PRECESSIONAL GEAR TRANSMISSION

Abstract

The p recessional gear transmission comprises a body, a satellite wheel with two bevel gear rings driven by a crankshaft in sphero-spatial motion around a fixed point, two central bevel wheels, one immobile fixed in the body and the other mobile mounted on a driven shaft. The teeth of the gear rings have a circular arc flank profile, those of the central bevel wheels are variable curvilinear. The configuration of the parameters of angles, the number of teeth, the ratio of the numbers of teeth of the mating wheels in the gears and the radius of the circular arc of the teeth profile of the gear rings determines the geometry and the kinematics of the contact of the teeth, the degree of frontal overlap, expressed by the number of simultaneously engaged pairs of teeth and defines the pressure angle between the mating flanks.

Claims

1. A precessional gear transmission including a body, a satellite wheel with two bevel gear rings installed on the inclined portion of the crankshaft between two central bevel wheels, one immobile fixed in the body and the other mobile mounted on the driven shaft, wherein the teeth gearing is performed in contacts with convex-concave geometry, wherein the central bevel wheels are made with curvilinear flank profiles with variable curvature with one tooth less than the satellite wheel gear rings made with circular are flank profiles, the teeth flanks mate with frontal overlap sf within the limits 1.5≤ε.sub.f≤4.0 simultaneously engaged pairs of teeth, at the same time the gearwheels are made with the conical axoid angle within the limits 0°≤δ≤30°, with the angle between the axes of the crank and the central bevel wheels within the limits 1.5°≤θ≤7°, and the circular arc radius of the flank profile of the Z-toothed satellite wheel gear rings is within the limits (1.0-1.57) D/Z [mm], which generally provides a reduction of the difference in the curvatures of the flank profiles in the section with diameter D of up to (0.02-1.5) D/Z [mm] and a decrease in the pressure angle α between the flanks of up to 15°, as well as a decrease in the relative sliding velocity between the mating flanks.

2. A precessional gear transmission of claim 1, wherein the wheel teeth are made inclined, which provides an increase in the total length of the contact lines with their gradual entry into the gear field and an increase in the share of pure rolling of the engaged teeth flanks with sphero-spatial interaction.

3. A precessional gear transmission of claims 1 and 2, wherein one of the satellite wheel bevel gear rings with the conical axoid angle δ=0° is made of bolts with one less or more than the number of central bevel wheel teeth with which it engages, which provides the pressure angle between the mating flanks α≤45° and the extension of kinematic possibilities.

4. A precessional gear transmission of claims 1 and 2, wherein one of the satellite wheel gear rings with the conical axoid angle δ>0° is made of conical bolts with one less than the number of central bevel wheel teeth and with a profile angle α>45°, which provides the transformation of motion and the transmission of load by rolling the conical bolts on the flank profile of the central wheel teeth with inclined slope effect and, respectively, the operation of the transmission in multiplier mode.

5. A precessional gear transmission of claims 1 and 2, wherein the satellite wheel is installed on a spherical support placed on the driven shaft in its center of precession and coaxially with the mobile central bevel wheel, at the same time the satellite wheel is equipped with a semi-axle, at the end of which is mounted a bearing, kinematically coupled with the crankshaft.

Description

BRIEF DESCRIPTION OF DRAWINGS

[0023] Summary of the invention is explained by the drawings that represent:

[0024] FIG. 1. Precessional toothed gear transmission.

[0025] FIG. 2. Kinematic precessional toothed gear transmission: with wheels injected from plastics (a) and pressed from metal powders by sintering (b).

[0026] FIG. 3. Path of motion of the circular arc radius G origin.

[0027] FIG. 4. Description of the flank profile of the central wheel teeth.

[0028] FIG. 5. Teeth contact in the precessional toothed gear with frontal reference multiplicity ε.sub.f=100% (a) and ε.sub.f=66.6% (b).

[0029] FIG. 6. Evolution of the contact point of teeth in the gears with frontal reference multiplicity ε.sub.f=100% (a) and ε.sub.f=66.6% (b).

[0030] FIG. 7. Geometry of the modified teeth by making the tips in the gear with frontal reference multiplicity ε.sub.f=100% (a) and ε.sub.f=66.6% (b).

[0031] FIG. 8. Convex-concave gear with slight difference in the curvatures of the mating flank profiles with frontal multiplicity ε.sub.f=27.58%.

[0032] FIG. 9. Evolution of the teeth contact geometry variation with three, two and four simultaneously engaged pairs of teeth.

[0033] FIG. 10. Kinematics and tooth profiles of the gears (Z.sub.1-Z.sub.2) and (Z.sub.3-Z.sub.4) with reduction gearbox (a) and multiplier (b) operation modes.

[0034] FIG. 11. Geometry of the teeth contact and relative positioning of the mating flanks in contact k.sub.1 for: a) Z.sub.1=Z.sub.1−1, δ=22.5°; b) Z.sub.1=Z.sub.2+1, δ=0°, c) Z.sub.1=Z.sub.2+1, δ=22.5°.

[0035] FIG. 12. Profiles and pressure angle between the flanks of the central wheel teeth with reference multiplicity ε.sub.f=100% (a) and ε.sub.j=73% (b).

[0036] FIG. 13. Kinematics and geometry of the teeth contact in the gear (Z.sub.3-Z.sub.4) with (Z.sub.4=Z.sub.3−1) (a) and gear (Z.sub.1-Z.sub.2) with (Z.sub.1=Z.sub.2−1) (c) with three simultaneously engaged pairs of teeth.

[0037] FIG. 14. Linear velocities in the contact point V.sub.E1, V.sub.E2, V.sub.a1 (a) and difference in the radii of curvature (ρ.sub.ki-r) (b) of the mating profiles in contact k.sub.1 (c) depending on ψ for Z.sub.1=Z.sub.2−1 and δ=22.5°, Z.sub.1=24, Z.sub.2=25, θ=3.5°, δ=22.5°, r=6.27 mm and R=75 mm.

[0038] FIG. 15. Evolution of the total contact line variation of the inclined teeth depending on ψ.

[0039] FIG. 16 Contact of the inclined teeth with the angle β.sub.g placed in the gear field with a pair of frontal mating teeth (a) and with two pairs (b).

[0040] FIG. 17. Precessional transmission with the gear (Z.sub.3-Z.sub.4) of bolts (δ=0) with the pressure angle between the flanks α≤45° for the ratio of the numbers of mating teeth Z.sub.4=Z.sub.3−1.

[0041] FIG. 18. Precessional transmission with the gear (Z.sub.3-Z.sub.4) of bolts (δ>0) with the pressure angle between the flanks for the ratio of the numbers of mating teeth Z.sub.4=Z.sub.3+1 (for multiplier operation mode).

[0042] FIG. 19. Bolt-tooth interaction in the gear Z.sub.4=Z.sub.3+1 with multiplier operation mode.

[0043] FIG. 20. Precessional toothed gear transmission with convex-concave contact and difference in the numbers of mating teeth Z.sub.1(4)==Z.sub.1(3)−1 according to claim 5.

