JP5746665B2 - Gear device - Google Patents

Description

  The present invention relates to a gear device including three or more gears.

  Conventionally, it has been considered to use a planetary gear device for a transmission used in an automobile or the like provided with an engine and / or a motor as a drive source of the vehicle. The planetary gear device includes a sun gear, a plurality of pinion gears meshing around the sun gear, and a ring gear arranged coaxially with the sun gear. The plurality of pinion gears are supported by a carrier so as to be able to rotate and revolve around a central axis of the sun gear. In the case of a simple single planetary type planetary gear device, the plurality of pinion gears mesh with the inner teeth of the ring gear. In this case, the pinion gear that is a gear may be rotatably supported around the pinion shaft that is a gear shaft by, for example, a needle roller bearing (hereinafter also simply referred to as “needle bearing”).

  For example, FIG. 12 shows a cross-sectional view of a part of the conventional structure of the planetary gear device described in Patent Document 1. The planetary gear device of FIG. 12 includes two carrier plates 12 disposed on both sides of the pinion gear 10, a pinion shaft 14 that is supported and fixed to each carrier plate 12 and supports the pinion gear 10, and is press-fitted into the two carrier plates 12. And a thrust ring 16 fixed by casting. Each thrust ring 16 is made of a steel that has been subjected to a quenching process and has an outer diameter smaller than the inner diameter of the pinion gear 10 and larger than the center diameter of the needle bearing 18. In the planetary gear device, a thrust washer 20 that thrust-supports the end face of the pinion gear 10 with respect to the carrier plate 12 is press-fitted to the thrust ring 16. The plurality of pinion gears 10 mesh with the sun gear 22 and an internal gear 24 corresponding to the ring gear. In such a structure, since the thrust ring 16 receives the thrust force from the needle bearing 18, the local force from the needle bearing 18 is applied to the thrust washer 20 interposed between the pinion gear 10 and the carrier plate 12. Does not work. For this reason, it is said that the abrasion resistance of the thrust washer 20 can be improved.

  Note that there are Non-Patent Documents 1 and 2 as prior art documents related to the present invention. Non-Patent Document 1 describes the influence of friction and thrust force due to needle skew of a needle roller bearing. Non-Patent Document 1 describes a relational expression of an inclination angle with respect to an axis of a needle constituting the bearing, that is, a skew angle, a bearing contact load, and a thrust force in a needle roller bearing.

  Non-Patent Document 1 describes that the thrust angle changes in the range of 0 to 1 °, and the thrust force generated in the bearing increases linearly as the skew angle increases.

  Non-Patent Document 2 describes an efficiency map showing the gear meshing efficiency corresponding to changes in pinion pitch diameter and pitch linear velocity.

JP-A-10-103418

Nobutamune Hamada, Koichi Funabashi, “Fundamental Study of Friction and Thrust Based on Skew of Needle Roller Bearing”, Lubrication, Vol. 16, No. 3, 1971, 196-209 John J. Coy, Dennis P. Townsend and Erwin V. Zaretsky, “Gearing”, NASA Reference Publication 1152, AVSCOM Technical Report 84-C-15,1985

  Conventionally, in a gear device such as the above planetary gear device, it has been desired to reduce the thrust force applied to the gear and reduce the loss of the gear device. On the other hand, in the prior art described in Patent Document 1 described above, the thrust ring 16 receives a thrust force from the needle bearing 18 to improve the wear resistance of the thrust washer 20. However, the thrust ring 16 receives the thrust force generated by the needle skew. For this reason, it is not a fundamental solution to the problem that it is necessary to reduce the thrust force applied to the gear. For example, a loss due to friction occurs between the thrust ring 16 that receives a thrust force and the pinion gear 10 that is a gear, and as a result, the rotation transmission efficiency between the gears 22, 10, and 24 may deteriorate. Further, since the thrust ring 16 receives a thrust force, there is a possibility that wear may occur between the opposing member facing in the axial direction. For this reason, it is desired to reduce the thrust force applied to the gear. None of the non-patent documents 1 and 2 discloses a means for solving such a problem.

  An object of the present invention is to reduce a thrust force applied to a gear and reduce a loss of the gear device in the gear device.

  The gear device according to the present invention includes three or more gears, each of which is a helical gear, and an intermediate gear that is at least one of the gears meshes with the other two gears, Among the three axes parallel to or coaxial with each other, the shafts are arranged so as to be rotatable around the centers of the corresponding axes, and the twist angle of the coast side tooth surface of the intermediate gear and the twist of the drive side tooth surface The gear device is characterized in that the sizes of the corners are different from each other.

