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.

Description

TECHNICAL FIELD

The invention relates to mechanical engineering, namely to precessional planetary transmissions.

BACKGROUND ART

It is known a precessional gear transmission wherein the satellite wheel has teeth with rectilinear profile, and the central wheel—with circular arc profile with the radius origin placed on the normal raised from the contact point of the mating teeth passing through the intersection point of the tooth evolute slope line with circular arc profile with rectilinear profile equidistance[1].

The disadvantage of the studied transmission consists in the execution of the tooth profiles with approximation, which leads to the diminution of the kinematic precision of the transmission and the mating of teeth with the presence of frictional sliding between the flanks of the mating teeth, which implies an increase of the energetic losses in the gear and a decrease in the mechanical efficiency.

It is also known the precessional gear transmission of conical bolts, including a body, a crankshaft and a coaxially driven shaft, 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 satellite wheel rings are made of conical bolts, assembled with springs between them, which provide axial and radial flotation to gear rings, and, consequently, the diminution of the impact of the execution and mounting errors on the load distribution in the bolt gear.

The embodiment of the precessional transmission with bolt gear is characterized by disadvantages, which limit the extension of their use by the following:

• • The load-bearing capacity of the “conical bolt-tooth” contact is limited by the median radius of the conical bolts, which cannot exceed half of the teeth pitch, and the teeth contact in most bearing pairs is convex-convex and convex-rectilinear.
  • The conical bolts require high precision of fabrication of the dimensions and of individual axial positioning thereof, on which the uniformity of load distribution between the simultaneously engaged pairs of teeth depends.
  • The satellite rings made of conical bolts make it irrational, difficult and sometimes impossible to manufacture gears with diameters of ≤50 mm.
  • The production and assembly cost of the gear with bolts is relatively higher and requires high precision of execution and assembly.

The technical problem consists in creating a precessional gear transmission, which would provide an increase in the load-bearing capacity and mechanical efficiency, extension of kinematic and functional possibilities, as well as extension of the scope of transmission.

DISCLOSURE OF INVENTION

The invention, according to the first embodiment, removes the aforesaid disadvantages by comprising a body, a crankshaft and a coaxially driven shaft, 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. The novelty consists in that the teeth engage in contacts with convex-concave geometry, wherein the central bevel wheels are made with straight teeth and have curvilinear flank profiles with variable curvature with one tooth less than the satellite wheel rings made with circular arc flank profiles, the teeth flanks mate with frontal overlap sf within the limits 1.5≤ε_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≤97° and the circular are radius of the flank profile of the Z-toothed satellite wheel gear ring 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, of the pressure angle α between the flanks of up to 15° as well as a decrease in the relative friction velocity between the mating flanks.

Second, 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.

Third, one of the satellite wheel bevel 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.

Fourth, one of the satellite wheel 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 for the rolling of the conical bolts on the flank profile of the central wheel teeth with inclined slope effect and the operation of the transmission in multiplier mode.

Fifth, 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.

The precessional gear transmission, according to the second embodiment, comprises a body, a coaxial crankshaft and driven shaft, a satellite wheel with two bevel gear rings, mobile and immobile central bevel wheels.

The transmission with precessional gear, according to the second variant, includes housing, crank shaft and coaxial driven shaft, satellite rate with two conical gear crowns, movable and immobile center conical wheels.

The novelty consists in that the transmission consists of at least two satellite wheels kinematically interconnected in series by means of at least an intermediate crankshaft installed in cantilever on bearings in the body, which is laterally equipped with an offset seat at the nutation angle θ with the common axis of the central wheels, the first satellite wheel by means of a bearing mounted on the end of its semi-axle is kinematically coupled with the crankshaft, and the second satellite wheel by means of a bearing mounted on the end of its semi-axle is kinematically coupled with the offset seat of the intermediate crankshaft at the nutation angle θ with the common axis of the central wheels.

The technical result consists in:

• • Increasing the load-bearing capacity of the transmission by engaging the teeth in contacts with the convex-concave geometry and the minimum difference in the curvatures of the mating flanks, including by increasing the length of the total contact line of the inclined teeth;
  • Increasing the mechanical efficiency by changing the tooth shape, reducing the pressure angle between the flanks and at the expense of increasing the rolling share of the engaging teeth by decreasing the relative frictional sliding between the flanks with a reduction in the frontal overlap degree and a compensatory increase in the longitudinal overlap degree with pure rolling of teeth in the sphero-spatial interaction of the mating wheels with the nutation angle θ;
  • Extending the kinematic and technological possibilities.

