KR101486880B1 - Flexible engagement gear device and method for determining shape of gear tooth of flexible engagement gear device - Google Patents

Abstract

The impact resistance is improved, and the transmission torque and transmission efficiency are further increased. 120A and 120B having flexibility and bending with the internal gear 130A for deceleration and the internal gear 130B for output having rigidity with which the external gears 120A and 120B respectively engage with each other, The tooth gears of the gear teeth 120A and 120B at the portions engaged with the internal gear 130A for deceleration and the internal gear 130B for output are the same as those of the external gears 120A and 120B, The deceleration internal gear 130A and the internal gear 130B for output are arranged so that the simultaneous engagement number Nph of the external gear 120A and the internal gear 130A for deceleration and the internal gear 120B for output and the internal gear 130B for output, And the number of simultaneous meshing (Npl) with respect to each tooth is two or more.

Description

BACKGROUND OF THE INVENTION 1. Field of the Invention [0001] The present invention relates to a method of determining a tooth shape of a bending gear and a bending gear,

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a bending gear type gear device and a method of determining a tooth profile of a bending gear type gear device.

The bending-mesh type gear device shown in Patent Document 1 includes a vibrating body, a tubular external gear disposed on the outer periphery of the vibrating body and having flexibility that is flexibly deformed by rotation of the vibrating body, And a second internal gear which is rigid in the axial direction and engages with and meshes with the external gear, the first internal gear having a stiffness in which the external gear is in contact with the internal gear, and the second internal gear in axial alignment with the first internal gear.

Therefore, when the first internal gear is fixed to the casing, the external gear which is deflected and deformed by the rotation of the vibrator is in contact with the first internal gear and is engaged with the first internal gear, do. Then, the output of the decelerated external gear can be taken out from the second internal gear.

(Patent Literature)

Patent Document 1: JP-A-2006-29508

However, in the warping gear type gear device as shown in Patent Document 1, it is necessary to realize meshing with the internal gear by bending the external gear, and in the case of the cylindrical external gear, It is difficult to theoretically engage the two internal gears and the external gears, and the number of theoretical engagements as the rigid gears is very small. As a result, the bending-mesh type gear unit using the conventional cylindrical external gear has low impact resistance, low transmission torque, and low transmission efficiency.

Accordingly, the present invention has been made to solve the above problems, and it is an object of the present invention to provide a method of determining the tooth profile of a warping gear type gear device and a warping gear type gear device which are improved in impact resistance and capable of further increasing transmission torque and transmission efficiency .

According to the present invention, there is provided a vibration damping device comprising: a vibration body; a cylindrical external gear disposed on the outer periphery of the vibration body and flexibly deformed by rotation of the vibration body; And a second internal gear having a stiffness that is in contact with and meshes with the external gear, the first internal gear and the second internal gear, the first internal gear and the second internal gear, The internal teeth and the first internal gear and the internal teeth of the internal gear are the same, and the external teeth gear, the first internal gear and the second internal gear form the number of simultaneous meshing of the external gear and the first internal gear, And the number of simultaneous meshing with the second internal gear is 2 or more, respectively.

The present invention is characterized in that a tooth profile is set so that both the internal gear and the two internal gears (the first internal gear and the second internal gear) are simultaneously engaged with each other so that the external gear, the first internal gear and the second internal gear It is. As a result, the impact resistance is improved and the surface pressure applied to the tooth surface of the engagement is dispersed, so that a large torque can be transmitted. According to the present invention, as a basic structure, there is provided a structure in which a cylindrical external gear is engaged with an internal gear having two rigidity, and also a latching resistance can be improved, The stress generated in the external gear can be reduced as compared with that of the cup-shaped external gear, so that the load capacity can be increased. Thus, the present invention can increase transmission torque and increase transmission efficiency.

Since the tooth profile of the external gear is the same at the portions engaging with the first internal gear and the second internal gear, it is possible to easily machine the external gear, reduce the machining cost, It is possible to do.

According to the present invention, there is also provided a vibration damping device comprising: a vibration body; a cylindrical external gear disposed on the outer periphery of the vibration body and flexibly deformed by rotation of the vibration body; A first internal gear and a second internal gear provided in the axial direction on the first internal gear and having a rigidity in contact with the external gear in a meshing manner, A straight line passing through an eccentric shaft which is the center of the engagement radius of the external gear at the time of engagement with the first internal gear or the second internal gear and a contact point formed by engagement of the external gear with the first internal gear and the second internal gear When the external tooth of the external gear is made into a cylindrical pin or between a pitch point which is an intersection point with each common normal line and a cylindrical pin, And the center of the pin is disposed when the internal tooth of the internal gear or the second internal gear is formed into a cylindrical pin or when the inner tooth is imaginary with a cylindrical pin. Specifically, when the internal teeth of the first internal gear and the second internal gear are imagined by cylindrical pins, specifically, the external teeth are determined based on the fins, and based on the external teeth, the first internal gear and the second internal gear The inner tooth is formed as an envelope.

The present invention is characterized in that when the outer tooth of the external tooth gear is formed into a cylindrical pin between the two pitch points and the center of the pin or the inner tooth of the first internal gear or the second internal gear is assumed to be a cylindrical pin The center of the pin is disposed. As a result, the load applied to the external teeth of the cylindrical external gear at the time of engagement with the first internal gear and the load applied to the external teeth of the cylindrical external gear at the time of engaging with the second internal gear form have opposite directions And the two load regions applied to the external gear can be brought close to each other in the circumferential direction of the external gear. That is, when viewed from the axial direction, in the engagement operation, the two internal gears can be configured such that only a small number of external teeth are sandwiched. As a result, the phenomenon that the meshing of the external gear and the internal gear is shifted to an excessive torque (latching phenomenon) can be prevented in particular. In other words, the present invention makes it possible to increase permissible transmission torque and increase transmission efficiency, in particular focusing on improvement of ratchetting property.

According to the present invention, the impact resistance is improved, and it becomes possible to increase the transmission torque and the transmission efficiency.

1 is an exploded perspective view showing an example of the entire configuration of a bending gear capable of engaging with gear according to a first embodiment of the present invention.
2 is a cross-sectional view showing an example of the entire configuration in the same manner.
Fig. 3 is a diagram showing a vibrator.
Fig. 4 is a diagram showing a vibrator.
Fig. 5 is a schematic view showing a combination of a vibrator and a vibrator bearing.
Fig. 6 is a diagram showing engagement between the external gear and the internal gear in the same manner.
Fig. 7 is an enlarged view of engaging the external gear, the internal gear for deceleration, and the internal gear for output.
Fig. 8 is a diagram showing the positions of teeth-like bodies of the external gear, the internal gear for deceleration, and the internal gear for output.
Fig. 9 similarly shows the tooth profile of the external gear.
10 is a view for defining the teeth of the internal gear for deceleration and the internal gear for output.
11 is a view for defining the teeth of the internal gear for deceleration and the internal gear for output.
12 is a view for defining the teeth of the internal gear for deceleration and the internal gear for output.
13 is a table showing the relationship between the circumferential length, the number of teeth, and the pitch of the internal gear for deceleration, the internal gear for output, and the external gear.
14 is a diagram showing the relationship between the pitch point and the position of the actual body of the external gear.
15 is a diagram showing the relationship between the pitch point and the position of the actual body of the external gear.
Fig. 16 is a diagram showing the modification of the teeth of the internal gear for deceleration and the internal gear for output.
17 is a table showing the number of simultaneous meshing in the internal gear for deceleration when the reduction ratio and the diameter of the internal gear are changed in the first embodiment.
18 is a table showing the number of simultaneous meshing in the internal gear for output when the reduction ratio and the diameter of the internal gear are changed in the first embodiment.
19 is a diagram showing the relationship between the position of the actual gear of the external gear and the pitch point in the first embodiment.
20 is an exploded perspective view showing an example of the overall configuration of a warping gear type gear apparatus according to a second embodiment of the present invention.
21 is a cross-sectional view showing an example of the entire configuration in the same manner.
22 is a view for defining the tooth profile of the external gear.
23 is a view for defining the teeth of the internal gear for deceleration and the internal gear for output.
24 is a diagram showing the relationship between the pitch point and the position of the actual body of the internal gear in the same manner.
25 is a diagram showing the relationship between the pitch point and the position of the actual body of the internal gear.
26 is a table showing the number of simultaneous meshing in the internal gear for deceleration when the reduction gear ratio and the diameter of the internal gear are changed in the second embodiment.
Fig. 27 is a table showing the number of simultaneous meshing in the internal gear for output when the reduction gear ratio and the diameter of the internal gear are changed in the second embodiment. Fig.
28 is a view showing the relationship between the position of an actual gear of the internal gear and the pitch point in the second embodiment.
29 is a view showing an effect of preventing ratchet in the second embodiment.
30 is a diagram for finding a contact line between the external gear and the internal gear for output and the internal gear for output according to the first embodiment.
31 is a view showing the contact line in the same manner.