  According to the said structure, the thrust force added to a gearwheel can be reduced and the loss of a gear apparatus can be reduced. That is, the intermediate gear meshes with two other gears and receives thrust forces in opposite directions from the respective gears. Further, a first thrust force is applied to the intermediate gear due to bearing behavior or the like. On the other hand, by providing a difference in the magnitude of the twist angle between the coast-side tooth surface and the drive-side tooth surface of the intermediate gear, the first thrust force according to the difference in the twist angle due to the meshing with the counter gear The thrust force applied to the gear can be reduced by generating the second thrust force in the reverse direction and canceling the second thrust force.

Preferably, the intermediate gear is rotatably supported on a corresponding shaft by a bearing provided between the shaft and including a plurality of rolling elements, and θ ₁ is a twist angle of the coast side tooth surface. A is 1/2 of the contact width between the corresponding shaft and each rolling element in Hertz contact theory, and P _m is the average contact pressure between the corresponding shaft and each rolling element in Hertz contact theory Yes, when Eg is the rotational efficiency of the intermediate gear and k is 2.56 × 10 ⁴ kg / mm ³ , the twist angle between the coast side tooth surface and the drive side tooth surface of the intermediate gear The difference α satisfies βmin ≦ | α | ≦ βmax, βmin satisfies sinθ ₁ = Eg × sin (θ ₁ + βmin), and βmax is sinθ ₁ + 2ak × sin1 ° / Pm = Eg × sin (θ ₁ + βmax) and Eg satisfies 0.94 ≦ Eg ≦ 0.998.

  According to the above configuration, in a gear device having an intermediate gear and two gears meshed with the intermediate gear, a configuration capable of reducing or eliminating the thrust force applied to the gear under normal use conditions and reducing the loss of the gear device. Can be realized more effectively.

  Preferably, a sun gear that is the gear, a plurality of pinion gears that are intermediate gears that are arranged around the sun gear and mesh with the sun gear, and are arranged coaxially with the sun gear on the outside of the plurality of pinion gears, A ring gear that is another gear that meshes with the plurality of pinion gears, and the plurality of pinion gears are supported by the gear shaft that is the shaft supported by the first carrier and the second carrier, and are used as a planetary gear device. Is done.

  Preferably, the front sun gear and the rear sun gear, which are the rotatable gears arranged coaxially with each other and shifted in the axial direction, and a plurality of the intermediate gears arranged around the front sun gear and meshing with the front sun gear. A plurality of second pinion gears that are separate intermediate gears that are arranged around the rear sun gear and mesh with the rear sun gear; and a plurality of second pinion gears that are arranged around the plurality of second pinion gears. A ring gear that is another gear that meshes with a pinion gear, and the plurality of first pinion gears are respectively supported by a first gear shaft that is the shaft supported by a first carrier and a second carrier, and The second pinion gears mesh with the second pinion gears, and the plurality of second pinion gears are respectively connected to the first pinion gears. Is another of the shaft supported by Yaria and the second carrier is supported by the second gear shaft is used as a Ravigneaux type planetary gear device.

  According to the gear device of the present invention, the thrust force applied to the gear can be reduced and the loss of the gear device can be reduced.

It is a schematic diagram which shows the gear apparatus of 1st Embodiment of this invention. It is AA sectional drawing of FIG. It is a partial expansion perspective view of the 2nd gearwheel which comprises the apparatus of FIG. FIG. 3 is a diagram exaggeratingly showing a difference in torsion angles between tooth surfaces on both the coast side and the drive side by taking out only the second gear from FIG. 1. It is a schematic diagram for demonstrating the thrust force which acts on a 2nd gearwheel resulting from the torsion angle of a tooth | gear in an example of the conventional gear apparatus. It is a figure for taking out only the 2nd gear from Drawing 5, and explaining the thrust force generated by bearing behavior. In the gear device of the first embodiment, the minimum value βmin and the maximum value βmax for obtaining the difference in torsion angle between the coast side and the drive side of the second gear using the gears Ga and Gb having different gear rotation efficiencies It is a figure which shows the result of having calculated. It is a perspective view of the gear apparatus of 2nd Embodiment of this invention. It is a schematic sectional drawing of the apparatus of FIG. It is a block diagram of the apparatus of FIG. It is a schematic sectional drawing of the gear apparatus of 3rd Embodiment of this invention. It is sectional drawing of a part of conventional structure of a planetary gear apparatus.

