve as shown in FIG. 7 can be drawn. Looking at the line of sigma = 1.4 GPa in this figure, mu _t near slip of 1% is the maximum, then because of heat generation due to sliding, it decreases. Elastic area on the right ascending part,
The portion where the creep is greater than the maximum value is called the plastic zone,
Traction drive elastic range proportional creep and mu _t a (a range indicated by A in FIG. 7) is used.

From the above, in order for traction drive to be established, the traction curve drawn in (2) must be expressed by σ =
It has a maximum value as the curve of 1.4 GPa, and have a traction coefficient mu _t and the intersection defined by (1) a condition. That is, this intersection is the operating point DP of the rolling bearing clutch.

As described above, according to the present invention, the roller 4
The traction oil having the property of behaving like a solid under high pressure is interposed in the contact portion between the roller 4 and the raceway surfaces 5 and 6 of the outer ring 3 and the inner ring 2 as viewed from the axial direction, and the inner and outer rings are relatively rotated in the locking direction. In this case, a normal load (pressing force) Fc is generated in the traction oil by causing the roller 4 to approach the raceway surfaces 5 and 6 of the outer ring 3 and the inner ring 2 at the contact portion, and a slight shear strain is applied to the traction oil. And a tangential force Ft generated by a shear stress proportional to the strain as a traction force between the roller 4 and the inner ring 2 and the roller 4.
And the outer ring 3 enables torque transmission, that is, traction drive. Therefore, when the clutch is turned on, while the clutch is turned on and transmitting torque, and while the clutch is turned off, the inner and outer wheels 2, 3
Since the traction oil is always present between the contact surfaces between the roller and the roller 4, the durability of the repeated on / off of the clutch can be dramatically improved, and the reliability in high torque transmission can be improved. . 7 while the clutch is on and transmitting torque.
A slight slip (1% or less) of the operating point DP shown in FIG.

Next, the overall operation of the rolling bearing clutch shown in FIG. 1 will be described. The input shaft 20 is a prime mover (not shown)
It is connected to. The input shaft 20 is a rolling bearing clutch 1
The case of relative rotation in the lock direction of the inner and outer races will be described.
When the input shaft 20 rotates, the outer ring 3 rotates together with the outer ring 10 in the locking direction, and the rollers 4 spirally roll in the biting direction between the raceway surfaces. When the roller 4 bites between the inner and outer rings 2 and 3, the traction oil is sealed under high pressure, so that traction drive is established, the clutch is turned on, and a slight slippage occurs between the roller 4 and the inner and outer rings 2 and 3. (1% or less) is generated and torque is transmitted.

On the other hand, the input shaft 20 is connected to the rolling bearing clutch 1
When the roller 4 is relatively rotated in the rolling direction opposite to the locking direction of the inner and outer wheels, the roller 4 rolls while sliding in the direction in which the roller 4 comes off, the clutch is turned off, no torque is transmitted, and only the outer wheel 3 rotates freely.

FIG. 8 is a sectional view showing a second embodiment of the present invention. Members having the same functions as in the example shown in FIG. 1 will be described using the same reference numerals. The rolling bearing clutch 1 includes an inner ring 2, an outer ring 3, a number of cylindrical rollers 4 interposed between the outer circumference of the inner ring 2 and the inner circumference of the outer ring 3, and a retainer 7 for holding the rollers 4 at predetermined positions. It has. The inner ring 2 is formed integrally with the output shaft 8. The outer ring 3 is connected to an outer ring 10 via a ball spline mechanism 9, and the outer ring 3 is relatively movable in the axial direction with respect to the outer ring 10 and integrally rotates in the rotation direction. The outer ring 10 has the input shaft 2
0 is fixed.

A preload spring 12 is interposed between the opposing surfaces of the outer ring 3 and the outer ring 10, and applies a preload to the roller 4 by pressing the outer ring 3 against the inner ring 2 in the axial direction. Have been. Further, a pressing force acts on the outer ring 3 so that the raceway surfaces 5 and 6 of the inner ring 2 and the outer ring 3 approach each other. This pressing force can be generated by various pressing force generation means such as a hydraulic pressure, a pneumatic pressure, and an electric actuator. Further, the entirety of the inner and outer rings 2 and 3 and the roller 4 is immersed in traction oil as in the first embodiment.

