JP2011007339A - Eccentric oscillating-type planetary gear device - Google Patents

Abstract

PROBLEM TO BE SOLVED: To extend the life of a tooth surface of an external tooth 19 by suppressing the elastic deformation of a bridge section 30 in an externally toothed gear 18 and the external tooth 19, and to improve vibration characteristics.SOLUTION: Since a ratio obtained by dividing the diameter (D) of a pin constructing an internal tooth 14 by the pitch P of the internal tooth 14 is made smaller to an extent where the tooth top 19a of the external tooth 19 exceeds radially outside the inner periphery 15a of an internally toothed gear 15, the diameter D of the internal tooth (the pin) 14 becomes smaller than before. Therefore, the tooth bottom 19b of the external tooth 19 of the externally toothed gear 18 is moved radially outward. As a result, the thickness (the minimum thickness) J of the bridge section 30 becomes thicker than before, and the bending rigidity is enhanced. Thus, the elastic deformation of the bridge section 30 and the external tooth 19 due to the reaction force of a driving component force is suppressed.

Description

    The present invention relates to an eccentric oscillating planetary gear device in which an external gear meshing with an internal gear is eccentrically oscillated by a crankshaft.

    As a conventional eccentric oscillating planetary gear device, for example, one described in Patent Document 1 below is known.

JP-A-7-299791

  This has an internal gear composed of a large number of cylindrical pins on the inner circumference and a plurality of crankshaft holes and through holes formed on the outer circumference, and has a trochoidal tooth profile on the outer circumference. An external gear that meshes with the teeth and has external teeth that have only one fewer teeth than the internal teeth, a crankshaft that is inserted into each crankshaft hole and rotates to eccentrically swing the external gear, The crankshaft is rotatably supported, and includes a support body having a plurality of pillar portions loosely fitted in the respective through holes.

  In this case, as shown in FIG. 9, the external teeth 02 of the external gear 01 that are in contact with each other to the internal teeth (pins) 04 of the internal gear 03 on the tooth surface at the contact point. A driving force in the vertical direction is applied, and a reaction force K of the driving force is applied from the inner teeth (pins) 04 to the outer teeth 02 as a reaction.

    However, in the planetary gear device described above, of the external gear 01, the bridge portion 06 located on the radially outer side of the through hole 05 has a thickness (from the radial outer end of the through hole 05 to the external tooth 02. (The radial distance to the root 07 is the minimum wall thickness.) Since J is considerably smaller than the wall thickness in other parts and the bending rigidity is low, the reaction force K as described above is almost radial with respect to the bridge portion 06. , The external teeth 02 in the vicinity of the bridge portion 06 and the bridge portion 06 are elastically deformed so that the external teeth 02 and the internal teeth (pins) 04 come into contact with each other, and the tooth surface life of the external teeth 02 is shortened. There was a problem of ending up.

  In addition, if the bending rigidity of the bridge portion 06 is low as described above, the natural frequency becomes low when a torque load exists when the planetary gear device is applied to a robot, a machine tool, or the like. There was also a problem that the vibration characteristics deteriorated and the controllability deteriorated.

  The present invention provides an eccentric oscillating planetary gear device capable of extending the tooth surface life of the external teeth by suppressing elastic deformation of the bridge portion of the external gear and the external teeth and improving the vibration characteristics. With the goal.

    The purpose of this is to form an internal gear in which inner teeth composed of a large number of cylindrical pins are provided at a constant pitch P on the inner periphery, at least one crankshaft hole and a plurality of through holes. The external gear has a trochoidal tooth shape and meshes with the internal teeth, and has only one external tooth with a smaller number of teeth than the internal teeth, and is inserted into each crankshaft hole and rotated to rotate the external gear eccentrically. In an eccentric oscillating planetary gear device comprising a crankshaft to be moved and a support body having a plurality of column portions inserted into each through hole while rotatably supporting the crankshaft, internal teeth are configured. The ratio obtained by dividing the diameter D of the pin by the constant pitch P of the internal teeth is reduced until the tooth tips of the external teeth exceed the inner circumference of the internal gear radially outward, and the internal gear between adjacent internal teeth The inner circumference is more than the amount that the outer teeth exceed the inner circumference. Only excised can be achieved by an eccentric oscillating-type planetary gear device which is adapted to avoid interference with the inner periphery of the outer teeth and the internal gear is.