[First Embodiment]
Hereinafter, embodiments according to the present invention will be described in detail with reference to the drawings. 1 to 4 and 7 show a first embodiment of the present invention. FIG. 1 is a schematic diagram showing a gear device according to a first embodiment of the present invention. FIG. 2 is a cross-sectional view taken along the line AA of FIG. FIG. 3 is a partially enlarged perspective view of a second gear constituting the apparatus of FIG. FIG. 4 is a diagram exaggerating the difference in the twist angles of the tooth surfaces on both the coast side and the drive side by taking out only the second gear from FIG. In FIG. 1, the portions surrounded by I and J respectively show the torsion angle θ ₁ of the drive side tooth surface 28 of each tooth T1 of the first gear 26 and the coast side tooth surface of each tooth T3 of the third gear 30. 32 twist angles θ ₂ (= θ1 + α). 4, the torsion angle θ ₁ of the coast side tooth surface 36 of each tooth T2 of the second gear 34 and the drive side tooth surface 38 of each tooth T2 of the second gear 34 by the portions surrounded by K and L in FIG. The twist angle θ ₂ (= θ1 + α). In FIG. 4, the tip of each tooth T2 is hatched.

  As shown in FIG. 1, the gear device includes a first gear 26, a second gear 34, and a third gear 30, which are three gears supported by three shafts S1, S2, and S3 arranged in parallel to each other. The gears 26, 34, and 30 adjacent to each other are engaged with each other. Each gear 26, 34, 30 is a helical gear, and is rotatably supported around a corresponding shaft S 1, S 2, S 3 via a needle bearing 66 (see FIGS. 9 and 11). That is, each of the gears 26, 34, and 30 is rotatably arranged around the central axis of the corresponding axis S1 (or S2 or S3) among the three axes S1, S2, and S3 parallel to each other. That is, the gear device is a helical gear device having three parallel axes. Each gear 26, 34, 30 is provided by a needle bearing 66 including a needle as a plurality of rolling elements provided between the corresponding shaft S 1 (or S 2 or S 3) and the corresponding shaft S 1 (or S 2 or S 3). It is rotatably supported.

As shown in FIG. 1, a second gear 34 that is an intermediate gear is disposed between the first gear 26 and the third gear 30, the first gear 26 and the second gear 34 mesh, and the second gear 34 and the second gear 34. Three gears 30 are engaged. The first gear 26 is on the driving side, and the third gear 30 is on the driven side. That is, when the gears 26, 34, and 30 are rotated forward, that is, when the gears 26, 34, and 30 are rotated in the directions of arrows U, V, and W in FIG. Each tooth T2 is in contact with the coast side tooth surface 36 (FIG. 2), and the drive side tooth surface 38 (FIG. 2) of each tooth T2 of the second gear 34 is the coast side tooth surface 32 of each tooth T3 of the third gear 30. To touch. In particular, as shown in FIG. 4, the magnitude of the torsion angle θ ₁ of the coast side tooth surface 36 of the second gear 34 and the torsion angle θ ₂ of the drive side tooth surface 38 opposite to the coast side tooth surface 36. Are different from each other.

Specifically, as shown in FIG. 4, the twist angle theta ₂ of the drive side tooth surfaces 38 of the second gear 34 relative to the helix angle theta ₁ of coast side tooth surfaces 36 of the second gear 34, the angle difference α (Θ ₂ = θ ₁ + α).

  On the other hand, in the gear device, the upward first direction of FIG. 1 indicated by the arrow Fb in FIG. 1 is caused by the behavior of the needle bearing 66 (see FIGS. 9 and 11) provided between the second gear 34 and the shaft S2. One thrust force is generated. On the other hand, if such a first thrust force Fb is maintained without being eliminated by the gear device, the second gear 34 is pressed against another member by the first thrust force Fb, and a large friction is generated. There is a possibility of loss due to such reasons. In this case, the loss at the time of rotation transmission between the gears 26, 34, and 30 increases, and the efficiency of rotation transmission may deteriorate.

  For example, FIG. 5 is a schematic diagram for explaining how a thrust force acting on the second gear 34 is generated due to the torsion angle of a tooth in an example of a conventional gear device. FIG. 6 is a view for explaining the thrust force Fb generated by the bearing behavior by taking out only the second gear 34 from FIG. The basic configuration of the gear device shown in FIGS. 5 and 6 is the same as the basic configuration of the present embodiment shown in FIGS. 1 to 4 described above. However, in the case of the gear device of FIGS. 5 and 6, the twist angles θ of the tooth surfaces on both the coast side and the drive side of each gear 26, 34, 30 are the same.

  In such a gear device of FIG. 5, the second gear 34 is in the direction of arrow V in FIG. 5 and the third gear 30 is in the direction of arrow W in FIG. Rotate forward in each direction. When the tooth surface load acting on the tooth surfaces of the first gear 26 and the second gear 34 is Fg, the torsion angle is θ, and the gear rotation efficiency is Eg, it is caused by the torsion of the teeth T1 and T2. The upward thrust force in FIG. 5 generated by the engagement of the drive-side tooth surface 28 of the first gear 26 and the coast-side tooth surface 36 of the second gear 34 is Fgsinθ. Also, in the same case, a diagram that occurs in the second gear 34 due to the engagement of the drive side tooth surface 38 of the second gear 34 and the coast side tooth surface 32 of the third gear 30 due to the twist of the teeth T2, T3. The downward thrust force of 5 is Eg × Fgsinθ.

