Internal meshing gear system

Abstract

To provide a gear device capable of realizing high efficiency and low wear of a gear.SOLUTION: An inscription engagement type gear device comprises an inner gear, and an outer gear engaged with the internal gear. When a linear line passing through an outer gear center O1 of the outer gear and an inner gear center O2 of the internal gear is designated as a gear center line L1, seen from an axial direction of the internal gear, a pitch point P of the outer gear and the inner gear is positioned on a radial inner side than an outer gear circle C22 of the outer gear and an inner gear circle C24 of the inner gear. An intersection point K between a normal line L3 of a contact point trajectory Lc of the outer gear and the inner gear and the gear center line is positioned between the pitch point P and the outer gear center O1. The contact point trajectory Lc is constituted by one or both of a curve and a linear line positioned in an engagement region R1 between the outer gear circle C22 and the inner gear circle 24C, and is positioned in an angle range, in which an angle θ to the gear center line L1 with the pitch point P as a center is 0-π / 2[rad].SELECTED DRAWING: Figure 3

Description

[Technical Field]

[0001] This disclosure relates to an internal meshing type gear system. [Background technology]

[0002] Patent Document 1 discloses an internal gear system comprising an internal gear and an external gear that meshes with the internal gear. [Prior art documents] [Patent Documents]

[0003] [Patent Document 1] Japanese Patent Publication No. 2018-155263 [Overview of the project] [Problems that the invention aims to solve]

[0004] In conventional gear design methods, the tooth profile curve of one gear in a gear pair (external gear and internal gear) is determined first, and then the tooth profile curve of the other gear is set based on that first tooth profile curve. The inventors of this invention have come up with a new idea to improve the efficiency of gear systems and reduce gear wear by taking an approach different from this conventional tooth profile design method.

[0005] One of the purposes of this disclosure is to provide an internal meshing gear device that can achieve high efficiency and low gear wear. [Means for solving the problem]

[0006] The gear apparatus of the present disclosure is an internal gear apparatus comprising an internal gear and an external gear that meshes with the internal gear, wherein, when viewed from the axial direction of the internal gear, a straight line passing through the external gear center O1 of the external gear and the internal gear center O2 of the internal gear is defined as the gear centerline, the pitch point P of the external gear and the internal gear is located radially inward from the external tooth tip circle of the external gear and the internal tooth tip circle of the internal gear, the intersection point K of the normal of the contact point trajectory of the external gear and the internal gear and the gear centerline is located between the pitch point P and the external gear center O1, the contact point trajectory is composed of one or both of a curve and a straight line located in the meshing region between the external tooth tip circle and the internal tooth tip circle, and is located within an angular range of 0 to π / 2 [rad] with respect to the gear centerline centered on the pitch point P. [Effects of the Invention]

[0007] According to this disclosure, it is possible to improve the efficiency of gear systems and reduce gear wear. [Brief explanation of the drawing]

[0008] [Figure 1] This is a side cross-sectional view of the gear mechanism of the embodiment. [Figure 2] This is a front cross-sectional view of the gear apparatus according to the embodiment. [Figure 3] This is an explanatory diagram of the contact point trajectory. [Figure 4] This is an explanatory diagram of the meshing equation. [Figure 5] Figure 5(A) is an explanatory diagram of overtaking meshing, and Figure 5(B) is an explanatory diagram of passing meshing. [Figure 6] This diagram shows the position of intersection point K to achieve overtaking engagement. [Figure 7] This diagram shows the movement trajectory of the meshing portion of the external gear and the internal gear. [Modes for carrying out the invention]

[0009] The embodiments will be described below. The same components are denoted by the same reference numerals, and redundant descriptions will be omitted. In each drawing, for the sake of convenience of explanation, components are appropriately omitted, enlarged, or reduced. The drawings are to be viewed in accordance with the orientation of the reference numerals.

[0010] An explanation will be given starting from the background that led to the conception of the gear device of the present embodiment. As described above, in a general gear design method, after first determining the tooth profile curve of one of the gears in a gear pair, the tooth profile curve of the other gear is set based on that tooth profile curve. However, gear characteristics such as frictional loss and wear in gears are greatly affected by the shape of the tooth profile curve. Therefore, when using such a general design method, due to the fact that the tooth profile curve of one of the gears in the gear pair has been determined first, the design freedom with respect to these gear characteristics becomes low.