MODES OF CARRYING OUT OF INVENTION

[0044] The precessional gear transmission according to claim 1, shown in FIG. 1 comprises a body 1, a satellite wheel 2 with two bevel gear rings 3 and 4 installed on the inclined portion of the crankshaft 5, two central bevel wheels 6 and 7, one immobile 6, fixed in the body 1 and the other mobile 7, mounted on the driven shaft 8. The novelty of the invention consists in that the teeth engage in contacts with convex-concave geometry with the small difference in the curvatures of the flank profiles, wherein the central bevel wheels 6 and 7 are made with straight teeth and curvilinear flank profiles with variable curvature with one tooth less than the gear rings 3 and 4 of the satellite wheel 2 made with circular arc flank profiles, the flanks of the teeth mate with frontal overlap ε.sub.f within the limits 1.5≤ε.sub.j≤4.0 simultaneously engaged pairs of teeth, at the same time the gearwheels are made with the conical axoid angle within the limits 0°≤δ≤30°, with the nutation angle θ between the axes of the crank 5 and the central bevel wheels 6 and 7 within the limits 1.5°≤θ≤<7°, and the circular arc radius of the flank profile of the Z-toothed satellite wheel gear ring 2 is within the limits (1.0-1.57) D/Z mm, which generally provides for the mating of teeth in convex-concave contacts with the difference in the in the curvatures of the flank profiles in the section with diameter D of up to (0.02-1.5) D/Z mm and of the pressure angle α between the flanks of up to 15°, as well as the decrease in the relative sliding velocity between the mating flanks.

[0045] The bevel gear rings 3 and 4 of the satellite wheel 2 have teeth with circular arc flank profiles, and the central bevel wheels 6 and 7—variable curvilinear, depending on the angles θ and δ, the circular arc radius r, the number and the ratio of the numbers of teeth of the gears (Z.sub.1-Z.sub.2) and (Z.sub.3-Z.sub.4), the configuration of the numerical values of which influence the change of the teeth profile shape, determines their degree of frontal overlap, expressed by the number of simultaneously engaged pairs of teeth ε.sub.f, the size of the pressure angle α between the mating flanks and the frictional sliding velocity between the flanks.

[0046] The following approaches to the creation of precessional gear with gearwheels, claimed in FIG. 1, are valid both for the kinematic transmissions with wheels injected from plastics and the kinematic transmissions with wheels pressed by sintering from metallic powders shown respectively in FIGS. 2 (a) and (b).

[0047] The precessional gear transmission according to claim 1, operates in the following way: Upon rotation of the crankshaft 5, the satellite wheel (FIGS. 1 and 2) is communicated a sphero-spatial motion around a fixed point, which through its bevel gear rings 3 and 4 (and/or the gear ring 4 made of bolts) interacts with the fixed bevel gearwheel 6 and the mobile gearwheel 7 respectively.

[0048] The difference in the number of teeth of the engaged wheels is only one tooth, and the numerical ratio of teeth is:

Z.sub.1=Z.sub.2−1 and Z.sub.4=Z.sub.3−1  (1)

[0049] Due to the fact that the central bevel wheel 6 is fixed in the body 1, and the central bevel wheel 7 is mounted on the driven shaft 8, when rotating the crank 5 with the electromotor rotational frequency, the driven shaft 8 will rotate with reduced rotational frequency with the transmission ratio i.sub.HV.sup.b:

[00001]$\begin{array}{cc}{i}_{\mathrm{HV}}^b=-\frac{Z_2{Z}_4}{Z_1{Z}_3-Z_2{Z}_4}.& \left(2\right)\end{array}$

[0050] Generally, when transmitting the motion and load through the gears (Z.sub.1-Z.sub.2) and (Z.sub.3-Z.sub.4) with the ratio of the numbers of teeth Z.sub.1(4)=Z.sub.2(3)±1, the direction of rotation of the driven shaft 8 coincides or not with the direction of the input shaft 5.

[0051] If Z.sub.2>Z.sub.3, the crankshaft 5 and the driven shaft 8 rotate counterclockwise, and if Z.sub.2<Z.sub.3—in the same direction.

[0052] Frontal multiplicity of the mating wheel teeth gearing in the precessional transmission is determined by three interdependent constructive-kinematic conditions. [0053] the satellite wheel performs a sphero-spatial motion with a fixed point, in which the extensions of the teeth generators of the engaged wheels intersect; [0054] the difference between the number of teeth of the engaged wheels is Z.sub.1=Z.sub.2±1 and Z.sub.4=Z.sub.3±1, and the difference between the number of teeth of the satellite gear rings can be Z.sub.2=Z.sub.3±1; [0055] compliance with the continuity of the rotational motion transformation function, therefore ω.sub.1/ω.sub.8=const.

[0056] It was found that the absolute multiplicity of gearing (100%) with the compliance of the three conditions can only occur when using the variable convex/concave profile of the teeth flanks, usually of the central wheels, depending on the values of the conical axoid δ and nutation θ angles of radius r of the curvature of the teeth profiles of the satellite wheel gear rings, as well as on the number of teeth of the wheels Z and their ratio±1 (see FIGS. 1 and 2).

[0057] The load-bearing capacity and mechanical efficiency of the precessional gear transmissions, according to the invention, are proposed to be increased by achieving the following technical solutions stipulated in claim 1: [0058] creating contacts between the teeth flanks with convex-concave geometry with small difference of curvatures; [0059] providing the minimum pressure angles between the flanks of the engaged teeth; [0060] providing the minimum relative frictional sliding velocity between the mating flanks; [0061] decreasing the frontal gear multiplicity and increasing the degree of longitudinal overlap with pure rolling of the teeth in the sphero-spatial interaction of the mating wheels.

[0062] The constructive-kinematic conditions and the distinctive technical solutions mentioned above, constitute the basis for the development of the precessional gear transmission according to claim 1, for both the power transmissions shown in FIG. 1 and the kinematic transmissions shown in FIG. 2.

[0063] The elaboration of the precessional gear transmission in the embodiment according to claim 1 covered the following approaches and technical solutions:

[0064] 1. Creation of the Contact with Convex-Concave Geometry Between the Flanks of the Teeth with Small Difference of Curvatures.

[0065] In accordance with claim 1 for creating the convex-concave contact of the engaged teeth with sphero-spatial motion, the profile of the satellite wheel teeth 2 is described by an arbitrary curve LEM, for example, in circular arc of radius r with the origin in point G (FIG. 3), which belongs to the satellite wheel teeth 2.