BRIEF DESCRIPTION OF DRAWINGS

Summary of the invention is explained by the drawings that represent:

FIG. 1. Precessional toothed gear transmission.

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

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

FIG. 4. Description of the flank profile of the central wheel teeth.

FIG. 5. Teeth contact in the precessional toothed gear with frontal reference multiplicity ε_f=100% (a) and ε_f=66.6% (b).

FIG. 6. Evolution of the contact point of teeth in the gears with frontal reference multiplicity ε_f=100% (a) and ε_f=66.6% (b).

FIG. 7. Geometry of the modified teeth by making the tips in the gear with frontal reference multiplicity ε_f=100% (a) and ε_f=66.6% (b).

FIG. 8. Convex-concave gear with slight difference in the curvatures of the mating flank profiles with frontal multiplicity ε_f=27.58%.

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

FIG. 10. Kinematics and tooth profiles of the gears (Z₁-Z₂) and (Z₃-Z₄) with reduction gearbox (a) and multiplier (b) operation modes.

FIG. 11. Geometry of the teeth contact and relative positioning of the mating flanks in contact k₁ for: a) Z₁=Z₁−1, δ=22.5°; b) Z₁=Z₂+1, δ=0°, c) Z₁=Z₂+1, δ=22.5°.

FIG. 12. Profiles and pressure angle between the flanks of the central wheel teeth with reference multiplicity ε_f=100% (a) and ε_j=73% (b).

FIG. 13. Kinematics and geometry of the teeth contact in the gear (Z₃-Z₄) with (Z₄=Z₃−1) (a) and gear (Z₁-Z₂) with (Z₁=Z₂−1) (c) with three simultaneously engaged pairs of teeth.

FIG. 14. Linear velocities in the contact point V_E1, V_E2, V_a1 (a) and difference in the radii of curvature (ρ_ki-r) (b) of the mating profiles in contact k₁ (c) depending on ψ for Z₁=Z₂−1 and δ=22.5°, Z₁=24, Z₂=25, θ=3.5°, δ=22.5°, r=6.27 mm and R=75 mm.

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

FIG. 16 Contact of the inclined teeth with the angle β_g placed in the gear field with a pair of frontal mating teeth (a) and with two pairs (b).

FIG. 17. Precessional transmission with the gear (Z₃-Z₄) of bolts (δ=0) with the pressure angle between the flanks α≤45° for the ratio of the numbers of mating teeth Z₄=Z₃−1.

FIG. 18. Precessional transmission with the gear (Z₃-Z₄) of bolts (δ>0) with the pressure angle between the flanks for the ratio of the numbers of mating teeth Z₄=Z₃+1 (for multiplier operation mode).

FIG. 19. Bolt-tooth interaction in the gear Z₄=Z₃+1 with multiplier operation mode.

FIG. 20. Precessional toothed gear transmission with convex-concave contact and difference in the numbers of mating teeth Z₁₍₄₎==Z₁₍₃₎−1 according to claim 5.

MODES OF CARRYING OUT OF INVENTION

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_f within the limits 1.5≤ε_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.

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 (Z₁-Z₂) and (Z₃-Z₄), 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_f, the size of the pressure angle α between the mating flanks and the frictional sliding velocity between the flanks.

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).

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.

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

Z ₁ =Z ₂−1 and Z₄=Z₃−1  (1)

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_HV^b:

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

Generally, when transmitting the motion and load through the gears (Z₁-Z₂) and (Z₃-Z₄) with the ratio of the numbers of teeth Z₁₍₄₎=Z₂₍₃₎±1, the direction of rotation of the driven shaft 8 coincides or not with the direction of the input shaft 5.

If Z₂>Z₃, the crankshaft 5 and the driven shaft 8 rotate counterclockwise, and if Z₂<Z₃—in the same direction.

Frontal multiplicity of the mating wheel teeth gearing in the precessional transmission is determined by three interdependent constructive-kinematic conditions.

• • 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;
  • the difference between the number of teeth of the engaged wheels is Z₁=Z₂±1 and Z₄=Z₃±1, and the difference between the number of teeth of the satellite gear rings can be Z₂=Z₃±1;
  • compliance with the continuity of the rotational motion transformation function, therefore ω₁/ω₈=const.