However, in the present embodiment, unlike the first embodiment, the raceway surfaces 5 of the inner race 2 and the outer race 3 are different.
Roller 4 so that roller 4 does not roll between 6
The condition is that the contact portions between the and the raceway surfaces 5 and 6 are slipped. As a condition, at least one of the rolling wedge angles θi and θo of the contact portions formed between the roller 4 and the raceway surfaces 5 and 6 is larger than the friction angles λi and λo corresponding to the static friction coefficients μsi and μso of the contact portions. It is set large. Here, the friction angle λ is an angle at which the object starts to slide when an object is placed on a flat slope and gradually inclined, and tanλi = μ
si, tanλo = μso.

The contact of the roller 4 does not bite if at least one of the contact portions on the inner ring 2 side and the outer ring 3 side slips, so that θi> λi or θo> λo. Of course, θi> λi and θo> λo may be satisfied. In other words, at least one of the two conditions of tanθi> μsi and tanθo> μso is satisfied.

FIG. 2C shows a composite angle (θi) of θi and θo.
+ Θo) is represented as a model. The inner and outer rings 2 and 3 are made of the same material, and the coefficient of static friction at the contact portion between the roller 4 and the two raceway surfaces 5 and 6 is equal, and the rolling contact angle θ between the inner race 2 and the raceway surfaces 5 and 6 of the outer ring 3.
Since i and θo are considered to be substantially equal, the bite angle (ψ) is set so as to satisfy the relationship of the expression tan (ψ / 2)> μs. There is no practical problem with this setting. In the figure, N is the drag acting on the roller 4 from the contact surface, F is the frictional force, and P is the resultant force.

However, it is difficult to measure and form such a rolling contact angle θ, and in practice, the contact angle θ is set based on the relationship between the torsion angle β of the roller and the socket angle φ.
The contact angle θ, the roller twist angle β and the socket angle φ are
It has a certain geometric relationship.

In the case of the tapered screw model shown in FIG. 4A, the following relationship is obtained. tanθ = sinβ · tanφ That is, in the drawing, h = ltanβ, Δ = htanφ, Δ = l
tanβ and tanφ. Also, l ₁ = 1 / cos β, tan θ = Δ / l
_{From 1} , it is proved that tanθ = cosβ · tanβ · tanφ = sinβ · tanφ. Therefore, the roller twist angle β and the socket angle φ may be set so that sin β · tan φ> μs. The value of μs varies depending on conditions such as the material of the inner ring 2, the outer ring 3, and the roller 4, and the lubrication state. μs is 0.
When the roller twist angle β is about 21 ° to 2
It is preferable to set 4 ° and the socket angle φ in the range of 8 ° to 10 °.

When the torsion angle β of the roller is less than 21 °, locking is easy, and when it is more than 25 °, rolling becomes difficult. Further, when the socket angle φ is less than 8 °, locking is easy, and when it is more than 10 °, slip tends to occur and stability is deteriorated. When sinβ · tanφ is calculated, about 0.057 for β: 24 ° and φ: 8 °,
β: 21 °, φ: 10 °, about 0.063, β: 2
In the case of 1 °, φ: 8 °, it is about 0.0503, which is larger than the maximum static friction coefficient μs, and satisfies the slip condition.
The bite angle 食 at this time is around (ψ / 2) = 3 °.

On the other hand, conventionally, the roller twist angle β
The angle was 15 to 18 °, and the socket angle φ was in the range of 4 to 4.5 °. Calculating sin β · tan φ in the conventional case, it is approximately 0.01 for β: 15 ° and φ: 4 °.
8, β: 18 °, φ: 4.5 °, about 0.024,
In the case of β: 18 ° and φ: 4 °, it is about 0.021, which is smaller than the maximum static friction coefficient, and the lock condition is established.

Of course, the roller torsion angle β and the socket angle φ are relative, and β is 15 ° to 18 ° in the related art.
The socket angle φ may be selected so that sinβ · tanφ is within the range of 0.05 or less, and the socket angle is set to the conventional range of 4 to 4.5 ° in accordance with that.
What is necessary is just to set so that sinβ · tanφ is in the range of 0.05 or less.

