JP2004044685A - Inscribed engagement planetary gear mechanism - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To make a pressure angle smaller and increase transmission efficiency of an outer and inner toothed gears because the shapes of conventional toothed wheels have large pressure angles with low transmission efficiency due to formation by curves such as a prolate epicycloid, a prolate hypocycloid, a curtate epicycloid, or a curtate hypocycloid. <P>SOLUTION: The inscribed engagement planetary gear mechanism is equipped with a first shaft 1, an outer toothed gear 3 eccentrically rotatable relative to the first shaft 1 through an eccentric part 2 of the first shaft 1, an inner toothed gear 4 inscribed and engaged to the outer toothed gear 3, and a second shaft 6 connected through a transmitting means 5 for transmitting only the rotation component of the outer toothed gear 3. The forms of both the teeth of the outer toothed gear 3 and the inner toothed gear 4 are provided with the tooth form of the inside of a pitch circle as a hypocycloid, and the form of the outside of the pitch circle as an epicycloid. Thereby, the pressure angles of the outer toothed gear 3 and the inner toothed gear 4 can be made smaller, and transmission efficiency can be increased. <P>COPYRIGHT: (C)2004,JPO

Description

[0001]
TECHNICAL FIELD OF THE INVENTION
The present invention relates to an internally meshing planetary gear mechanism used for a speed reducer or a speed increaser, and particularly to a tooth profile.
[0002]
[Prior art]
The tooth profile of the external gear and the internal gear in the conventional internal meshing planetary gear mechanism is formed by a prolate epicycloid curve, a prolate hypocycloid curve, a curate epicycloid curve or a curate hypocycloid curve.
[0003]
[Problems to be solved by the invention]
However, the internal meshing planetary gear mechanism, which has a tooth profile formed by a prolate epicycloid curve, a prolate hypocycloid curve, a curate epicycloid curve or a curate hypocycloid curve, has a large pressure angle, and therefore has a low transmission efficiency. There was a defect.
[0004]
[Object of the invention]
The present invention has been made in view of the above circumstances, and an object of the present invention is to provide an internally meshing planetary gear mechanism that can increase the transmission efficiency by reducing the pressure angle between an external gear and an internal gear. is there.
[0005]
[Means for Solving the Problems]
[Means of claim 1]
The tooth profile of the external gear and the internal gear is provided by a hypocycloid curve on the inner side of the pitch circle, and the tooth profile on the outer side of the pitch circle is represented by the epicycloid curve. By providing such a structure, the pressure angle can be reduced, and the transmission efficiency can be increased.
[0006]
In the following, the number of teeth of the external gear is N,
The diameter of the pitch circle of the external gear is φD1,
The number of teeth of the internal gear is M,
The diameter of the pitch circle of the internal gear is φD2,
The diameter of the rolling circle for drawing a hypocycloid curve forming the tooth profile curve of the external gear is φD1H,
The diameter of the rolling circle for drawing the epicycloid curve forming the tooth profile curve of the external gear is φD1E,
The diameter of a rolling circle for drawing a hypocycloid curve forming the tooth profile curve of the internal gear is φD2H,
A diameter of a rolling circle for drawing an epicycloid curve forming a tooth profile curve of the internal gear will be described as φD2E.
[0007]
[Means of Claim 2]
The means of claim 2 may be adopted and provided so as to satisfy the relationship of φD1 / N = φD2 / M.
[0008]
[Means of Claim 3]
The means of claim 3 may be employed so that the relationship of φD1H> φD1E is satisfied and the relationship of φD1H + φD1E = φD1 / N is satisfied.
With such provision, the clearance is formed in the external gear, and interference with the internal gear can be avoided.
[0009]
[Means of Claim 4]
The means of claim 4 may be employed so as to satisfy the relationship of φD2H <φD2E and to satisfy the relationship of φD2H + φD2E = φD2 / M.
With such provision, the clearance is formed in the internal gear, and interference with the external gear can be avoided.
[0010]
[Means of claim 5]
Adopting the means of claim 5,
satisfies the relationship φD1H> φD1E,
The relationship of φD1 H + φD1 E = φD1 / N is satisfied,
satisfy the relationship of φD2 H <φD2 E,
The relationship of φD2H + φD2 E = φD2 / M is satisfied,
satisfies the relationship φD1H = φD2E,
It may be provided so as to satisfy the relationship of φD1E = φD2H.
With such provision, the clearance is formed on both the external gear and the internal gear, and interference between the external gear and the internal gear can be avoided.