  In the above invention, the ratio of the diameter D of the pins constituting the internal teeth divided by the constant pitch P of the internal teeth is reduced until the tooth tips of the external teeth exceed the inner periphery of the internal gear radially outward. The diameter D of the internal teeth (pins) is smaller than that of the conventional one, which causes the roots of the external teeth of the external gear to move outward in the radial direction. As a result, the bridge portion positioned radially outward of the through hole The wall thickness (minimum wall thickness) becomes thicker than before, and the bending rigidity increases. As a result, the elastic deformation of the bridge portion and the external teeth when the reaction force of the driving component force acts in the radial direction is suppressed, the tooth surface life of the external teeth can be extended, and the natural frequency is increased. In addition, vibration characteristics and controllability can be improved.

  Here, when configured as described above, the external teeth interfere with the inner periphery of the internal gear, but in the present invention, the external teeth exceed the inner periphery of the internal gear between adjacent internal teeth. By cutting away by a depth more than that amount, interference between the external teeth and the internal circumference of the internal gear is avoided.

Moreover, if comprised as described in Claim 2, the Hertzian stress in the contact point of an internal tooth and an external tooth can be maintained at a low value, and the tooth surface life of an external tooth can further be extended.
Furthermore, if comprised as claimed in claim 3, it is possible to increase the output torque while preventing the occurrence of a pointed portion on the tooth surface.

It is side surface sectional drawing which shows an example of an eccentric rocking | swiveling planetary gear apparatus. It is II sectional view taken on the line of FIG. It is explanatory drawing which shows the reaction force K given to an external tooth, and its action line S. FIG. It is an enlarged view of the U section of FIG. It is side surface sectional drawing similar to FIG. 1 which shows Embodiment 1 of this invention. It is sectional drawing similar to FIG. 2 which shows Embodiment 1 of this invention. It is a graph which shows the relationship between the diameter D of an internal tooth (pin), and Hertzian stress ratio. It is a graph which shows the relationship between the value of L / R, and a load ratio. It is sectional drawing similar to FIG. 2 which shows an example of background art.

Hereinafter, an example of an eccentric oscillating planetary gear device capable of avoiding the above-described interference will be described with reference to the drawings.
1 and 2, reference numeral 11 denotes an eccentric oscillating planetary gear unit used for a robot or the like. The planetary gear unit 11 has a substantially cylindrical shape attached to, for example, a robot arm or hand not shown. A rotating case 12 is provided. A large number of pin grooves 13 having a semicircular cross section are formed in the central portion in the axial direction at the inner periphery of the rotating case 12, and these pin grooves 13 extend in the axial direction and are equidistant from each other in the circumferential direction. Here, they are spaced apart by a fixed pitch P. 14 is an internal tooth composed of a number of cylindrical pins (the same number as the pin groove 13), and almost half of these internal teeth (pins) 14 are inserted into the pin groove 13 so that the rotating case 12 The inner circumference is provided at equal distances (by a constant pitch P) in the circumferential direction.

  Here, the above-described constant pitch P is a value obtained by dividing the circumferential length of the pin circle V passing through the centers of the pins constituting all the internal teeth 14 by the number of internal teeth (pins) 14. For example, it is the arc length when the centers of any two adjacent internal teeth (pins) 14 are connected by an arc segment. The rotary case 12 and the internal teeth (pins) 14 described above constitute an internal gear 15 in which internal teeth 14 formed of a plurality of cylindrical pins are provided on an inner periphery 15a. As a result, the inner periphery 15a of the internal gear 15 (fixed case 12) is located on the pin circle V or in the vicinity of the pin circle V that can hold at least the internal teeth (pins) 14. Yes.

  Here, about 25 to 100 inner teeth (pins) 14 are arranged, but preferably in the range of 30 to 80. The reason is that the number of the internal teeth (pins) 14 is within the above-mentioned range, and the reduction ratio of the external gears 40 and 42 described later is 1 before the meshing of the external gear 18 and the internal gear 15 described later. If a reduction gear ratio of 1/1 to 1/7 is provided and the reduction ratio of the front and rear stages is combined, a high reduction ratio can be easily obtained, and a planetary with a high reduction ratio and a high natural frequency can be obtained. This is because a gear device can be configured.