[0066] From the Euler equations, taking into account the kinematic relation between the angles φ and ψ expressed by φ=−Z.sub.1/Z.sub.2ψ (2), we obtain the coordinates of the origin G of the circular arc radius X.sub.G, Y.sub.G, Z.sub.G depending on the rotation angle ψ of the crankshaft:

[00002]$\begin{array}{cc}{X}_G=R\cos \delta \left[-\cos \mathrm{\psi sin}\left(\psi \frac{Z_1}{Z_2}\right)+\sin \mathrm{\psi cos}\left(\psi \frac{Z_1}{Z_2}\right)\cos \theta \right]-R\sin \mathrm{\delta sin}\mathrm{\psi sin}\theta ,& \left(3\right)\end{array}$$\begin{array}{cc}{Y}_G=-R\cos \delta \left[\sin \mathrm{\psi sin}\left(\psi \frac{Z_1}{Z_2}\right)+\cos \mathrm{\psi cos}\left(\psi \frac{Z_1}{Z_2}\right)\cos \theta \right]+R\sin \mathrm{\delta cos}\mathrm{\psi sin}\theta ,& \left(4\right)\end{array}$$\begin{array}{cc}{Z}_G=-R\cos \mathrm{\delta cos}\left(\psi \frac{Z_1}{Z_2}\right)\sin \theta -R\sin \delta \cos \theta .& \left(5\right)\end{array}$

[0067] The origin G of the circular are radius, with which the teeth of the satellite wheel 2 gear rings 3 and 4 are arbitrarily described (see FIG. 2), moves on the sphere surface with radius R with the origin in the center of precession O, describing the path ζ.sub.1=f(ξ.sub.1), expressed by the coordinates X.sub.G, Y.sub.G, Z.sub.G (FIG. 3).

[0068] The path of motion G of the circular arc LEM on the sphere with radius R is projected on the plane P.sub.1 using the rules of spherical trigonometry. Thus, it is obtained the path T.sub.G, of motion of the origin of the circular arc G radius on the plane P.sub.1, expressed by the dependence ζ=f(ξ.sub.1).

[0069] Knowing the path of motion of the origin of the circular arc G radius, expressed in the coordinates X.sub.G, Y.sub.G, Z.sub.G (FIG. 4), one can determine the position of the contact point E of the flank profiles of the mating teeth in the gears (Z.sub.1-Z.sub.2) and (Z.sub.3-Z.sub.4) for any angular position ψ of the crankshaft 5.

[0070] The family of contact points E obtained in a precession cycle 0<ψ<2πZ.sub.2/Z.sub.1 represents the profile of the teeth of the immobile 6 or mobile 7 central wheels.

[0071] To describe the flank profiles of the central wheel teeth 6 and 7, the projections of the velocity vector V.sub.G on the coordinate axes of the mobile system OX.sub.1Y.sub.1Z.sub.1, are determined depending on the angular velocity of the crankshaft 5 (see FIGS. 1 and 2).

[0072] To determine the position of the contact point E of the teeth on the spherical surface, we identify the equation of a plane P.sub.2 drawn perpendicular to the velocity vector V.sub.G, passing through the center of precession O and the origin of the circular arc radius G. The equation of plane P.sub.2 can be written by the expression:

[OG×OC]×V.sub.G=0,  (6)

where OG and OCare vectors that determine the position of the origin of the circular arc radius of curvature of the satellite tooth G and, respectively, of an arbitrary point C of plane P.sub.2 with respect to the origin of the immobile coordinate system OX Y Z (FIG. 3).

[0073] The vectorial product [OG×OC](6) is expressed as a third-order determinant and, by opening it according to the elements of the first line, we obtain:

[OG×OC]=i(Y.sub.GZ−Z.sub.GY)+j(Z.sub.GX−X.sub.GZ)+k(X.sub.GY−Y.sub.GX),  (7)

wherein X.sub.G, Y.sub.G, Z.sub.G are the coordinates of the origin of the radius of curvature G of the circular arc profile of the satellite wheel teeth; X,Y,Z—the coordinates of the arbitrary point C on the plane P.sub.2.

[0074] If the contact point of the teeth E is placed on the sphere with the radius R, then its coordinates satisfy its equation:

X.sub.E.sup.2+Y.sub.E.sup.2+Z.sub.K.sup.2−R.sup.2=0.  (8)

[0075] From FIG. 4 we observe that the angle between the position vectors of the origin of the circular arc radius of curvature OG of the satellite tooth and the position vector of the contact point E of the teeth OE represents the angle of contact β from the center of precession O of the radius r of the circular arc profile of the satellite wheel teeth, from which results:

OG.Math.OE=R.sup.2 cos β  (9)

or

X.sub.EZ.sub.G+X.sub.EY.sub.G+Z.sub.EZ.sub.G−R.sup.2 cos β=0.  (10)

[0076] From the equation (12) we determine:

X.sub.E=(R.sup.2 cos β−Y.sub.EY.sub.G−Z.sub.EZ.sub.G)/X.sub.G.  (11)

[0077] To determine the coordinate Y.sub.E of the contact point of the teeth E, we substitute (11) in (8) and obtain:

Y.sub.E=k.sub.1Z.sub.E−d.sub.1,  (12)

and by substituting (12) into (11), we obtain the expression of the contact point coordinate X.sub.E:

X.sub.E=k.sub.2Z.sub.E+d.sub.2,  (13)

where

k.sub.1=[X.sub.G(X.sub.G.Math.{dot over (X)}.sub.G+Y.sub.G{dot over (Y)}.sub.G+Z.sub.G.sup.2{dot over (X)}.sub.G]/(X.sub.G{dot over (Y)}.sub.G−Y.sub.G{dot over (X)}.sub.G)Z.sub.G

d.sub.1=R.sup.2 cos β{dot over (X)}.sub.G(X.sub.G{dot over (Y)}.sub.G−Y.sub.G{dot over (X)}.sub.G)

k.sub.2=−(k.sub.1Y.sub.G+Z.sub.G)/X.sub.G

d.sub.2=(R.sup.2 cos β+d.sub.1Y.sub.G)X.sub.G.  (14)

[0078] Substituting (12) and (13) in (8) and, considering that the profile curve of the central wheel teeth is equidistant from the path of motion of the origin G of the circular arc radius, and for any rotation angle ψ of the crankshaft, the condition Z.sub.E<Z.sub.G must be met, the coordinate Z.sub.E can be determined by the relation:

[00003]$\begin{array}{cc}{Z}_E=\frac{\left(k_1{d}_1-k_2{d}_2\right)-{\left[{\left(k_1{d}_1-k_2{d}_2\right)}^2+\left(k_1^2+k_2^2+1\right)\left(R^2-d_1^2-d_2^2\right)\right]}^{1/2}}{\left(k_1^2+k_2^2+1\right)}.& \left(15\right)\end{array}$

[0079] Relationships (12), (13) and (15) determine the coordinates X.sub.E, Y.sub.E and Z.sub.E of the contact point E of the teeth, the set of which in a precession cycle represents the flank profile of the central wheel teeth, placed on the sphere of radius R.

[0080] From the analysis of equations (12), (13) and (15), we state that the flank profile of the central wheel teeth is variable depending on the number of teeth Z.sub.2, the ratio of the numbers of teeth of the engaged wheels Z.sub.1=Z.sub.2−1 or Z.sub.1=Z.sub.2+1, the conical axoid δ, nutation θ and contact angles at the center of precession of the radius of curvature of the circular arc profile β of the satellite wheel teeth.

[0081] The precessional gear being bevel, with the extensions of the generators intersected in the center of precession, it is appropriate to render the teeth profile in normal section, for example, in the plane P.sub.1 drawn by the points E.sub.1 and E.sub.2 perpendicular to the plane OE.sub.1E.sub.2(FIG. 4).

