WO2012101777A1 - 撓み噛合い式歯車装置及び撓み噛合い式歯車装置の歯形の決定方法 - Google Patents

撓み噛合い式歯車装置及び撓み噛合い式歯車装置の歯形の決定方法 Download PDF

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Publication number
WO2012101777A1
WO2012101777A1 PCT/JP2011/051464 JP2011051464W WO2012101777A1 WO 2012101777 A1 WO2012101777 A1 WO 2012101777A1 JP 2011051464 W JP2011051464 W JP 2011051464W WO 2012101777 A1 WO2012101777 A1 WO 2012101777A1
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WIPO (PCT)
Prior art keywords
gear
internal gear
external
internal
meshing
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PCT/JP2011/051464
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English (en)
French (fr)
Japanese (ja)
Inventor
真司 吉田
正昭 芝
安藤　学
Original Assignee
住友重機械工業株式会社
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Application filed by 住友重機械工業株式会社 filed Critical 住友重機械工業株式会社
Priority to DE112011104783T priority Critical patent/DE112011104783T5/de
Priority to PCT/JP2011/051464 priority patent/WO2012101777A1/ja
Priority to KR1020137014112A priority patent/KR101486880B1/ko
Priority to CN201180061524.6A priority patent/CN103270335B/zh
Publication of WO2012101777A1 publication Critical patent/WO2012101777A1/ja

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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H49/00—Other gearings
    • F16H49/001—Wave gearings, e.g. harmonic drive transmissions
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H1/00—Toothed gearings for conveying rotary motion
    • F16H1/28—Toothed gearings for conveying rotary motion with gears having orbital motion
    • F16H1/34—Toothed gearings for conveying rotary motion with gears having orbital motion involving gears essentially having intermeshing elements other than involute or cycloidal teeth
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H1/00—Toothed gearings for conveying rotary motion
    • F16H1/28—Toothed gearings for conveying rotary motion with gears having orbital motion
    • F16H1/32—Toothed gearings for conveying rotary motion with gears having orbital motion in which the central axis of the gearing lies inside the periphery of an orbital gear
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H55/00—Elements with teeth or friction surfaces for conveying motion; Worms, pulleys or sheaves for gearing mechanisms
    • F16H55/02—Toothed members; Worms
    • F16H55/08—Profiling
    • F16H55/0833—Flexible toothed member, e.g. harmonic drive

Definitions

  • the present invention relates to a flexure meshing gear device and a method for determining a tooth profile of a flexure meshing gear device.
  • a flexure meshing gear device shown in Patent Document 1 includes a vibrating body and a cylindrical external tooth that is arranged on the outer periphery of the vibrating body and has the flexibility to bend and deform by the rotation of the vibrating body.
  • the external gear flexed and deformed by the rotation of the vibration generator is in meshed with the first internal gear, and the first internal gear and The external gear is decelerated based on the difference in the number of teeth. And the output of the decelerated external gear can be taken out from the second internal gear.
  • the present invention has been made to solve the above-described problems, and a flexure meshing gear device and a flexure meshing gear capable of improving impact resistance and further increasing transmission torque and transmission efficiency. It is an object of the present invention to provide a method for determining a tooth profile of a device.
  • the present invention relates to a vibrating body, a cylindrical external gear that is disposed on the outer periphery of the vibrating body and is flexible and deformed by the rotation of the vibrating body, and the external gear A first internal gear having rigidity for intermeshing engagement; and a second internal gear having rigidity for intermeshing engagement with the external gear, which is arranged in parallel in the axial direction on the first internal gear.
  • tooth shapes of the external gears at the portions meshing with the first internal gear and the second internal gear are the same, and the external gear, the first internal gear, and the first gear are the same.
  • the two internal gears have tooth shapes such that the number of simultaneous meshes of the external gear and the first internal gear and the number of simultaneous meshes of the external gear and the second internal gear are both 2 or more, respectively.
  • the external gear and the first internal gear have a tooth profile in which the number of simultaneous meshes of the external gear and the two internal gears (the first internal gear and the second internal gear) is 2 or more. And the second internal gear. Therefore, the impact resistance is improved, the surface pressure applied to the meshing tooth surfaces is dispersed, and a large torque can be transmitted. And in the present invention, as its basic configuration, coupled with the configuration of meshing a cylindrical external gear with two rigid internal gears, ratcheting resistance can be improved, Further, the stress generated in the external gear at no load can be reduced as compared with the cup-type external gear, and the load capacity can be increased. For this reason, the present invention can increase the transmission torque and increase the transmission efficiency.
  • the tooth profile of the external gear is the same at the portions that mesh with the first internal gear and the second internal gear, the external gear can be easily processed and the processing cost can be kept low. Shape processing can be performed with high accuracy.