[0011] Therefore, the inventor of the present application, when designing a gear pair, instead of first determining the gear curve of one of the gears, as conditions for defining the gear curves of the two gears, an approach is taken of first determining the locus of the contact points of the two gears (hereinafter referred to as the contact point locus) and the pitch point P. As a result, compared with the general design method of first determining the tooth profile curve of one of the gears, the gear characteristics can be set with a high degree of design freedom without being restricted by the shape of the gear curve of the gear. In addition, in the process of proceeding with the design of the gear pair by such an approach, the inventor of the present application newly found conditions suitable for achieving high efficiency of the gear device and low wear of the gears as conditions such as the contact point locus.

[0012] Before describing the characteristics of the gear pair (external gear and internal gear), we will first describe the overview of the gear device in which the gear pair is used. Refer to Figures 1 and 2. The gear device 10 comprises an input shaft 12 to which input rotation is input from a drive source, a gear mechanism 14 that transmits the rotation of the input shaft 12, and an output member 16 that outputs the output rotation transmitted from the gear mechanism 14 to a driven device. The specific example of the drive source is not particularly limited and may be a motor, gear motor, engine, etc. The specific example of the driven device is not particularly limited and may be a part of a driven machine such as a conveyor, wheel, machine tool, or robot (industrial robot, service robot, etc.). In addition, the gear device 10 of this embodiment comprises a casing 18 that houses the gear mechanism 14 and carriers 20A and 20B arranged axially on the sides of the external gear 22 that constitutes the gear mechanism 14.

[0013] The gear device 10 of this embodiment is an eccentric oscillating gear device comprising an external gear 22 and an internal gear 24 as the gear mechanism 14. This type of gear device 10 can transmit output rotation to the output member 16 by oscillating one of the external gear 22 and the internal gear 24 (in this case, the external gear 22) by an eccentric body 30 of the crankshaft 26. In this embodiment, an example in which the carrier 20A is the output member 16 is described, but the casing 18 may also be the output member 16.

[0014] In this embodiment, the input shaft 12 is a crankshaft 26. The crankshaft 26 comprises a shaft body 28 and at least one (two in this case) eccentric body 30 that can rotate integrally with the shaft body 28. The eccentric body 30 is eccentric with respect to the rotation center C26 of the crankshaft 26. The shaft body 28 and the eccentric body 30 may be separate components or may be provided as part of the same component.

[0015] The external gear 22 is supported on the input shaft 12 via a bearing 32 so as to be rotatable relative to it. The external gear 22 comprises a plurality of external teeth 22a provided on the outer circumference of the external gear 22 and a through hole 22b through which the input shaft 12 passes. The internal gear 24 comprises an internal gear body 24a integrated with a casing 18 that houses the external gear 22, and a plurality of internal teeth 24b provided on the inner circumference of the internal gear body 24a. In this embodiment, the internal teeth 24b are formed directly on the inner surface of the internal gear body 24a.

[0016] In this embodiment, the carriers 20A and 20B are individually positioned on both axial sides of the external gear 22. The carriers 20A and 20B include a first carrier 20A positioned on the non-input side and a second carrier 20B positioned on the input side. In this embodiment, the first carrier 20A is formed by integrating a first carrier member 20a positioned on the input side and a second carrier member 20b positioned on the non-input side. The carriers 20A and 20B are synchronized with the rotational component of the external gear 22 when the external gear 22 oscillates, by an internal pin 34 that penetrates the external gear 22.

[0017] The bearing 32 supporting the external gear 22 is equipped with multiple rolling elements 32a. In this embodiment, the bearing 32 does not have a dedicated outer ring; the inner circumferential surface of the through hole 22b of the external gear 22 serves as the outer ring, and the outer rolling surface is provided on this inner circumferential surface. In this embodiment, the bearing 32 does not have a dedicated inner ring; the outer circumferential surface of the input shaft 12 (in this case, the outer circumferential surface of the eccentric body 30) serves as the inner ring, and the inner rolling surface is provided on this outer circumferential surface. Alternatively, the bearing 32 may be equipped with a dedicated outer ring and inner ring, and the outer rolling surface and inner rolling surface may be provided on them.

[0018] The operation of the gear system described above will now be explained. When the input shaft 12 (in this case, the crankshaft 26) rotates, the external gear 22 oscillates due to the eccentric body 30 of the crankshaft 26. As the external gear 22 oscillates, the meshing position of the external gear 22 and the internal gear 24 changes sequentially in the circumferential direction. As a result, with each rotation of the crankshaft 26, either the external gear 22 or the internal gear 24 (in this case, the external gear 22) rotates by the difference in the number of teeth between the two. This rotational component is transmitted to the output member (in this case, the carrier 20A via the internal pin 34) and then output to the driven member as output rotation.

[0019] As described above, the gear device 10 of this embodiment is an internal meshing type gear device comprising an internal gear 24 and an external gear 22 that meshes with the internal gear 24 as a gear pair. We will now move on to the details of the features of this gear pair.