[0082] The coordinates X.sub.E, Y.sub.E, and Z.sub.E of points E.sub.1 and E.sub.2 on the teeth profile on the sphere are determined from the relations (12), (13) and (15) for the angles of precession ψ=0 and ψ=2πZ.sub.2/Z.sub.1, corresponding to a precession cycle.

[0083] Using the rules of spherical trigonometry, we design the teeth profile on the sphere with the radius R on the plane P.sub.1.

[0084] To design the profile of the central wheel teeth in two coordinates ζ and ξ in the plane P.sub.1 we draw the coordinate system E.sub.1ξζ with the origin in point E.sub.1=, whose axis E.sub.1ξ passes through point E.sub.2 (FIG. 5). From coordinates X.sub.N, Y.sub.N and Z.sub.N we pass to coordinates ζ and ξ using the relations:

[00004]$\begin{array}{cc}\xi =\frac{\left[{\left(E_1{E}_2\right)}^2+v_1^2-v_2^2\right]}{2\left(E_1{E}_2\right)},\zeta =\sqrt{v_1^2-{\xi }^2}.& \left(16\right)\end{array}$

[0085] The expressions (16) represent the coordinates of the curve points, whose family constitute the flank profile of the central wheel teeth, designed on the plane P.sub.1, expressed in parametric form with the variation of the precession angle from ψ=0 to ψ=2πZ.sub.2/Z.sub.1.sup.2.

[0086] To design the path of motion of the origin of the circular arcs G in 2D, we pass from coordinates X.sub.N, Y.sub.N and Z.sub.N to Cartesian coordinates ξ.sub.1, ζ.sub.1 using the relations:

[00005]$\begin{array}{cc}{\xi }_1=\frac{\left[{\left(E_1{E}_2\right)}^2+S_1^2-S_2^2\right]}{2\left(E_1{E}_2\right)},{\zeta }_1=\sqrt{S_1^2-{\xi }_1^2}.& \left(17\right)\end{array}$

[0087] Function ξ.sub.1 of ζ.sub.1 (17) represents the projection of the path of motion of the origin of the circular arcs G on the plane P.sub.1, and function ξ of ζ (16) represents the flank profile of the central wheel teeth projected on the plane P.sub.1.

[0088] The value configuration of parameters Z, r, δ and θ influences the shape of the flank profile of the central wheel teeth and provides for the teeth front reference gear of up to 100% simultaneously engaged pairs of teeth. In the precessional transmission shown in FIG. 1 the teeth gears (Z.sub.1-Z.sub.2) and (Z.sub.3-Z.sub.4) may be with the same or different reference multiplicity of the teeth gearing.

[0089] 2. Transformation of Teeth Contact Geometry into Precessional Gear Depending on the Angle of Precession Vi and Distinctive Solutions for Creating the Convex-Concave Contact with Small Difference of Curvatures.

[0090] The profiles of the central wheel teeth are presented by the functions ζ=f(ξ) constructed according to the relations (17), and of the satellite teeth are prescribed in circular arc with radius r.

[0091] The generalizing shape parameters of the teeth contact in the gears of the mechanical transmissions are the radius of equivalent curvature of the teeth profiles and the difference parameters of the curvatures of the mating flanks.

[0092] In designing the teeth contact geometry in the precessional gear, it was admitted that LEM is a circular arc shaped-curve (FIG. 5 a, b), which prescribes the teeth profile of the satellite gear rings with sphero-spatial motion with a fixed point, and the curve E.sub.1ECE.sub.2 (FIG. 6 a) represents the flank profile of the central wheel teeth, expressed by the evolutes of the circular arc families LEM of radius r with the origin G located on the path of its motion within a precession cycle 0<ψ<2π.

[0093] To address the degree of influence of the gear geometric and kinematic parameters on the teeth contact geometry and the kinematics of their contact point in the following analyzes, analyzes for gears with concrete parameters will be presented.

[0094] FIG. 5 shows the profilogram of the flank profiles contact of the mating teeth projected on the plane P.sub.1, in which concomitantly engage 100% (FIG. 5 a) and 66.6% respectively pairs of teeth (FIG. 5 b) called frontal reference multiplicity of gear.

[0095] We admit that in the sphero-spatial movement of the satellite wheel, in the position of the crankshaft with the precession angle ψ=0, the satellite teeth circular arc profile LEM comes in contact with the active profile of the central wheel teeth E.sub.1EC in point E (FIG. 5 a) or with the active profile of the teeth E.sub.1EE.sub.N (FIG. 5 b). As the precession angle 0<ψ<π increases, the contact point E of the circular arcs LEM and of the active profile E.sub.1EC of the central wheel teeth migrate from point E.sub.1, when ψ=0, to point C, when ψ=π (FIG. 5 a), or to point E.sub.N (FIG. 5 b).

[0096] Geometrically, the location of the contact points E (FIG. 5) of the satellite wheel teeth profiles on the active profile of the central wheel teeth is defined by the precession angle ψ of the crankshaft with the location shown in FIG. 6: (a)—for the gear with frontal reference multiplicity ε.sub.f=100% (a) and (b)—for the gear with ε.sub.f=66.6%. On the curves ζ.sub.1=f(ξ.sub.1) are located the origins of the circular arcs G of the satellite teeth profile, and on the curves ζ=f(ξ)—the contacts k.sub.1,k.sub.2,k.sub.3 . . . k.sub.n of the pairs of simultaneously engaged satellite—central wheel teeth at different angular positions of the crankshaft.

[0097] The position of the origins of the circular arcs G placed on the curve ζ.sub.1=f(ξ.sub.1) denoted by p. 1, 2, 3 . . . i, correspond to the precession angles ψ of the crankshaft increasing from one pair of teeth to another with the angular pitch ψ=360.Math.Z.sub.2/Z.sub.1.sup.2.

[0098] Depending on the satellite precession phase, determined by the precession angle ψ of the crankshaft, each pair of satellite—central wheel teeth passes through three geometrical contact forms, namely from convex-concave in contacts k.sub.0, k.sub.1 and k.sub.2, located in the dedendum area of the central wheel teeth, to convex-rectilinear in contacts k.sub.3 and k.sub.4, located in the passage area of the central wheel teeth profile from concave curvature to convex and convex-convex curvature in contacts k.sub.5 . . . k.sub.14 (FIG. 6 a) and k.sub.5 . . . k.sub.8 respectively (FIG. 6 b), located in the tip area of the central wheel teeth.

[0099] According to the claims of the invention, for increasing the load-bearing capacity of the teeth contact, the convex-concave geometrical shape is proposed, and considering the classical theory of contact between deformable bodies, the difference of the radii of curvature of the conjugated tooth flank profiles must be minimal. This claim in the precessional gear transmissions is achievable by two interdependent solutions: first—by varying, selecting the configuration of parameters Z.sub.1, Z.sub.2, δ, θ and r, which determines the shape of the central wheel tooth profile, and second—by excluding from the gear the pairs of teeth with convex-convex and/or convex-rectilinear geometrical contact, with extension of the teeth contact area with convex-concave geometry.