  • the present invention also provides a vibrating body, a cylindrical external gear that is arranged on the outer periphery of the vibrating body and has a flexibility that is bent and deformed by the rotation of the vibrating body, and the external gear.
  • a first internal gear having a rigidity for internal meshing
  • a second internal gear having a rigidity for axially juxtaposing the first internal gear and axially meshing with the external gear.
  • the rotating shaft of the vibration generator and the eccentric shaft that is the center of the meshing radius of the external gear when meshed with the first internal gear or the second internal gear.
  • the center of the pin or the internal tooth of the first internal gear or the second internal gear is cylindrical.
  • the center of the pin is arranged. Therefore, the load applied to the external teeth of the cylindrical external gear when meshed with the first internal gear and the load applied to the external teeth of the cylindrical external gear when engaged with the second internal gear.
  • the two load regions applied to the external gear can be brought close to each other in the circumferential direction of the external gear. In other words, when viewed from the axial direction, the two internal gears can sandwich only a small number of external teeth during the meshing operation.
  • the present invention makes it possible to increase the allowable transmission torque and increase the transmission efficiency, particularly focusing on improving the ratchetability.
  • impact resistance is improved, and transmission torque and transmission efficiency can be increased.
  • the disassembled perspective view which shows an example of the whole structure of the bending meshing type gear apparatus which concerns on 1st Embodiment of this invention.
  • sectional drawing which shows an example of whole composition
  • the figure which also shows a vibration body The figure which also shows a vibration body Schematic diagram of a combination of a vibrator and a vibrator bearing Similarly, meshing diagram of external gear and internal gear Similarly, an enlarged view of the meshing of the external gear, the internal gear for reduction, and the internal gear for output
  • the figure which similarly shows the position of the entity of the tooth profile of the external gear, the internal gear for reduction, and the internal gear for output The figure which defines the tooth profile of the external gear
  • the figure which similarly defines the tooth profile of the internal gear for reduction and the internal gear for output The figure which similarly defines the tooth profile of the internal gear for reduction and the internal gear for output
  • the figure which similarly defines the tooth profile of the internal gear for reduction and the internal gear for output The figure which similarly defines the tooth profile of the internal gear for reduction and the internal gear for
  • Table showing the number of simultaneous meshes in the internal gear for output when the reduction ratio and the diameter of the internal gear are changed in the first embodiment The figure which shows the relationship between the position of the entity of the external gear in 1st Embodiment, and a pitch point.
  • the disassembled perspective view which shows an example of the whole structure of the bending meshing type gear apparatus which concerns on 2nd Embodiment of this invention.
  • the flexure meshing gear device 100 is arranged on the outer periphery of the vibrating body 104 and the vibrating body 104 and has flexible external gears 120A and 120B (simply simply deformed by the rotation of the vibrating body 104).
  • External gear 120 and a reduction internal gear 130 ⁇ / b> A that is a first internal gear having rigidity with which the external gear 120 internally meshes, and an output internal gear that is a second internal gear. 130B.
  • the internal gear 130A for reduction and the internal gear 130B for output are collectively referred to simply as the internal gear 130.
  • the vibrator 104 has a column 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 provided in the input shaft hole 106 so that the vibrator 104 rotates integrally with the input shaft when the input shaft is inserted and rotated.
  • the vibration body 104 is configured in a shape (two arc shapes) obtained by connecting two arc portions (first arc portion FA, second arc portion SA).
  • the first arc portion FA is an arc having a radius of curvature r1 centered on a point B (referred to as an eccentric shaft), and an arc portion (also referred to as a meshing range) for meshing the external gear 120 and the internal gear 130. ).
  • the second arc portion SA is an arc having a radius of curvature r2 centered on the point C, and constitutes an arc portion (also referred to as a non-meshing range) in a range where the external gear 120 and the internal gear 130 do not mesh with each other.
  • the length of the first arc portion FA is determined by a meshing angle ⁇ that is an angle formed between the major axis x and the normal line N at the point A.
  • the radius of curvature r1 of the first arc portion FA is expressed by equation (1), where L is the eccentricity. .
  • the tangent line T (normal line N) is common to the connecting portion A between the first arc portion FA and the second arc portion SA.
  • the radius of curvature r2 of the second arc portion SA is (the radius of curvature r1 + the length BC), and therefore is expressed by Expression (2).
  • the vibration body bearing 110A is a bearing disposed between the outside of the vibration body 104 and the inside of the external gear 120A. As shown in FIGS. 2 and 5, the inner ring 112, the cage 114A, and the rolling element. As a roller 116A and an outer ring 118A. The inner side of the inner ring 112 abuts on the vibrating body 104, and the inner ring 112 rotates while deforming integrally with the vibrating body 104.
  • the roller 116A has a cylindrical shape (including a needle).
  • the outer ring 118A is disposed outside the roller 116A.