[0011]
[Means of claim 6]
By adopting the means of claim 6, the tooth profile of the portion of the external gear that is drawn with the hypocycloid curve may have a relief portion that widens in the radial direction from the pitch circle for drawing the hypocycloid curve. good.
[0012]
[Means of claim 7]
The means according to claim 7, wherein the tooth profile of the portion of the internal gear that is drawn by the epicycloid curve has a relief portion that increases in width from the pitch circle for drawing the epicycloid curve toward the outer diameter direction. Is also good.
[0013]
[Means of claim 8]
The tooth profile of the portion of the external gear that is drawn by the hypocycloid curve is provided with a relief portion that increases in width from the pitch circle for drawing the hypocycloid curve toward the inner diameter direction. The tooth profile of the portion of the toothed gear drawn by the epicycloid curve may have a relief portion that becomes wider from the pitch circle for drawing the epicycloid curve toward the outer diameter direction.
[0014]
By adopting the above-mentioned means, it is possible to reliably prevent wear loss due to contact of the tooth tip portion which does not contribute to torque transmission.
Further, since the relief portion can be used as a grease pool, grease can be prevented from running out.
Further, since the epicycloid curve and the hypocycloid curve can be smoothly connected, stress concentration due to contact of the protruding portion can be reduced, and uneven wear of the tooth surface can be prevented.
[0015]
[Means of claim 9]
Employing the means of claim 9, the eccentric amount α of the eccentric portion is
{(ΦD1 / N) / 2} × (M−N) = {(φD2 / M) / 2} × (M−N).
[0016]
[Claim 10 means]
Employing the means of claim 10, the eccentric amount α of the eccentric portion,
{(ΦD1 / N) / 2} × (MN) or more,
{(ΦD1 / N) / 2} × (M−N) + (φD1H−φD1E) or less.
[0017]
[Means of claim 11]
The means of claim 11 is adopted, and the eccentric amount α of the eccentric portion is
{(ΦD1 / N) / 2} × (MN) or more,
{(ΦD1 / N) / 2} × (M−N) + (φD2E−φD2H).
[0018]
[Means of claim 12]
The eccentric amount α of the eccentric portion is obtained by adopting the means of claim 12.
{(ΦD1 / N) / 2} × (MN) or more,
{(ΦD1 / N) / 2} × (M−N) +2 (φD2 E−φD2H) or less.
[0019]
By adopting the means of the ninth to twelfth aspects, the amount of eccentricity α of the eccentric portion becomes optimal with respect to the amount of clearance of the tooth tip.
For this reason, a reliable meshing state can be realized at a portion where the tooth surfaces mesh with each other, and interference of the tooth tip can be prevented at a portion where the tooth tip passes without contacting.
[0020]
[Means of claim 13]
The means of claim 13 is adopted, wherein the transmission means is provided with a plurality of inner pin holes provided on the same circumference of a flange which rotates integrally with the second shaft, and each of the transmission means is loosely fitted into the inner pin holes, and one end is provided outside. And a plurality of inner pins fixed to the tooth gear and provided on the same circumference of the external gear.
[0021]
[Means of claim 14]
A transmission means is freely fitted in each of a plurality of inner pin holes provided on the same circumference of the external gear, and one end is rotated integrally with the second shaft. And a plurality of inner pins provided on the same circumference of the flange.
[0022]
[Means of claim 15]
The diameter φDPDC-hole of the pitch circle forming the plurality of inner pin holes and the diameter φDPDC-pin of the pitch circle forming the plurality of inner pins are provided to have the same size. Is also good.
[0023]
[Means of claim 16]
The means of claim 16 is employed, wherein a plurality of inner pin holes are provided at an equal angle pitch with respect to a pitch circle forming the plurality of inner pin holes, and the plurality of inner pins are provided with the plurality of inner pins. The pitch circles may be provided at equal angular pitches.
[0024]
[Means of claim 17]
The diameter of the inner pin hole φDhole may be set to be equal to or more than the eccentric amount α of the eccentric portion + the diameter of the inner pin φDpin.
With such provision, the dimensional relationship between the inner pin diameter φDpin and the inner pin hole diameter φDhole can be optimized with respect to the eccentric amount α of the eccentric portion. Therefore, it is possible to prevent the occurrence of torque unevenness due to twisting between the inner pin and the inner pin hole.
[0025]
[Means of Claim 18]
Employing the means of claim 18, a relief portion for avoiding contact with the inner pin hole may be provided on the inner diameter side and the outer diameter side of the inner pin.
[0026]
[Means of claim 19]
Employing the means of claim 19,
A relief portion for avoiding contact with the inner pin may be provided on the inner diameter side and the outer diameter side of the inner pin hole.