[0020] Refer to Figure 3. Hereinafter, the radial direction with the internal gear center O2 of the internal gear 24 as the center of the circle will simply be referred to as the "radial direction". Also, when viewed from the axial direction of the internal gear 24, the line passing through the external gear center O1 of the external gear 22 and the internal gear center O2 of the internal gear 24 will be called the gear centerline L1. Furthermore, the Cartesian coordinate system fixed to the gear centerline L1 will be called the reference coordinate system. The origin of the reference coordinate system will be the gear center of one of the gears, and one of its coordinate axes (y-axis) will be the gear centerline L1. In Figure 3, the reference coordinate system will remain stationary even as the meshing of the external gear 22 and the internal gear 24 progresses.

[0021] The tip circle of the external gear 22 is called the external tooth tip circle C22, and the tip circle of the internal gear 24 is called the internal tooth tip circle C24. The external tooth tip circle C22 is the circle formed by connecting the tips of multiple external teeth 22a on the external gear 22, and the internal tooth tip circle C24 is the circle formed by connecting the tips of multiple internal teeth 24b on the internal gear 24. In this embodiment, both the external tooth tip circle C22 and the internal tooth tip circle C24 are circular in shape.

[0022] The region enclosed by the external tooth tip circle C22 and the internal tooth tip circle C24 is called the meshing region R1. The meshing region R1 is formed radially inside the external tooth tip circle C22 and radially outside the internal tooth tip circle C24. The meshing of the external gear 22 and the internal gear 24 takes place within this meshing region R1. The contact point C is the point where the tooth profile curves of the external gear 22 and the internal gear 24, which mesh with each other, come into contact. In the reference coordinate system, the contact point C moves in direction D1 (clockwise in this case) as the meshing of the external gear 22 and the internal gear 24 progresses. In the reference coordinate system, the trajectory traced by the contact point C as it moves with the progression of meshing is called the contact point trajectory Lc. The range within the meshing region R1 where the contact point trajectory Lc is located is the actual meshing range.

[0023] The pitch point P of the external gear 22 and the internal gear 24 is the intersection of the common normal L2 of the two tooth profile curves and the gear centerline L1 at the contact point C. The pitch point P is the instantaneous center of the relative motion between the external teeth of the external gear 22 and the internal teeth of the internal gear 24 that mesh with each other. In the case of an eccentric oscillating gear system and a simple planetary gear system, the pitch point P is a fixed point in the reference coordinate system that does not move during the process of the meshing of the external gear 22 and the internal gear 24.

[0024] Let K be the intersection point of the normal L3 (hereinafter referred to as the trajectory normal L3) of the contact point trajectory Lc at contact point C and the gear center line L1, and let k be the length from intersection K to pitch point P. The trajectory normal L3 is a straight line perpendicular to the tangent to the contact point trajectory Lc at contact point C. The sign of k is positive when intersection K is on the opposite side of the internal gear center O2 from the pitch point P, as shown in Figure 3, and negative when intersection K is on the side of the internal gear center O2 from the pitch point P.

[0025] In the reference coordinate system, the angle θ is the angle between the pitch point P and the gear centerline L1. The angle θ is considered positive when the contact point C moves away from the gear centerline L1 in the direction of movement D1.

[0026] In this embodiment, the gear assembly 10 is provided that the external gear 22 and internal gear 24 satisfy the following three conditions. The first condition is that the pitch point P is located radially inward from the respective tooth tips C22 and C24. The second condition is that the intersection point K is located between the pitch point P and the center O1 of the external gear. This second condition must be satisfied over the entire range of the process in which the contact point C moves within the meshing region R1. The third condition is that the contact point trajectory Lc is located within the angular range in which the angle θ is 0 to π / 2 [rad] within the meshing region R1. The third condition can also be said to be that the contact point trajectory Lc does not extend outside the angular range in which the angle θ is 0 to π / 2 (the angular range in which θ is less than 0 and greater than π / 2) within the meshing region R1. Figure 3 shows an example where the contact point trajectory Lc is within the angular range from position C0 where angle θ is 0 to position C1 where angle θ is π / 2 (i.e., the entire angular range where angle θ is from 0 to π / 2). The reasons for setting these conditions are explained below.