  • the outer ring 118 ⁇ / b> A is bent and deformed by the rotation of the vibrating body 104, and deforms the external gear 120 ⁇ / b> A disposed outside the outer ring 118 ⁇ / b> A.
  • the vibration body bearing 110B is comprised from the inner ring
  • the vibration body 104 and the inner ring 112 are common to the vibration body bearings 110A and 110B.
  • the retainer 114B, the roller 116B, and the outer ring 118B are the same as the retainer 114A, the roller 116A, and the outer ring 118A as a single member (component).
  • the external gear 120A meshes internally with the reduction internal gear 130 ⁇ / b> A.
  • the external gear 120A includes a base member 122 and external teeth 124A.
  • the base member 122 is a flexible cylindrical member that supports the external teeth 124 ⁇ / b> A, and is disposed outside the vibration body bearing 110 ⁇ / b> A.
  • the external teeth 124A are cylindrical pins having a radius ⁇ 1 (for this reason, the external teeth 124A (124B), the external gear 120A (120B), and the flexibly meshing gear device 100 of this embodiment are simply pin types. Called).
  • the external teeth 124A are held on the base member 122 by a ring member 126A.
  • the external gear 120B meshes internally with the output internal gear 130B as shown in FIG.
  • the external gear 120B is comprised from the base member 122 and the external tooth 124B similarly to the external gear 120A.
  • the external teeth 124B have the same number as the external teeth 124A and are configured by the same cylindrical pin, and are held by the base member 122 by the ring member 126B. That is, the base member 122 supports the external teeth 124A and the external teeth 124B in common. That is, the external gears 120A and 120B have the same shape.
  • the eccentric amount L of the vibrator 104 is transmitted to the external teeth 124A and the external teeth 124B in the same phase.
  • the external teeth 124A and 124B are collectively referred to as external teeth 124.
  • the internal gear 130A for deceleration is formed of a rigid member as shown in FIG.
  • the internal gear 130A for reduction has a number of teeth that is a multiple of two than the number of teeth 124A of the external gear 120A (the number of teeth will be described in detail later).
  • a casing (not shown) is fixed to the reduction internal gear 130A via a bolt hole 132A.
  • the internal gear 130A for deceleration contributes to the deceleration of rotation of the vibration body 104 by meshing with the external gear 120A.
  • FIG. 6A shows how the external gear 120A meshes with the reduction internal gear 130A
  • FIG. 7A shows the appearance of the external teeth 124A and the internal teeth 128A on the x-axis.
  • the output internal gear 130B is also formed of a rigid member, like the reduction internal gear 130A.
  • the output internal gear 130B has the same number of teeth of the internal teeth 128B as the number of teeth of the external teeth 124B of the external gear 120B (constant speed transmission).
  • an output shaft (not shown) is attached to the output internal gear 130B via a bolt hole 132B, and the same rotation as the rotation of the external gear 120B is output to the outside.
  • FIG. 6B shows how the external gear 120B meshes with the output internal gear 130B
  • FIG. 7B shows the appearance of the external teeth 124B and internal teeth 128B on the x-axis.
  • the inner teeth 128A and 128B are collectively referred to as the inner teeth 128.
  • the simultaneous meshing number Nph of the external gear 120A and the reduction internal gear 130A and the simultaneous meshing number Npl of the external gear 120B and the output internal gear 130B are both set to 2 or more and
  • the meshing is the theoretical meshing. For this reason, torque transmission efficiency is not lowered, smooth torque transmission can be realized, and transmission torque can be increased.
  • the tooth profile of the external gear 120 is defined.
  • the locus of the tooth profile of the external gear 120 is represented by a trochoid curve equation, and the tooth profile of the internal gear 130 is defined using the trochoid curve equation.
  • a plurality of parameters defining the tooth shapes of the external gear 120 and the internal gear 130 are associated with each other from the size and the number of teeth of the external gear 120 and the internal gear 130.
  • the correction range of the tooth tip and the tooth root of the tooth profile of the internal gear 130 is determined.
  • the tooth profile portion outside the correction range is obtained using the associated parameters, and the number of simultaneous meshes is obtained from the tooth shape portion.
  • optimal parameters are determined so that the number of simultaneous meshes is 2 or more. In determining the parameters, trial and error are performed so that the target values such as the torque, the allowable surface pressure of the tooth surface, the main stress of each portion, and the bearing life are satisfied at the same time.
  • the tooth profile of the external gear 120 is defined.
  • the internal tooth 128 of the internal gear 130 is a cylindrical pin having a radius ⁇ 2 (including a case where the internal tooth 128 is simply assumed in design)
  • the internal tooth is determined from the rotation axis Fc (point on the axial direction O) of the vibration generator 104.
  • the relationship between the radius R and the radius R1 is expressed by Expression (3).
  • the external gear 120 is disposed on the outer periphery of the vibration generator 104 via the vibration generator bearing 110.