[0027]
By adopting the means of claims 18 and 19, the inner pin and the inner pin hole do not come into contact with each other in a positional relationship where the pressure angle becomes large, and the mechanical efficiency can be improved.
Further, since the relief portion can be used as a grease pool, grease can be prevented from running out.
[0028]
[Means of claim 20]
21. The sliding bearing according to claim 20, wherein the second shaft is rotatably supported on the housing via a sliding bearing, and one end of the first shaft is rotatably supported on the second shaft via the first bearing. May be provided so as to overlap the axial support section x of the first axis and the axial support section y of the first axis.
With such provision, the inclination of the first shaft caused by the reaction force of the load can be avoided.
[0029]
BEST MODE FOR CARRYING OUT THE INVENTION
Embodiments of the present invention will be described using examples and modifications.
〔Example〕
An internally meshing planetary gear mechanism to which the present invention is applied will be described with reference to FIGS. In this embodiment, for convenience, the left side in FIG. 1 will be described as a front (front), and the right side as a rear (rear).
[0030]
First, the schematic structure of the internally meshing planetary gear mechanism will be described with reference to FIGS.
The internally meshing planetary gear mechanism according to this embodiment is used as a small and thin reduction gear. The first shaft 1 and an eccentric portion 2 provided on the first shaft 1 are used to drive the first shaft 1. Via an external gear 3 attached in such a manner as to be capable of eccentric rotation with respect to the internal gear, an internal gear 4 with which the external gear 3 meshes internally, and a transmission means 5 for transmitting only the rotation component of the external gear 3. And a second shaft 6 connected to the second shaft 6.
[0031]
The transmission means 5 of this embodiment has a plurality of inner pin holes 11 provided on the same circumference of a flange 10 which rotates integrally with the second shaft 6, and each of the transmission means 5 is loosely fitted into the inner pin holes 11, and one end of the transmission means 5 has an outer end. The external gear 3 is fixed to the tooth gear 3 and includes a plurality of inner pins 12 provided on the same circumference.
Different from this embodiment, the transmission means 5 has a plurality of inner pin holes 11 provided on the same circumference of the external gear 3, and each of the transmission means 5 is loosely fitted into the inner pin holes 11, and one end is provided on the second shaft 11. 6 and a plurality of inner pins 12 fixed to a flange 10 that rotates integrally with the flange 6 and provided on the same circumference of the flange 10.
[0032]
When used as a reduction gear, the first shaft 1 is an input shaft, and is provided at a first rolling bearing 13 (corresponding to a first bearing) provided at an end on the front side and at an end on the rear side. The second rolling bearing 14 rotatably supports the second rolling bearing 14. The first rolling bearing 13 is supported inside the rear end of the second shaft 6, and the second rolling bearing 14 is supported by the rear housing 15.
The eccentric portion 2 is a shaft that eccentrically rotates with respect to the rotation center of the first shaft 1 and swings and rotates the external gear 3. The eccentric portion 2 is externally rotated via a third rolling bearing 16 disposed on the outer periphery of the eccentric portion 2. The gear 3 is rotatably supported.
[0033]
Like the first shaft 1, when used as a speed reducer, the second shaft 6 is an output shaft and is rotatably supported by a front housing 18 via a slide bearing 17.
Here, the support section x in the axial direction of the slide bearing 17 is provided so as to overlap the support section y in the axial direction of the first rolling bearing 13. With such provision, the inclination of the first shaft 1 due to the reaction force of the load applied to the engagement between the external gear 3 and the internal gear 4 can be avoided.
[0034]
As described above, the external gear 3 is rotatably supported by the eccentric portion 2 of the first shaft 1 via the third rolling bearing 16, and the internal gear 4 is rotated by the rotation of the eccentric portion 2. It is configured to rotate in a state of being pressed against.
The internal gear 4 is fixed to the front housing 18 by a fixing projection 19 formed on the front surface side.
[0035]
The plurality of inner pins 12 are provided so as to protrude toward the front surface of the external gear 3.
The plurality of inner pin holes 11 are provided on a flange 10 provided at the rear end of the second shaft 6, and the rotation of the external gear 3 is restricted by the engagement between the inner pins 12 and the inner pin holes 11. It is configured to be transmitted to the two shafts 6.
[0036]
Next, the definition of a cycloid curve will be described with reference to FIG.
As shown by a, b, c, a ', b', and c 'in FIG. 8, the cycloid curve rolls on an arc of a pitch circle (base circle) without slipping on an abduction or adduction circle. This is a locus drawn by one point in the radial direction of the rolling circle when it is touched.
