Ideal contact geometry The calculation of the modern MCVT kinematics does not include elastic effects, as the change in size, such as after the roller radius, indicates a small change. The elastic effects and loading and lubrication conditions should be taken into account together as the definite contact condition, which determines the contact area. However, this is outside the scope of this study.
The contact angle θ between the outer and inner ring surfaces of the planetary roller and the outer ring can be calculated as the gap δ between the two outer rings. The free half outer race eccentricity Xoffset, and its outermost outer raceway radius Router and planetary roller radius Rplanet are as follows: X=Rplanetsinθ and X=Routersinθ-Xoffset+δ2 such that sinθ=δ2-Xoffset(Rplanet-Router)(1) Relative to the planetary roller rotation axis contact radius Rplant, out = Rplantcosθ (2) relative to the outer ring axis radius Z = Routercosθ (3) and thus the outer contact radius can be determined as Rcont, out = Rout, datum + Routercosθ (4) Therefore, the orbital radii of the planetary rollers are Rorbit=Rcont, out-Rplanltcosθ(5)2. The outer ring can contact the planetary roller contact geometry. The planetary roller can analyze the inner ring contact geometry in the same way. Cos thus eliminates Rcont,in the angular displacement of cos corresponding roller spiral is Φinner,free=Xinner,free2πPitchin(11) where Pitchin is the inner loop length. The radial relative motion of the planetary roller and the driven wheel and the idler/driven wheel mounted in the planetary carrier with respect to the planetary roller causes an increase in complexity, as shown in FIG.
The cosine law can be used to determine the lengths of the triangles of the interior angles α, β and γ, namely Rcarrier, Rorbit and Rplanet+Ridler. The displacement coordinates X, Y, X′ and Y′ are given as X=(Rplanet+Ridler)sinαY=(Rplanet+Ridler)cosαX′=RidlersinαY′=Ridlercosα. If the origin of the coordinate system is on the transmission shaft, planetary rollers The center is (0, Rorbit), and the planetary idler contact position is ((XX′), (Rorbit-Y+Y′)) and the contact radius with respect to the drive shaft is Rplanet, idler=(XX′)2+ (Rorbit-Y+Y′)2(12)6 Steady-State Kinematics In this analysis, it is assumed that slippage will occur at each contact zone. The slip rate cannot be determined here, but it can be determined by the Contact Elastohydrodynamic Lubrication (EHL) analysis. This is not part of this article. The direction of sliding is indicated by the direction of contact torque transmission rather than the speed vector.
Rotational kinematics In traction contact analysis, pure scrolling is considered to be very unlikely. The cause of sliding is clearly divided into three parts: (i) longitudinal relative motion occurs each time power is transmitted; (ii) relative motion is perpendicular to the axis (side axis); and (iii) normal contact area (rotation) around the axis Relative rotation, which is incidental to movement around the contact point.
According to the autogenous rotation of the pressure distribution in the contact zone, it is minimal in design due to the disappearance of its additional energy. However, as the literature points out, it has little effect on oil film thickness. The rotation speed in the inner and outer contact is ωspin, in = (ω in-ωorbit) sin Free Body Method In this section, we will study the force and torque that can be applied to the components together with the inertial force that can separate the components. Unlike ordinary gear drives, the traction drive's motion is transmitted by friction or oil film shear rather than normal forces acting on the surface of the component, so it is possible to maintain relative movement between the components in the desired traction direction. When working in a steady state due to environmental changes the relative influence of the elimination of free bodies between the various parts permits the performance licensing simulation of the MCVT. For example, if the gap in the inner raceway is greater than the possibility of gripping the planetary rollers, the input shaft and its associated two parts (free half inner raceway and roller spiral) will be separated by the output of the transmission. The various movements of these components will be determined by the torques and forces that they affect on them.
In general, there are three possible conditions for the inner ring contact. The first case is that the traction force can transmit the input torque acting on the shaft in the contact area, and thus the force calculated by the planetary roller torque can be eliminated. The second possibility is that most of the input torque can be transmitted through the contact area, so this condition cannot be met. Under these conditions, the lubricant in the contact zone will reach its ultimate shear conditions. This excess input shaft torque will start to accelerate the input shaft and inner race instead of all transmission components and output loads.
The inner roller screw thread shown in Fig. 7(e) is the work of the inner roller screw. The roller screw torque creates an axial load and establishes a normal contact force on the inner raceway. The mechanism is likely to withstand a lot of friction, which is included in the model, and μ is the coefficient of friction.
The initial experimental demonstration of the model is shown. This shows that an MCVT model forms a simulation result of a continuously variable transmission reaching a zero speed "neutral" condition. This is achieved by combining the input and output shaft speeds in a differential epicyclic train arrangement. In this particular case, the input shaft drives a sun gear and the output shaft attaches to the rotating arm of the epicyclic gear train. The epicyclic gear train is a branch roller design. The roller 1 meshes with the input sun gear, the roller 2 meshes with the ring gear, and the IVT is output from the ring gear.
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