  • the radial thicknesses of the vibrator bearing 110 and the external gear 120 are both constant. For this reason, since the vibrating body 104 has a two-arc shape, the external gear 120 also has a two-arc shape.
  • the radius of the entity of the tooth profile in the meshing range of the external gear 120 corresponding to the curvature radius r1 of the meshing range of the oscillator 104 is R1.
  • Equations (2) and (3) are used.
  • the radius R2 can be expressed by the equation (4).
  • R2 R1-L / cos ⁇ ... (4)
  • the tooth profile of the external gear 120 is defined by the radius ⁇ 1, the eccentric amount L, the radius R, and the meshing angle ⁇ .
  • the reduction gear ratio when the external gear 120 is a circular gear (referred to as a virtual gear) having a tooth profile entity radius R1 is referred to as a virtual reduction ratio n.
  • the external gear 120 is revolved by an angle ⁇ around the rotation axis Fc of the vibration body 104. That is, the eccentric shaft B rotates ⁇ .
  • the coordinate (x1, y1) of the position of the tooth profile of the external gear 120 rotates in the opposite direction by the angle ⁇ / n by the virtual reduction ratio n and moves to the coordinate (x2, y2).
  • the coordinates (x pfc , y pfc ) indicating the locus of the position of the tooth profile of the external gear 120 are expressed by equations (5) and (6).
  • the coordinate of the position of the tooth profile of the internal gear 130 is expressed by an internal trochoid curve equation (hypotrochoid). (Curve formula). That is, the radius b1 of the base circle BA fixed around the rotation axis Fc, the radius a1 of the rolling circle AA that rotates without sliding along the circumference of the base circle BA, the radius L1 of the drawing point, and the rotation angle ⁇ 1.
  • the coordinates (x pfc , y pfc ) of the position of the tooth profile of the internal gear 130 are expressed by equations (7) and (8).
  • Formula (14) is calculated
  • radius R, ⁇ 1, eccentricity L, virtual reduction ratio n (virtual reduction ratio n h for making the tooth profile of the internal gear 130A for reduction, virtual reduction ratio n for making the tooth shape of the internal gear 130B for output
  • the coordinates (x fc , y fc ) of the tooth profiles of the reduction internal gear 130A and the output internal gear 130B can be obtained.
  • the table shows LC (circumference length) and the pitch P (the circumferential length of one tooth period) when using the virtual gear reduction ratio n and the number of teeth NT. Can do.
  • Expression (19) and Expression (20) can be derived from the table of FIG.
  • a parameter Gp (referred to as pin type pitch coefficient) is introduced.
  • pin type pitch coefficient an intersection of a straight line passing through the eccentric shaft B and the rotation shaft Fc and a common normal of a contact point generated by meshing of the external gear 120 (external teeth 124) and the internal gear 130 (internal teeth 128) is defined.
  • the pitch point is defined by the external gear 120 and the internal gear 130.
  • the pin type pitch coefficient Gp can easily grasp the relative positional relationship between the positions of the tooth forms of the external gear 120 and the internal gear and the pitch points, and can easily adjust the parameters. It was introduced as follows.
  • a pin type pitch coefficient Gph (referred to as a pin type deceleration-side pitch coefficient) obtained at this time is defined in Expression (22) based on Expression (21).
  • a pin type pitch coefficient Gpl (referred to as a pin type output side pitch coefficient) obtained at this time is defined in Expression (24) based on Expression (21).
  • the value of the pin type output side pitch coefficient Gpl> 1 is obtained by substituting the pin type deceleration side pitch coefficient Gph ⁇ 1.
  • the result of obtaining each tooth form is as follows. Is a more preferable condition.
  • an angle when an angle ⁇ formed by a straight line connecting the coordinates of the inner teeth 128 and the center Oc of the outer teeth 124 (pins) and the x axis is about 45 degrees is ⁇ s.
  • the angle ⁇ is from 0 to ⁇ s, there is a possibility of interference with the external teeth 124 of the external gear 120, and therefore, correction is performed on the root of the internal teeth 128 of the internal gear 130 within that range.
  • the angle ⁇ when the distance ⁇ between the tooth tip of the outer tooth 124 and the tooth tip of the inner tooth 128 is about 15% of the radius ⁇ 1 of the pin is defined as an angle ⁇ f.
  • angles ⁇ s to ⁇ f are the effective range in which the theoretical engagement is achieved.
  • the simultaneous meshing 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 (a value obtained by dividing 2 ⁇ by the number of teeth NT).
  • the angles ⁇ fh and ⁇ sh are angles in the deceleration internal gear 130A
  • the angles ⁇ fl and ⁇ sl are angles in the output internal gear 130B.
  • the rotation angles determined by the angles ⁇ fh, ⁇ sh, ⁇ fl, ⁇ sl from the relationship of the equation (14) are ⁇ fh, ⁇ sh, ⁇ fl, ⁇ sl, respectively. That is, by using the equation (14), the simultaneous meshing number Nph of the reduction internal gear 130A is obtained by the equation (26), and the simultaneous meshing number Npl of the output internal gear 130B is obtained by the equation (27). .