[0027] Refer to Figure 4. Hereafter, subscript 1 of each symbol indicates information relating to the external gear 22, and subscript 2 indicates information relating to the internal gear 24. The gear centers O of the external gear 22 and internal gear 24. i We will examine the representation of a contact point C on the contact point trajectory Lc using tangent coordinates with poles (i=1,2). These tangent coordinates are the gear center O i Half-line O from point O to point C i The length of C is the radial ρ. i And, the half-line O i The tangent angle σ is the angle of the tangent line to the tooth profile curve at the point of contact C (which is also the common tangent line L4 between the two tooth profile curves). i This consists of the above. In other words, the contact point C can be defined as (ρ1, σ1) when using tangential coordinates with the gear center O1 of the external gear 22 as the pole, and as (ρ2, σ2) when using tangential coordinates with the gear center O2 of the internal gear 24 as the pole.

[0028] Next, we consider the representation of the contact point C on the contact point trajectory Lc using polar coordinates with the pitch point P as the pole. These polar coordinates consist of the radial r, which is the length of the ray PC from the pitch point P to the contact point C, and the deflection θ, which is the angle of the ray PC with respect to the gear centerline L1. In other words, the contact point C can be defined as (r,θ) when using polar coordinates with the pitch point P as the pole. This ray PC is also the common normal L2 of the two tooth profile curves at the contact point C. The angle θ is the same as the one used in the third condition mentioned above. The angle θ can also be said to be the angle made by the common normal L2 of the two tooth profile curves at the contact point C with respect to the gear centerline L1 centered at the pitch point P.

[0029] It is known that the meshing equation for an internally meshing gear pair can be described by the following equation (1). This expresses the tooth profile curve of the gear pair in tangential coordinates (ρ,σ) on the left side and the contact point trajectory Lc in polar coordinates (θ) on the right side. R is the pitch circle radius of each gear 22 and 24.

[0030]

number

[0031] The relative radius of curvature ρ between infinitesimal line elements on the tooth profile curve of the gear pair at the contact point C. rc Consider the radius of curvature ρ of a small line element on the tooth profile curve of the external gear 22. c1 This can be expressed by the following equation (2).

[0032]

number

[0033] Here, the length k in equation (2) can be expressed by the following equation (3). In equation (3), dr and d(r·cosθ) are the infinitesimal changes in r and (r·cosθ) when the tooth profile curve is displaced by an infinitesimal linear element from point C to point C'.

[0034]

number

[0035] Similarly, the radius of curvature ρ of the infinitesimal element on the tooth profile curve of the internal gear 24 c2 can be expressed by the following formula (4).

[0036]

Equation

[0037] Relative radius of curvature ρ rc can be obtained from formulas (2) and (4) by the following formula (5).

[0038]

Equation

[0039] As this relative radius of curvature ρ rc increases, it is known that the contact stress at the contact point C between the external gear 22 and the internal gear 24 can be reduced and the lubricating oil film thickness can be increased. This contributes to improving efficiency and reducing wear due to reduced frictional losses in a gear pair with the same load capacity because the oil film thickness can be increased without reducing the load capacity of the gear pair. Note that the radius of curvature ρ of the external gear 22 c1 is positive when the tooth profile curve of the external gear 22 is a convex surface and negative when it is a concave surface. Also, the radius of curvature ρ of the internal gear 24 c2 is positive when the tooth profile surface of the internal gear 24 is a concave surface and negative when it is a convex surface.

[0040] Next, the first type of interference is examined using the relative radius of curvature ρ rc The first type of interference means that as the meshing of the external gear 22 and the internal gear 24 progresses, the substantial sides (solid parts) of the external gear 22 and the internal gear 24 overlap. It is known that this first type of interference occurs in the case of convex-concave meshing described later when, as the meshing progresses, the radius of curvature of one tooth profile curve approaches the radius of curvature of the other tooth profile curve and then the magnitudes of their respective radii of curvature are reversed. The relative radius of curvature ρ in formula (5) rcIn relation to this, as the meshing progresses, the relative radius of curvature ρ rc First-kind interference occurs when the sign of changes. From equation (5), in order to avoid first-kind interference, it is necessary that "(k+R1)·(k+R2)" and "cosθ" always have the same sign or opposite signs with respect to the change in θ.

[0041] In this embodiment, it is assumed that the intersection point K is located between the pitch point P and the center O1 of the external gear (second condition). This means that (k+R1)·(k+R2) in equation (5) is always positive. Furthermore, in this embodiment, it is assumed that the contact point trajectory Lc is located within the angular range where the angle θ is 0 to π / 2 (third condition). This means that cosθ in equation (5) is always positive. Therefore, according to this embodiment, meshing between the external gear 22 and the internal gear 24 can be achieved without causing first-class interference.