The roller torsion angle β other than the above values is 15 ° or less other than the above range, 18 to 21 ° range,
Further, the present invention can be applied to a range of 24 ° or more. Also for the socket angle φ, a range of 4 ° or less other than the above, a range of 4.5 to 8 °, and a range of 10 ° or more can be applied.

Further, the maximum static friction coefficient μ is set to about 0.05, but the static friction coefficient μ can be adjusted, and various values such as 0.1 and 1.5 can be taken. That is, the slip condition can be set by changing the maximum static friction coefficient while keeping the roller twist angle β and the socket angle φ at the conventional angles. That is, even if the conventional roller twist angle is in the range of 15 to 18 ° and the socket angle φ is in the range of 4 to 4.5 °, the slip condition is satisfied if the maximum static friction coefficient μs is smaller than 0.02. Coefficient of static friction μ
And the rolling friction angle θ or the bite angle 食 of the roller 4, and the static friction coefficient is not limited to 0.05.

The relationship between the roller twist angle β and the socket angle φ
In the meantime, as shown in FIG.
There is a clerk. From FIG. 3C, the following relationship is established. tanφ = xtan²β / (F²+ X²・ Tan²β)^1/2 tanθ = sinβ ・ tanφ = sinβ ・ ｛x ・ tan²β / (F²
+ X²・ Tan²β)^{1 / 2}｝ Here, the bite angle ロ ー ラ of the roller 4 is
Is the sum of the rolling contact angles θi and θo of
Indicated by φ = tan^-1[｛(Y_i+ Y_o) Xtan²βsinβ｝ /
｛Y_i・ Y_o− (Xtan ²βsinβ)²｝] F: Arbitrary value given at the time of track design y_i: Inner ring raceway radius at x (F_i ²+ X²・ Tan²
β)^1/2 y_o: Outer ring raceway radius at x (F_o ²+ X²・ Tan²
β)^1/2 F_i= FrF_o= F + r r: roller radius

The relationship will be described below.
As shown in (1), the track cross-sectional shape is determined by the friction angle θ and the twist angle β of the roller 4. The socket angle φ is determined by which range of the determined track cross section is actually used as the track. That is, the X coordinate of the roller contact point P (x in the above equation)
0) and a predetermined value of F. The roller contact is located at the axial center of the roller.

Next, the operation of the rolling bearing clutch constructed as described above will be described. The input shaft 20 is connected to a prime mover (not shown). The case where the input shaft 20 relatively rotates in the locking direction of the inner and outer races of the rolling bearing clutch 1 will be described. Even if the input shaft 20 rotates, the rolling bearing clutch 1
Is off, the outer ring 3 rotates in the locking direction integrally with the outer ring 10, but the inner ring 2 does not rotate because the roller 4 rolls while sliding without spiraling into the raceway. In this state, a pressing force is applied to the outer ring 3 so that the outer ring 3
Against the rightward force of the preload spring 12, and the raceway surfaces of the outer ring 3 and the inner ring 2 approach each other to apply high pressure to the traction oil between the inner and outer rings 2, 3 and the roller 4. Traction drive is established, and the clutch is turned on. That is, the input shaft 20 and the output shaft 8 are coupled with a slight slip, and the input shaft 20
Is transmitted to the output shaft 8. The conditions for establishing the traction drive are as described in the first embodiment. However, in the first embodiment, the inner and outer rings are relatively rotated in the locking direction to apply a normal load (F
Although c) is generated, in the second embodiment, a normal load (Fc) is generated in the traction oil by applying a pressing force to either the inner ring 2 or the outer ring 3.

When the pressing force applied to the outer ring 3 is released during the torque load, the rollers 4 rotate the raceway surfaces 5 of the inner ring 2 and the outer ring 3.
Only the outer ring 3 starts rotating at the contact with the inner ring 2 and the inner ring 2
Does not rotate, and the rolling bearing clutch 1 is immediately turned off. That is, the coupling between the input shaft 20 and the output shaft 8 is immediately released.