  • FIG. 18 shows the simultaneous meshing number Npl of the output internal gear 130B.
  • the reduction gear ratio of the internal gear 130 of this embodiment is determined.
  • the external gear 120A is bent and deformed via the vibrator bearing 110A according to the rotation state.
  • the external gear 120B is also bent and deformed in the same phase as the external gear 120A via the vibration body bearing 110B.
  • the bending deformation of the external gear 120 is made according to the shape of the radius of curvature r1 of the vibrating body 104. Since the curvature is constant at the position of the first arc portion FA of the vibrating body 104 shown in FIG. 4, the bending stress is constant. Since the tangent line T is the same at the position of the first arc part FA and the second arc part SA in the connecting part A, sudden deformation at the connecting part is prevented. At the same time, since there is no abrupt position change of the rollers 116A and 116B in the connecting portion A, the rollers 116A and 116B are less slipped and torque transmission loss is small.
  • the tooth shape of the inner tooth 128 is a tooth shape based on a trochoid curve with respect to the outer tooth 124 which is a cylindrical pin. For this reason, since the external teeth 124 and the internal teeth 128 are perfectly meshed with each other, loss can be reduced and high torque transmission efficiency can be realized.
  • rollers 116A and 116B have a cylindrical shape, the load capacity is large and there are many portions that come into contact with the inner ring 112 and the outer rings 118A and 118B with respect to ball bearings having the same size ball. Can be bigger.
  • the external teeth 124 are divided into a portion (external teeth 124A) that meshes with the internal gear 130A for deceleration and a portion (external teeth 124B) that meshes with the internal gear 130B for output. Therefore, when the external gear 120A meshes with the reduction internal gear 130A, even if the external teeth 124B are deformed, the deformation does not cause the external teeth 124A to be deformed. Similarly, when the external gear 120B meshes with the output internal gear 130B, even if the external teeth 124A are deformed, the external teeth 124B are not deformed by the deformation.
  • the meshing position between the external gear 120 ⁇ / b> A and the reduction internal gear 130 ⁇ / b> A rotates and moves as the vibration body 104 moves in the long axis direction x.
  • the rotation phase of the external gear 120A is delayed by a difference in the number of teeth from the internal gear 130A for deceleration.
  • the reduction ratio by the reduction internal gear 130A is ((the number of teeth of the external gear 120A (N * k) ⁇ the number of teeth of the reduction internal gear 130A ((N + 1) * k)) / the external gear 120A.
  • the number of teeth (N * k)) ⁇ 1 / N.
  • both the external gear 120B and the output internal gear 130B have the same number of teeth (N * k)
  • the external gear 120B and the output internal gear 130B are engaged with each other without moving.
  • the same teeth are meshed with each other.
  • the same rotation as the rotation of the external gear 120B is output from the output internal gear 130B.
  • an output obtained by reducing the rotation of the vibrating body 104 based on the reduction ratio 1 / N by the reduction internal gear 130A can be extracted from the output internal gear 130B.
  • the present embodiment includes, as its basic configuration, a configuration in which the cylindrical external gear 120 is engaged with two internal gears 130 having a rigidity (a reduction internal gear 130A and an output internal gear 130B), and
  • the external gear 120 and the internal gear 130 are configured so that the external gear 120 and the internal gear 130 have tooth shapes in which the simultaneous meshing numbers Nph and Npl are both 2 or more, and a trochoid curve is used. So, the theoretical engagement is realized. For this reason, the impact resistance is improved, the surface pressure applied to the meshing tooth surfaces is dispersed, and a large torque can be transmitted, and the local stress generated in the external gear 120 is particularly affected by the conventional general cup. Compared to the type of flexure meshing gear device, the number can be significantly reduced.
  • the conical area is increased and the surface is not constricted due to the bending of the vibration generating body, and there is no stress concentration at the bottom of the cup. Since the pressure can be dispersed, the load capacity can be greatly increased.
  • the load applied to the external teeth 124A of the external gear 120A when meshing with the internal gear 130A for reduction and the load applied to the external teeth 124B of the external gear 120B when meshing with the internal gear 130B for output are:
  • the two load regions applied to the external gear 120 can be brought close to each other in the circumferential direction of the external gear 120 while having components opposite to each other. That is, when viewed from the axial direction O, the two internal gears 130 can sandwich only a small number of external teeth 124 during the meshing operation. For this reason, the phenomenon (ratcheting phenomenon) that the engagement between the external gear 120 and the internal gear 130 is shifted by excessive torque can be prevented. That is, the ratcheting resistance can be improved.