  A plurality of (in this case, two) external gears 18 in the form of a ring are accommodated in the axial direction in the internal gear 15, and a trochoidal tooth profile, more specifically a peritrochoid, is provided on the outer periphery of these external gears 18. A large number of external teeth 19 each having a tooth shape are formed. The number of teeth Z of the external teeth 19 of the external gear 18 is one less than the number of teeth of the internal teeth (pins) 14 (the difference in the number of teeth is 1). In this way, the difference in the number of teeth between the internal teeth (pins) 14 and the external teeth 19 is set to 1, compared to the case where the difference in the number of teeth is a value G of 2 or more, a high reduction ratio can be obtained. This is because the processing cost can be reduced.

  Here, the external gear whose tooth number difference is a value G of 2 or more means that the outer contour of the trochoid external gear is shifted in the circumferential direction by a distance obtained by dividing the pitch between the external teeth 19 by the value of G, This is an external gear obtained by taking out a portion where G outlines deviated in the circumferential direction overlap each other as a tooth profile (see Japanese Patent Laid-Open No. 3-181641). The external teeth 19 and the internal teeth (pins) 14 are meshed with each other while the external gear 18 and the internal gear 15 are inscribed, but the maximum meshing portion (the most meshed portion of the two external gears 18). The deep part) is 180 degrees out of phase.

  Each external gear 18 is formed with at least one, in this case, three crankshaft holes 21 penetrating in the axial direction, and the plurality of crankshaft holes 21 are arranged in the radial direction from the central axis of the external gear 18. Along with the distance, they are equidistant in the circumferential direction. Reference numeral 22 denotes a plurality of (three as many as the crankshaft holes 21) through-holes formed in each external gear 18. These through-holes 22 are alternately arranged in the circumferential direction with the crankshaft holes 21. , Are arranged equidistantly in the circumferential direction. The through-hole 22 has a substantially base shape with a circumferential width increasing toward the outside in the radial direction.

  Reference numeral 25 denotes a support body (carrier) that is loosely fitted in the rotary case 12 and is attached to a fixed robot member (not shown). The support body 25 is a pair of substantially arranged externally in the axial direction of the external gear 18. Ring-shaped end plate portions 26, 27, and a plurality of (through-hole 22 and through-hole 22), one end of which is integrally connected to end plate portion 26 and the other end is detachably connected to end plate portion 27 by a plurality of bolts 28. 3) which is the same number). The column portion 29 that connects the end plate portions 26 and 27 extends in the axial direction, and is inserted (freely fitted) into the through hole 22 of the external gear 18 while maintaining a slight gap.

  Since the pillar portion 29 is loosely fitted in the through hole 22 in this way, the external gear 18 located at the radially outer side of the through hole 22 is not supported from the inside. The bridge portion 30 has a thickness J (the radial distance from the radial outer end of the through hole 22 to the root 19b of the external tooth 19 of the external gear 15 is the minimum thickness). It is much smaller than the wall thickness at the part, and the bending rigidity is low.

  31 is a pair of bearings interposed between the support body 25, specifically, the outer periphery of the end plate portions 26 and 27 and the inner periphery of both end portions in the axial direction of the rotary case 12, and the internal gear 15 is supported by these bearings 31. Is rotatably supported by the support 25. Reference numeral 35 denotes at least one (three as many as the crankshaft holes 21) crankshafts arranged at an equal angle in the circumferential direction. The plurality of crankshafts 35 are externally fitted to one axial end portion thereof. A tapered roller bearing 36 and a tapered roller bearing 37 fitted on the other end in the axial direction thereof are rotatably supported by the support body 25, specifically the end plate portions 26 and 27.

  The crankshaft 35 has two eccentric cams 38 that are eccentric by an equal distance from the central axis of the crankshaft 35 at the center in the axial direction, and these eccentric cams 38 are out of phase with each other by 180 degrees. Here, the eccentric cam 38 of the crankshaft 35 is loosely fitted in the crankshaft hole 21 of the external gear 18, and a needle roller bearing 39 is interposed between them, and as a result, the outer cam 38 is inserted. Relative rotation between the toothed gear 18 and the crankshaft 35 is allowed. An external gear 40 is fixed to one end of each crankshaft 35 in the axial direction, and an external gear 42 provided at one end of an output shaft 41 of a drive motor (not shown) meshes with these external gears 40. .