Of these, the trajectory drawn by rolling the abduction circle is generally called an epicycloidal curve (a, b, c), and the trajectory drawn by rolling the abduction circle is generally a hypocycloidal curve (a ′, b ′). , C ').
[0037]
Specifically, the points where the trajectory is drawn inside the turning circle (inside in the radial direction) are called a prolate epicycloid curve (a) and a prolate hypocycloid curve (a '). Is taken outside of the rolling circle (outside in the radial direction) is referred to as a “curate epicycloid curve (c)” and a “curtain hypocycloid curve (c ′)”.
And, the points where the trajectory is drawn on the arc of the turning circle are also simply called the epicycloid curve (b) and the hypocycloid curve (b ').
Here, the epicycloid curve and the hypocycloid curve referred to in the present invention indicate the epicycloid curve (b) and the hypocycloid curve (b ') in which the points of the locus are drawn on the arc of the turning circle. .
[0038]
Next, the tooth shapes of the external gear 3 and the internal gear 4 will be described.
The tooth profile of the external gear 3 and the internal gear 4 is such that the tooth profile inside the pitch circle is provided by a hypocycloid curve and the tooth profile outside the pitch circle is provided by an epicycloid curve. .
[0039]
Specific tooth shapes of the external gear 3 and the internal gear 4 are as follows.
The number of teeth of the external gear 3 is N,
The diameter of the pitch circle of the external gear 3 is φD1,
The number of teeth of the internal gear 4 is M,
The diameter of the pitch circle of the internal gear 4 is φD2,
The diameter of a rolling circle for drawing a hypocycloid curve forming the tooth profile curve of the external gear 3 is φD1H,
The diameter of the rolling circle for drawing the epicycloid curve forming the tooth profile curve of the external gear 3 is φD1E,
The diameter of a rolling circle for drawing a hypocycloid curve forming the tooth profile curve of the internal gear 4 is φD2H,
When the diameter of a rolling circle for drawing an epicycloid curve forming the tooth profile curve of the internal gear 4 is φD2E,
satisfying the relationship of φD1 / N = φD2 / M,
satisfies the relationship φD1H> φD1E,
The relationship of φD1 H + φD1 E = φD1 / N is satisfied,
satisfy the relationship of φD2 H <φD2 E,
The relationship of φD2H + φD2 E = φD2 / M is satisfied,
satisfies the relationship φD1H = φD2E,
It is provided so as to satisfy the relationship of φD1E = φD2H.
[0040]
That is, the relationship of φD1 / N = φD2 / M is satisfied,
satisfies the relationship φD1H> φD1E,
By satisfying the relationship φD1 H + φD1 E = φD1 / N,
The tooth profile of the external gear 3 is as shown by the solid line A in FIG. 9, and a predetermined clearance Δx is formed with respect to the tooth profile of the clearance 0 shown by the solid line B in FIG.
[0041]
Further, the relationship of φD1 / N = φD2 / M is satisfied,
satisfy the relationship of φD2 H <φD2 E,
By satisfying the relationship of φD2 H + φD2 E = φD2 / M,
The tooth profile of the internal gear 4 is as shown by the solid line C in FIG. 9, and a predetermined clearance Δy is formed with respect to the tooth profile of the clearance 0 shown by the solid line B in FIG.
[0042]
Further, the relationship of φD1 H = φD2 E is satisfied,
By satisfying the relationship φD1E = φD2H,
The clearances Δx and Δy respectively formed in the external gear 3 and the internal gear 4 become equal (Δx = Δy).
With such provision, the clearance is formed on both the external gear 3 and the internal gear 4, and the interference between the external gear 3 and the internal gear 4 can be reliably avoided.
[0043]
Here, the solid line A in FIG. 9 showing the tooth profile of the external gear 3 is an epicycloid curve outside the pitch circle (see A1 in the figure) and a hypocycloid curve inside the pitch circle (see A2 in the figure). The solid line C in FIG. 9 showing the tooth profile of the internal gear 4 is an epicycloid curve outside the pitch circle (see C1 in the figure) and a hypocycloid inside the pitch circle. The curve (see C2 in the figure) is continuous.
[0044]
2 to 7 disclose an example in which the number N of teeth of the external gear 3 is 60 and the number M of teeth of the internal gear 4 is M = 61.
However, in FIG. 10, the number of teeth N of the external gear 3 is described in order to explain the contact point β between the external gear 3 and the internal gear 4 (a portion provided for torque transmission, hereinafter referred to as a contact point β). = 9 and the number of teeth M of the internal gear 4 is set to M = 10.