  • Deflection-meshing gear device using cup-type external gears that are actually commercialized with an internal gear tooth shape having a radius of about 26 mm and a reduction ratio of 1/50 (referred to as a comparative example)
  • the ratchet resistance is greatly improved (approximately 4 times or more) from the actual measurement value of the comparative example. It has been confirmed that.
  • the rated torque in the comparative example was 3.3 kgfm
  • the flexure meshing gear device 100 of the present embodiment it can be confirmed by theoretical calculation and tests that the rated torque is 6.6 kgfm. It was. That is, it can be confirmed by theoretical calculation that the rated torque is about twice, and it can be confirmed by a test.
  • the flexure meshing gear device 100 can be made more compact.
  • the tooth shape of the external gear 120 is the same at the portions that mesh with the reduction internal gear 130A and the output internal gear 130B, respectively, so that the processing of the external gear 120 is easy.
  • the processing cost can be kept low and the shape can be processed with high accuracy.
  • Second Embodiment An example of the second embodiment according to the present invention will be described in detail with reference to FIGS.
  • a tooth shape based on a trochoid curve is adopted for the external gear, and the external teeth of the external gear are formed integrally with the base member (solid Called type).
  • solid Called type the base member
  • the external gear 220 ⁇ / b> A meshes internally with the reduction internal gear 230 ⁇ / b> A.
  • the external gear 220A includes a base member 222 and external teeth 224A.
  • the base member 122 is a flexible cylindrical member, is disposed outside the vibration body bearing 210A, and is integrally formed with the external teeth 224A. For this reason, the external teeth 224A can be made small and highly accurate processing can be performed. That is, the external gear 220A of this embodiment is suitable for a small flexure meshing gear device with a small load capacity.
  • the external teeth 224A are formed based on a trochoid curve.
  • the external gear 220B meshes internally with the output internal gear 230A as shown in FIGS. And the external gear 220B is comprised from the base member 222 and the external tooth 224B similarly to the external gear 220A.
  • the external teeth 224B are formed in the same number and shape as the external teeth 224A.
  • the external teeth 224A and the external teeth 224B are separated in the axial direction, but the base member 222 is common. That is, the external gears 220A and 220B have the same shape.
  • the eccentric amount L of the vibrator 204 is transmitted to the external teeth 224A and the external teeth 224B in the same phase.
  • the external teeth 224A and 224B are collectively referred to as external teeth 224.
  • the locus of the position of the tooth profile of the external gear is obtained, and the tooth profile of the internal gear is defined from the locus.
  • a plurality of parameters defining the tooth shapes of the external gear 220 and the internal gear 230 are associated with each other based on the size and the number of teeth of the external gear 220 and the internal gear 230.
  • the correction range of the tooth tip and the tooth root of the tooth profile of the internal gear 230 is determined.
  • the tooth profile portion outside the correction range is obtained using the associated parameters, and the number of simultaneous meshes is obtained from the tooth shape portion.
  • optimal parameters are determined so that the number of simultaneous meshes is 2 or more.
  • trial and error are performed so that the target values such as the torque, the allowable surface pressure of the tooth surface, the main stress of each portion, and the bearing life are satisfied at the same time.
  • the tooth profile of the external gear 220 is defined.
  • a cylindrical pin having a radius ⁇ 2 is virtually arranged as the internal tooth 228A of the reduction internal gear 230A (for convenience, the reduction internal gear 230A is provided, but it may be arranged on the output internal gear 230B. ),
  • the locus of the position of the tooth profile of the internal gear 230A for reduction with the pin radius ⁇ 2 0 (synonymous with the center of the pin). Then, the tooth profile of the external gear 220 is moved to the inside (external gear 220 side) by the pin radius ⁇ 2. More specific description will be given below.
  • the virtual reduction ratio n (n h , n l ) has the same definition as in the first embodiment.
  • the external gear 220 has a two-arc shape, and the relationship between the radii R1 and R2 is expressed by Expression (3) and Expression (4).
  • the external gear 220 theoretically meshes with a reduction internal gear 230A that virtually includes a pin. For this reason, as shown in FIG. 22, the locus drawn when the center of the pin of the deceleration internal gear 230A moves from the coordinate (x4, y4) to the coordinate (x5, y5) in a static space centered on the eccentric shaft B. (X p , y p ) are expressed by an external trochoid curve equation (epitrochoid curve equation) as the coordinates of the position of the tooth profile of the external gear 220.
  • the coordinates (x kfc , y kfc ) of the external gear 220 can be obtained by changing the angle ⁇ by substituting the radius R, ⁇ 2, the eccentricity L, and the virtual reduction ratio n h .
  • the tooth profile of the internal gear 230 is defined.
  • the envelope of the coordinate (x p , y p ) of the position of the tooth profile of the external gear 220 is obtained, and the envelope is moved inward (internal gear 230 side) by a radius ⁇ 2 to determine the tooth profile of the internal gear 230.