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).

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:

• • creating contacts between the teeth flanks with convex-concave geometry with small difference of curvatures;
  • providing the minimum pressure angles between the flanks of the engaged teeth;
  • providing the minimum relative frictional sliding velocity between the mating flanks;
  • 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.

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.

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

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

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.

From the Euler equations, taking into account the kinematic relation between the angles φ and ψ expressed by φ=−Z₁/Z₂ψ (2), we obtain the coordinates of the origin G of the circular arc radius X_G, Y_G, Z_G depending on the rotation angle ψ of the crankshaft:

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

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 ζ₁=f(ξ₁), expressed by the coordinates X_G, Y_G, Z_G (FIG. 3).

The path of motion G of the circular arc LEM on the sphere with radius R is projected on the plane P₁ using the rules of spherical trigonometry. Thus, it is obtained the path T_G, of motion of the origin of the circular arc G radius on the plane P₁, expressed by the dependence ζ=f(ξ₁).

Knowing the path of motion of the origin of the circular arc G radius, expressed in the coordinates X_G, Y_G, Z_G (FIG. 4), one can determine the position of the contact point E of the flank profiles of the mating teeth in the gears (Z₁-Z₂) and (Z₃-Z₄) for any angular position ψ of the crankshaft 5.

The family of contact points E obtained in a precession cycle 0<ψ<2πZ₂/Z₁ represents the profile of the teeth of the immobile 6 or mobile 7 central wheels.

To describe the flank profiles of the central wheel teeth 6 and 7, the projections of the velocity vector V_G on the coordinate axes of the mobile system OX₁Y₁Z₁, are determined depending on the angular velocity of the crankshaft 5 (see FIGS. 1 and 2).

To determine the position of the contact point E of the teeth on the spherical surface, we identify the equation of a plane P₂ drawn perpendicular to the velocity vector V_G, passing through the center of precession O and the origin of the circular arc radius G. The equation of plane P₂ can be written by the expression:

[OG×OC]×V_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 P₂ with respect to the origin of the immobile coordinate system OX Y Z (FIG. 3).

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_GZ−Z_GY)+j(Z_GX−X_GZ)+k(X_GY−Y_GX),  (7)

wherein X_G, Y_G, Z_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 P₂.

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

X _E ² +Y _E ² +Z _K ² −R ²=0.  (8)

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·OE=R ² cos β  (9)

or

X _E Z _G +X _E Y _G +Z _E Z _G −R ² cos β=0.  (10)

From the equation (12) we determine:

X _E=(R² cos β−Y_EY_G−Z_EZ_G)/X_G.  (11)

To determine the coordinate Y_E of the contact point of the teeth E, we substitute (11) in (8) and obtain:

Y _E =k ₁ Z _E −d ₁,  (12)

and by substituting (12) into (11), we obtain the expression of the contact point coordinate X_E:

X _E =k ₂ Z _E +d ₂,  (13)

where

k ₁=[X_G(X_G·{dot over (X)}_G+Y_G{dot over (Y)}_G+Z_G²{dot over (X)}_G]/(X_G{dot over (Y)}_G−Y_G{dot over (X)}_G)Z_G

d ₁ =R ² cos β{dot over (X)}_G(X_G{dot over (Y)}_G−Y_G{dot over (X)}_G)

k ₂=−(k₁Y_G+Z_G)/X_G

d ₂=(R² cos β+d₁Y_G)X_G.  (14)

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_E<Z_G must be met, the coordinate Z_E can be determined by the relation:

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

Relationships (12), (13) and (15) determine the coordinates X_E, Y_E and Z_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.

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 Z₂, the ratio of the numbers of teeth of the engaged wheels Z₁=Z₂−1 or Z₁=Z₂+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.

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 P₁ drawn by the points E₁ and E₂ perpendicular to the plane OE₁E₂(FIG. 4).

The coordinates X_E, Y_E, and Z_E of points E₁ and E₂ 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=2πZ₂/Z₁, corresponding to a precession cycle.

Using the rules of spherical trigonometry, we design the teeth profile on the sphere with the radius R on the plane P₁.