  Therefore, a thrust force of Fgsinθ−Eg × Fgsinθ, that is, (1-Eg) Fgsinθ, acts on the second gear 34 due to the twist of the tooth T2.

  On the other hand, the second gear 34 is also subjected to a thrust force due to bearing behavior such as a needle skew of the needle bearing, and this is set as an upward thrust force Fb. As a result, an upward thrust force of (1-Eg) Fgsinθ + Fb acts on the second gear 34, which may not be eliminated during use. For this reason, friction due to the thrust force is generated between the second gear 34 and the counterpart member as described above, and the efficiency of rotation transmission may be deteriorated.

On the other hand, in the present embodiment shown in FIGS. 1 to 4, as described above, the twist angles θ ₁ and θ ₂ of the tooth surfaces 36 and 38 on both the coast side and the drive side of the second gear 34 (FIG. 4). ) Are different from each other. Preferably, the absolute value | α | (= | θ ₁ −θ ₂ |) of the difference α between the twist angles θ ₁ and θ ₂ of the tooth surfaces 36 and 38 on both the coast side and the drive side is regulated as follows. . In the following, although the twist angle theta ₂ of the drive side tooth surfaces 38 of the second gear 34 will be described a case than the helix angle theta ₁ of coast side tooth surface 36 α (> 0) content greater, the theta ₂ theta It can be smaller than ₁ .

1 to 4, the tooth surface load acting on the tooth surfaces of the first gear 26 and the second gear 34 is Fg, the twist angle of the coast side tooth surface 36 of the second gear 34 is θ _1, and the second The twist angle of the drive side tooth surface 38 of the gear 34 is θ ₂ (= θ ₁ + α), the gear rotation efficiency is Eg, and the upward first thrust force in FIG. And In this case, as is apparent from FIG. 1, if the difference α between the twist angles θ ₁ and θ ₂ of the coast-side and drive-side tooth surfaces 36 and 38 is defined so as to satisfy the following equation (1), The thrust force acting on the two gears 34 can be eliminated.

Fg × sin θ ₁ + Fb = Eg × Fg sin (θ ₁ + α) (1)

  Equation (1) indicates that the total thrust force acting on the second gear 34 is zero as a whole. Further, the first thrust force Fb due to the bearing behavior is known from the above-described conventional research of Non-Patent Document 1 to be the following equation (2).

Fb = ak × sin φ × Pb / Pm (2)

Here, φ is a rolling element constituting the corresponding needle bearing, and is an inclination angle with respect to the axis of the needle, that is, a skew angle, and a is a contact width in a short side direction between the corresponding axis and each needle according to Hertz contact theory. Pm is an average contact pressure between the corresponding shaft and each needle in Hertz contact theory, and Pb is a bearing load (contact load). K is a constant of 2.56 × 10 ⁴ kg / mm ³ .

  On the other hand, the first thrust force Fb generated by the bearing behavior is a skew angle φ, which is an inclination angle with respect to the axis of the needle constituting the needle bearing, from a conventional study such as Non-Patent Document 1 described above in a general use state. It has been found that the maximum is about 1 ° and the minimum is 0 °. In addition, a ball bearing can be used instead of the needle bearing, and in that case, the inclination with respect to the axis of the ball, which is a rolling element constituting the bearing, becomes zero. For this reason, in the range normally used, the 1st thrust force Fb which generate | occur | produces by a bearing behavior becomes the range of following (3) Formula.

ak × sin (φmin) × Pb / Pm ≦ Fb ≦ ak × sin (φmax) × Pb / Pm (3)

  Here, φmin is 0 ° and φmax is 1 °.

  Generally, since the bearing load Pb is about twice the tooth surface load Fg, the expression (3) is replaced with the following expression (4).

ak × sin (φmin) × 2Fg / Pm ≦ Fb ≦ ak × sin (φmax) × 2Fg / Pm (4)

  Further, from the conventional research such as Non-Patent Document 2 described above, it has been found that the meshing efficiency of the gear is about 94.0 to 99.8% within the normal use range. That is, the gear rotation efficiency Eg satisfies the following formula.

0.94 ≦ Eg ≦ 0.998 (5)

As a result, using the above equations (1), (4), and (5), the tooth surfaces 36 and 38 on both the coast side and the drive side of the second gear 34 (see FIG. The difference α between the twist angles θ ₁ and θ ₂ is preferably regulated so as to satisfy all of the following relational expressions.