  • the locus Q of the tooth profile of the external gear 220 on the xd-yd coordinate centered on the eccentric axis B of the external gear 220 is as shown in FIG. Draw an envelope (solid line portion shown in FIG. 23).
  • the coordinates (x pfc , y pfc ) of the position of the tooth profile of the internal gear 230 with the rotation axis Fc as the origin are expressed by the equations (40), (41) using the equations (30), (31). It is represented by 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.
  • the shape of the external gear 220 is a two-arc shape defined by the radii R1 and R2. That is, also in this embodiment, the relationship of Formula (19) and Formula (20) is established.
  • a parameter Gs (referred to as a solid type pitch coefficient) is introduced.
  • a solid type pitch coefficient an intersection of a straight line passing through the eccentric shaft B and the rotation shaft Fc and a common normal of a contact point generated by meshing of the external gear 220 (external teeth 224) and the internal gear 230 (internal teeth 228) is obtained.
  • the pitch point is defined by the external gear 220 and the internal gear 230 (that is, the definition of the pitch point is the same as in the first embodiment).
  • the solid type pitch coefficient Gs can easily grasp the relative positional relationship between the positions of the tooth forms of the external gear 220 and the internal gear 230 and the pitch points, and These parameters are introduced so as to facilitate the adjustment of these parameters.
  • the solid type pitch coefficient Gs is determined by the radius R and the distance (n + 1) * L from the rotation axis Fc to the pitch point by the external gear 220 and the reduction internal gear 230. It is expressed as a ratio.
  • FIG. 25 shows the relationship between the radius R of the tooth profile of the internal gear 230 and the virtual reduction ratio n 1 .
  • the solid type pitch coefficient Gsl (referred to as a solid type output side pitch coefficient) obtained at this time is defined in Expression (49) based on Expression (46).
  • the present embodiment substitutes the solid type deceleration side pitch coefficient Gsh ⁇ 1 to obtain the value of the solid type output side pitch coefficient Gsl> 1, as shown in FIGS. Yes.
  • the present embodiment may further include a case where the meshing angle ⁇ 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. It is a more preferable condition from the result of obtaining each tooth profile.
  • the tips and roots of the internal teeth 228 are corrected.
  • the angles ⁇ s to ⁇ f in which the tooth profile has not been corrected are the effective range in which the theoretical engagement is performed.
  • the simultaneous meshing 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, using the relationship of the equation (43), the simultaneous meshing number Nsh of the reduction internal gear 230A is obtained by the equation (52), and the simultaneous meshing number Nsl of the output internal gear 230B is obtained by the equation (53). It is done.
  • FIG. 27 shows the simultaneous meshing number Nsl of the output internal gear 230B.
  • the tooth profile of the internal gear 230 in the present embodiment is determined by the condition of the diameter (2 * R) and the reduction ratio (1 / N) that realize both of these simultaneous meshing numbers Nsh and Nsl are 2 or more.
  • the reduction ratio (1 / N) is 1/30, which is not the tooth profile of the present embodiment, and is reduced to 1/50 or less (greater than 1/50).
  • the reduction gear ratio determines the tooth profile of the internal gear of this embodiment.
  • the external gear 220 can be easily processed and the processing can be performed with high accuracy.
  • the solid type deceleration side pitch coefficient Gsh ⁇ 1 and the solid type output side pitch coefficient Gsl> 1 are set.
  • Formula (54) is materialized. That is, as shown in equation (55), the distance from the axis of rotation Fc until the pitch point P h by a gear 230A in the deceleration and the external gear 220A ((n h +1) * L) and the external teeth from the rotation axis Fc
  • the center of the pin when the internal tooth 228 of the internal gear 230 is virtually assumed to be a pin between the distance ((n 1 +1) * L) to the pitch point P 1 by the gear 220B and the output internal gear 230B ( The position of the tooth profile entity) is arranged.
  • the load Fd applied to the external teeth 224A of the external gear 220A when meshing with the internal gear 230A for reduction and the load Fo applied to the external teeth 224B of the external gear 220B when meshing with the internal gear 230B for output. Can have components opposite to each other and can bring the regions of the two loads Fd and Fo applied to the external gear 220 close to each other in the circumferential direction of the external gear 220. That is, as shown in FIG. 29, when viewed from the axial direction O, the two internal gears 230 have only a small number of external teeth 224 while the regions of the load Fd and the load Fo are brought close to each other during the meshing operation. It can be set as the aspect inserted. For this reason, the ratcheting resistance can be improved as in the first embodiment.
  • the tooth profile of the external gear or the internal gear is obtained based on the trochoid curve.
  • the contact line that is the locus of the contact point generated by the meshing of the external gear and the internal gear can be uniquely determined from the coordinates of the tooth profile of the internal gear that is obtained, so that it can also be used.
  • the unique relationship between the coordinates of the tooth profile of the internal gear 130 and the contact line in the case of the first embodiment will be specifically described.