  When the drive motor operates and the external gear 40 rotates, the crankshaft 35 rotates around its own central axis, and as a result, the eccentric cam 38 of the crankshaft 35 is moved into the crankshaft hole 21 of the external gear 18. As a result, the external gear 18 rotates eccentrically in the direction of the arrow. At this time, as shown in FIGS. 2, 3, and 4, contact points between the internal teeth (pins) 14 and the external teeth 19 that are meshed with each other are connected to the corresponding internal teeth (pins) 14 from the external teeth 19. A driving component force in the direction of action S is applied, and a reaction force K of driving force in the direction of action S is applied from the internal teeth (pins) 14 to the external teeth 19 as a reaction.

  Here, the action line S of each reaction force K described above is located on a line perpendicular to the tooth surface at the contact point, but the plurality of action lines S have the inner teeth (pins) 14 in a cylindrical shape as described above. Since the external teeth 19 are formed of a trochoidal tooth profile, they are set (intersected) at one point on the external gear 18, that is, at the set point C. Then, the sum of the tangential direction components of the driving component force is applied to the internal gear 15 as a rotational driving force.

  Further, a part of the reaction force K of the driving component force acts on the bridge portion 30 having the low bending rigidity described above, and the reaction portion K causes the bridge portion 30 and the vicinity of the bridge portion 30 to be external. The teeth 19 are elastically deformed so that the outer teeth 19 and the inner teeth (pins) 14 come into contact with each other, and the tooth surface life of the outer teeth 19 is shortened, or the natural frequency is lowered and the vibration characteristics and controllability are reduced. May fall.

  For this reason, in this example, the tooth tip 19a indicated by the phantom line of the external tooth 19 is a ratio B obtained by dividing the diameter D of the pin constituting the internal tooth 14 by the constant pitch P of the internal tooth 14. For example, when the number of teeth of the inner teeth (pins) 14 is 40, the inner teeth 15a are reduced from about 0.55 to about 0.32, so that the inner teeth ( The diameter D of the pin 14 is made smaller than before, and the root 19b of the external tooth 19 of the external gear 18 is moved outward in the radial direction.

  As described above, when the root 19b of the external tooth 19 moves radially outward, the radial distance from the radial outer end of the through hole 22 to the root 19b of the external tooth 19, that is, the bridge portion 30 As a result, the thickness J becomes thicker and the bending rigidity becomes higher. As a result, the elastic deformation of the bridge portion 30 and the external teeth 19 when the reaction force K is applied is suppressed, and the tooth surface life of the external teeth 19 is reduced. In addition to being able to extend, even when a torque load is present, the natural frequency is increased and the vibration characteristics and controllability can be improved.

  Here, when the diameter D of the internal teeth (pins) 14 becomes small as described above, both the tooth surfaces (the front tooth surfaces and the rear tooth surfaces in the rotation direction) contact the adjacent internal teeth (pins) 14 respectively. The tooth thickness and height of the tooth 19 increase, but if the ratio B is decreased until the tooth tip 19a exceeds the inner periphery 15a radially outward as described above, the outer tooth 19 having an increased tooth height is formed on the inner periphery. Interfering with 15a. For this reason, at least a portion of the external tooth 19 beyond the internal circumference 15a of the internal gear 15 is cut away so that interference between the external teeth 19 and the internal circumference 15a of the internal gear 15 is avoided.

  In this example, in the maximum meshing portion between the internal gear 15 and the external gear 18, the excision is made to such an extent that a slight gap is generated between the distal end of the external tooth 19 after the excision and the inner periphery 15a of the internal gear 15. By doing so, interference between the external teeth 19 and the inner periphery 15a of the internal gear 15 is avoided. Then, when the distance between the rotation direction front edge 44a and the rotation direction rear edge 44b of the external tooth 19 after cutting in this way is A, the diameter D of the pin constituting the internal tooth 14 is the distance A. It is preferable to make it smaller.