[0042] Next, we examine the slip ratio Σ at the contact point C between the external gear 22 and the internal gear 24. When ds1 and ds2 are infinitesimal line elements on the tooth profile curves of the external gear 22 and the internal gear 24 at the contact point C, it is known that the infinitesimal line elements ds1 and ds2 can be expressed by the following equations (6) and (7).

[0043]

number

[0044]

number

[0045] The slip ratio Σ1 of the external gear 22 can be expressed by the following equation (8).

[0046]

number

[0047] From equation (8), we can use equations (1), (3), (6), and (7) to derive the following equation (9). From this equation (9), we can find the slip ratio Σ1 of the external gear 22.

[0048]

number

[0049] Similarly, the slip ratio Σ2 of the internal gear 24 can be expressed by the following equation (10).

[0050]

number

[0051] Refer to Figure 5. Two types of gear meshing configurations are known: overtaking meshing and passing meshing. Overtaking meshing refers to a meshing configuration where the movement directions D22 and D24 of each gear 22 and 24 at the contact point C of the two gears are in the same direction, while passing meshing refers to a meshing configuration where the movement directions D22 and D24 are in opposite directions. It is known that overtaking meshing and passing meshing can be expressed as follows using the product of the slip ratio Σ1 of the external gear 22 and the slip ratio Σ2 of the internal gear 24. Σ1×Σ2<0: Overtaking meshing Σ1×Σ2>0: Passing meshing

[0052] The position of intersection K that satisfies the conditions for this overtaking mesh is given by equations (9) and (10) as (A1)k>0, (A2)k<0 and k 2 <R1 2 , (A3)k<0 and R2 2 <k 2 It can be seen that it is one of the following. If the position of intersection K does not satisfy any of these conditions, the meshing configuration of the external gear 22 and the internal gear 24 will be a passing mesh. The position of intersection K that satisfies this condition is as shown in Figure 6. Note that in the case of an internally meshing gear pair, the relationship between the pitch circle radii of the external gear 22 and the internal gear 24 is R1 <R2である。

[0053] It is known that overtaking meshing is more advantageous than passing meshing in order to reduce wear by forming an oil film at the contact point C of two gears. In this embodiment, the condition is that the intersection point K is located between the pitch point P and the center O1 of the external gear (second condition). In relation to the conditions for overtaking meshing, this means that condition (A2) is satisfied. As a result, wear can be reduced by realizing overtaking meshing.

[0054] As in the first condition, when the pitch point P is inside the tip circles C22 and C24 of each tooth, geometrically, it is easier to lengthen the contact point trajectory in the meshing region R1 when the position of the intersection point K is k<0 than when the position of the intersection point K is k>0. In other words, by adopting the condition (A2) where the position of the intersection point K is k<0, it is easier to lengthen the contact point trajectory in the meshing region R1 compared to the condition (A1) where the position of the intersection point K is k>0. This means that the number of meshing teeth that can mesh simultaneously in the meshing region R1 can be increased, which contributes to a reduction in the load applied to each tooth, i.e., reduced wear. Also, by adopting the condition (A2), the k value can be reduced compared to the condition (A3). This can be seen from equation (5), where the relative radius of curvature ρ rc This means that the contact stress can be increased, contributing to a reduction in contact stress and an increase in the lubricating oil film thickness. In particular, an increase in the lubricating oil film thickness contributes to higher efficiency and lower wear.

[0055] Next, we will examine the conditions for realizing concave-concave meshing between the external gear 22 and the internal gear 24. It is known that when the external gear 22 and the internal gear 24 mesh concave-concave, contact stress can be reduced by increasing the relative radius of curvature compared to convex-concave meshing. Here, concave-concave meshing refers to the meshing of a convex surface and a concave surface, while convex-concave meshing refers to the meshing of two convex surfaces. In order to realize concave-concave meshing, it is necessary that the radii of curvature of the two tooth profile curves at the contact point C have the same sign. Here, the same sign means that the radii of curvature of both tooth profile curves are positive (when the external gear 22 is a convex surface and the internal gear 24 is a concave surface) or negative (when the external gear 22 is a concave surface and the internal gear 24 is a convex surface).