  • the contact line CL is a locus viewed from the XY coordinate system shown in FIG. 30 in which the coordinates of the tooth profile (x fc , y fc ) of the internal gear 130 are rotated by an angle ⁇ . Therefore, the coordinates (x cfc , y cfc ) of the contact line are given by the equations (57) and (58) obtained by rotating the tooth shape coordinates (x fc , y fc ) of the internal gear 130 by an angle ⁇ .
  • the contact line CL obtained by the above equation is shown in FIG.
  • the contact line CL is drawn between a plurality of tooth tips and tooth roots of the external gear 120 and the internal gear 130, and it can be seen that a plurality of simultaneous meshing numbers Nph and Npl can be secured.
  • 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.
  • the present invention is not necessarily limited to such a relationship.
  • the deceleration side pitch coefficients Gph and Gsh may be larger than 1 and the output side pitch coefficients Gpl and Gsl may be smaller than 1.
  • 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 profiles of the external gear and the internal gear can be obtained by determining not only the parameters that define the pitch coefficient but also the adjustment of many parameters through trial and error.
  • the flexure meshing gear device of the present invention can be used for various applications, but can be suitably used for precision control applications such as industrial robot joint (wrist) drive devices and machine tools.
  • precision control applications such as industrial robot joint (wrist) drive devices and machine tools.

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PCT/JP2011/051464 2011-01-26 2011-01-26 撓み噛合い式歯車装置及び撓み噛合い式歯車装置の歯形の決定方法 WO2012101777A1 (ja)

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Application Number Priority Date Filing Date Title
DE112011104783T DE112011104783T5 (de) 2011-01-26 2011-01-26 Getriebevorrichtung der flexibel eingreifenden Bauart und ein Bestimmungsverfahren für ihr Zahnprofil
PCT/JP2011/051464 WO2012101777A1 (ja) 2011-01-26 2011-01-26 撓み噛合い式歯車装置及び撓み噛合い式歯車装置の歯形の決定方法
KR1020137014112A KR101486880B1 (ko) 2011-01-26 2011-01-26 휨 맞물림식 기어장치 및 휨 맞물림식 기어장치의 치형의 결정방법
CN201180061524.6A CN103270335B (zh) 2011-01-26 2011-01-26 挠曲啮合式齿轮装置及挠曲啮合式齿轮装置的齿形确定方法

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PCT/JP2011/051464 WO2012101777A1 (ja) 2011-01-26 2011-01-26 撓み噛合い式歯車装置及び撓み噛合い式歯車装置の歯形の決定方法

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CN105065580A (zh) * 2015-06-26 2015-11-18 吴小杰 搬运焊接机器人行星摆线减速器
EP3165795A1 (en) * 2015-10-16 2017-05-10 Hamilton Sundstrand Corporation Flex spine for use with a compound harmonic generator

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JP6175381B2 (ja) * 2014-02-21 2017-08-02 住友重機械工業株式会社 撓み噛合い式歯車装置
WO2016198673A1 (de) * 2015-06-12 2016-12-15 Ovalo Gmbh Radial bauraumsparendes spannungswellengetriebe
DE102015109426A1 (de) 2015-06-12 2016-12-15 Ovalo Gmbh Radial bauraumsparendes Spannungswellengetriebe

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JPH0526305A (ja) * 1991-07-25 1993-02-02 Takeda Haguruma Kogyo Kk 撓み噛み合い式駆動伝達装置
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JP4645086B2 (ja) 2004-07-20 2011-03-09 株式会社ジェイテクト 波動歯車装置、伝達比可変装置、及び波動歯車装置の製造方法
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JPS63130949A (ja) * 1986-11-21 1988-06-03 Sumitomo Heavy Ind Ltd 遊星歯車の歯形を利用したハ−モニツクドライブ増減速機
JPH0526305A (ja) * 1991-07-25 1993-02-02 Takeda Haguruma Kogyo Kk 撓み噛み合い式駆動伝達装置
WO2007116756A1 (ja) * 2006-03-31 2007-10-18 Jtekt Corporation 撓み噛み合い式歯車装置及び車両用ステアリング装置

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Publication number Priority date Publication date Assignee Title
CN105065580A (zh) * 2015-06-26 2015-11-18 吴小杰 搬运焊接机器人行星摆线减速器
EP3165795A1 (en) * 2015-10-16 2017-05-10 Hamilton Sundstrand Corporation Flex spine for use with a compound harmonic generator
US9915334B2 (en) 2015-10-16 2018-03-13 Hamilton Sundstrand Corporation Flex spline for use with a compound harmonic generator

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KR20130084313A (ko) 2013-07-24
CN103270335A (zh) 2013-08-28
KR101486880B1 (ko) 2015-01-28
CN103270335B (zh) 2015-12-09

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