As shown in FIG. 10, the contact point β is located on the innermost side of the pitch circle of the external gear 3 in FIG. 10 (the part where the external gear 3 and the internal gear 4 are closest to each other in the radial direction). ) Except for the contact point β, and the contact point β approaches the outer diameter direction of the external gear 3 as the distance from the uppermost contact point β decreases, so that the influence of torque transmission is reduced. Then, on the lower side in FIG. 10, the contact between the external gear 3 and the internal gear 4 is lost, and the influence of torque transmission is eliminated.
Note that the internal gear 4 does not have a contact point β in the outer diameter direction from the pitch circle of the internal gear 4.
[0045]
Here, the dashed line A shown in FIG. 10 is a line segment connecting the tooth center of the internal gear 4 formed by the hypocycloid curve and the center of the pitch circle of the internal gear 4. A dashed line B shown in FIG. 10 is a line segment connecting the tooth center of the external gear 3 formed by the epicycloid curve and the center of the pitch circle of the external gear 3.
Further, a solid line C shown in FIG. 10 is a line segment connecting the intersection of the one-dot chain line A and the pitch circle of the internal gear 4 and the intersection of the one-dot chain line B and the pitch circle of the external gear 3.
The contact point β imparted to the torque transmission is the intersection of the solid line C and the tooth, and when the center point of the pitch circle of the external gear 3 rotates eccentrically clockwise in the figure, the external gear 3 Exists only on the right side of a line segment passing through the center point of the pitch circle of the internal gear 4 and the center point of the pitch circle of the internal gear 4.
[0046]
As shown in FIG. 11, the tooth profile of the portion of the external gear 3 drawn by the hypocycloid curve has a relief portion 21 (see FIG. 11) whose width increases from the pitch circle for drawing the hypocycloid curve toward the inner diameter direction. (Indicated by medium hatching).
In addition, the tooth profile of the portion of the internal gear 4 drawn by the epicycloid curve also has a relief portion 22 (a hatched portion in the figure) which increases in width from the pitch circle for drawing the epicycloid curve toward the outer diameter direction. ) Is formed.
[0047]
By providing the relief portions 21 and 22 in both the external gear 3 and the internal gear 4 in this way, it is possible to reliably prevent wear loss due to contact of the tooth tip portion that does not contribute to torque transmission.
In addition, since the relief portions 21 and 22 can be used as a grease pool, it is possible to prevent the grease from running out of the meshing portion between the external gear 3 and the internal gear 4.
Further, since the epicycloid curve and the hypocycloid curve can be smoothly connected, stress concentration due to contact of the protruding portion can be reduced, and uneven wear of the tooth surface can be prevented.
[0048]
Next, FIG. 12 shows an enlarged view of a portion indicated by an arrow A in FIG.
The eccentric amount α of the rotation center of the eccentric part 2 with respect to the rotation center of the first shaft 1 is
α = {(φD1 / N) / 2} × (M-N)
When the external gear 3 and the internal gear 4 are closest to each other in the radial direction (see FIG. 12), the radial clearance (Δx + Δy) between the external gear 3 and the internal gear 4 is:
Δx + Δy = (φD1H−φD1E) + (φD2E−φD2H).
Accordingly, it can be seen that there is no problem even if the eccentricity α is α = α + (Δx + Δy).
[0049]
Next, FIG. 13 shows an enlarged view of a portion indicated by an arrow B in FIG.
The eccentric amount α of the rotation center of the eccentric part 2 with respect to the rotation center of the first shaft 1 is
α = {(φD1 / N) / 2} × (M-N)
, The closest distance (Δx + Δy) / 2 between the external gear 3 and the internal gear 4 in the portion where the external gear 3 and the internal gear 4 are farthest (see FIG. 13) is
(Δx + Δy) / 2 = {(φD1H−φD1E) + (φD2E−φD2H)} / 2.
It can be seen that there is no problem even if the amount of eccentricity α is α = α− (Δx + Δy). However, in this case, the engagement ratio at the engagement portion shown in FIG.
[0050]
The eccentricity α is
It is not less than {(φD1 / N) / 2} × (M−N) and not more than {(φD1 / N) / 2} × (M−N) +2 (φD2 E−φD2H).
With this arrangement, the amount of eccentricity α becomes optimal with respect to the amount of clearance at the tooth tip.
For this reason, a reliable meshing state can be realized at a portion where the tooth surfaces mesh with each other, and interference of the tooth tip can be prevented at a portion where the tooth tip passes without contact (see FIG. 13).
[0051]
The relationship between the inner pin hole 11 and the inner pin 12 will be described with reference to FIGS.