  Here, the cutting position of the external gear 18 on the external teeth 19 is radially outward from the line M connecting the inflection points H of both tooth surfaces of the external teeth 19 (the front tooth surface and the rear tooth surface in the rotational direction). Thus, the diameter D of the pin constituting the inner tooth 14 is changed from the linear distance Y between the centers of two adjacent pins constituting the inner tooth 14 to the outer tooth 19 after being cut along the line M. It is preferable that the distance F between the rotation direction front edge 45a and the rotation direction rear edge 45b is not less than a value obtained by subtracting the distance F. The reason for this is that, as described above, the inflection point H that plays the largest role in torque transmission (the contact pressure with the internal teeth 14 is the maximum value) can be left without being cut, and the transmission torque This is because it is possible to suppress the reduction of the above. Here, the line M is an arc line passing through both inflection points H with the central axis of the external gear 18 as the center of curvature.

  In addition, the cutting position of the external gear 18 on the external teeth 19 is set radially inward from the boundary N (the height position of 1/2 of the tooth height) between the end portion and the root portion of the external teeth 19. Accordingly, the diameter D of the pin constituting the inner tooth 14 is determined from the rotational distance front edge 46a and the rotational rear edge 46b of the external tooth 19 after being cut at the boundary N from the linear distance Y between the centers. It is preferable that the distance E be equal to or less than the value obtained by subtracting the distance E. The reason is that a large slip occurs when the outer teeth 19 and the inner teeth (pins) 14 radially outward from the boundary N are engaged with each other. However, as described above, the outer teeth 19 and the inner teeth (pins) 14 This is because it is possible to leave a portion where there is little slippage, thereby reducing noise and heat generation.

  For this reason, the distance between the rotation direction front edge 45a and the rotation direction rear edge 45b of the external tooth 19 after the external tooth 19 is excised by a line M connecting the inflection points H of both tooth surfaces. F, and the distance between the rotational direction front edge 46a and the rotational direction rear edge 46b of the external tooth 19 after the external tooth 19 has been excised at the boundary N between the end of the tooth and the base of the tooth. Then, by cutting the outer teeth 19 radially outward from the line M and radially inward from the boundary N, the diameter D of the pins constituting the inner teeth 14 is changed to the linear distance Y between the centers. It is preferable that the distance be equal to or greater than the value obtained by subtracting the distance F from the distance, and not greater than the value obtained by subtracting the distance E from the center-to-center linear distance Y.

  Then, as described above, when each external tooth 19 is excised from the tooth tip 19a by a predetermined amount, the internal teeth (pins) 14 and the external teeth 19 mesh with each other only, so that the remaining internal teeth (pins) ) 14 tries to come out of the pin groove 13 without contacting the external teeth 19. Therefore, in this example, as shown in FIG. 1, as a restricting means in which insertion holes 49 into which both end portions of the internal teeth (pins) 14 are inserted are formed between the bearing 31 and the external gear 18. Two pin presser rings 50 are interposed, and these two pin presser rings 50 are fixed to the internal gear 15 so as not to rotate, thereby restricting the movement of the internal teeth (pins) 14 described above. .

  As the aforementioned restricting means, an insertion hole formed on the inner end surface of the outer race of the bearing 31 and into which both end portions of the inner teeth (pins) 14 are inserted, or the inner end surface of the outer race of the bearing 31 is used. A circumferential groove having the same width as the diameter of the inner teeth (pins) 14 may be used.

Next, the operation of the example will be described.
Now, the drive motor operates and the crankshaft 35 rotates. At this time, the eccentric cam 38 of the crankshaft 35 rotates eccentrically in the crankshaft hole 21 of the external gear 18 to rotate the external gear 18 eccentrically, and the number of teeth of the external teeth 19 of the external gear 18 is increased. Is one less than the number of internal teeth (pins) 14, the rotating case 12, the robot arm, and the like rotate at a low speed due to the eccentric oscillation rotation of the external gear 18.

  Here, as described above, the ratio B obtained by dividing the diameter D of the internal teeth (pins) 14 by a constant pitch P is used until the tooth tip 19a of the external tooth 19 exceeds the inner periphery 15a of the internal gear 15 outward in the radial direction. Since the diameter D of the internal teeth (pins) 14 is smaller than before, the bottom 19b of the external teeth 19 of the external gear 18 moves outward in the radial direction. The wall thickness J (minimum wall thickness) becomes thicker than before, and the bending rigidity increases.