[0100] From the analysis of FIGS. 5 and 6, we state that the convex-convex and convex-rectilinear contacts are characteristic for the flank mating with the tip area of the central wheel teeth. Using this geometrical aspect, it is possible to change the tooth shape, implicitly of the performance characteristics of the contact, by shortening its height to a level that would only provide a convex-concave contact (FIG. 7).

[0101] Modifying the shape of the central wheel tooth by shortening its height (FIG. 8), the teeth flanks mate in convex-concave contact to the limit in point k.sub.2(FIG. 8), and in the area between it and the tip of the modified tooth, the flanks mate in convex-rectilinear contact. Therefore, depending on the modified height of the central wheel teeth and the parametric configuration Z, δ, β, θ, ±1 that would provide the transformation of motion with constant transmission ratio, we can provide single, two-pair, three-pair gear, etc., i.e. we can intervene on the frontal and reference gear multiplicity.

[0102] Based on the computer simulations on virtual models, it was found that when varying the precession angle of the crankshaft 0<ψ<37°, the convex-concave contact is provided in the engaged pairs of teeth in the contacts k.sub.0, k.sub.1, k.sub.2 and k.sub.3, presented in the tooth profilogram evolute in FIG. 8.

[0103] Thus, for example, for the gear with geometric parameters Z.sub.1=29, Z.sub.2=30, R=75 mm, r=5.0 mm, θ=2.5°, δ=30°, β=3.8°, the teeth contact is characterized by the following geometry (FIG. 9): in contact k.sub.0 corresponding to the crankshaft precession angle ψ=0°, the difference of the radii of curvature between the central wheel and satellite wheel profiles ρ.sub.1−r=5.26−5.0=0.26 mm in the contact point k.sub.1 corresponding to the precession angle ψ=12.84°; ρ.sub.1−r=5.78−5.0=0.78 mm (first pair of engaged teeth); in the contact point k.sub.2 corresponding to the precession angle ψ=5.68°, ρ.sub.1−r=11.3−5.0−6.3 mm (second pair of engaged teeth); in the contact point k.sub.3 corresponding to the precession angle ψ=38.53°, ρ.sub.1−r=225−5.0=220 mm (third pair of engaged teeth, etc.).

[0104] We see that by varying the parameters Z, β, β, θ and the tooth ratio±1 by changing the shape of the central wheel teeth, we can design single, two-pair, three-pair or four-pair precessional toothed gear. In the three-pair gear shown in FIG. 9, when the crankshaft rotates, the contact point of each pair of teeth improvise an oscillatory motion along a path with the amplitude A=R.sub.1gθ, the period P=2πRZ.sub.2/Z.sub.1 and the origin in point k.sub.0, and the concomitant gear area of the load-bearing teeth extends from contact k.sub.0 to k.sub.1.

[0105] When the crankshaft rotates, each pair of teeth in contacts k.sub.i performs an improvised motion along the same path, moving imaginary, for example, from contact k.sub.0 of the satellite tooth on the bottom of the central wheel tooth (FIG. 8) to contact k.sub.0 formed by the pair of teeth (preceding) after crankshaft rotation with the angle ψ=360.Math.Z.sub.2/Z.sub.1.sup.2. In this evolution, while the position angle of the crankshaft y increases in the interval O<ψ<360.Math.Z.sub.1/Z.sub.1.sup.2, contact k.sub.1 moves to the position of contact k.sub.0 (see k.sub.0 from the previous pair of teeth, mated on the bottom of the central wheel tooth), contact k.sub.2- to k.sub.1, contact k.sub.3- to k.sub.2, and the pair of teeth preceding the first three forms a new contact k.sub.3 and so on, so that in the concomitant gear a constant number of pairs of teeth is kept. The simultaneously engaged pairs of teeth, in the precessional motion of the satellite, are kept as a constant (predetermined) number, and their contacts migrate, following the principle of similarity between them according to ψ.

[0106] In classical mechanical transmissions, to provide the transformation of motion with constant transmission ratio, it is necessary that when one pair of teeth disengages, the preceding pair is already engaged, thus the degree of overlap ε>1 is provided.

[0107] In the precessional toothed gear shown in FIG. 9, four pairs of load transmitting teeth and four passive pairs of teeth (do not transmit load), located on both sides of the contact, are concomitantly engaging. When the crankshaft rotates, the engaged pair of teeth in contact k.sub.0 disengages, and the pair with position 5 forms a new load-bearing contact k.sub.4, thus constantly maintaining four pairs of load-bearing teeth.

[0108] According to FIG. 9, each of the four simultaneously engaged pairs of teeth has angular coordinates expressed by crankshaft positioning according to the center angles ψ.sub.k.sub.1, . . . ψ.sub.k.sub.4, rising from contact to contact with the pitch ψ=360Z.sub.2/Z.sub.1.sup.2. All four pairs of teeth required with load rotate around the axis Z with the angular velocity ψ and the starting coordinate located in the plane P passing through the contact k.sub.0.

[0109] FIG. 9 shows the positions of contacts k.sub.0 . . . k.sub.4 and point 5 on the satellite teeth profile corresponding to the positioning angles ψ.sub.k, =0°, ψ.sub.k1=15.6°, ψ.sub.k2=31.2°, ψ.sub.k346.8°, ψ.sub.k4=62.4° and ψ.sub.k5=78.0°, determined from the relation ψ.sub.k=360iZ.sub.2/Z.sub.1.sup.2, where i=0,1,2,3,4 . . . is the contact order number. The difference in the radii of curvature of the engaged flanks is calculated by alternation, varying the geometric parameters Z, δ, β, θ and the teeth ratio±1.

[0110] It is worth mentioning that analogously with the precessional toothed gear with four simultaneously engaged pairs of teeth shown in FIG. 9, gears with three, two and one pair of engaging teeth can be designed, correspondingly changing the shape of the central wheel and satellite teeth profile by respectively shortening the height of the teeth of both engaged wheels.

[0111] 3. Influence of the Ratio of the Numbers of Teeth of the Mating Wheels on the Kinematics of the Contact Point and Shape of the Tooth Flank Profile.

[0112] In precessional toothed gears, unlike those with bolts, the transformation and transmission of motion and load occur with the presence of relative frictional sliding between the teeth flanks, depending on the kinematics of the teeth contact point, in particular on the ratio of the numbers of teeth of the mating gear rings Z.sub.1=Z.sub.2−1 or Z.sub.1=Z.sub.2+1.

[0113] Therefore, the calculation and design of precessional toothed gears, unlike classical, including precessional with bolts, include a separate algorithm for designing the teeth contact geometry, which generally defines the load-load-bearing capacity and mechanical efficiency of the transmission.

[0114] The design of the teeth contact geometry from the precessional toothed gear is limited to the identification of the contact form (see FIG. 10 a, b) and the parameters of its geometry, determination of the kinematics of the flanks contact point considered as tribosystem—all being subjected to the purpose of increasing the load-load-bearing capacity and mechanical efficiency of the teeth contact.