To design the profile of the central wheel teeth in two coordinates ζ and ξ in the plane P₁ we draw the coordinate system E₁ξζ with the origin in point E₁=, whose axis E₁ξ passes through point E₂ (FIG. 5). From coordinates X_N, Y_N and Z_N we pass to coordinates ζ and ξ using the relations:

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

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 P₁, expressed in parametric form with the variation of the precession angle from ψ=0 to ψ=2πZ₂/Z₁².

To design the path of motion of the origin of the circular arcs G in 2D, we pass from coordinates X_N, Y_N and Z_N to Cartesian coordinates ξ₁, ζ₁ using the relations:

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

Function ξ₁ of ζ₁ (17) represents the projection of the path of motion of the origin of the circular arcs G on the plane P₁, and function ξ of ζ (16) represents the flank profile of the central wheel teeth projected on the plane P₁.

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 (Z₁-Z₂) and (Z₃-Z₄) may be with the same or different reference multiplicity of the teeth gearing.

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.

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.

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.

In designing the teeth contact geometry in the precessional gear, it was admitted that LEM is a circular arc shaped-curve (FIG. 5a, b), which prescribes the teeth profile of the satellite gear rings with sphero-spatial motion with a fixed point, and the curve E₁ECE₂ (FIG. 6a) 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.

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.

FIG. 5 shows the profilogram of the flank profiles contact of the mating teeth projected on the plane P₁, in which concomitantly engage 100% (FIG. 5a) and 66.6% respectively pairs of teeth (FIG. 5b) called frontal reference multiplicity of gear.

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 E₁EC in point E (FIG. 5a) or with the active profile of the teeth E₁EE_N (FIG. 5b). As the precession angle 0<ψ<π increases, the contact point E of the circular arcs LEM and of the active profile E₁EC of the central wheel teeth migrate from point E₁, when ψ=0, to point C, when ψ=π (FIG. 5a), or to point E_N (FIG. 5b).

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_f=100% (a) and (b)—for the gear with ε_f=66.6%. On the curves ζ₁=f(ξ₁) are located the origins of the circular arcs G of the satellite teeth profile, and on the curves ζ=f(ξ)—the contacts k₁,k₂,k₃ . . . k_n of the pairs of simultaneously engaged satellite—central wheel teeth at different angular positions of the crankshaft.

The position of the origins of the circular arcs G placed on the curve ζ₁=f(ξ₁) 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·Z₂/Z₁².

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 k₀, k₁ and k₂, located in the dedendum area of the central wheel teeth, to convex-rectilinear in contacts k₃ and k₄, located in the passage area of the central wheel teeth profile from concave curvature to convex and convex-convex curvature in contacts k₅ . . . k₁₄ (FIG. 6a) and k₅ . . . k₈ respectively (FIG. 6b), located in the tip area of the central wheel teeth.

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 Z₁, Z₂, 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.

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).

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 k₂(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.

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 k₀, k₁, k₂ and k₃, presented in the tooth profilogram evolute in FIG. 8.

Thus, for example, for the gear with geometric parameters Z₁=29, Z₂=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 k₀ corresponding to the crankshaft precession angle ψ=0°, the difference of the radii of curvature between the central wheel and satellite wheel profiles ρ₁−r=5.26−5.0=0.26 mm in the contact point k₁ corresponding to the precession angle ψ=12.84°; ρ₁−r=5.78−5.0=0.78 mm (first pair of engaged teeth); in the contact point k₂ corresponding to the precession angle ψ=5.68°, ρ₁−r=11.3−5.0−6.3 mm (second pair of engaged teeth); in the contact point k₃ corresponding to the precession angle ψ=38.53°, ρ₁−r=225−5.0=220 mm (third pair of engaged teeth, etc.).

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=R₁gθ, the period P=2πRZ₂/Z₁ and the origin in point k₀, and the concomitant gear area of the load-bearing teeth extends from contact k₀ to k₁.

When the crankshaft rotates, each pair of teeth in contacts k_i performs an improvised motion along the same path, moving imaginary, for example, from contact k₀ of the satellite tooth on the bottom of the central wheel tooth (FIG. 8) to contact k₀ formed by the pair of teeth (preceding) after crankshaft rotation with the angle ψ=360·Z₂/Z₁². In this evolution, while the position angle of the crankshaft y increases in the interval O<ψ<360·Z₁/Z₁², contact k₁ moves to the position of contact k₀ (see k₀ from the previous pair of teeth, mated on the bottom of the central wheel tooth), contact k₂- to k₁, contact k₃- to k₂, and the pair of teeth preceding the first three forms a new contact k₃ 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.