[0056] The conditions for achieving this interlocking can be determined using equations (2) and (4). The case where the radius of curvature of both tooth profile curves is positive is when the left side of equation (2) is ρ c1 , and the left side ρ of equation (4) c2 This means that the value is positive. In this embodiment, the second condition (that the intersection point K is located between PO1) is satisfied, so both "k+R1" in the numerator of the right-hand side of equation (2) and "k+R2" in the numerator of the right-hand side of equation (4) are positive. The left-hand side of equation (2) is ρ c1 If the result is positive, and the numerator of the right-hand side, "k+R1", is positive, then the denominator of the right-hand side must be positive. Similarly, the left-hand side of equation (4), ρ c2 If the radius of curvature of both tooth profile curves is positive, then the denominator of the right-hand side must be positive. In other words, if the radius of curvature of both tooth profile curves is positive, then a concave-concave meshing can be achieved when the denominator of the right-hand side of both equations (2) and (4) is positive. The same applies when the radius of curvature of both tooth profile curves is negative. The case where the radius of curvature of both tooth profile curves is negative is when the left-hand side ρ of equation (2) is negative. c1 ρ on the left side of equation (4) c2 This means that the value is negative. Therefore, following the same reasoning as above, when the radius of curvature of both tooth profile curves is negative, it means that a concave-concave meshing can be achieved when the denominator of the right-hand side of both equations (2) and (4) is negative.

[0057] In this embodiment, the condition is that the intersection point K is located between the pitch point P and the center O1 of the external gear (second condition). This means that the conditions k+R1>0 and k+R2>0 are satisfied. This also means that k is always negative, as shown in Figure 3.

[0058] If the radius of curvature of both tooth profile curves is positive, when the angle θ is between π / 2 and π, the denominator on the right side of equations (2) and (4) will take on a negative sign depending on the position of the intersection point K, which may make it impossible to achieve concave-convex meshing. Also, if the radius of curvature of both tooth profile curves is negative, when the angle θ is between 0 and π / 2, the denominator on the right side of equations (2) and (4) will take on a positive sign depending on the position of the intersection point K, which may make it impossible to achieve concave-convex meshing.

[0059] In contrast, if the radii of curvature of both tooth profile curves are positive, then if the angle θ is in the range of 0 to π / 2 (i.e., if the third condition is met), the sign of the denominator in equations (2)(4) will be positive regardless of the magnitude of k. In other words, in this case, if the position of the intersection K satisfies the second condition, then a concave-concave meshing can be achieved regardless of the position of the intersection K.

[0060] To summarize: By satisfying the second condition (the intersection point K is located between PO1) and the third condition (the contact point trajectory Lc is within the angular range where θ is 0 to π / 2), meshing between the external gear 22 and the internal gear 24 can be achieved without causing type I interference. Furthermore, by satisfying the second condition, gear wear can be reduced by achieving overtaking meshing. At this time, by also satisfying the first condition (the pitch point P is inside each tooth tip circle C22, C24), contact stress can be reduced and wear reduced by lengthening the contact point trajectory. Furthermore, by satisfying the second and third conditions, gear wear can be reduced and efficiency can be increased by achieving uneven meshing, regardless of the position of the intersection point K. Due to these features, friction loss in the gears can be reduced in conjunction with gear wear, and the efficiency of the gear device 10 can be increased.

[0061] Let's consider the case where the second condition is not met. For example, if the intersection point K is on the opposite side of the internal gear center O2 from the pitch point P, as mentioned above, the contact point trajectory Lc in the meshing region R1 becomes shorter, leading to increased contact stress and wear. Also, if the intersection point K is between the external gear center O1 and the internal gear center O2, it results in swerving meshing, making it impossible to achieve low gear wear. Furthermore, if the intersection point K is on the opposite side of the pitch point P from the internal gear center O2, the increase in the k value leads to a decrease in the relative radius of curvature and an increase in the slip ratio, resulting in increased contact stress, wear, and friction loss.

[0062] Let's consider the case where the third condition is not met. If the contact point trajectory Lc extends beyond the angular range where angle θ is less than 0, or beyond the angular range where angle θ is greater than π / 2, then from equation (5), the sign of cosθ changes with respect to the change in θ, resulting in first-kind interference or misalignment.

[0063] The contact point trajectory Lc described above is composed of one or both of a curve and a straight line located in the cleavage region R1. The contact point trajectory Lc may be composed of only one of a curve and a straight line, or it may be composed of a combination of a curve and a straight line. When the contact point trajectory Lc is composed of a combination of a curve and a straight line, the number of at least one of the curves and straight lines that constitute the contact point trajectory Lc may be multiple. When the contact point trajectory Lc includes a curve, it may include multiple curves with different curvatures. The terms "curve" and "straight line" here are not limited to geometrically precise curves and straight lines, but may include approximate curves and approximate straight lines. It is preferable that the contact point trajectory Lc be composed of a curve with its center of curvature inside the internal tooth tip circle C24, as in this embodiment. This makes it possible to lengthen the cleavage range in the cleavage region R1 compared to when the contact point trajectory Lc is a straight line, or when the contact point trajectory Lc is composed of a curve with its center of curvature outside the external tooth tip circle C22. As mentioned above, this means that the number of meshing teeth can be increased, contributing to a reduction in contact stress and wear. In particular, when using an internally meshing external gear 22 and internal gear 24 with an involute tooth profile, the contact point trajectory becomes a straight line, which is advantageous because it allows for a larger number of meshing teeth compared to an involute tooth profile.