βmin ≦ | α | ≦ βmax (6)
sinθ ₁ = Eg × sin (θ ₁ + βmin) (7)
sinθ ₁ + 2ak × sin ₁ ° / Pm = Eg × sin (θ ₁ + βmax) (8)
0.94 ≦ Eg ≦ 0.998 (9)

Thus, in the present embodiment, the second gear 34 is a helical gear that meshes with the first gear 26 and the third gear 30, and each of the gears 26, 34, and 30 is arranged in parallel with each other. It is supported by shafts S1, S2, and S3. Further, the torsion angle magnitudes θ ₁ and θ ₂ of the coast side tooth surface 36 and the drive side tooth surface 38 of the second gear 34 are different. For this reason, the second gear 34 receives thrust forces in opposite directions from the first and third gears 26, 30. Further, the upward first thrust force Fb in FIG. 1 is applied to the second gear 34 by the bearing behavior as described above. On the other hand, by providing a difference α in the magnitudes of the torsion angles θ ₁ and θ ₂ between the coast side tooth surface 36 and the drive side tooth surface 38 of the second gear 34, the torsion angle is caused by meshing with the counter gear. A second thrust force in the direction opposite to the first thrust force Fb can be generated according to the difference between θ ₁ and θ ₂ . For this reason, the thrust force applied to the second gear 34 can be reduced by canceling the first thrust force Fb. That is, as the second thrust force, a downward force having a magnitude of {Eg × Fgsin (θ ₁ + α) −Fg × sin θ ₁ } in FIG. 1 is generated, whereby the thrust force applied to the second gear 34 can be reduced. .

Further, according to the gear device of this embodiment that regulates α as in the above formulas (6) to (8), the second gear 34 that is a helical gear and the second gear 34 that meshes with the second gear 34. In a gear device having two first gears 26 and a third gear 30, it is possible to more effectively realize a configuration in which the thrust force applied to the second gear 34 can be eliminated and the loss of the gear device can be reduced under normal use conditions. . That is, by appropriately setting the difference α between the twist angles θ ₁ and θ ₂ of the tooth surfaces 36 and 38 on both the coast side and the drive side as described above, the thrust force resulting from the twist of the tooth surfaces 36 and 38 is obtained. Thus, the thrust force Fb caused by the bearing behavior can be effectively canceled out and eliminated.

  In the above formulas (6) to (9), it is not necessary to consider the influence of the axial length of the second gear 34. Therefore, a helical gear that is normally used and has any parallel axis that does not limit the axial length. It can be applied to the device.

  FIG. 7 shows the minimum value βmin for obtaining the twist angle difference α between the coast side and the drive side of the second gear using the gear Ga and the gear Gb having different gear rotation efficiencies in the gear device of the first embodiment. It is a figure which shows the result of having calculated and maximum value (beta) max. In the following description, the same elements as those shown in FIGS. 1 to 4 are denoted by the same reference numerals. The above βmin and βmax are used to obtain a preferable torsion angle difference α between the tooth surfaces 36 and 38 on both the coast side and the drive side of the second gear 34 (FIG. 1 and the like). In this calculation, a gear Ga having a gear rotation efficiency of 0.94 and a gear Gb having a gear rotation efficiency of 0.998 used as the second gear 34 were used. Moreover, the diameter of the shaft that supports the gears Ga and Gb was set to about 50 mm, and the diameter of the needle that is the rolling element of the bearing was set to about 5 mm. In this case, when the maximum value of the thrust force Fb due to the bearing behavior is obtained from the above equation (2), this maximum value is about 12% of the tooth surface load Fg. The minimum value of the thrust force Fb is Fb = 0 when the skew angle φ is 0 °.

These were substituted into the above equations (7) and (8), and βmax and βmin were calculated for two gears Ga and Gb having different gear rotation efficiencies. The calculation result is shown in FIG. Therefore, in the second gear 34, a preferable difference α between the twist angles θ ₁ and θ ₂ can be defined using βmax and βmin in FIG.

  In the above description, the case where the present invention is applied to a triaxial parallel helical gear device has been described. However, the present invention can be applied to various gear devices. For example, the present invention can be applied to the following planetary gear devices.

[Second Embodiment]
FIG. 8 is a perspective view of the gear device according to the second embodiment of the present invention. FIG. 9 is a schematic cross-sectional view of the apparatus of FIG. The gear device of the present embodiment is used as a Ravigneaux type planetary gear device. First, a basic configuration of the planetary gear device that is the gear device will be described.