[0115] FIG. 11 (a) shows the profilogram of the mating wheel teeth for the configuration of parameters Z.sub.1=24, Z.sub.2=25, θ=3.5°, δ=22.5°, r=6.27 mm, R=75 mm and the ratio of the number of teeth Z.sub.1=Z.sub.2−1, in FIG. 11 (b) in the configuration of parameters, the ratio of the number of teeth Z.sub.1=Z.sub.2+1 differs, i.e. Z.sub.1=25 and Z.sub.2=24 and δ=0°, and in FIG. 11 (c), the ratio of the number of teeth Z.sub.1=Z.sub.2+1 and the conical axoid angle δ=22.30° differs.

[0116] 4. Reduction of the Pressure Angle Between the Mating Flank Profiles.

[0117] From FIG. 12 we state that in the precessional gear depending on the parameters Z, δ, θ, r and Z.sub.1=Z.sub.2−1 (a) and the reference multiplicity ε.sub.f=100% the contact points k.sub.0 . . . k.sub.5 of the teeth flanks are placed on the portion of the central wheel teeth profile with the pressure angle between the flanks α=31°, and for parameters (b) and ε.sub.f=73% α=14°. Decreasing the pressure angle α leads to a decrease in the static and dynamic load from the shaft and satellite wheel bearings.

[0118] So, unlike the classical ones, in the precessional transmission the profile of the central wheel teeth is variable, which leads to the variation of the teeth contact geometry in one and the same gear, passing from one form to another, namely from convex-concave at the dedendum of central wheel tooth to convex-rectilinear towards the middle of the tooth and convex-convex towards the tip of tooth.

[0119] 5. Relative Sliding Between the Teeth Flanks in Gear.

[0120] The kinematics of the teeth contact point in precessional gear and the geometric shape of the mating flanks are two determining characteristics of the mechanical efficiency and the load-load-bearing capacity of the contact.

[0121] The mechanical efficiency of the gear is the expression of energy losses generated by the frictional sliding forces between the mating flanks, and the load-bearing capacity of the convex-concave contact results from the size of the difference in their radii of curvature.

[0122] For these reasons, the gear contact kinematics and geometry (FIG. 13) are examined for gears with different parametric configurations Z, δ, β, θ between them only by the ratio of the numbers of teeth Z.sub.1=Z.sub.2±1 and the conical axoid angles δ≥0°. From the aforesaid, the generalized configuration can be expressed by the parameters Z.sub.1=24(25), Z.sub.2=25(24), θ=3.5°, δ=22.5° (0), r=6.27 mm and R=75 mm.

[0123] The analysis of kinematics in the contact points k.sub.0, k.sub.1, k.sub.2 . . . k.sub.i corresponding to the crankshaft positioning angles takes place by varying the linear velocities of the contact points E.sub.1 on the central wheel teeth profile and E.sub.2 on the satellite teeth profile and the relative sliding velocity between the flanks V.sub.ol, and the teeth contact geometry is presented through the radii of curvature ρ.sub.k.sub.i of the central wheel teeth profile and the satellite teeth profile r and their difference (ρ.sub.1−r). Analysis of the teeth contact kinematics is performed for the crankshaft speed n.sub.1=0.3000 min.

[0124] Thus, in the gear Z, δ, β, θ with the ratio of the numbers of teeth Z.sub.1=Z.sub.2−1 and the conical axoid angle δ=22.5°, shown in FIG. 14 (a), in the teeth contact k.sub.0 the linear velocity is V.sub.ε.sub.1=9.83 mA, V.sub.ε.sub.2=9.69 m/s, V.sub.alkθ=0.14 m/s and the radius of curvature of the central wheel teeth profile is ρ.sub.k0=6.43 mm of the satellite teeth profile r=6.27 mm and their difference (ρ.sub.k0−r)=0.16 mm (FIG. 14 b).

[0125] As the angular coordinate increases from one mating pair to the other with the pitch ψ=360iZ.sub.1/Z.sub.1, for example, from the angular coordinate ψ.sub.k.sub.o=0° up to ψ=15,6° assigned to contact k.sub.1, the linear velocities V.sub.E.sub.1 and V.sub.E.sub.2 decrease, registering in contact k.sub.1 the difference V.sub.alki V.sub.al=, V.sub.ε.sub.1.sub.k1−V.sub.E.sub.2.sub.k2=0.34 m/s and the difference of the radii of curvature of the mating flanks in (ρ.sub.k2−r)=1.17 mm in contacts k.sub.2 corresponding to ψ=31.2° V.sub.al4=0.67 m/s and the difference of the radii of curvature (ρ.sub.k2−r)=9.55 mm; in the contact k corresponding to ψ=46.8° m/s and the teeth contact geometry passes from convex-concave to convex-convex, with the radius of external curvature of the central wheel teeth profile ρ.sub.k3=57.66 mm. FIG. 14 (c) shows the evolution of the geometry from contact k.sub.0 to contact k.sub.4.

[0126] Table 1 presents the argumentation and justification of the limits of variation of the frontal overlap degree values ε.sub.f of the pairs of teeth that are concomitantly in the gear field, of the conical axoid angle δ, of the nutation angle θ between the axes of the crank and central conical wheels, as well as of the circular arc radius r of the flank profile of Z teeth of the satellite wheel gear ring in the section with diameter D, which generally provides for the mating of teeth in convex-concave contact and the reduction in the difference of curvatures of the mating flanks and the relative sliding velocity in the teeth contacts.

Argumentation of the Limits of Variation of the Precessional Gear Parameters According to Claim 1

[0127]

TABLE-US-00001 TABLE 1 Parameter Lower limit Upper limit Note Degree of ε.sub.f = 1.5 pairs of teeth. ε.sub.f = 4.0 pairs of teeth. frontal Decreasing the ε.sub.f < 1,5 leads to Increasing the ε.sub.f > 4,0 leads to overlap ε.sub.f of sensitization of the influence of the increase of relative teeth that are teeth deformability (other frictional sliding in the teeth concomitantly elements of the gear) and contacts and the difference in in the gear technological errors of curvatures of the mating field. execution (of the teeth profile flanks, which favors the and pitch, etc.) on the kinematic increase of energetic losses in precision of the gear, as well as gear and the diminution of on following the basic principle mechanical efficiency. of the fundamental law of gearing ω.sub.5/ω.sub.8 = const. Gear conical δ = 0°, degrees. If Z.sub.1(4) = Z.sub.2(3) − 1, δ = 30°, degrees. If axoid angle δ. decreasing the bevel axoid angle Z.sub.1(4) = Z.sub.2(3) − 1, increasing bevel by δ < 0° leads to an increase in axoid angle by δ > 30° leads to the radius of curvature of the the interference of central flank profiles of the central wheel teeth profiles and wheel teeth in the contact points trajectories of the origin of the and, respectively, to an increase radius of curvature of the in the difference of flank teeth circular arc profiles of curvatures in the contact points, the satellite wheel gear rings. because r = const, and the load- bearing capacity and mechanical efficiency are decreased by θ = 1.5°, degrees. Nutation Decreasing the nutation angle θ = 1.5°, degrees. Increasing angle θ θ < 1,5° leads to an increase in the nutation angle leads to an between the the pressure angle between the increase in the radius of axes of the mating flanks, favoring the curvature of the flank profile crank and increase of load in the bearings of the central wheel teeth in central bevel of the satellite wheel, drive and the contact points of the first wheels. driven shafts, including energy four pairs of teeth in the gear losses in gears. field and to an increase in the dynamic of the load in gear. The circular r = 1.0 D/Z, mm Exceeding the value of the arc radius r of Decreasing the radius of radius r > 1.57 leads to the the flank curvature r = 1.0 D/Z leads to non-compliance with the ratio profile of Z the transformation of the teeth of the teeth pitch lengths of teeth of the flanks of the four pairs of teeth the satellite wheel and central satellite in the gear field with contact wheel rings, proceeding from wheel gear with convex-concave geometry the condition of the ratio of ring in the in contact with convex- the number of teeth section with rectilinear or convex-convex Z.sub.1(4) = Z.sub.2(3) − 1. diameter D. geometry.