On the other hand, the input shaft 20 is connected to the rolling bearing clutch 1
When the roller 4 is relatively rotated in the rolling direction opposite to the locking direction of the inner and outer rings, when there is no pressing force on the outer ring 3, the roller 4 rolls while sliding in the direction in which it comes off, the clutch is turned off, and torque is transmitted. However, only the outer ring 3 rotates freely. If a pressing force is applied to the outer race 3 in this state, the raceways of the outer race 3 and the inner race 2 approach each other and the clutch is turned on. That is, according to the rolling bearing clutch of the present embodiment, bidirectional rotational torque can be transmitted.

Next, a third embodiment of the present invention will be described. Since the structure of the third embodiment is the same as that of the embodiment shown in FIG. 8, the description of the structure is omitted. However, in the present embodiment, unlike the second embodiment, the structures of the raceway surfaces 5 and 6 of the inner ring 2 and the outer ring 3 have the same hyperboloidal shape as the conventional one, and a description thereof will be omitted.

Next, the operation of the rolling bearing clutch according to the third embodiment configured as described above will be described. Input shaft 2
0 is connected to a prime mover (not shown). In the present embodiment, the input shaft 20 relatively rotates only in the rolling direction opposite to the locking direction of the inner and outer races of the rolling bearing clutch 1. Even if the input shaft 20 rotates, when the rolling bearing clutch 1 is off, the outer ring 3 rotates integrally with the outer ring 10, but the rolling in the direction in which the inner and outer rings 2, 3 are separated from each other by the rolling of the roller 4. Since a force is generated and the roller 4 rolls and rotates in a free state, the inner ring 2 does not rotate. When a pressing force is applied to the outer race 3 shown in FIG. 8 in this state, the outer race 3 is pushed to the left, and the raceways of the outer race 3 and the inner race 2 approach each other, so that traction between the inner and outer races 2, 3 and the rollers 4 is achieved. Since high pressure acts on the oil, traction drive is established, the rolling bearing clutch 1 is turned on, and the input shaft 2
0 and the output shaft 8 are coupled. When the pressing force on the outer ring 3 is released during the torque load, the rolling of the roller 4 generates a component force in a direction of separating the inner and outer rings 2 and 3 from each other, and the roller 4 rotates in a free state, and the rolling bearing The clutch 1 is immediately turned off.

[0048]

As described above, according to the present invention,
By interposing traction oil having the property of behaving like a solid under high pressure at the contact portion between the roller and the raceway surface of the outer race and the inner race as viewed from the axial direction of the roller, and relatively rotating the inner and outer races in the locking direction, or By applying a pressing force to bring the inner and outer rings closer to each other in a state where the inner and outer rings are relatively rotated in the locking direction or the rolling direction opposite to the locking direction, the rollers are brought closer to the outer and inner rings at the contact portion, and traction is achieved. Normal load (pressing force) Fc on oil
That gives a slight shear strain to the traction oil by generating traction oil, and enables traction drive, that is, traction drive between the roller and the inner ring and between the roller and the outer ring as a traction force using a tangential force Ft generated by a shear stress proportional to the strain. It is. Therefore, when turning on the clutch, while the clutch is on and transmitting torque, and while the clutch is off, traction oil is always interposed between the contact surfaces between the inner and outer wheels and the rollers, The durability of the repeated on / off operation of the clutch can be remarkably improved, and the reliability in high torque transmission can be improved.

[Brief description of the drawings]

FIG. 1 is a sectional view showing a first embodiment of a rolling bearing clutch according to the present invention.

FIG. 2 is an explanatory view of the rolling bearing clutch shown in FIG.

FIG. 3 is an explanatory diagram of a raceway surface of a rolling bearing clutch.

FIG. 4 is an explanatory diagram of a wedge angle of a roller of a rolling bearing clutch.

FIG. 5 is a diagram illustrating a relationship between forces acting on rollers.

FIG. 6 is a diagram illustrating a relationship between forces acting on rollers.

FIG. 7 is a graph showing a relationship between a traction coefficient and creep (slip ratio).

FIG. 8 is a sectional view showing a second embodiment of the rolling bearing clutch according to the present invention.

FIG. 9 is a view showing a conventional rolling bearing clutch.