Hereinafter, an embodiment of the present invention will be described in detail with reference to the drawings.

≪ First Embodiment >

<Configuration>

First, the overall configuration of the present embodiment will be schematically described mainly with reference to Figs. 1 and 2. Fig.

The flexural-mesh type gear device 100 includes a vibrator 104 and a plurality of external gears 120A and 120B which are disposed on the outer periphery of the vibrator 104 and are flexible to be flexibly deformed by rotation of the vibrator 104 And the output internal gear 130B, which is the second internal gear, is a first internal gear having a stiffness in which the external gear 120 meshes with and engages with the external gear 120, ). Hereinafter, the internal gear 130A for deceleration and the internal gear 130B for output are collectively referred to simply as the internal gear 130.

Hereinafter, each component will be described in detail.

As shown in Figs. 3A and 3B, the vibrating body 104 has a columnar shape and an input shaft hole 106 into which an input shaft (not shown) is inserted is formed at the center. A keyway 108 is formed in the input shaft hole 106 so that the vibrating body 104 rotates integrally with the input shaft when the input shaft rotates.

As shown in Figs. 3 and 4, the vibrator 104 is formed in a shape (two-arcuate shape) in which two circular arc portions (a first circular arc portion FA and a second circular arc portion SA) are connected . The first circular arc portion FA is an arc having a radius of curvature r1 centered on a point B (referred to as an eccentric axis) and is a circular arc portion for engaging the external gear 120 and the internal gear 130 Quot;). The second circular arc portion SA is a circular arc having a radius of curvature r2 centered on the point C and an arc portion (also referred to as a non-engagement range) in a range in which the external gear 120 and the internal gear 130 are not engaged Respectively. The length of the first arcuate portion FA is determined by the engagement angle? That is an angle formed by the long axis x and the normal line N at the point A.

4, when the radius of the major axis x of the vibrating body 104 is r, the radius of curvature r1 of the first circular arc FA is expressed by Expression (1) with the amount of eccentricity as L. In this case, Respectively.

r1 = r-L ... (One)

As shown in Fig. 4, the tangent line (T, normal line N) is common to the connecting portion A between the first circular arc portion FA and the second circular arc portion SA. Therefore, the radius of curvature r2 of the second circular arc portion SA is (radius of curvature r1 + length BC), which is expressed by Equation (2).

r2 = r1 + length (BC)

  = r1 + L / cos? (2)

The vibrator bearing 110A is a bearing that is disposed between the outer side of the vibrating body 104 and the inner side of the external gear 120A. As shown in Figs. 2 and 5, the inner ring 112, A roller 114A as a rolling body, a roller 116A as a rolling body, and an outer ring 118A. The inner side of the inner ring 112 abuts against the vibrating body 104, and the inner ring 112 rotates integrally with the vibrating body 104. The roller 116A is a cylindrical shape (including a needle). As a result, the portion where the roller 116A contacts the inner ring 112 and the outer ring 118A is increased as compared with the case where the rolling body is a sphere, so that the load capacity can be increased. In other words, by using the roller 116A, the transmission torque of the vibrator bearing 110A can be increased and the life time can be increased. The outer ring 118A is disposed outside the roller 116A. The outer ring 118A is bent and deformed by the rotation of the vibrating body 104 and deforms the external gear 120A disposed outside the outer ring 118A.

2, the vibrator bearing 110B includes the inner ring 112, the support device 114B, the roller 116B, and the outer ring 118B as well as the vibrator bearing 110A . The vibrator 104 and the inner ring 112 are common to the vibrator bearings 110A and 110B. The supporting unit 114B, the roller 116B and the outer ring 118B are the same as the supporting unit 114A, the roller 116A and the outer ring 118A as a single member (part).

As shown in Fig. 2, the external gear 120A is in contact with and engages with the internal gear 130A for deceleration. The external gear 120A is composed of a base member 122 and a foreign tooth 124A. The base member 122 is a flexible tubular member that supports the external teeth 124A and is disposed outside the vibrator bearings 110A. The external teeth 124A and 124B of the present embodiment and the external gears 120A and 120B and the flexural mesh type gear device 100 of the present embodiment are simply formed into a cylindrical shape having a radius rho 1 Also referred to as a pin type). The external teeth 124A are supported on the base member 122 by a ring member 126A.

As shown in Fig. 2, the external gear 120B meshes with the internal gear 130B for output. The shout gear 120B is constituted by a base member 122 and an external tooth 124B in the same manner as the shout gear 120A. The external teeth 124B are formed of the same cylindrical fins in the same number as the external teeth 124A and are supported by the base members 122 with the ring members 126B. That is, the base member 122 commonly supports the external teeth 124A and external teeth 124B. That is, the external gears 120A and 120B have the same tooth shape. The eccentric amount L of the vibrating body 104 is transmitted to the external teeth 124A and external teeth 124B in the same phase. Hereinafter, the external values 124A and 124B will be referred to collectively as the external value 124. [

As shown in Fig. 2, the internal gear 130A for deceleration is formed of a rigid member. The internal gear 130A for deceleration has a dimension larger than the dimension of the external teeth 124A of the external gear 120A by a multiple of 2 (the dimensions will be described later). A casing (not shown) is fixed to the internal gear 130A through a bolt hole 132A. The internal gear 130A for deceleration is engaged with the external gear 120A, thereby contributing to the deceleration of the rotation of the vibrating body 104. [ 6A shows a state in which the external gear 120A and the internal gear 128A are in engagement with the internal gear 128A on the x axis in the state in which the deceleration internal gear 130A is engaged.

On the other hand, the internal gear 130B for output is also formed of a rigid member like the internal gear 130A for deceleration. The output internal gear 130B has a dimension of the internal tooth 128B equal to the dimension of the external teeth 124B of the external gear 120B (constant velocity transmission). However, an output shaft (not shown) is mounted on the internal gear 130B for output via the bolt hole 132B, and the same rotation as the rotation of the external gear 120B is output to the outside. 6B shows a state in which the external gear 120B and the internal gear 128B for output are engaged with each other and FIG. 7B shows a state in which the external tooth 124B and the internal tooth 128B on the x axis are engaged. Then, the internal teeth 128A and 128B are collectively referred to as the internal teeth 128.