[0128] Variation of the frontal overlap within the limits 1.5≤ε.sub.f≤4.0 pairs of teeth that are concomitantly in the gear field, of the bevel axoid angle within the limits 0°≤δ≤30° and of the nutation angle within the limits 1.5≤θ≤7°, as well as of the circular are radius r of the tooth flank profiles of the satellite wheel gear rings within the limits 1.0 D/Z, mm≤r≤1.57 D/Z, mm, provides for the existence of convex-concave geometry in the contacts of the pairs of teeth located in the gear area with the decrease in the difference of curvatures by up to (0.02-1.5) D/Z, mm and of the pressure angle α between the flanks by up to 15°, as well as the decrease of the relative sliding velocity between the mating flanks.

[0129] These technical solutions favor the increase of the load-bearing capacity and mechanical efficiency of the transmission.

[0130] Another difference of the transmission according to claim 2 consists in that the teeth of the fixed 6 and mobile 7 central wheels, as well as of the gear rings 3 and 4 of the satellite wheel 2 are inclined, which provides for the increase of the pure rolling share of the engaged teeth flanks with sphero-spatial interaction dependent on the nutation θ and inclination β angles, and the increase of the total length of the contact lines, with their gradual entry into the gear field.

[0131] According to claim 2, the total teeth contact line l.sub.Σ in the gear with inclined teeth is determined from the condition of frontal gear ε.sub.j of a certain number of pairs of teeth (ε.sub.f=1,2,3 . . . ), but not less than one pair (ε.sub.f,min=1). In the case of ε.sub.f,min=1 it turns out that a pair of teeth engages, while the previous pair disengages.

[0132] According to the condition of providing continuity of gear and the slow course of the transmission, it is necessary that the tooth overlap degree to be ε.sub.m>1. Thus, in the case of ε.sub.f,min=1 it is proposed to incline the teeth at the angle β.sub.g, which would ensure a degree of longitudinal (axial) overlap.

[00006]$\begin{array}{cc}{\epsilon }_f^{\beta }=\frac{b_w{Z}_1\sin {\beta }_g}{2\pi Z_2}.& \left(19\right)\end{array}$

[0133] FIG. 15 shows the length, variation and positioning of the contact lines of the inclined engaged teeth within the overlap area, which extends to the center angle α.

[0134] From the analysis of the succession of the entry and exit of the tooth pairs from the gear area, we state that the degree of overlap of the engaging teeth and, respectively, the total length of the contact lines of the engaged teeth depend on the frontal overlap ε.sub.f.sup.B, determined by the frontal gear multiplicity ε.sub.f and the longitudinal overlap ε.sub.a dependent on the teeth inclination angle β.sub.2, including the configuration parameters Z, δ, θ and the ratio± of the mating teeth, and on the modification of the teeth height. It is also observed that the contact lines between the inclined teeth are positioned in space so that their extensions are tangent to the cylinder with radius e.

[0135] It should be mentioned that the inclination of teeth leads to the diminution of the frictional sliding in the engaged teeth contact, because the teeth mating for the same parameters of the configuration Z, δ, β, θ and Z.sub.1=Z.sub.2±1 takes place with an increased share of pure rolling of teeth depending on the angle θ.

[0136] Unlike straight teeth, the inclined ones do not engage concomitantly along the entire length, but gradually with a certain angle offset ψ depending on the inclination angle β and the tooth length b.sub.w.

[0137] The position of contact lines in the gear with concave-concave contact of the teeth inclined within the limits of the gear field is shown in FIG. 16. Upon rotation of the crankshaft ω.sub.1, the contact lines of the engaged teeth move in the gear field in the direction indicated by arrows A and B.

[0138] In FIG. 16 (a) with a pair of teeth in frontal gear ε.sub.f=1, in the gear field are covered three pairs of inclined teeth, where pair 2 contacts along the entire length 2-2′ of the teeth, pair 1—along the length 1-1′, and pair 3—3-3′.

[0139] When the crankshaft positioning angle Δψ is increased (FIG. 16, a), the length of the contact line 3-3′ increases by Δl by moving point 3′ to point 3″, and the length of the contact line 1-1′ decreases by the same length Δl by moving point 1 to 1′. The evolution of the total length of the contact lines 1-1′, 2-2′ and 3-3′ for any value ψ remains constant, l.sub.Σ=const.

[0140] In the case of gearing with two pairs of teeth in frontal gear ε.sub.f=2 shown in FIG. 16 (b), the teeth pairs with contact 2 and 3 along the entire teeth length are present in the gear field. When the crankshaft rotates with the angular value Δψ, the length of the contact line 4-4′increases by Δl by moving point 4′ to 4″, and the length of the contact line 1-1′ decreases by the same length Δl by moving point 1 to 1′. The total length of the contact lines for any angle value ψ is constant, l.sub.Σ=const.

[0141] In the precessional gear, the inclined teeth are loaded gradually, as they enter the gear field, and in permanent gear there are at least two pairs of teeth:

ε.sub.m=ε.sub.f.sup.β+ε.sub.a.sup.β.  (20)

[0142] The precessional gear with inclined teeth can also work without frontal overlap, thus with ε.sub.f.sup.β>1, if the axial overlap ε.sub.β is ensured, i.e. b.sub.w>(2πrZ.sub.3)/(Z.sub.1 tgβ). In the precessional gear with inclined teeth, the load between simultaneously engaged teeth is distributed proportionally to the contact line lengths of the required teeth pairs with load.

[0143] Obviously, the specific teeth load q decreases with the increase of the total length of the contact lines l.sub.Σ=ε.sub.mb.sub.w sin δ/cos β, and l.sub.Σ does not change over time, because decreasing the length of the teeth contact line 1-1′ in any position ψ of the crankshaft is compensated by an equal increase in the length of the contact line 3-3′ (FIG. 16). Obviously, in case of compliance with l.sub.Σ=const, the teeth load will not change over time, and the noise emission and dynamic loads will decrease.

[0144] At the same time, we can state that in gear the convex-concave contact of the mating teeth in the frontal gear (FIG. 14) is formed of flank profiles with the small difference in the radii of curvature (FIG. 14 b), and for the same tooth width, the length of the contact lines increases, which leads to the decrease of the specific teeth load.

[0145] The maximum effect produced by the inclined teeth of the gear consists in the essential decrease of the relative sliding velocity V.sub.al, between the flanks (FIG. 15), due to its replacement with the pure rolling of teeth in shares provided by the angle ψ, (FIG. 16 a, b) dependent on the teeth inclination angle β, teeth length b.sub.w and nutation angle θ of the sphero-spatial motion of the satellite wheel.