  As a result, the elastic deformation of the bridge portion 30 and the external teeth 19 when the reaction force K of the driving component force acts is suppressed, and the tooth surface life of the external teeth 19 can be extended, and the natural frequency is increased. Thus, vibration characteristics and controllability can be improved. Here, when configured as described above, the external teeth 19 interfere with the inner periphery 15a of the internal gear 15, but at least by cutting out the external teeth 19 at a portion beyond the inner periphery 15a of the internal gear 15, Such interference between the external teeth 19 and the inner periphery 15a of the internal gear 15 is avoided.

    5 and 6 are views showing Embodiment 1 of the present invention. In this embodiment, the external teeth 19 are not removed as in the above example, and the internal circumference of the internal gear 15 (rotary case 12) between the adjacent internal teeth (pins) 14 and the internal teeth (pins) 14 The outer circumference of the outer teeth 19 is cut by a depth equal to or greater than the amount of the outer teeth 19 exceeding the inner circumference, here a depth equal to approximately half the diameter D of the inner teeth (pins) 14. And the inner periphery 15a of the internal gear 15 (rotating case 12) after cutting are avoided.

  As a result, the radially outer end of each internal tooth (pin) 14 is in line contact with the inner periphery 15a of the internal gear 15 after resection, and thereby the driving force applied to each internal tooth (pin) 14 is reduced. The rotating case 12 receives the radial component. At this time, since the pin groove 13 does not exist, each internal tooth (pin) 14 can freely move. However, the pin pressing ring 50 similar to that described above allows the internal teeth (pin) 14 to move. The movement is restricted. Other configurations and operations are the same as in the above example.

  Here, the diameter D of the pins constituting the internal teeth 14 described above is in the range of 2R / Z ± 1.5 mm, where R is the radius of the pin circle V and Z is the number of external teeth 19 of the external gear 18. It is preferable to be inside. The reason for this is that when the diameter D is within the above-mentioned range, the Hertz stress at the contact point between the inner teeth (pins) 14 and the outer teeth 19 rapidly increases as is apparent from the graph shown in FIG. This is because it is maintained at a lower value inside the starting point, and the tooth surface life of the external teeth 19 can be extended.

  The graph shown in FIG. 7 is obtained by performing a simulation in the following specifications using the planetary gear device of the above example, but the same effect can be obtained with this embodiment. That is, the number of teeth (number) of the internal teeth (pins) of each planetary gear device is 40, the radius R of the pin circle V is 120 mm, the number of external teeth is 39, and the eccentric amount of the external gear 18 with respect to the internal gear 15 While making Q a constant value of 2.7 mm, the Hertz stress at the contact point between the external tooth 19 and the internal tooth (pin) 14 was determined while changing the diameter D of the internal tooth (pin) 14. Here, FIG. 7 shows the Hertzian stress value when the diameter D is equal to 2R / Z as an index 1.

  As described above, as the method of moving the root 19b of the external tooth 19 outward in the radial direction as the diameter D decreases, the eccentric amount Q of the external gear 18 with respect to the internal gear 15 is kept constant without being changed. However, a method of increasing the diameter of the root circle passing through the entire tooth bottom 19b of the external gear 18, a method of increasing the eccentric amount Q while keeping the root circle unchanged, and a root circle and There is a method of increasing both of the eccentric amounts Q. In this example, the eccentric amount Q is increased while keeping the root circle constant.

  When the amount of eccentricity Q is increased in this way, the distance L from the center O of the internal gear 15 to the set point C (obtained by multiplying the amount of eccentricity Q by the number of teeth of the inner teeth 14) is larger than before. That is, the position of the set point C can be moved radially outward. At this time, the ratio between the distance L and the radius R of the pin circle V, that is, the value of L / R is within the range of 0.86 to 1.00. It is preferable to do.