FIG. 10 is a view showing a conventional rolling bearing clutch.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Rolling bearing clutch 2 Inner ring 3 Outer ring 4 Roller 5, 6 Track surface 7 Cage 8 Output shaft 9 Ball spline mechanism 10 Outer ring 12 Preload spring 20 Input shaft 23 Radial bearing

 ──────────────────────────────────────────────────続 き Continuing on the front page (72) Inventor Kazuhiko Sugiyama 11-1 Haneda Asahimachi, Ota-ku, Tokyo Ebara Corporation (72) Inventor Hirohisa Tanaka 1-1-15-3 Ookayama, Meguro-ku, Tokyo

Claims (6)

[Claims] An inner ring, an outer ring, and a number of rollers interposed between raceways of an inner ring outer circumference and an outer ring inner circumference, wherein the rollers have a predetermined twist angle (β) with respect to a center axis of the inner and outer rings. ), The raceways of the inner ring and the outer ring are formed in a hyperboloidal shape which is a rotation locus when the roller is rotated with respect to the center axis, and locked when the inner and outer rings are relatively rotated in one direction. In a rolling bearing clutch configured to rotate relatively freely in the opposite direction, traction oil is interposed in a contact portion between the roller and the raceway surfaces of the outer ring and the inner ring as viewed from the axial direction of the roller, and the roller is formed by the outer ring and the outer ring. By bringing the traction oil close to the raceway surface of the inner ring, a normal load (Fc) is generated in the traction oil, a shear strain is applied to the traction oil, and a tangential force (Ft) generated by a shear stress proportional to the strain is defined as a traction force. A rolling bearing clutch characterized in that torque can be transmitted between a roller and an inner ring and between a roller and an outer ring. 2. A normal load (Fc) is applied to the traction oil by bringing the rollers close to the raceways of the outer ring and the inner ring.
2. The rolling bearing clutch according to claim 1, wherein the rotation is generated by relatively rotating the inner and outer rings in a locking direction. 3. A normal load (Fc) is applied to the traction oil by bringing the rollers close to the raceways of the outer ring and the inner ring.
The rolling bearing clutch according to claim 1, wherein the generation of the frictional force is performed by applying a pressing force to one of the outer ring and the inner ring. 4. The roller according to claim 1, wherein a relative speed difference (slip ratio) between the outer ring and the inner ring is 1% (0.01) or less, and a lubricating state is elastohydrodynamic lubrication (EHL). Item 7. A rolling bearing clutch according to item 1. 5. A condition for the roller to slip at a relative speed difference of 1% or less with respect to an outer ring and an inner ring, wherein a ratio of a tangential force (Ft) to a normal load (Fc) is determined as a traction coefficient (friction coefficient). and mu _t, .theta.i the angle of the common tangent to the inner ring axis (hereinafter rolling referred to as a wedge angle)
And then, the distance between the inner ring common tangent and the center axis of the contact portion when the ri, torque generated by the oil film represented by _{Ti = μ t · ri · cosθi} ( traction force)
Tiw = Ftb / (ri + rb) where Ftb is the force that the roller bites into the inner and outer rings, and rb is the radius of the roller.
· Rolling bearing clutch according to claim 1, wherein the torque generated by the roller bites into the inner and outer rings is characterized by having a traction coefficient mu _t such that the relation of Ti> Tiw represented by cos .theta.i. 6. A condition for the roller to slide at a relative speed difference of 1% or less with respect to an outer ring and an inner ring, wherein a ratio of a tangential force (Ft) to a normal load (Fc) is determined as a traction coefficient (friction coefficient). and mu _t, the angle of the common tangent to the axis of the outer ring .theta.o (hereinafter rolling referred to as a wedge angle)
And then, when the distance between the outer common tangential line and the center axis of the contact portion with the ro, torque generated by the oil film represented by _{To = μ t · ro · cosθo} ( traction force)
When the force of the roller biting into the inner and outer rings is Ftb, and the radius of the roller is rb, Tow = Ftb / (ro-rb)
2. The rolling bearing clutch according to claim 1, wherein the clutch has a traction coefficient [mu] _t such that the relationship between the torque generated when the roller bites into the inner and outer races represented by cos [theta] o and To> To is satisfied.