The simultaneous engagement number Nph between the external gear 120A and the internal gear 130A for deceleration and the simultaneous engagement number Npl between the internal gear 120B and the internal gear 130B for output are both 2 or more And the engagement is made the theoretical engagement. As a result, the transmission efficiency of the torque is not reduced, smooth torque transmission can be realized, and the transmission torque can be increased.

<Determination method of tooth profile>

A method of determining the tooth profile of the external gear 120, the internal gear 130A for deceleration, and the internal gear 130B for output will be described.

First, a method of obtaining a tooth profile will be described below.

First, the tooth profile of the shout gear 120 is defined. Next, the trajectory of the tooth profile of the external gear 120 is represented by a trochoidal curve, and the tooth profile of the internal gear 130 is defined by using the trochoidal curve. Next, a plurality of parameters defining the tooth profile of the external gear 120 and the internal gear 130 are related to each other from the size and the dimension of the external gear 120 and the internal gear 130. Next, the correction range of the tooth line of the tooth profile of the internal gear 130 and the tooth circle is determined. Next, using the associated parameters, a tooth portion outside the correction range is obtained, and the number of simultaneous meshing in the tooth portion is obtained. Then, the optimal parameters are determined so that the number of simultaneous meshing is all 2 or more. In the determination of the parameters, trial and error are performed so as to satisfy simultaneously the target values such as torque, allowable surface pressure of the tooth surface, principal stress at each point, and bearing life.

This will be described in detail below.

First, the tooth profile of the external gear 120 is defined.

When the external tooth 124 is formed into a cylindrical pin having a radius of r1, the center position of the pin, which is the external tooth 124 in the engagement range of the external gear 120 from the eccentric shaft B, Is referred to as the radius of the tooth-like substance in the engagement range of the external gear 120. When the inner tooth 128 of the internal gear 130 is formed into a cylindrical pin having a radius of rho 2 (including a case where the design is simulated in a simple manner), the rotational axis Fc of the vibrating body 104, ) To the center position (r2 = 0) of the fin (including the fictitious part) to the internal tooth 128 is referred to as the radius of the teeth of the internal gear 130. Then, as shown in Fig. 8, the relationship between the radius R and the radius R1 is expressed by the equation (3).

R1 = R-L ... (3)

In this embodiment, the external gear 120 is disposed on the outer periphery of the vibrating body 104 through the vibrating body bearing 110. [ The radial thicknesses of the vibrator bearing 110 and the shout gear 120 are all constant. Thus, from the point that the vibrating element 104 has a two-arcuate shape, the external gear 120 also has a two-arcuate shape. The radius of the tooth-like substance in the engagement range of the external gear 120 corresponding to the radius of curvature r1 of the engagement range of the vibrator 104 is R1. Therefore, when the radius of the tooth-like substance in the non-meshing range of the external gear 120 corresponding to the radius of curvature r2 of the non-meshing range of the vibrator 104 is R2, 3), the radius R2 can be expressed by equation (4).

R2 = R1-L / cos? (4)

9, the external teeth 124 are cylindrical fins having a radius r1 on the circumference of the radius R1 (= RL) from the eccentric shaft B in the engagement range And the eccentric shaft B becomes the center of the engagement radius of the external gear 120 when the external gear 120 and the internal gear 130 are engaged.

Therefore, the tooth profile of the external gear 120 is defined by the radius? 1, the eccentric amount L, the radius R and the engagement angle?.

Next, the tooth profile of the internal gear 130 is defined. The locus of the position of the tooth-like substance of the external gear 120 (radius rho 1 = 0 position) is obtained and then shifted inward by the radius rho 1 as the tooth shape of the internal gear 130. Hereinafter, this will be described in more detail. However, the reduction ratio when the external gear 120 is a circular gear (referred to as a virtual gear) having a radius R1 of a tooth-shaped entity is referred to as a virtual reduction ratio n.

As shown in Fig. 10, the external gear 120 is angularly revolved about the rotational axis Fc of the vibrating body 104. As shown in Fig. That is, the eccentric shaft B rotates by?. At this time, the coordinate (x1, y1) of the position of the teeth of the teeth gear 120 rotates in the opposite direction by the angle? / N by the virtual reduction ratio n, . Therefore, the coordinates (x _pfc , y _pfc ) indicating the locus of the position of the teeth-like substance of the external gear 120 are expressed by the equations (5) and (6).

[Number 1]

11, the coordinate of the position of the tooth-like substance of the internal gear 130 is the same as that of the external tooth gear 120, ) It appears as a trochoid curve (hypotrochoid curve). That is, the radius b1 of the base circle BA fixed around the rotation axis Fc, the radius a1 of the rotation circle AA that does not slip along the circumference of the base circle BA, The coordinates (x _pfc , y _pfc ) of the position of the tooth-like substance of the internal gear 130 are expressed by the equations (7) and (8) using the radius L1 and the rotation angle?

[Number 2]

Here, by using the relationships of the equations (9) to (11), the relationship of the equations (12) and (13) is obtained.

[Number 3]

However, since Equation (5) and Equation (12) (Equations (6) and (13)) represent the same coordinates (x _pfc , y _pfc ), Equation (14) is obtained.

α = n * β ... (14)

Next, as shown in Fig. 12, the coordinate (x _pfc , y _pfc ) of the position of the tooth- _shaped substance of the internal gear 130 is set to the inside (the internal gear 130 side) by the radius rho 1 of the external tooth 124 The coordinates (x _fc , y _fc ) of the teeth of the internal gear 130 can be expressed by the equations (15) to (17).

[Number 4]

In other words, the radius (R), ρ1, eccentric amount (L), a virtual gear ratio (n, a virtual gear ratio for making the teeth of the reduction internal gear (130A) for the (n _h), Virtual for making the teeth of the output internal gear (130B) (X _fc , y _fc ) of the tooth profile of the internal gear for acceleration 130A and the internal gear for output 130B can be obtained by changing the angle β by substituting the speed reduction ratio n ₁ .

Next, the parameters defining the internal gear 130 and external gear 120 are related to each other.

As described above, the shape of the external gear 120 is a two-arc shape defined by radii R1 and R2. Therefore, when the parameter i (i, the internal gear for deceleration 130A) for deriving the parameter (k, 2 or more) representing the dimensional difference between the external gear 120A and the internal gear 130A for deceleration and the reduction gear ratio N (the radius (R, R1) of the tooth-like substance of each tooth of the external gear 120 and the internal gear 130 shown in FIG. 13 (i = 0 for i = 1 and i = 0 for the internal gear for output 130B) (P, a circumferential length of a cycle of one tooth (齒)) when the circumferential length (LC) (circumferential length) obtained from the virtual gear and the virtual reduction ratio (n) of the virtual gear are used, ) Can be expressed as a table. Here, since the pitch P due to the virtual gear and the pitch (= LC / NT) due to the shout gear 120 are the same, the relationship of the equation (18) exists.

NT = LC / P ... (18)

Using equation (18), equation (19) and equation (20) can be derived from the table of FIG.