The diameter φDPDC-hole of the pitch circle of the inner pin hole 11 and the diameter φDPDC-pin of the pitch circle of the inner pin 12 are provided in the same size (φDPDC-hole = φDPDC-pin).
In addition, a plurality of inner pin holes 11 are provided at an equal angle pitch with respect to a pitch circle of the inner pin holes 11, and a plurality of inner pins 12 are also provided at an equal angle pitch with respect to the pitch circle of the inner pins 12. ing.
The diameter φDhole of the inner pin hole 11 is provided to be larger than the eccentricity α + the diameter φDpin of the inner pin 12 (φDhole> α + φDpin).
With the provision as described above, the dimensional relationship between the inner pin 12 diameter φDpin and the inner pin hole 11 diameter φDhole can be optimized with respect to the amount of eccentricity α. The occurrence of torque unevenness due to twisting can be prevented.
[0052]
17 to 21 show the torque transmitted from the inner pin 12 to the inner pin hole 11.
The contact point γ at which torque is transmitted from the inner pin 12 to the inner pin hole 11 passes through the center point of the pitch circle of the external gear 3 and the center point of the pitch circle of the internal gear 4 in FIG. It exists only on the right side of the line segment XX.
The contact point γ at the portion indicated by arrow A in FIG. 17 is such that the inner pin 12 generates a load FA on the inner pin hole 11 as shown in FIG. 18, and the contact point γ at the portion indicated by arrow B in FIG. As shown in FIG. 19, the inner pin 12 generates a load FB on the inner pin hole 11, and the contact point γ at the portion indicated by the arrow C in FIG. As shown in FIG. 21, the inner pin 12 generates a load FD with respect to the inner pin hole 11 at the contact point γ at the portion indicated by the arrow D in FIG.
[0053]
Further, as shown in FIG. 22, the inner pin 12 of this embodiment is provided with relief portions 23 on the inner diameter side and the outer diameter side of the inner pin 12 for avoiding contact with the inner pin hole 11. .
By providing the relief portions 23 on the inner diameter side and the outer diameter side of the inner pin 12 in this manner, a structure in which the inner pin 12 and the inner pin hole 11 do not come into contact with each other in a positional relationship where the pressure angle becomes large can be achieved, and the mechanical efficiency can be improved. be able to. Further, since the relief portion 23 can be used as a grease pool, it is possible to prevent the grease from running out at the contact portion between the inner pin 12 and the inner pin hole 11.
In this embodiment, an example in which the relief portions 23 are provided on the inner diameter side and the outer diameter side of the inner pin 12 has been described. However, unlike this embodiment, the inner pin holes 11 are provided on the inner diameter side and the outer diameter side, A relief portion 23 for avoiding contact with the inner pin 12 may be provided.
[0054]
(Effects of the embodiment)
The tooth shapes of the external gear 3 and the internal gear 4 in the above-mentioned internal meshing planetary gear mechanism are provided such that a tooth shape inside the pitch circle is provided by a hypocycloid curve and a tooth shape outside the pitch circle is formed. By providing an epicycloidal curve, the pressure angle at the contact point γ between the external gear 3 and the internal gear 4 can be reduced, and the transmission efficiency can be increased.
[0055]
(Modification)
In the above-described embodiment, an example is described in which the internal gear 4 is fixed, the first shaft 1 is used as the input shaft, and the second shaft 6 is used as the output shaft. However, the second shaft 6 is fixed and the first shaft 1 is used. The reduction gear may be configured by using the input shaft and the internal gear 4 as the output shaft. Further, the speed increaser may be configured by reversing the input shaft and the output shaft.
[Brief description of the drawings]
FIG. 1 is a sectional view of an internally meshing planetary gear mechanism.
FIG. 2 is a diagram viewed from a direction A (rear side) in FIG.
FIG. 3 is a diagram viewed from a direction B (front side) in FIG. 1;
FIG. 4 is a perspective view of the internally meshing planetary gear mechanism as viewed from the rear side.
FIG. 5 is a perspective view of the internally meshing planetary gear mechanism as viewed from the front side.
FIG. 6 is an exploded perspective view of the internally meshing planetary gear mechanism as viewed from the rear side.
FIG. 7 is an exploded perspective view of the internally meshing planetary gear mechanism as viewed from the front side.
FIG. 8 is an explanatory diagram of various cycloid curves.
FIG. 9 is a diagram showing tooth shapes of an external gear and an internal gear.
FIG. 10 is an explanatory diagram of a contact point between an external gear and an internal gear.