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.

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 k₀ disengages, and the pair with position 5 forms a new load-bearing contact k₄, thus constantly maintaining four pairs of load-bearing teeth.

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_k1, . . . ψ_k4, rising from contact to contact with the pitch ψ=360Z₂/Z₁². 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 k₀.

FIG. 9 shows the positions of contacts k₀ . . . k₄ and point 5 on the satellite teeth profile corresponding to the positioning angles ψ_k, =0°, ψ_k1=15.6°, ψ_k2=31.2°, ψ_k346.8°, ψ_k4=62.4° and ψ_k5=78.0°, determined from the relation ψ_k=360iZ₂/Z₁², 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.

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.

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.

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 Z₁=Z₂−1 or Z₁=Z₂+1.

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.

The design of the teeth contact geometry from the precessional toothed gear is limited to the identification of the contact form (see FIG. 10a, 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.

FIG. 11 (a) shows the profilogram of the mating wheel teeth for the configuration of parameters Z₁=24, Z₂=25, θ=3.5°, δ=22.5°, r=6.27 mm, R=75 mm and the ratio of the number of teeth Z₁=Z₂−1, in FIG. 11 (b) in the configuration of parameters, the ratio of the number of teeth Z₁=Z₂+1 differs, i.e. Z₁=25 and Z₂=24 and δ=0°, and in FIG. 11 (c), the ratio of the number of teeth Z₁=Z₂+1 and the conical axoid angle δ=22.30° differs.

4. Reduction of the Pressure Angle Between the Mating Flank Profiles.

From FIG. 12 we state that in the precessional gear depending on the parameters Z, δ, θ, r and Z₁=Z₂−1 (a) and the reference multiplicity ε_f=100% the contact points k₀ . . . k₅ 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_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.

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.

5. Relative Sliding Between the Teeth Flanks in Gear.

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.

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.

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 Z₁=Z₂±1 and the conical axoid angles δ≥0°. From the aforesaid, the generalized configuration can be expressed by the parameters Z₁=24(25), Z₂=25(24), θ=3.5°, δ=22.5° (0), r=6.27 mm and R=75 mm.

The analysis of kinematics in the contact points k₀, k₁, k₂ . . . k_i corresponding to the crankshaft positioning angles takes place by varying the linear velocities of the contact points E₁ on the central wheel teeth profile and E₂ on the satellite teeth profile and the relative sliding velocity between the flanks V_ol, and the teeth contact geometry is presented through the radii of curvature ρ_ki of the central wheel teeth profile and the satellite teeth profile r and their difference (ρ₁−r). Analysis of the teeth contact kinematics is performed for the crankshaft speed n₁=0.3000 min.

Thus, in the gear Z, δ, β, θ with the ratio of the numbers of teeth Z₁=Z₂−1 and the conical axoid angle δ=22.5°, shown in FIG. 14 (a), in the teeth contact k₀ the linear velocity is V_ε1=9.83 mA, V_ε2=9.69 m/s, V_alkθ=0.14 m/s and the radius of curvature of the central wheel teeth profile is ρ_k0=6.43 mm of the satellite teeth profile r=6.27 mm and their difference (ρ_k0−r)=0.16 mm (FIG. 14b).

As the angular coordinate increases from one mating pair to the other with the pitch ψ=360iZ₁/Z₁, for example, from the angular coordinate ψ_ko=0° up to ψ=15,6° assigned to contact k₁, the linear velocities V_E1 and V_E2 decrease, registering in contact k₁ the difference V_alki V_al=, V_ε1k1−V_E2k2=0.34 m/s and the difference of the radii of curvature of the mating flanks in (ρ_k2−r)=1.17 mm in contacts k₂ corresponding to ψ=31.2° V_al4=0.67 m/s and the difference of the radii of curvature (ρ_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_k3=57.66 mm. FIG. 14 (c) shows the evolution of the geometry from contact k₀ to contact k₄.