[Number 5]

Next, a parameter (Gp, referred to as a pin type pitch coefficient) is introduced. Here, a straight line passing through the eccentric shaft B and the rotation axis Fc and a common point of contact points generated by engagement of the tooth gear 120 (the tooth outer tooth 124) and the inner tooth gear 130 (the tooth inner tooth 128) The point of intersection with the normal is called the pitch point by the external gear 120 and internal gear 130. The pin-type pitch coefficient Gp can easily grasp the relative positional relationship between the position of the teeth of the external tooth gear 120 and the tooth point of each tooth of the internal gear and the pitch point, . More specifically, as shown by the equation (21), the pin type pitch coefficient Gp is determined by the radius R1 (= RL) from the eccentric shaft B and the pitch And the distance to the point (n * L).

[Number 6]

14 shows the relationship between the radius RL of the tooth profile of the external gear 120 and the virtual reduction ratio n (n) of the teeth of the external gear 120. In the case where the point P _h indicates a pitch point by the external gear 120A and the internal gear 130A for deceleration, _h ). The pin-type pitch coefficient (Gph, referred to as pin-type deceleration-side pitch coefficient) obtained at this time is defined by equation (22) based on equation (21). In Expression (19) and Expression (20), Expression (23) is obtained by summarizing Expression (22) with parameter (i) = 1.

[Numeral 7]

Point (P _l) the external tooth gear (120B) and the output internal gear radius (RL) and the virtual gear ratio of the substance of the teeth of the case showing the pitch point by (130B), shouting in Fig gear (120) (n _l ). The pin-type pitch coefficient (Gpl, referred to as pin-type output side pitch coefficient) obtained at this time is defined by equation (24) based on equation (21). In Expression (19) and Expression (20), Expression (25) is obtained by summarizing Expression (24) with parameter (i) = 0.

[Numeral 8]

Therefore, when the radius R, the reduction ratio N, the pin type reduction pitch coefficient Gph and the engagement angle? Are given, the virtual reduction ratio n _h and the eccentricity L are determined, pitch coefficient (Gpl), can be determined a virtual gear ratio (n _l).

In this embodiment, as shown in Fig. 14 and Fig. 15, the value of the pitch type output side pitch coefficient Gpl &gt; 1 is obtained by substituting the pin type reduction pitch coefficient Gph <1. In this embodiment, when the engagement angle? Is 40 to 65 degrees and the value of cos- ¹ of the pin type reduction pitch coefficient Gph is 15 to 30 degrees, It is a desirable condition.

Next, the correction range of the tooth profile of the internal gear 130 is determined.

16, the angle when the angle? Formed by the straight line connecting the coordinates of the internal tooth 128 and the center Oc of the external tooth 124 (pin) becomes approximately 45 degrees is? S do. If the angle beta is in the range of from zero to s, there is a risk of interference with the external teeth 124 of the external gear 120, and the internal teeth 128 of the internal gear 130 are corrected in the range. The angle? When the distance delta between the tooth line of the external teeth 124 and the tooth line of the internal teeth 128 becomes about 15% of the radius? 1 of the pins is referred to as an angle? F. There is a possibility that a high surface pressure occurs between the angle f of the angle? With the external teeth 124 of the external gear 120 and the external surface 124 of the external gear 120 from? F to? 130 by the tooth line of the inner tooth 128. That is, the angle (? S to? F, the region of the tooth shape that has not been corrected) at which the tooth profile is not corrected is the effective range at which theoretical engagement is performed.

Next, the number of simultaneous meshes (Nph, Npl) is obtained.

The simultaneous engagement numbers Nph and Npl can be obtained by dividing the effective range determined by the rotation angle? Of the external gear 120 by the pitch angle (2? Divided by the dimension NT). Here, the angles? Fh and? Sh are angles in the internal gear for acceleration 130A and the angles? Fl and? Sl are angles in the internal gear for output 130B. The rotation angles obtained by the angles (? Fh,? Sh,? Fl,? Sl) from the relationship of the expression (14) are? Fh,? Sh,? Fl and? That is, by using the equation (14), the simultaneous engagement number Nph of the internal gear for acceleration 130A and the simultaneous engagement number Npl of the internal gear 130B for output in the equation (27) Is obtained.

[Number 9]

Calculate the number of simultaneous engagements along Eqs. (26) and (27). The simultaneous engagement number Nph of the internal gear 130A for deceleration obtained when k = 2 is shown in Fig. 17 and the number of simultaneous engagement Npl of the internal gear 130B for output is shown in Fig.

The gear teeth of the internal gear 130 in the present embodiment are formed so as to satisfy the condition of the diameter (2 * R) and the reduction gear ratio (1 / N) for achieving two or more simultaneous engagement numbers (Nph, Npl) . In other words, when the difference in dimension is 2 (k = 2), the reduction ratio (1 / N) is 1/20 and does not become the tooth form of the present embodiment, but is reduced to 1/30 or less The tooth profile of the internal gear 130 of the embodiment is determined.

<Operation>

The operation of the flexural-mesh type gear device 100 will be mainly described with reference to Fig.

When the vibrating body 104 is rotated by the rotation of the input shaft (not shown), the external gear 120A is deflected via the vibrating body bearing 110A in accordance with the rotation state thereof. At this time, however, the external gear 120B also undergoes bending deformation in phase with the external gear 120A through the vibrator bearing 110B.

The bending deformation of the external gear 120 is made according to the shape of the curvature radius r1 of the vibrating body 104. [ At the position of the first circular arc portion FA of the vibrating body 104 shown in Fig. 4, the curvature is constant, so that the bending stress becomes constant. Since the tangent line T is the same at the connecting portion A between the first circular arc portion FA and the second circular arc portion SA, abrupt bending deformation in the connecting portion is prevented. At the same time, since there is no abrupt positional fluctuation of the rollers 116A and 116B in the connecting portion A, the slippage of the rollers 116A and 116B is small and the transmission loss of the torque is small.

The external tooth 124 is radially outwardly moved in the portion of the first circular arc portion FA (engagement range) as the external tooth gear 120 is flexed and deformed in the vibrating body 104, 128). Since the external teeth 124 are rotatable pins, the external teeth 124 are moved closer to the rolling surfaces on the engagement surface, and the external teeth 124 slide on the side of the base member 122 where the surface pressure is lower than the engagement surface . As a result, the loss of the transfer efficiency is small. The teeth of the internal teeth 128 are formed in a tooth shape based on the trochoidal curve with respect to the external tooth 124, which is a cylindrical pin. As a result, the external teeth 124 and the internal teeth 128 are completely theoretical meshing, and the loss can be reduced to realize a high torque transmission efficiency.

A load (direction and size) different from the external teeth 124B is applied to the external teeth 124A (different from the external teeth 120 of the present embodiment, see Fig. 29). However, the vibrator bearings 110A and 110B are arranged in the axial direction O, except for the inner ring 112, with respect to the external teeth 124A engaged with the internal gears for deceleration 130A and the internal gears 130B for output And the external teeth 124B which are engaged with the external teeth 124B. The skew of the roller 116B caused by the meshing of the deceleration internal gear 130A with the external teeth 124A and the skew of the roller 116A caused by the engagement of the internal gears 130B for output and the external teeth 124B Each of the skew is prevented.

Since the rollers 116A and 116B are cylindrical, ball bearings having the same size of balls have a large load bearing capacity and many parts that contact the inner ring 112 and the outer rings 118A and 118B, The capacity can be increased.