FIG. 11 is an explanatory view of a relief portion provided on the external gear and the internal gear.
FIG. 12 is an enlarged view of a portion A in FIG. 2;
FIG. 13 is an enlarged view of a portion B in FIG. 2;
FIG. 14 is a view of the external gear seen from the front side.
FIG. 15 is a front view of a second shaft provided with a flange.
FIG. 16 is a front view of the internal meshing planetary gear mechanism showing an inner pin pitch circle and an inner pin hole pitch circle.
FIG. 17 is a diagram showing contact points between the inner pin and the inner pin hole, a load applied to the inner pin hole, and a direction of component force.
18 is an enlarged view of a portion A in FIG.
19 is an enlarged view of a portion B in FIG.
20 is an enlarged view of a portion C in FIG.
FIG. 21 is an enlarged view of a portion D in FIG. 17;
FIG. 22 is an explanatory diagram of a relief portion provided on an inner pin.
[Explanation of symbols]
DESCRIPTION OF SYMBOLS 1 First shaft 2 Eccentric part 3 External gear 4 Internal gear 5 Transmission means 6 Second shaft 10 Flange 11 Inner pin hole 12 Inner pin 13 First rolling bearing (first bearing)
17 Sliding bearing 21 Relief portion of external gear 22 Relief portion of internal gear 23 Relief portion of inner pin α Eccentric amount of eccentric portion x Support section in the axial direction of slide bearing y Support section N of first rolling bearing N External gear The number of teeth M of the internal gear The number of teeth φD1 of the external gear The diameter φD2 of the pitch circle of the external gear The diameter φD1 of the pitch circle of the internal gear H The roll for drawing a hypocycloid curve which forms the tooth profile curve of the external gear Diameter of circle φD1 E Diameter of rolling circle φD2 H for drawing epicycloid curve forming tooth profile curve of external gear Gear Inside diameter of rolling circle φD2 E for drawing hypocycloid curve forming tooth profile curve of internal gear Diameter φDPDC-hole of a rolling circle for drawing an epicycloid curve forming a tooth profile curve of a tooth gear Diameter φDPDC-pin of a pitch circle circle of a plurality of inner pin holes Pin pitch circle circle of diameter φDhole in the pin hole with a diameter of φDpin within the diameter of the pin of

Claims (20)

A first axis,
An external gear that is eccentrically rotatable with respect to the first shaft via an eccentric portion provided on the first shaft;
An internal gear in which the external gear meshes internally,
An internally meshing planetary gear mechanism comprising: a second shaft connected via a transmission unit that transmits only the rotation component of the external gear.
The tooth profile of the external gear and the internal gear is
An internally meshing planetary gear mechanism, wherein the tooth profile inside the pitch circle is provided by a hypocycloid curve, and the tooth profile outside the pitch circle is provided by an epicycloid curve. The internally meshing planetary gear mechanism according to claim 1,
The number of teeth of the external gear is N,
The diameter of the pitch circle of the external gear is φD1,
The number of teeth of the internal gear is M,
The diameter of the pitch circle of the internal gear is φD2,
The diameter of a rolling circle for drawing a hypocycloid curve forming the tooth profile curve of the external gear is φD1H,
The diameter of a rolling circle for drawing an epicycloid curve forming the tooth profile curve of the external gear is φD1E,
The diameter of a rolling circle for drawing a hypocycloid curve forming the tooth profile curve of the internal gear is φD2H,
When a diameter of a rolling circle for drawing an epicycloid curve forming a tooth profile curve of the internal gear is φD2E,
An internally meshing planetary gear mechanism which satisfies a relationship of φD1 / N = φD2 / M. The internally meshing planetary gear mechanism according to claim 2,
An internally meshing planetary gear mechanism which satisfies the relationship of φD1H> φD1E and the relationship of φD1H + φD1E = φD1 / N. The internally meshing planetary gear mechanism according to claim 2,
An internally meshing planetary gear mechanism, which satisfies a relationship of φD2H <φD2E and a relationship of φD2H + φD2E = φD2 / M. The internally meshing planetary gear mechanism according to claim 2,
satisfies the relationship φD1H> φD1E,
The relationship of φD1 H + φD1 E = φD1 / N is satisfied,
satisfy the relationship of φD2 H <φD2 E,
The relationship of φD2H + φD2 E = φD2 / M is satisfied,
satisfies the relationship φD1H = φD2E,
An internally meshing planetary gear mechanism which satisfies a relationship of φD1 E = φD2 H. The internally meshing planetary gear mechanism according to any one of claims 2, 3, and 5,
