New split-type gears are particularly desirable when using small axle angles (≤15°) because they are beneficial in terms of product, design features, total input, and so on. The new type of open-type gears are mainly used for cross-shaft speed change devices, inclined-axis speed change devices or non-backlash parallel shaft speed change devices. The new type of split-type gears are spur gears and helical gears with variable tooth tip correction (tooth thickness) that intersect each other with a wide tooth surface. They can all be used on the helical gear of the development shaft of the new type of gear-changing gear of the same body speed change device (the transmission amount is about 50).
The oblique drive shaft is an indispensable beneficiary of this design. With the parallel shaft mounting scheme, the size of the passage is so large that it cannot be used on a right-hand drive body. The speed change device designed by a bevel gear and a matching bevel gear (at an angle of 85°) is likely to be used on the right-hand drive body. The transfer gear shares a common oil path with the bevel gear (at an angle of 85° axis).
Basic Concepts of New Gears 1. Macroscopic Geometric Features In a nutshell, a bevel gear refers to a spur gear with variable crest correction (tooth thickness) that intersects the tooth flanks. During the process, the gear cutting tool axis and the gear shaft are inclined at a delta angle. Due to the different base circles of the left and right flank surfaces, a tooth profile with asymmetrical auger gears is produced. The new root angle produced by the rack-type tool manufacturing should be equal to the new angle δ of the design. The designed apex angle should avoid interference with the tip of the meshing gear teeth to achieve the maximum axial meshing coefficient. Due to the geometric design constraints of the undercut of the gear and the tip of the tooth tip, as the new angle increases, the tooth surface width decreases. An absolutely perfectly symmetrical gear may have a new angle of approximately 150°.
Since most cone-tooth corrected cone gear tooth contacts are in point contact mode, the above-mentioned approximate linear contact is necessarily brought about by the tooth surface correction. The change in the wide tooth profile along the flank causes a change in the load capacity along the flank width of the individual flank sections. The main effects include changes in the root load capacity in the opposite direction of the tooth surface width. The load capacity of the same tooth root along the tooth surface width is mutually aimed. The tooth surface load capacity also varies with the tooth surface width. The load capacity of the entire gear can be seen from the load capacity of the tooth center section. Therefore all gear geometry parameters are determined in the center of the tooth. However, the center distance must be converted to an alternative center distance as, and the alternative center distance as can be obtained by adding the pitch circle radius: as = rw1 rw2 The contact area quality can be estimated based on the tooth width factor. (For example, KHβ and KFβ values ​​in the DIN/ISO standard), this estimate should be as accurate as possible based on the exposure.
Car automatic speed change device cone gear 1. Macro geometry According to the design experience of the marine speed change device and the robot speed change device gear, a new six-speed speed change device cone gear can be designed. However, in the car speed change device, the requirements for macroscopic geometric characteristics are very strict, and the requirements for low noise design are higher. In order to minimize the error of the speed change device as much as possible, the pressure angle of the gear is set at 17.5° and the helix angle is set at 29.5°. Thus, it can be concluded that the transverse meshing coefficient is 1.7 on the non-working tooth surface and 1.8 on the working tooth surface. The end face meshing coefficient is 1.89 on the non-working tooth surface and 1.63 on the working tooth surface.
Due to the relatively wide helix angle, the transverse pressure angle is 15.3° on the working tooth surface and 23.9° on the non-working tooth surface. To match with the 8° axis angle, a cone gear (input) can be used with Cylindrical gears (output) combined. This results in a root angle of 8.6° on the cone gear. This combination is necessary mainly because the output shaft gear must be axially movable and the adjustment range of the mushroom drive pinion on the output shaft is relatively wide. If two cone gears are used, the lateral movement of the two gears will have a serious effect due to the toothed backlash. Due to the limitation of the large end teeth of the cone gear and the undercut of the small end of the cone gear, the choice of the tooth width of the cone gear is also limited. Due to the maximum load capacity of the working face, a relatively large undercut is accepted on the non-working face of the small end of the cone gear, and this undercut phenomenon is rare for cylindrical gears.
Micro-geometric properties When the finite element analysis design enters the actual design goal, the point meshing of the cone gear is a big problem. This unfavorable meshing condition has a negative impact on the load capacity, and a high running deviation is generated at low loads, which causes irregular high noise.
ZF has developed a new load analysis calculation program LVR that analyzes the effect of pressure on the tooth gap (meshing gap). In the analysis and calculation, the gear unit is approximately used as a parallel shaft drive model, and the meshing gap is considered It is the topography of the tooth surface of a tooth on the gear unit. This requires a sufficient representation of the meshing conditions. By comparing the calculated load distribution with the measured meshing contact sample, it can be concluded that the expression of the meshing condition is correct.
By increasing the meshing contact width to greatly improve the meshing condition, it is necessary to calculate the correction amount of the cylindrical tooth surface of the gear device. The purpose of this is to compensate for the existing gap as much as possible, without completely eliminating the concave phenomenon. Special attention should be paid to the helical tooth profile compensation because the helical tooth profile produces a diagonal meshing contact pattern and a similar concave shape and prevents the meshing contact.
The noise of the topology correction gear is also significantly reduced. In the procedure, the changed flank shape forms a conical shape of the cone gear and determines the conjugate shape as a sample to determine the method of grouping the product. At the same time, in the beginning of mass production, the optimal tooth profile has been very tightly determined. In the final shape adjustment, more attention should be paid to noise interference, especially noise interference caused by local loads.
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