The external tooth 124 has a portion (external tooth 124B) where the portion to be engaged with the internal gear 124A for speed reduction and the internal gear for output 130B are engaged with each other in the axial direction O, . Thereby, when the external gear 124A is deformed, for example, when the internal gear 120A and the internal gear 124A for deceleration are engaged with each other, the external tooth 124A is not deformed by the deformation. Similarly, when external gear 124A is deformed, external gear 124B is not deformed by deformation when external gear 120B and output internal gear 130B are engaged with each other. That is, by dividing the external teeth 124, it is possible to prevent a reduction in the transmitted torque that deforms the other external teeth 124B and 124A due to the deformation of one of the external teeth 124A and 124B, thereby deteriorating the meshing relation.

The engagement position of the external gear 120A with the internal gear for acceleration 130A is rotationally moved in accordance with the movement of the vibrator 104 in the longitudinal direction x. Here, when the vibrator 104 makes one revolution, the rotation phase is delayed by the difference in dimension between the external gear 120A and the internal gear for acceleration 130A. That is, the reduction ratio by the deceleration internal gear 130A is calculated as follows: ((Dimension N * k) of the external gear 120A - Dimension ((N + 1) * k) of the internal gear 130A for deceleration / (N * k)) = - 1 / N of the dimension 120A.

Since both the external gear 120B and the internal gear 130B for output are the same in size (N * k), the external gear 120B and the internal gear for output 130B are not in contact with each other, Respectively. As a result, rotation equivalent to rotation of the external gear 120B is output from the internal gear 130B for output. As a result, the decelerated output can be extracted from the internal gear 130B for output based on the reduction gear ratio 1 / N by the internal gear for acceleration 130A for rotation of the vibrator 104. [

The present embodiment has a configuration in which cylindrical external gear 120 is engaged with internal gear 130 (two internal gears 130 for deceleration, internal gear 130B for output and 130B for output) The internal gear 130 and the external gear 120 are provided so that the number of simultaneous engagement Nph and Npl between the external gear 120 and the internal gear 130 is two or more, In addition, theoretical coupling is realized by using the trochoid curve. As a result, the impact resistance is improved and the surface pressure applied to the tooth surface of the engagement is dispersed, so that a large torque can be transmitted. Thus, the local stress generated in the external gear 120 can be transmitted to, Can be significantly reduced. That is, in the bending gear type gear device according to the present embodiment, the conical deformation does not occur due to the warping of the vibrator, and the increase in the engagement area and the dispersion of the surface pressure can be achieved without stress concentration at the bottom of the cup It is possible to greatly increase the load capacity.

In this embodiment, as shown in Figs. 14, 15, and 19, the pin type reduction pitch coefficient Gph <1 and the pin type output pitch coefficient Gpl> 28) is established. That is, as shown in the equation (29), in this embodiment, the distance n _h * from the eccentric shaft B to the pitch point P _h by the external gear 120A and the deceleration internal gear 130A, between the L) and the pitch point (P _l) the distance (n _l * L) of up by the gear (120B) and the output internal gear (130B) shouting from the eccentric axis (B), outer tooth from the eccentric axis (B) gear The position of the center of the pin (the tooth-shaped substance) of the pin 120 is disposed.

[Number 10]

The load applied to the external teeth 124A of the external teeth gear 120A and the external teeth 124B of the external teeth gear 120B when engaged with the internal gear for output 130B when engaged with the deceleration internal gear 130A The loads applied to the gears 120 are provided with components opposite to each other and the two load regions applied to the gear 120 can be brought close to each other in the circumferential direction of the external gear 120. That is, when viewed from the axial direction O, in the engaging operation, the two internal gears 130 may be configured to sandwich only a small number of external teeth 124. This can prevent a phenomenon in which the engagement of the external gear 120 and the internal gear 130 with an excessive torque (latching phenomenon) occurs. That is, it is possible to improve the latching resistance.

(A comparative example in which the radius of the tooth-shaped substance of the internal gear is about 26 mm and the reduction ratio is 1/50 (comparative example)) using a cup-shaped external gear which is actually manufactured, and a reduction gear ratio (Approximately four times or more) with respect to the measured value of the comparative example with respect to the latching performance in the flexural-meshing gear device 100 according to the present embodiment. At the same time, in the comparative example, the rated torque was 3.3 kgfm. In the bending gear type gear unit 100 of the present embodiment, however, it was confirmed by theoretical calculation and test that the rated torque became 6.6 kgfm. That is, it was confirmed by the theoretical calculation that the rated torque was doubled, which was confirmed by the test.

In this manner, in the present embodiment, it is possible to increase the transmission torque and increase the transmission efficiency. However, instead of improving the transmission torque, the bending gear type gear device 100 may be made more compact.

In this embodiment, the gear teeth of the external gear 120 are the same as those of the internal gear for speed reduction 130A and the internal gear for output 130B, so that the external gear 120 can be easily machined The machining cost can be suppressed to a low level, and it is possible to perform the shape processing with high accuracy.

That is, according to the present invention, by increasing the number of simultaneous meshing (Nph, Npl) between the external gear 120 and the internal gear 130, it is possible to increase the transmission torque and transmission efficiency.

&Lt; Embodiment 2 &gt;

An example of the second embodiment of the present invention will be described in detail with reference to Figs. 20 to 29. Fig. In this embodiment, instead of the cylindrical pin according to the first embodiment, a tooth formed by a trochoid curve is used as a gear for external gear, and the external teeth of the external gear are integrally formed with a base member (referred to as a solid type). However, if the parameters used in the first embodiment are the same as the definitions, the parameters used in the present embodiment are also matched.

A configuration different from that of the first embodiment and a method of determining a tooth profile will be described, and the remaining parts are denoted by the same reference numerals as the second two digits, and redundant description is omitted.

<Configuration>

As shown in Figs. 20 and 21, the external gear 220A is in contact with and engages with the internal gear 230A for deceleration. The external gear 220A is composed of a base member 222 and a foreign tooth 224A. The base member 222 is a tubular member having flexibility and is disposed on the outside of the vibrator bearing 210A and integrally molded with the external teeth 224A. As a result, the external teeth 224A can be made smaller and high-precision machining can be performed. That is, the external gear 220A of the present embodiment is suitable for a small bending gear capable of small load capacity. The external tooth 224A is formed based on the trochoid curve.

As shown in Figs. 20 and 21, the external gear 220B is in contact with and engages with the internal gear 230B for output. Like the external gear 220A, the external gear 220B is composed of a base member 222 and a foreign tooth 224B. The external teeth 224B are the same number as the external teeth 224A and are formed into the same shape. Here, as shown in Fig. 20, the external teeth 224A and 224B are divided in the axial direction, but the base member 222 is common. That is, the external gears 220A and 220B have the same tooth shape. The eccentric amount L of the vibrating body 204 is transmitted to the external teeth 224A and 224B in the same phase. Hereinafter, the external teeth 224A and 224B are referred to collectively as the external teeth 224.

<Determination method of tooth profile>

A method of determining the tooth profile of the external gear 220, the internal gear 230A for deceleration, and the internal gear 230B for output will be described.

First, a method of obtaining a tooth profile will be described below.

First, the trajectory of the position of the tooth-shaped substance of the internal gear in the case where the internal tooth of the internal gear is assumed to be a cylindrical pin and the radius of the pin (r2 = 0) is represented by a trochoidal curve, 220). Next, the locus of the position of the tooth-like substance of the external gear is obtained, and the tooth profile of the internal gear is defined from the locus thereof. Next, a plurality of parameters defining the tooth profile of the external gear 220 and the internal gear 230 are associated with each other from the size and the dimension of the external gear 220 and internal gear 230. Next, the tooth line of the teeth of the internal gear 230 and the correction range of the teeth are determined. Next, using the associated parameters, a tooth portion outside the correction range is obtained, and the number of simultaneous meshing in the tooth portion is obtained. Then, the optimal parameters are determined so that the number of simultaneous meshing is all 2 or more. In the determination of the parameters, trial and error are performed so as to satisfy simultaneously the target values such as torque, allowable surface pressure of the tooth surface, principal stress at each point, and bearing life.