Table 1 presents the argumentation and justification of the limits of variation of the frontal overlap degree values ε_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

TABLE 1
Parameter	Lower limit	Upper limit	Note
Degree of	εf = 1.5 pairs of teeth.	εf = 4.0 pairs of teeth.
frontal	Decreasing the εf < 1,5 leads to	Increasing the εf > 4,0 leads to
overlap εf 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 ω5/ω8 = const.
Gear conical	δ = 0°, degrees. If Z1(4) = Z2(3) − 1,	δ = 30°, degrees. If
axoid angle δ.	decreasing the bevel axoid angle	Z1(4) = Z2(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	Z1(4) = Z2(3) − 1.
diameter D.	geometry.

Variation of the frontal overlap within the limits 1.5≤ε_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.

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

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.

According to claim 2, the total teeth contact line l_Σ in the gear with inclined teeth is determined from the condition of frontal gear ε_j of a certain number of pairs of teeth (ε_f=1,2,3 . . . ), but not less than one pair (ε_f,min=1). In the case of ε_f,min=1 it turns out that a pair of teeth engages, while the previous pair disengages.

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_m>1. Thus, in the case of ε_f,min=1 it is proposed to incline the teeth at the angle β_g, which would ensure a degree of longitudinal (axial) overlap.

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

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.

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_f^B, determined by the frontal gear multiplicity ε_f and the longitudinal overlap ε_a dependent on the teeth inclination angle β₂, 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.

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 Z₁=Z₂±1 takes place with an increased share of pure rolling of teeth depending on the angle θ.

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_w.

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 ω₁, the contact lines of the engaged teeth move in the gear field in the direction indicated by arrows A and B.

In FIG. 16 (a) with a pair of teeth in frontal gear ε_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′.

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_Σ=const.

In the case of gearing with two pairs of teeth in frontal gear ε_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_Σ=const.

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_m=ε_f^β+ε_a^β.  (20)

The precessional gear with inclined teeth can also work without frontal overlap, thus with ε_f^β>1, if the axial overlap ε_β is ensured, i.e. b_w>(2πrZ₃)/(Z₁ 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.

Obviously, the specific teeth load q decreases with the increase of the total length of the contact lines l_Σ=ε_mb_w sin δ/cos β, and l_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_Σ=const, the teeth load will not change over time, and the noise emission and dynamic loads will decrease.

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. 14b), and for the same tooth width, the length of the contact lines increases, which leads to the decrease of the specific teeth load.

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

The optimal choice (see FIG. 14) of the mating reference pair of teeth in the point with the crankshaft positioning angle ψ_i is based on three considerations, namely: the difference in the radii of curvature of the flank profiles in contact (ρ_k−r)=min, the sliding velocity between the flanks in contact V_al=min, the pressure angle α_w of the teeth flank profile of the central wheels α_w=min. All these geometry and kinematics parameters of the teeth contact are according to the precession angle ψ.

From the analysis of FIG. 14 we observe that the conditions (ρ_k−r)=min and V_al=min can be achieved by reducing the angle ψ, and α_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.

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.

According to claim 3, in FIG. 17, the gear (Z₃-Z₄) 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 Z₄=Z₃±1.

Thus, the difference of the transmission (FIG. 17) consists in the constructive specific character of the satellite wheel 2 in which the gear (Z₁-Z₂) is geometrically analogous to the gear (Z₁-Z₂) of the transmission in FIG. 2 (b), and the gear (Z₃-Z₄) 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 Z₄=Z₃+1 or Z₄=Z₃−1 do not influence the shape of the central wheel teeth profile 7 and respectively the teeth contact geometry. In the case of Z₄=Z₃+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.

Claim 4.

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.

It is worth mentioning that, according to FIG. 18, the gear (Z₃-Z₄) 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.

This configuration with δ>0 and Z₄=Z₃+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.

Thus, the solution according to FIG. 18 with δ>0° and Z₄=Z₃+1 provides for the operation of the transmission in multiplier operation mode by multiplying the revolutions from shaft 8 to shaft 5.

Claim

The fifth difference of the claimed transmission consists in that the satellite wheel 2 (FIG. 20a) 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.

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

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

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

INDUSTRIAL APPLICABILITY

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.

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

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:

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 (R₁-R₂) m_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_ki−r) respectively is 0.16 mm, 1.17 mm, 9.55 mm (see FIG. 14).

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

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).

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

BIBLIOGRAPHICAL REFERENCES

• 1. SU 1455094 A1 1989.01.30
• 2. SU 1758322 A1 1992.08.30

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≤εf≤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.