[0146] The optimal choice (see FIG. 14) of the mating reference pair of teeth in the point with the crankshaft positioning angle ψ.sub.i is based on three considerations, namely: the difference in the radii of curvature of the flank profiles in contact (ρ.sub.k−r)=min, the sliding velocity between the flanks in contact V.sub.al=min, the pressure angle α.sub.w of the teeth flank profile of the central wheels α.sub.w=min. All these geometry and kinematics parameters of the teeth contact are according to the precession angle ψ.

[0147] From the analysis of FIG. 14 we observe that the conditions (ρ.sub.k−r)=min and V.sub.al=min can be achieved by reducing the angle ψ, and α.sub.w=min—by increasing the angle ψ. These three conditions define geometrically and kinematically the contact parameters of the gear flanks, which would provide high efficiency and load-bearing capacity and minimum static demand for the crankshaft and satellite wheel supports.

[0148] The third difference of the claimed transmission (FIG. 17) consists in that one of the bevel gear rings 3 or 4 of the satellite wheel 2 has the conical axoid angle δ=0° and is made of bolts with one less or more than the number of teeth of the central bevel wheel with which it engages, which provides for the pressure angle between the mating flanks α≤45° and the extension of kinematic possibilities.

[0149] According to claim 3, in FIG. 17, the gear (Z.sub.3-Z.sub.4) is made flat with the conical axoid angle δ=0° from the toothed crown 4 executed in the form of bolts with one less or more than the number of teeth of the central wheel 7 with which it engages Z.sub.4=Z.sub.3±1.

[0150] Thus, the difference of the transmission (FIG. 17) consists in the constructive specific character of the satellite wheel 2 in which the gear (Z.sub.1-Z.sub.2) is geometrically analogous to the gear (Z.sub.1-Z.sub.2) of the transmission in FIG. 2 (b), and the gear (Z.sub.3-Z.sub.4) is made of toothed crown 4 made in the form of bolts placed in a plane ring with δ=0. Theoretically and by computer simulations based on mathematical models, it was found that in the plane gears with δ=0 the ratio of the numbers of teeth Z.sub.4=Z.sub.3+1 or Z.sub.4=Z.sub.3−1 do not influence the shape of the central wheel teeth profile 7 and respectively the teeth contact geometry. In the case of Z.sub.4=Z.sub.3+1 it is noticed the greater presence of the relative frictional sliding in contact, which is excluded from the teeth contact area by using the toothed crown 4 executed in the form of bolts.

[0151] Claim 4.

[0152] The fourth difference of the claimed transmission consists in that at least one of the bevel gear rings 3 or 4 of the satellite wheel 2 is made of bolts with one less than the number of teeth of the engaged bevel central wheel and has the conical axoid angle δ3>0° (FIG. 18), which provides for the pressure angle between the mating flanks α>45° and the operation of the transmission in multiplier mode.

[0153] It is worth mentioning that, according to FIG. 18, the gear (Z.sub.3-Z.sub.4) is also made of bolts, but with the conical axoid angle δ>0 and with a bolt less than the number of teeth of the central wheel with which it mates.

[0154] This configuration with δ>0 and Z.sub.4=Z.sub.3+1 provides for the increase of the pressure angle α between the flanks of the central wheel teeth 7 and the toothed crown 4 made in the form of bolts of the satellite wheel 2, which favors, from the point of view of energy losses, the transformation of the rotational motion of the driving shaft 5 (which replaces the function of the crankshaft) in sphero-spatial motion of the satellite wheel 2 by using the inclined slope effect.

[0155] Thus, the solution according to FIG. 18 with δ>0° and Z.sub.4=Z.sub.3+1 provides for the operation of the transmission in multiplier operation mode by multiplying the revolutions from shaft 8 to shaft 5.

[0156] Claim

[0157] The fifth difference of the claimed transmission consists in that the satellite wheel 2 (FIG. 20 a) is installed on a spherical support 9 placed on the driven shaft 8 in its precession center and coaxially with the mobile central bevel wheel 7, at the same time the satellite wheel 2 is equipped with a semi-axle 10, at the end of which is mounted a bearing 11, kinematically connected to the crank 5 installed on the driving shaft of the electric motor.

[0158] The precessional gear reducer shown in FIG. 20 (a) and the motor-reducer shown in FIG. 20 (b) operate in the following mode.

[0159] The rotational motion of the crank 5 (or the electric motor) is transformed in sphero-spatial motion of the satellite wheel 2 by means of the bearing 11, mounted on the end of the semi-axle 10 of the satellite wheel 2, which, in turn, is mounted in the seat of the crank 5. The satellite 2 involved in the sphero-spatial motion with the frequency of precession cycles respectively with the teeth of the immobile 6 and mobile 7 central wheels. As a result, the driven shaft will rotate with reduced rotational frequency with the transmission ratio

[00007]$\begin{array}{cc}{i}_{\mathrm{HV}}^b=-\frac{Z_2{Z}_4}{Z_1{Z}_3-Z_2{Z}_4}& \left(20\right)\end{array}$

INDUSTRIAL APPLICABILITY

[0160] Thus, the technical solutions set forth in claims 1-5 provide for the increase of the load-bearing capacity and mechanical efficiency, as well as the extension of the kinematic and functional possibilities.

[0161] The load-bearing capacity of the mechanical transmission gears depends on the degree of overlap and the contact geometry of the engaging teeth.

[0162] Based on these considerations, the analysis of the load-bearing capacity of the precessional gear transmission according to the invention, in comparison with the most efficient existing transmissions, for example Wildhaber-Novicov (W-N) shows the following:

[0163] 1. In case of compliance with the similarity of the “convex-concave” contact geometry with equal diameters of the gears, the difference in the radii of curvature in the gear (W-N) is estimated by (R.sub.1-R.sub.2) m.sub.n=(1.55−1.4) mm, m=0.75 mm, and in the claimed precessional gear, the difference in the curvatures of the flanks in the first three pairs of teeth (ρ.sub.ki−r) respectively is 0.16 mm, 1.17 mm, 9.55 mm (see FIG. 14).

[0164] It is also worth mentioning that in the gear (W-N) the frontal overlap of the teeth is ε.sub.f=(0.85-0.95) pairs of teeth, and in the precessional gear transmission according to the proposed invention is ε.sub.f=(1.5-4.0) pairs of teeth concomitantly in the gear field.

[0165] 2. The mechanical efficiency of a gear with gear wheels depends on the relative frictional sliding velocity between the mating flanks. From the analysis of the graphs presented in FIG. 14 it is obvious that the relative sliding velocity in the first three mating pairs of teeth in the precessional gear is lower than in the classical evolvent gears including in the gear (W-N).

[0166] 3. Concerning the kinematic possibilities, the precessional gear transmission at the present time has no analogues among the worldwide known transmissions.

BIBLIOGRAPHICAL REFERENCES

[0167] 1. SU 1455094 A1 1989.01.30 [0168] 2. SU 1758322 A1 1992.08.30