This will be described in detail below.

First, the tooth profile of the external gear 220 is defined.

Although the internal gear 228A of the deceleration internal gear 230A is provided with a cylindrical pin having a virtually radius r2 as the internal gear for deceleration 230A for the sake of convenience, ), The locus of the position of the tooth-like substance of the internal gear 230A for deceleration of the radius of the pin = 2 (= the same as the center of the pin) is obtained. Then, the teeth of the external gear 220 are shifted to the inner side (the external gear 220 side) by the radius rho 2 of the pin. Hereinafter, this will be described in more detail. However, the virtual reduction ratio n (n _h , n _l ) has the same definition as that of the first embodiment.

Similar to the first embodiment, the external gear 220 has a two-arcuate shape, and the relationship between the radii R1 and R2 is expressed by equations (3) and (4).

Shout gear 220 performs theoretical meshing with deceleration internal gear 230A having a virtually pin. 22, when the center of the pin of the internal gear 230A for deceleration moves from the coordinates (x4, y4) to the coordinates (x5, y5) in the stop space centered on the eccentric shaft (B) The coordinate (x _p , y _p ) of the trajectory to be drawn appears as an outer trocode curved (epitrochoid curved) as the coordinate of the position of the tooth-like substance of the external gear 220. That is, the radius b2 of the base circle BB fixed around the eccentric axis B, the radius a2 of the rotation circle AB that does not slip along the circumference of the base circle BB, The coordinates (x _p , y _p ) of the position of the teeth-like substance of the external gear 220 are expressed by the equations (30) and (31) using the radius L2 and the rotation angle?

[Number 11]

Here, by using the relationship of the expressions (32) to (34), the relationship of the expressions (35) and (36) is obtained.

[Number 12]

Next, the coordinate (x _p , y _p ) of the position of the tooth-shaped entity of the external gear 220 is moved to the inner side (the external gear 220 side) by the radius rho 2 of the pin assumed as the internal tooth 228 . The coordinates (x _kfc , y _kfc ) of the tooth profile of the external gear 220 having the rotation axis Fc as the origin can be expressed by the formulas (37) to (39).

[Num. 13]

That is, the coordinates (x _kfc , y _kfc ) of the tooth profile of the external gear 220 can be obtained by substituting the radius R, ρ2, the eccentric amount L and the virtual reduction ratio n _h to change the angle β have.

Next, the tooth profile of the internal gear 230 is defined. External tooth is moved to the inside (internal gear 230 side) obtain the envelope of the coordinates (x _p, y _p) of the position of the entity of the tooth, the envelope as long as the radius (ρ2) of the gear 220, internal gear 230, As shown in FIG. That is, with respect to the internal gear 230A for deceleration, the tooth profile of the internal gear 230A is obtained again. Hereinafter, this will be described in more detail.

The locus (Q, two broken line portions shown in Fig. 23) of the tooth profile of the external gear 220 on the xd-yd coordinate centered on the eccentric shaft B of the external gear 220 is set to 23 (A solid line portion shown in Fig. 23) as shown in Fig. Therefore, the coordinates (x _pfc , y _pfc ) of the position of the tooth-like substance of the internal gear 630 having the rotation axis Fc as its origin can be obtained by the formulas (40) and 41). Here, the relationship between the angles? And? Is expressed by the equation (43) by using the equation (42), which is a conditional expression of the envelope.

[Number 14]

Next, the coordinates (x _pfc , y _pfc ) of the position of the tooth- _shaped entity of the internal gear 230 are shifted to the inside (the internal gear 230 side) by the radius rho 2 of the pin assumed as the internal tooth 228 The coordinates (x _fc , y _fc ) of the teeth of the internal gear 630 having the rotation axis Fc as the origin can be obtained by the equations (44) and (45).

[Number 15]

That is, by changing the angle? By substituting the radius R,? 2, the amount of eccentricity L and the virtual reduction ratio n _h , n ₁ , the tooth profile of the internal gear for deceleration 230A and the internal gear for output 230B (X _fc , y _fc ) can be obtained.

Next, the parameters defining the internal gear 230 and external gear 220 are related to each other.

As described above, like the first embodiment, the shape of the external gear 220 is a two-arc shape defined by radii R1 and R2. That is, also in this embodiment, the relationship of the formulas (19) and (20) is established.

Next, a parameter (Gs, referred to as a solid type pitch coefficient) is introduced. Here, a straight line passing through the eccentric shaft B and the rotary shaft Fc, and a common point of contact points generated by engagement of (the external teeth 224 of the external teeth) 220 and the internal teeth 230 (the internal teeth 228 of) The point of intersection with the normal is referred to as a pitch point by the external gear 220 and the internal gear 230 (i.e., the definition of the pitch point is the same as in the first embodiment). The solid type pitch coefficient Gs can be easily grasped with respect to the relative positional relationship between the positions of the teeth of the external teeth gear 220 and the internal teeth of the internal gear 230 and the pitch point, similarly to the pin type pitch coefficient Gp And it is also possible to facilitate the adjustment between the parameters. More specifically, as shown by the equation (46), the solid type pitch coefficient Gs is a ratio of the radius R to the pitch point by the external gear 220 and the internal gear 230 for deceleration from the rotation axis Fc Distance ((n + 1) * L).

[Num. 16]

24 shows the relationship between the radius R of the teeth of the internal gear 230 and the virtual reduction ratio n _h . The solid type pitch coefficient (Gsh, referred to as a solid type deceleration side pitch coefficient) obtained at this time is defined in the equation (47) based on the equation (46). In Expression (19) and Expression (20), Expression (48) is obtained by summarizing Expression (47) with parameter (i) = 1.

[Number 17]

25 shows the relationship between the radius R of the teeth of the internal gear 230 and the virtual reduction ratio n ₁ . The solid type pitch coefficient (Gsl, referred to as solid type output side pitch coefficient) obtained at this time is defined in the equation (49) based on the equation (46). (50) and (51) can be obtained by summarizing the equation (49) with the parameter (i) = 0 in the equations (19) and (20).

[Number 18]

Therefore, when the radius R, the reduction ratio N, the solid type reduction pitch coefficient Gsh, and the engagement angle? Are given, the virtual reduction ratio n _h and the eccentricity L are determined, pitch coefficient (Gsl), can be determined a virtual gear ratio (n _l).

In this embodiment, similarly to the first embodiment, as shown in Figs. 24 and 25, the value of the solid-type output pitch coefficient Gsl &gt; 1 is obtained by substituting the solid-type deceleration pitch coefficient Gsh <1 have. The present embodiment is also similar to the first embodiment in that the case where the engagement angle θ is 40 to 65 degrees and the value of cos ^-1 of the pin type deceleration side pitch coefficient Gph is 15 to 30 degrees, Is a more preferable condition.

Next, the correction range of the tooth profile of the internal gear 230 is determined.

The teeth line and the teeth of the internal teeth 228 are corrected similarly to the first embodiment. Therefore, the angle (? S to? F, the area of the tooth shape that has not been corrected) at which the tooth profile is not corrected becomes the effective range at which theoretical engagement is performed.

Next, the number of simultaneous meshes (Nsh, Nsl) is obtained.