[0064] Refer to Figure 7. Figure 7 shows the movement trajectory of the external gear 22 and internal gear 24 during operation at the meshing portion in the reference coordinate system. At the meshing portion of the external gear 22 and internal gear 24, the tooth surface (tooth profile curve) of the external gear 22 is a convex surface, and the tooth surface (tooth profile curve) of the internal gear 24 is a concave surface. This makes it possible to achieve the aforementioned concave-convex meshing and reduce gear wear.

[0065] The definition of the external gear center O1 is explained below. In a cross-section perpendicular to the axial direction of the internal gear 24, if there is no bearing inside the external gear 22 and there is a rotating shaft that rotates integrally with the external gear 22, the rotation center of that rotating shaft is defined as the external gear center O1. This assumes, for example, that the planetary gears in a simple planetary gear mechanism are the external gears 22. Also, in a cross-section perpendicular to the axial direction of the internal gear 24, if there is only a single bearing 32 inside the external gear 22, the curvature center of the outer shape of that bearing 32 is defined as the external gear center O1. This assumes, for example, that the gear device 10 is a center-crank type eccentric oscillating gear device as in this embodiment, as well as a flexible meshing type gear device, etc. Also, in a cross-section perpendicular to the axial direction of the internal gear 24, if there are multiple bearings inside the external gear 22, the center of the circle passing through the curvature centers of the outer shapes of each of the multiple bearings is defined as the external gear center O1. This assumes, for example, that the gear mechanism 10 is a distribution-type eccentric oscillating gear mechanism.

[0066] Here, "center of curvature of the bearing's outer shape" refers to the center of curvature of the outer shape formed by the inner rolling surfaces of the bearing. These inner rolling surfaces are the surfaces on which the rolling elements, located radially outward from the rolling elements, roll. If the bearing has a dedicated outer ring, these inner rolling surfaces are located on the inner circumferential surface of that dedicated outer ring. Alternatively, if the bearing does not have a dedicated outer ring and the inner circumferential surface of an external gear serves as the outer ring, these inner rolling surfaces are located on the inner circumferential surface of the external gear.

[0067] This section describes the handling of a flexible meshing type gear system. This flexible meshing type gear system comprises a vibrator and a flexible gear (flex spline) that deforms into an elliptical shape by the vibrator. Here, we will describe the case where the flexible gear is an external gear 22. In the case of a flexible meshing type gear system, the external gear center O1 of the external gear 22 is known to be the instantaneous center of relative motion between one of the external teeth of the external gear 22 and the vibrator at the moment when the external gear 22 and the internal gear 24 are meshing. Furthermore, in this case, it is known that there is a different external gear center O1 for each external tooth of interest (for each external tooth meshing with the internal gear 24). This external gear center O1 coincides with the center of curvature of the curve at the intersection of the tooth center line of the external tooth meshing with the internal gear 24 and the curve formed by the inner rolling surface of the bearing. Assuming that the center O1 of the external gear does not overlap with the center O2 of the internal gear, it lies within a 90° angular range from the major axis to the minor axis of the elliptical shape formed by the external gear 22.

[0068] In the case of a flexible meshing gear system, it is known that, similar to the external gear center O1, the reference coordinate system determined by the gear centerline L1 passing through the external gear center O1 and the internal gear center O2 also differs for each external tooth of interest. In this case, similar to the external gear center O1, it is known that the pitch point P and intersection point K also differ for each external tooth of interest. In this case, the intersection point of the common normal vector L2 of the tooth profile curve at the contact point C between the external and internal teeth of interest and the gear centerline L1 corresponding to that external tooth becomes the pitch point P corresponding to that external tooth. Also, in this case, the intersection point of the trajectory normal vector L3 at the contact point C between the external and internal teeth of interest and the gear centerline L1 corresponding to that external tooth becomes the intersection point K corresponding to that external tooth. In this case, the external gear center O1 and pitch point P corresponding to the common external tooth move slightly relative to the internal gear center O2 during the process of meshing, but they remain on the gear centerline L1 corresponding to that external tooth. The center O2 of the internal gear is a fixed point on the internal gear 24, regardless of the external teeth being considered.

[0069] Thus, in the case of a flexible meshing gear system, each external tooth has a different reference coordinate system. Therefore, in this case, the aforementioned first to third conditions concerning the pitch point P and intersection point K derived from the contact point C between an external tooth and an internal tooth only need to be satisfied in relation to the reference coordinate system corresponding to that external tooth.