  As shown in FIGS. 8 and 9, the Ravigneaux type planetary gear device 46 is composed of a front sun gear 48 and a rear sun gear 50 (FIG. 9) provided coaxially with each other, and a plurality of intermediate gears meshing around the front sun gear 48. A first pinion gear 52, a second pinion gear 56 that is a plurality of other intermediate gears that mesh with the first pinion gear 52, and a plurality of second pinion gears 56 are arranged coaxially with the front sun gear 48 and the rear sun gear 50. A ring gear 54. The front sun gear 48 and the rear sun gear 50 are externally fitted to the input shaft 51, which is the central axis, so as to be offset from each other in the axial direction, and are respectively rotatable. Further, although the rear sun gear 50 is fixed to the input shaft 51, only the front sun gear 48 or both the front sun gear 48 and the rear sun gear 50 can be fixed to the input shaft 51.

  Around the front sun gear 48, a first pinion gear 52 corresponding to a plurality of short pinion gears is engaged, and around the rear sun gear 50, a second pinion gear 56 corresponding to a plurality of long pinion gears is engaged. Further, the ring gear 54 is arranged around one half of the plurality of second pinion gears 56. Each gear 48, 50, 52, 54, 56 is a helical gear.

  As shown in FIG. 9, each of the plurality of first pinion gears 52 is a plurality of first gear shafts having both ends supported and fixed to a front carrier 58 that is a first carrier and a rear carrier 60 that is a second carrier. The first pinion shaft 62 is rotatably supported and meshes with the second pinion gear 56 as shown in FIG. 8 and 9, only one first pinion gear 52 is shown, but a plurality of first pinion gears 52 are actually provided. Further, in FIG. 8, illustration of the front carrier and the rear carrier is omitted. As shown in FIG. 9, the plurality of second pinion gears 56 are rotatably supported by second pinion shafts 64 which are a plurality of second gear shafts whose both ends are supported and fixed to a front carrier 58 and a rear carrier 60, respectively. And meshes with the rear sun gear 50. Thus, the first pinion gear 52 meshes with the front sun gear 48 and the second pinion gear 56, and the second pinion gear 56 meshes with the rear sun gear 50 and the ring gear 54. The front carrier 58 and the rear carrier 60 are connected and integrated. As shown in FIG. 9, two rear carriers 60 are illustrated, but the position in the axial direction is shifted between the portion supporting the first pinion shaft 62 and the portion supporting the second pinion shaft 64. Can be integrated with each other.

  Each first pinion gear 52 is rotatably supported by a needle bearing 66 disposed around the first pinion shaft 62. Each second pinion gear 56 is rotatably supported by two needle bearings 66 disposed on both sides in the axial direction around the second pinion shaft 64. Each needle bearing 66 is a full roller bearing configured by arranging a plurality of needles side by side in the circumferential direction around the shaft. Each needle bearing 66 may have a configuration other than a full roller bearing. Thrust washers provided around the pinion shafts 62 and 64 corresponding to the front carrier 58 or the rear carrier 60 on both axial ends of the first pinion gear 52 and the second pinion gear 56, respectively. 68 is arranged. The first pinion gear 52, the front sun gear 48, and the second pinion gear 56 that mesh with each other are rotatably disposed about three corresponding shaft centers that are parallel to each other. The second pinion gear 56, the first pinion gear 52, and the rear sun gear 50 (or the ring gear 54) that mesh with each other are rotatably disposed around a plurality of corresponding axes that are parallel or coaxial with each other.

  Such a planetary gear unit 46 is used in a usage pattern as shown in FIG. 10, for example. FIG. 10 is a configuration diagram of the planetary gear device of FIG. In FIG. 10, the same elements as those shown in FIGS. 8 and 9 are denoted by the same reference numerals. The example of FIG. 10 is used, for example, as a transmission of an automobile, and a rotary shaft of a power source such as an engine is coaxially coupled to a rear sun gear 50 so that power can be input from the power source to the rear sun gear 50. Further, the ring gear 54 and the front sun gear 48 can be braked by the first brake 70 and the second brake 72, respectively. Further, the output shaft 74 is coaxially coupled to the rear carrier 60 so that the output can be taken out from the rear carrier 60. That is, the rear sun gear 50 is on the driving side, and the front sun gear 48 is on the driven side. Therefore, the tooth surface of the second pinion gear 56 that contacts the rear sun gear 50 is the coast side tooth surface, and the tooth surface of the second pinion gear 56 that contacts the first pinion gear 52 is the drive side tooth surface. Further, the tooth surface of the first pinion gear 52 that contacts the second pinion gear 56 is the coast side tooth surface, and the tooth surface of the first pinion gear 52 that contacts the front sun gear 48 is the drive side tooth surface.

  When such a planetary gear unit 46 is used, for example, the braking state of one or both of the ring gear 54 and the front sun gear 48 by one or both of the first brake 70 and the second brake 72 is controlled by a control unit (not shown). Thus, the shift can be switched. In addition, the usage example of the planetary gear apparatus 46 is not limited to such an example, It can be used by various methods. In any case, the Ravigneaux type planetary gear unit 46 is effective because it is small in size and can set many shift stages.