The simultaneous engagement numbers Nsh and Nsl can be obtained by dividing the effective range determined by the rotation angle? Of the external gear 220 by the pitch angle as in the first embodiment. That is, the simultaneous engagement number Nsh of the internal gear 230A for deceleration in the equation (52) and the simultaneous engagement number Nsl of the internal gear 230B for output in the equation (53) Respectively.

[Number 19]

(52) and (53), the number of simultaneous meshing is obtained. The simultaneous engagement number Nsh of the internal gear 230A for deceleration determined when k = 2 is shown in Fig. 26, and the number Nsl of simultaneous engagement of the internal gear 230B for output is shown in Fig.

The teeth of the internal gear 230 in the present embodiment are determined according to the conditions of the diameter (2 * R) and the reduction gear ratio (1 / N) for realizing two or more of these simultaneous engagement numbers Nsh and Nsl do. That is, when the dimension difference is 2 (k = 2), the reduction ratio (1 / N) does not become the tooth shape of the present embodiment at 1/30, but is reduced to 1/50 or less The tooth profile of the internal gear of the embodiment is determined.

In this embodiment, since the external teeth 224 are integrally formed with the base member 222, the external teeth gear 220 can be easily machined and the machining can be performed with high accuracy.

In other respects, substantially the same operational effect as that of the first embodiment can be obtained also in this embodiment.

For example, in the present embodiment, as shown in Figs. 24, 25 and 28, the solid-type deceleration side pitch coefficient Gsh <1 and the solid-type output side pitch coefficient Gsl> , Equation (54) is established. That is, the distance (n _h +1) * L from the rotation axis Fc to the pitch point P _h by the external gear 220A and the internal gear 230A for deceleration, as shown in equation (55) And the distance (n _l +1) * L from the rotation axis Fc to the pitch point P ₁ by the external gear 220B and the internal gear 230B for output, The position of the center of the pin (the tooth-like substance) when the pin 228 is imagined as a pin is disposed.

[Number 20]

The load Fd applied to the external teeth 224A of the external gear 220A and the external teeth 224A of the external gear 220B when the internal gear 230B is engaged with the output internal gear 230B The load Fo applied to the external gear 220 and the load Fo applied to the external gear 220 are set so that the areas of the two loads Fd and Fo applied to the external gear 220 in the circumferential direction of the external gear 220 Can be brought close. That is, as shown in Fig. 29, when the engagement operation is performed, the area of the load Fd and the load Fo are brought close to each other, and the two internal gears 230 are moved in a small number of external teeth 224 may be sandwiched. As a result, similar to the first embodiment, latching resistance can be improved.

In addition, both Expressions (29) and (55) can be transformed into Expression (56).

[Num. 21]

That is, in the above-described embodiment, the straight line passing through the rotation axis Fc and the eccentric shaft B and the intersection with each common normal line of the contact point generated by the engagement of the external gears 120 and 220 and the internal gears 130 and 230, the pitch point between the (P _h, P _l), outer tooth when the external tooth 124 of the gear 120 to the pin cylindrical pin center, or the internal teeth of the internal tooth 228 of the gear 230, a cylindrical pin on the (Virtual), since the center Oc of the pin is disposed, it is possible to improve the latching resistance.

While the present invention has been described with reference to the above embodiments, the present invention is not limited to the above embodiments. Needless to say, it is needless to say that improvements and design changes can be made without departing from the gist of the present invention.

For example, in the above embodiment, when the number of simultaneous engagement (Nph, Npl, Nsh, Nsl) is set to two or more, the tooth profile of the external gear or internal gear is obtained based on the trochoid curve. . For example, the contact line can be uniquely obtained from the coordinates of the tooth profile of the internal gear which is obtained, which is the trajectory of the contact point caused by engagement between the external gear and the internal gear. Hereinafter, a specific relationship between the coordinate of the teeth of the internal gear 130 and the contact line in the case of the first embodiment will be described concretely.

The contact line CL is a locus viewed from the XY coordinate system shown in Fig. 30 obtained by rotating the coordinate (x _fc , y _fc ) of the teeth of the internal gear 130 by an angle?. Therefore, the coordinates (x _cfc , y _cfc ) of the contact line are given by equations (57) and (58) obtained by rotating the coordinate (x _fc , y _fc ) of the teeth of the internal gear 130 by angle?

[Number 22]

The contact line CL obtained by the above formula is shown in Fig. The contact line CL is drawn between the teeth of the external gear 120 and the plurality of tooth lines of the internal gear 130 and the teeth and it is possible to secure a plurality of simultaneous engagement numbers Nph and Npl.

As a result, it is possible to obtain a contact line capable of securing a plurality of simultaneous engagement numbers (Nph, Npl) by using this, and the tooth profile of the internal gear from the contact line.

In the above embodiment, the deceleration side pitch coefficients Gph and Gsh are smaller than 1 and the output side pitch coefficients (Gpl and Gsl) are larger than 1. However, the present invention is not necessarily limited to such a relationship. For example, the deceleration-side pitch coefficients Gph and Gsh may be set to be larger than 1 and the output-side pitch coefficients Gpl and Gsl may be set to be smaller than 1. It should be noted that the case where any pitch coefficient is larger than 1 or any pitch coefficient is smaller than 1 is not denied. This is because the tooth profile of the external gear and the internal gear is obtained by determining not only the parameters defining the pitch coefficient but also a plurality of parameters by trial and error.

(Industrial availability)

The bending gear of the present invention can be used in various applications, but can be suitably used for precise control applications, such as a joint (wrist) drive device of an industrial robot or a machine tool.

100, 200 bending gear unit
104, 204 vibrator
110A, 110B, 210A, 210B vibrator bearings
114A, 114B, 214A, 214B.
116A, 116B, 216A, 216B rollers
120, 120A, 120B, 220, 220A, 220B,
122,
124, 124A, 124B, 224, 224A, 224B,
128, 128A, 128B, 228, 228A, 228B,
130, 130A, 130B, 230, 230A, 230B,
a1, a2 Radius of rotation circle
AA, AB Rotation circle
B eccentric shaft
b1, b2 Radius of base circle
BA, BB Foundation Circle
CL contact line
FA 1st arc part (engagement range)
Fc rotating shaft
Fd, Fo load
Gp, Gph, Gpl, Gs, Gsh, Gsl Pitch coefficients
L eccentricity
n, n _h , n _{l The} virtual reduction ratio (reciprocal of)
N Reduction ratio (reciprocal of)
Nph, Npl, Nsh, Nsl Number of simultaneous meshing
O axis direction
Center of Oc pin
P _h , P _l Pitch point
R The radius of the teeth of the internal gear
R1 Radius of the teeth of the teeth of the gear teeth of the external gear
R2 Radius of the teeth of non-meshing range of gear teeth
SA 2nd arc part (non-meshing range)
ρ1, ρ2 Radius of cylindrical pin

Claims (14)