[0070] The cross-sectional shape of the vibrator used in a flexible meshing gear system is elliptical, perpendicular to the axial direction. This elliptical shape is not limited to a geometrically strictly elliptical shape, but also includes a nearly elliptical shape. For example, the cross-sectional shape of the vibrator may be nearly elliptical, and an arc-shaped section with a single radius of curvature r1 may be provided along the major axis of the cross-sectional shape, forming an arc shape over the entire area where the external gear 22 and the internal gear 24 mesh. Here, the major axis refers to a straight line along the position where the distance from the rotation center of the vibrator to the circumferential surface of the vibrator is longest in the cross-section perpendicular to the axial direction. In this case, the instantaneous center of relative motion (external gear center O1) between one external tooth of the external gear 22 meshing with the internal gear 24 (the external tooth that meshes due to the arc-shaped section of the vibrator) and the arc-shaped section of the vibrator is a fixed point in the reference coordinate system corresponding to that external tooth.

[0071] Furthermore, in the case of flexible meshing gear systems, the tooth traces of the external gear 22 are sometimes inclined with respect to the axial direction such that they are radially offset in the axial direction. In this case, it is sufficient that the first to third conditions mentioned above are met in a cross section perpendicular to the axial direction that passes through any axial position where the external gear and the internal gear mesh.

[0072] Next, we will describe the transformation forms of each component described so far.

[0073] The type of gear system is not particularly limited. The gear system 10 may be, for example, an eccentric oscillating gear system, a simple planetary gear system, or a flexible meshing gear system. In the case of an eccentric oscillating gear system, the specific type is not particularly limited. This type may be a center crank type in which the crankshaft 26 is positioned on the center of the internal gear 24, as in this embodiment, or a distribution type in which multiple crankshafts are positioned radially offset from the axis of the internal gear 24. Also, in the case of an eccentric oscillating gear system, the internal gear may oscillate in addition to the external gear. In the case of a flexible meshing gear system, the specific type is not particularly limited. The type of flexible meshing gear system may be, for example, a cylindrical type with two internal gears, or a cup type or a top hat type with one internal gear. The gear system may also function as a reduction gear, a speed increaser, or a power transmission (power distribution) device.

[0074] When the gear device 10 is a flexible meshing type gear device, one of the external gear 22 and the internal gear 24 becomes the aforementioned flexible gear that deforms by bending as the meshing progresses, while the other becomes a rigid gear that does not deform by bending. In this case, the tip circle of the gear that becomes the flexible gear (one of the external gear tip circle C22 and the internal gear tip circle C24) may be elliptical in shape.

[0075] The specific examples of the contact point trajectory Lc are not particularly limited. Furthermore, the contact point trajectory Lc may be composed of a curve whose center of curvature is outside the external tooth tip circle C22.

[0076] The embodiments and variations described above are illustrative. The abstract technical ideas derived from them should not be interpreted restrictively to the content of the embodiments and variations. Many design changes, such as changes, additions, and deletions of components, are possible in the embodiments and variations. In the embodiments described above, the content that allows for such design changes is emphasized with the notation "embodiment." However, design changes are also permitted in content without such notation. The hatching applied to the cross-sections in the drawings does not limit the material to which the hatching is applied. The structures / numerical values ​​mentioned in the embodiments and variations naturally include those that can be considered identical when considering manufacturing tolerances, etc. [Explanation of Symbols]

[0077] L1...Gear centerline, O1...External gear center, O2...Internal gear center, 10...Gear assembly, 22...External gear, C22...External gear tip circle, 24...Internal gear, C24...Internal gear tip circle, 30...Eccentric body.

Claims

1. An internal gear mechanism comprising an internal gear and an external gear that meshes with the internal gear, Viewed from the axial direction of the internal gear, the external gear center O of the external gear １ and the center of the internal gear O ２ When the straight line passing through is defined as the gear centerline, The pitch point P of the external gear and the internal gear is located radially inward from the tip circle of the external gear and the tip circle of the internal gear. The intersection point K of the normal to the contact point trajectory of the external gear and the internal gear and the gear center line is the pitch point P and the center O of the external gear. １ Located between, The contact point trajectory is composed of one or both of a curve and a straight line located in the meshing region between the external tooth tip circle and the internal tooth tip circle, and is located within an angular range of 0 to π / 2 [rad] with respect to the gear center line centered on the pitch point P.

2. The internal meshing gear device according to claim 1, which has an eccentric oscillating body for oscillating the external gear.

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