Further, in each first pinion gear 52, the twist angle θ ₁ of the coast side tooth surface that contacts the tooth surface of the second pinion gear 56 during forward rotation, and the twist angle θ of the drive side tooth surface that contacts the tooth surface of the front sun gear 48. ₂ , the total thrust force acting on each first pinion gear 52 is set to zero. Further, a is 1/2 of the contact width in the short side direction between the corresponding axis of each first pinion gear 52 and each needle in the Hertz contact theory, and P _m is the average of the corresponding axis and each needle in the Hertz contact theory. It is assumed that the contact pressure is set, Eg is the rotational efficiency of each first pinion gear 52, and k is 2.56 × 10 ⁴ kg / mm ³ . In this case, the angle difference α1 between the twist angles θ ₁ and θ ₂ is set to α satisfying the above expressions (6) to (9).

Further, in each second pinion gear 56, the torsion angle θ ₃ of the coast side tooth surface that contacts one or both of the tooth surfaces of the ring gear 54 and the rear sun gear 50 during forward rotation, and the drive side that contacts the tooth surface of the first pinion gear 52. The total torsional force acting on each second pinion gear 56 is made zero by making the torsion angle θ ₁ of the tooth surface different. Further, a is ½ of the contact width in the short side direction between the corresponding axis of each second pinion gear 56 and each needle in the Hertz contact theory, and P _m is the average of the corresponding axis and each needle in the Hertz contact theory. It is assumed that the contact pressure is set, Eg is the rotation efficiency of each second pinion gear 56, and k is 2.56 × 10 ⁴ kg / mm ³ . In this case, the angle difference α2 between the twist angles θ ₁ and θ ₃ satisfies the expressions (6) to (9) by replacing θ ₁ with θ ₃ in the expressions (6) to (9) above. α is set. Therefore, the torsion angle of the front sun gear 48 is θ ₂ like the drive side tooth surface of the first pinion gear 52, and the torsion angle of one or both of the ring gear 54 and the rear sun gear 50 is the same as the coast side tooth surface of the second pinion gear 56. Is θ ₃ .

Thus, by appropriately setting the differences α1, α2 between the twist angles θ ₁ , θ ₂ , θ ₃ , the thrust force applied to each first pinion gear 52 and each second pinion gear 56 can be eliminated (can be set to 0). Loss due to friction between the pinion gears 52 and 56 and the axially opposing mating member such as the thrust washer 68 can be greatly reduced, and a highly efficient planetary gear unit 46 can be realized. Other configurations and operations are the same as those in the first embodiment.

[Third Embodiment]
FIG. 11 is a schematic cross-sectional view of a gear device according to a third embodiment of the present invention. In the present embodiment, the gear device of the present invention is used as a single planetary planetary gear device. That is, there is only one type of pinion gear. Specifically, the planetary gear device 84 that is a gear device includes a sun gear 88 that is fixed around the input shaft 86 that is a central shaft, a ring gear 90 that is disposed coaxially with the sun gear 88, and the sun gear 88 and the ring gear 90. A pinion gear 92 that is a plurality of intermediate gears arranged around the sun gear 88 and inside the ring gear 90 is provided. The plurality of pinion gears 92 are rotatably supported around a pinion shaft 94 by a needle bearing 66, and the pinion shafts 94 are supported and fixed to two carriers 96 provided on both sides. The pinion gear 92, the sun gear 88, and the ring gear 90 that mesh with each other are rotatably disposed around a plurality of corresponding axes that are parallel or coaxial with each other.

In the present embodiment, the sun gear 88 is the driving side and the ring gear 90 is the driven side. The torsion angle θ ₁ of the coast side tooth surface that contacts the tooth surface of the sun gear 88 during the forward rotation of each pinion gear 92 is different from the torsion angle θ ₂ of the drive side tooth surface that contacts the tooth surface of the ring gear 90. The total thrust force acting on the pinion gear 92 is set to zero. Further, a is 1/2 of the contact width in the short side direction between the corresponding axis of each pinion gear 92 and each needle in the Hertz contact theory, and P _m is the average contact pressure between the corresponding axis and each needle in the Hertz contact theory. Where Eg is the rotational efficiency of each pinion gear 92 and k is 2.56 × 10 ⁴ kg / mm ³ . In this case, the angle difference α between the twist angles θ ₁ and θ ₂ satisfies the above expressions (6) to (9). Therefore, the twist angle of the sun gear 88 is θ ₁ corresponding to the coast side tooth surface of the pinion gear 92, and the twist angle of the ring gear 90 is θ ₂ corresponding to the drive side tooth surface of the pinion gear 92.