An internal meshing planetary gear characterized in that the tooth profile of the portion of the external gear that is drawn with a hypocycloid curve is provided with a relief portion that increases in width from the pitch circle for drawing the hypocycloid curve toward the inner diameter direction. mechanism. In the internally meshing planetary gear mechanism according to any one of claims 2, 4, and 5,
The tooth profile of the portion of the internal gear that is drawn by the epicycloid curve is provided with a relief portion that becomes wider from the pitch circle for drawing the epicycloid curve toward the outer diameter direction. Gear mechanism. The internally meshing planetary gear mechanism according to any one of claims 2 to 5,
The tooth profile of the portion of the external gear that is drawn with a hypocycloid curve has a relief portion that widens in the radial direction from a pitch circle for drawing the hypocycloid curve,
The tooth profile of the portion of the internal gear that is drawn by the epicycloid curve is provided with a relief portion that becomes wider from the pitch circle for drawing the epicycloid curve toward the outer diameter direction. Gear mechanism. The internally meshing planetary gear mechanism according to claim 2,
The eccentric amount α of the eccentric portion is:
An internally meshing planetary gear mechanism, wherein {(φD1 / N) / 2} × (M−N) = {(φD2 / M) / 2} × (M−N). The internal meshing planetary gear mechanism according to claim 3 or claim 6,
The eccentric amount α of the eccentric portion is:
{(ΦD1 / N) / 2} × (MN) or more,
An internally meshing planetary gear mechanism provided below {(φD1 / N) / 2} × (M−N) + (φD1H−φD1E). The internal meshing planetary gear mechanism according to claim 4 or claim 7,
The eccentric amount α of the eccentric portion is:
{(ΦD1 / N) / 2} × (MN) or more,
An internally meshing planetary gear mechanism provided below {(φD1 / N) / 2} × (M−N) + (φD2E−φD2H). The internal meshing planetary gear mechanism according to claim 4 or claim 8,
The eccentric amount α of the eccentric portion is:
{(ΦD1 / N) / 2} × (MN) or more,
An internally meshing planetary gear mechanism provided below {(φD1 / N) / 2} × (M−N) +2 (φD2 E−φD2H). The internal meshing planetary gear mechanism according to any one of claims 1 to 12,
The transmission means,
A plurality of inner pin holes provided on the same circumference of a flange rotating integrally with the second shaft;
A plurality of inner pins which are loosely fitted in the inner pin holes, one end of which is fixed to the external gear, and which are provided on the same circumference of the external gear, respectively. Gear mechanism. The internal meshing planetary gear mechanism according to any one of claims 1 to 12,
The transmission means,
A plurality of inner pin holes provided on the same circumference of the external gear,
A plurality of inner pins which are loosely fitted into the inner pin holes, one ends of which are fixed to a flange which rotates integrally with the second shaft, and which are provided on the same circumference of the flange. Inner mesh planetary gear mechanism. The internal gear planetary gear mechanism according to claim 13 or claim 14,
A diameter φDPDC-hole of a pitch circle forming the plurality of inner pin holes,
The diameter of the pitch circle forming the plurality of inner pins, φDPDC-pin, is the same as that of the inner circle planetary gear mechanism. The internal gear planetary gear mechanism according to claim 14,
The plurality of inner pin holes are provided at an equal angular pitch with respect to a pitch circle forming the plurality of inner pin holes,
The internally meshing planetary gear mechanism, wherein the plurality of inner pins are provided at an equal angular pitch with respect to a pitch circle forming the plurality of inner pins. 17. The inner meshing planetary gear mechanism according to claim 13, wherein a diameter φDhole of the inner pin hole is provided to be equal to or larger than an eccentric amount α of the eccentric portion + a diameter φDpin of the inner pin. Features an internally meshing planetary gear mechanism. 18. The internal meshing planetary gear mechanism according to any one of claims 13 to 17, wherein the inner pin has a relief portion on its inner diameter side and outer diameter side to avoid contact with the inner pin hole. Features an internally meshing planetary gear mechanism. 18. The inner meshing planetary gear mechanism according to claim 13, wherein the inner pin hole includes a relief portion on an inner diameter side and an outer diameter side to avoid contact with the inner pin. Inscribed planetary gear mechanism. The internal meshing planetary gear mechanism according to any one of claims 1 to 19,
The second shaft is rotatably supported by a housing via a sliding bearing, and one end of the first shaft is rotatably supported by the second shaft via a first bearing.
An internal meshing planetary gear mechanism, wherein an axial support section x of the slide bearing and an axial support section y of the first shaft are provided so as to overlap with each other.

Images