  The reason is that if the value of L / R is 0.86 or more, the action line S is inclined in the tangential direction with respect to the external gear 18, and as a result, the thickness of the bridge portion 30 receiving the reaction force K is increased. Thus, the elastic deformation in the bridge portion 30 can be effectively suppressed, and as is apparent from FIG. 8, the torque ratio applied to the external teeth 19 is obtained so that the load ratio is substantially constant and the same torque is obtained. This is because the load regarding can be substantially constant and minimized. However, when the value of the ratio L / R exceeds 1.00, a sharp point may be generated on the tooth surface when the external tooth 19 is created. Therefore, the value of L / R is preferably 1.00 or less.

  Here, the above-mentioned graph is obtained by performing a simulation in the following specifications using the planetary gear device of the above example, but the same result is obtained even in this embodiment. That is, the number of teeth (pins) of the inner teeth (pins) of each planetary gear device is 40, the diameter D of the inner teeth (pins) is 10 mm, the radius R of the pin circle V is 120 mm, and the number of teeth of the outer teeth is 39. On the other hand, the value of L / R was changed in the range of 0.5 to 1.0, and the tangential direction component of the resultant force obtained by synthesizing the drive component acting on the gathering point C was obtained. Here, in FIG. 8, the tangential direction component when the value of L / R is 0.75 is displayed as a graph with the load ratio being index 1.

    In the above, a plurality (three) of the crankshaft holes 21 are formed in the external gear 18, and the crankshafts 35 that rotate at the same speed in the same direction are inserted into the respective crankshaft holes 21. The gear 18 is rotated in an eccentric manner. In the present invention, an eccentric cam of one crankshaft is inserted into one crankshaft hole formed on the central shaft of the external gear 18, The external gear may be rotated eccentrically by rotating the crankshaft. In this case, the pillar portion of the support needs to be in line contact with the inner periphery of the through hole. Further, in the above, the support 25 is fixed and the internal gear 15 is rotated at a low speed. However, in the present invention, the internal gear may be fixed and the support may be rotated at a low speed. .

  The present invention can be applied to an eccentric oscillating planetary gear device in which an external gear meshing with an internal gear is eccentrically oscillated by a crankshaft.

11 ... Planetary gear unit 14 ... Internal teeth (pin)
15 ... Internal gear 15a ... Inner circumference
18 ... External gear 19 ... External gear
21 ... Crankshaft hole 22 ... Through hole
25 ... Support 29 ... Column
35 ... Crankshaft 45a ... Front edge in the rotation direction
45b ... Rotational direction rear edge 46a ... Rotational direction front edge
46b ... Rear edge in rotation direction

Claims (3)

    An internal gear formed of a large number of cylindrical pins on the inner periphery and provided with a constant pitch P, at least one crankshaft hole and a plurality of through holes are formed, and has a trochoidal tooth profile on the outer periphery. An external gear that meshes with the internal teeth and has only one external tooth having a smaller number of teeth than the internal teeth, a crankshaft that is inserted into each crankshaft hole and rotates to eccentrically swing the external gear; In an eccentric oscillating planetary gear device including a support body having a plurality of pillar portions inserted into the through holes and rotatably supporting the crankshaft, a diameter D of a pin constituting an internal tooth is set to an inner diameter. The ratio of the teeth divided by the constant pitch P is reduced until the tooth tips of the external teeth exceed the inner periphery of the internal gear in the radial direction, and the inner periphery of the internal gear between adjacent internal teeth is the outer teeth. Excluded by a depth greater than the amount exceeding the inner circumference That the the to avoid interference with the inner circumference of the internal gear and the eccentric oscillating-type planetary gear device according to claim.     When the radius of the pin circle V passing through the center of all the pins constituting the internal teeth is R and the number of external teeth of the external gear is Z, the diameter D of the pins constituting the internal teeth is 2R. The eccentric oscillating planetary gear device according to claim 1, wherein the eccentric slewing planetary gear device is in a range of /Z±1.5 mm.     The radius of the pin circle V passing through the centers of all the pins constituting the internal teeth is R, and the reaction of the drive component force applied from the external teeth to the corresponding internal teeth from the center O of the internal gears. The eccentric oscillating type according to claim 1, wherein the radial distance L is set within a range of 0.86 to 1.00 times the radius R, where L is a radial distance to the set point C where the action lines S of the force K overlap. Planetary gear device.

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