Thus, by appropriately setting the difference α between the twist angles θ ₁ and θ ₂ , the thrust force applied to each pinion gear 92 can be eliminated (can be set to 0), and each pinion gear 92 and the thrust washer 68 are opposed to each other in the axial direction. Loss due to friction with the mating member can be significantly reduced, and a highly efficient planetary gear unit 84 can be realized. Other configurations and operations are the same as those in the first embodiment or the second embodiment.

  In the plurality of gears of the gear device of each of the above embodiments, the relationship between the drive side and the driven side during forward rotation can be reversed. In this case, in the second gear 34, the first pinion gear 52, the second pinion gear 56, or the pinion gear 92, the relationship between the coast side tooth surface and the drive side tooth surface is reversed. In the above description, the gear device having three or more gears has been described. However, in the gear device having two gears meshing with each other, the twist angle between the drive side tooth surface and the coast side tooth surface of one or both gears. A difference can also be provided. In the above-described embodiment, the case where the gear is supported on the shaft via the needle bearing has been described. However, the present invention can be similarly implemented when another bearing such as a ball bearing is used in addition to the needle bearing.

  10 pinion gear, 12 carrier plate, 14 pinion shaft, 16 thrust ring, 18 needle bearing, 20 thrust washer, 22 sun gear, 24 internal gear, 26 first gear, 28 drive side tooth surface, 30 third gear, 32 coast side Tooth surface, 34 Second gear, 36 Coast side tooth surface, 38 Drive side tooth surface, 46 Planetary gear device, 48 Front sun gear, 50 Rear sun gear, 51 Input shaft, 52 First pinion gear, 54 Ring gear, 56 Second pinion gear, 58 Front carrier, 60 Rear carrier, 62 1st pinion shaft, 64 2nd pinion shaft, 66 Needle bearing, 68 Thrust washer, 70 1st brake, 72 2nd brake, 74 Output shaft, 84 Planetary gear device, 86 Input shaft, 88 Sun gear, 90 Ring gear, 92 pinion gear, 94 pinion shaft, 96 carrier.

Claims (4)

Each including three or more gears, which are helical gears,
An intermediate gear, which is at least one of the gears, meshes with another two of the gears;
Each of the gears is disposed so as to be rotatable around the center of the corresponding axis among three axes that are parallel or coaxial with each other.
The gear device according to claim 1, wherein the torsion angle of the coast side tooth surface and the torsion angle of the drive side tooth surface of the intermediate gear are different from each other. The gear device according to claim 1, wherein
The intermediate gear is rotatably supported by a bearing including a plurality of rolling elements provided between the intermediate shaft and the shaft.
θ ₁ is the torsion angle of the coast side tooth surface, a is ½ of the contact width between the corresponding axis in the Hertz contact theory and each rolling element, and P _m is the corresponding in the Hertz contact theory. The average contact pressure between the shaft and each rolling element, Eg is the rotational efficiency of the intermediate gear, and k is 2.56 × 10 ⁴ kg / mm ³ , the coast side teeth of the intermediate gear Α, which is the difference in torsion angle between the surface and the drive-side tooth surface, satisfies βmin ≦ | α | ≦ βmax,
βmin satisfies sinθ ₁ = Eg × sin (θ ₁ + βmin),
βmax satisfies sinθ ₁ + 2ak × sin ₁ ° / Pm = Eg × sin (θ ₁ + βmax),
Eg satisfies 0.94 <= Eg <= 0.998. The gear apparatus characterized by the above-mentioned. The gear device according to claim 1 or 2,
A sun gear which is the gear;
A plurality of pinion gears arranged around the sun gear and meshing with the sun gear, each being the intermediate gear;
A ring gear that is arranged on the outside of the plurality of pinion gears and coaxially with the sun gear and that is another gear that meshes with the plurality of pinion gears;
The plurality of pinion gears are supported by a gear shaft that is the shaft supported by the first carrier and the second carrier,
A gear device that is used as a planetary gear device. The gear device according to claim 1 or 2,
A front sun gear and a rear sun gear, which are rotatable gears arranged coaxially with each other and shifted in the axial direction;
A plurality of first pinion gears arranged around the front sun gear and each being an intermediate gear meshing with the front sun gear;
A plurality of second pinion gears that are arranged around the rear sun gear and are separate intermediate gears that mesh with the rear sun gear;
A ring gear that is disposed around the plurality of second pinion gears and is another gear that meshes with the plurality of second pinion gears;
The plurality of first pinion gears are supported by a first gear shaft that is the shaft supported by the first carrier and the second carrier, respectively, and meshed with the second pinion gear,
The plurality of second pinion gears are supported by a second gear shaft, which is another shaft supported by the first carrier and the second carrier, respectively.
A gear device used as a Ravigneaux type planetary gear device.

Images