A first internal gear which is disposed on an outer periphery of the vibrator and has a flexible tubular external gear which is flexibly deformed by rotation of the vibrator and a stiffness which is in contact with the external gear, And a second internal gear disposed in the axial direction of the first internal gear and having rigidity in contact with the external gear,
The teeth of the external gears at the portions engaging with the first internal gear and the second internal gear are the same,
The external tooth gear, the first internal gear, and the second internal gear form teeth having a number of simultaneous engagement of the external tooth gear and the first internal gear and a simultaneous engagement of the external tooth gear and the second internal gear at two or more Respectively,
The external teeth of the external gear are cylindrical pins,
The tooth profile of the external gear, the first internal gear and the second internal gear is formed from an eccentric shaft which is the center of the engagement radius of the external gear at the time of engagement with the first internal gear or the second internal gear, And a pitch line which is an intersection of a straight line passing through the eccentric shaft from the eccentric shaft with a rotation axis of the vibrator and the eccentric shaft and a common line of a contact point generated by meshing of the external gear and the first internal gear or the second internal gear. And a ratio of a distance between the first gear and the second gear is given. The method according to claim 1,
Wherein the tooth profile of the external gear, the first internal gear, or the second internal gear is a shape based on a Trochoid curve. delete delete The method according to claim 1,
Wherein a reduction ratio obtained from a ratio of a dimension of the first internal gear to a dimension of the external gear and a ratio of a dimension of the second internal gear to a dimension of the external gear is 1/30 or less. Gear device. A first internal gear which is disposed on an outer periphery of the vibrator and has a flexible tubular external gear which is flexibly deformed by rotation of the vibrator and a stiffness which is in contact with the external gear, And a second internal gear disposed in the axial direction of the first internal gear and having rigidity in contact with the external gear,
The teeth of the external gears at the portions engaging with the first internal gear and the second internal gear are the same,
The external tooth gear, the first internal gear, and the second internal gear form teeth having a number of simultaneous engagement of the external tooth gear and the first internal gear and a simultaneous engagement of the external tooth gear and the second internal gear at two or more Respectively,
The tooth profile of the external tooth gear, the first internal gear, and the second internal gear extends from a rotational axis of the vibrator to a center position of the pin when the internal teeth of the first internal gear or the second internal gear is imagined as a cylindrical pin And a pitch point which is an intersection of a straight line passing through the rotary shaft and the eccentric shaft of the external gear from the rotary shaft and a common normal line of the contact point generated by meshing of the external gear and the first internal gear or the second internal gear. And a ratio of a distance between the first gear and the second gear is given. 3. The method according to claim 1 or 2,
The tooth profile of the external gear, the first internal gear and the second internal gear is set such that the ratio of the dimension of the first internal gear to the dimension of the external gear and the ratio of the dimension of the second internal gear to the dimension of the external gear Is determined on the basis of a reduction gear ratio determined by the gear ratio calculating means. A first internal gear which is disposed on an outer periphery of the vibrator and has a flexible tubular external gear which is flexibly deformed by rotation of the vibrator and a stiffness which is in contact with the external gear, And a second internal gear disposed in the axial direction of the first internal gear and having rigidity in contact with the external gear,
A first internal gear and a second internal gear; a straight line passing through an eccentric shaft which is the center of the engagement radius of the external gear when the rotary shaft of the vibrator is engaged with the first internal gear or the second internal gear; When the outer tooth of the external gear is formed into a cylindrical pin or between a center of the pin and a center of the pin when the external tooth of the external tooth gear is imaginary with a cylindrical pin, between the pitch point, which is an intersection point with each common normal line of the contact point, Wherein the center of the pin is disposed when the internal teeth of the internal gear or the second internal gear is formed into a cylindrical fin or when the internal tooth of the second internal gear is imaginary with a cylindrical pin. A first internal gear which is disposed on an outer periphery of the vibrator and has a flexible tubular external gear which is flexibly deformed by rotation of the vibrator and a stiffness which is in contact with the external gear, And a second internal gear disposed in the axial direction of the first internal gear and having rigidity in contact with the external gear, the method comprising the steps of:
The teeth of the external gears at the portions engaging with the first internal gear and the second internal gear are made the same,
A step of defining teeth of the external gear, the first internal gear, and the second internal gear;
A step of associating a plurality of parameters defining the tooth profile of the external gear, the first internal gear, and the second internal gear from the size and the dimension of each gear;
A step of determining a tooth line of each tooth of the first internal gear and the second internal gear and a correction range of the tooth,
A step of obtaining a tooth portion outside the correction range of each of the first internal gear and the second internal gear by using the plurality of parameters to obtain respective simultaneous engagement numbers,
And a step of determining the tooth profile of the external gear, the first internal gear, and the second internal gear by determining the plurality of parameters under the condition that the number of simultaneous engagement becomes two or more,
The tooth profile of the external tooth gear or the teeth of the first internal gear and the second internal tooth gear is defined as a trochoid curve defined by a base circle fixed around the rotation axis of the vibration body and a rotation source rotating without slipping along the circumference of the base circle Lt; / RTI &gt;
The external tooth of the external gear is made of a cylindrical pin and the trochoid curve is an internal trochoid curve and the trochoid curve is moved parallel to the radius of the pin to define a tooth profile of the first internal gear and the second internal gear ,
A distance from an eccentric axis which is a center of an engagement radius of the external gear to a center position of the pin when the internal gear is engaged with the first internal gear or the second internal gear, And a distance to a pitch point, which is an intersection of a straight line passing through the rotary shaft of the vibrator and the eccentric shaft, and a common normal line of a contact point occurring when the external gear and the first internal gear or the second internal gear meshes with each other Wherein said step (c) comprises the steps of: delete delete delete 10. The method of claim 9,
The inner tooth of the first internal gear or the second internal gear is assumed to be a cylindrical pin and the trochoid curve is defined as an outer trochoid curve and the trochoid curve is moved parallel to the radius of the pin to define a tooth profile of the external gear In addition,
Wherein an envelope of the trochoid curve is obtained and the envelope is moved parallel to the radius of the pin to define a tooth profile of the first internal gear and the second internal gear. A first internal gear which is disposed on an outer periphery of the vibrator and has a flexible tubular external gear which is flexibly deformed by rotation of the vibrator and a stiffness which is in contact with the external gear, And a second internal gear disposed in the axial direction of the first internal gear and having rigidity in contact with the external gear, the method comprising the steps of:
The teeth of the external gears at the portions engaging with the first internal gear and the second internal gear are made the same,
A step of defining teeth of the external gear, the first internal gear, and the second internal gear;
A step of associating a plurality of parameters defining the tooth profile of the external gear, the first internal gear, and the second internal gear from the size and the dimension of each gear;
A step of determining a tooth line of each tooth of the first internal gear and the second internal gear and a correction range of the tooth,
A step of obtaining a tooth portion outside the correction range of each of the first internal gear and the second internal gear by using the plurality of parameters to obtain respective simultaneous engagement numbers,
And a step of determining the tooth profile of the external gear, the first internal gear, and the second internal gear by determining the plurality of parameters under the condition that the number of simultaneous engagement becomes two or more,
The tooth profile of the external tooth gear or the teeth of the first internal gear and the second internal tooth gear is defined as a trochoid curve defined by a base circle fixed around the rotation axis of the vibration body and a rotation source rotating without slipping along the circumference of the base circle Lt; / RTI &gt;
The external tooth of the external gear is formed of a cylindrical pin and the trochoid curve is used as an internal trochoid curve to define a tooth profile of the first internal gear and the second internal gear by parallel movement of the trochoid curve by a radius of the pin ,
A distance between the rotational axis of the vibrator and the center position of the pin and a distance between the rotational axis of the vibrating member and the center of the pinion gear when the rotational shaft and the first internal gear or the second internal gear are meshed with each other, And a distance to a pitch point which is an intersection of a straight line passing through the eccentric shaft which is the center of the engaging radius and a common normal line of the contact point generated by the engagement of the external gear and the first internal gear, is taken into consideration. A method of determining the tooth profile of an.

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