Planetary gear device with two sun gears having different diameters

A planetary gear device having two sun gears with different diameters that engage with double planet gears having a first gear teeth area, which interacts with the fist sun gear, and a second gear teeth area, which interacts with the second sun gear. The double planet gears are continuous helical gear wheels. The crown circle diameters of the gear teeth areas of the double planet gears and the crown circle diameters of the sun gears are adapted such that the pitch circle diameters between the double planet gears and the sun gears are arranged in the center between the crown circle diameters and the effective root diameters of the gear teeth areas of the double planet gears and the sun gears.

FIELD OF THE INVENTION

The invention relates to a planetary gear device with two sun gears having different diameters.

BACKGROUND OF THE INVENTION

A differential gear unit is known from EP 0 844 416 A2, which has a bevel gear differential with an input shaft and two output shafts. The input shaft is positively connected with a carrier bearing of a bevel gear differential via a bevel gearing, to be able to transmit torque from the input shaft, via the carrier bearing, to the output shafts connected with the bevel gear wheels of further bevel gear wheel pairs. In this case, an outgoing supply of driving torque is distributed equally to both output shafts.

In addition, a power path is provided between the carrier bearing and both output shafts, each of which features a planetary gear device with two sun gears, which have different diameters. The sun gears mesh with the double planet gears that are constructed with a first gear teeth area that interacts with a first sun gear and a gear teeth area into which the second sun gear meshes. The planet carrier of the double planet gears are positively connected with frictionally engaged shifting elements, the rotational speed of the planet carriers being variable depending on the transmission capacity of the shifting elements.

In the present description, the double planet gears of the planetary gear devices are each constructed with two gear wheels which are spaced apart from one another, a shaft area designed with a reduced diameter being provided between both gear wheels of a double planet gear.

The disadvantage of this type of double planetary gear design is the technical complexity that results in high manufacturing costs.

Therefore, the task of the present invention is based on providing a planetary gear device that can be manufactured in a simple and cost-effective manner.

SUMMARY OF THE INVENTION

The planetary gear device according to the present invention is designed with two sun gears with different diameters, each of which meshes with double planet gears that are constructed with a first gear teeth area that interacts with the first sun gear and a second gear teeth area into which the second sun gear meshes.

According to the present invention, the double planet gears are designed as continuous helical gear wheels, and the crown circle diameters of the gear teeth area of the double planet gears as well as the crown circle diameters of the sun gears are adapted to each other, so that the pitch circle diameter between the double planet gears and the sun gears is arranged in the center between the crown circle diameters and the effective root diameters of the gear teeth areas of the double planet gears and the sun gears.

The double planet gears designed as continuous helical gear wheels can be manufactured in a simple and cost-effective manner in a single work process, since, based on the continuous design, the first gear teeth area and the second gear teeth area have the same number of teeth and the teeth of the teeth areas basically have the same tooth shape.

The adaptation of the crown circle diameters of the gear teeth areas of the double planet gears to the crown circle diameters of the sun gears with different diameters before manufacturing both gear teeth areas by lathing makes manufacturing especially cost-effective.

In addition, compared to prior art planetary gears, the efficiency of the planetary gear device according to the present invention is improved by the crown circle diameters of the gear teeth areas of the double planet gears and the sun gears being adapted to one another, since the sliding speed is the same in the area of the tooth root and in the area of the tooth crown because of the central position of the pitch circle with respect to the crown circle and the effective root diameters and the friction for a given contact length between the double planet gears and the sun gears is minimized.

In an advantageous further development of the planetary gear device according to the present invention, a total contact ratio between the teeth of the sun gears and the teeth of the gear teeth areas of the double planet gears is greater than 1 in order to achieve uniform power transmission of the torque between the sun gears and the double planet gears.

In order to further reduce friction loss appearing during operation, another advantageous embodiment of the planetary gear device according to the present invention provides that the profile contact ratio between the gear teeth areas of the double planet gears and the sun gears is smaller than 1, preferably 0.7, with which, in a transverse section less than one tooth of a sun gear mesh when transmitting torque with one of the gear teeth areas of the double planet gears and a sliding portion between the teeth of the sun gears meshing with one another and the double planet gears is minimized in a simple way.

An especially favorable embodiment of the planetary gear device according to the present invention with respect to friction can be achieved, if the overlap ratio, i.e., the number of teeth pairs in contact across the gear teeth area, is at least 1.

In a further advantageous embodiment of the planetary gear device according to the present invention, the number of gear teeth of the double planet gears and the number of gear teeth of the sun gears exceed the number of gear teeth at which the gearings between the double planet gears and the sun gears respectively show maximum load capacity, with which the slide movement in the area of the paired gear tooth flanks under stress is reduced because of a small gear teeth module compared to conventional gear teeth pairs of planetary gears.

In order to reduce a so-called breakaway torque, which must be applied to change gear from standstill to a rotating condition and which is greater than the friction torque acting in a rotating condition, the double planet gears are helical and in an advantageous further development of the planetary gear device according to the present invention a rotary meshing sequence is provided for this purpose between the sun gears and the double planet gears. The latter measure results in that torque can only be transmitted between the sun gears and one of the planet gears, the double planet gears respectively meshing to transmit torque in a respectively successive rotary sequence with the sun gears. The reduction of the breakaway torque is achieved in the rotary meshing sequence by means of the superimposition of the stiffness gradients of the individual teeth meshes because the superimposition is accompanied by a minimization of the stiffness variations.

DETAILED DESCRIPTION OF THE INVENTION

FIG. 1shows a vehicle1with a power train2. In the present description, the power train2includes a drive unit3and a main transmission4which can be a transmission of the type known from practice. In the exemplary embodiment shown inFIG. 1, the power engine3is designed as an internal combustion engine and in an advantageous further development it can be constructed as an electric motor or a hybrid power train. A gearing device6that is shown in more detail inFIG. 2is arranged between the main transmission4that is provided for illustrating various gear transmission ratios and a vehicle drive axle5, which is connected in a known manner with at least one drive wheel5A,5B on each side of the vehicle.

A driving torque of the drive unit3is transmitted via a longitudinal shaft7that is connected with a non-illustrated transmission output shaft of the main transmission4and can be introduced via a pair of bevel gear wheels8into the gearing device6, one bevel gear wheel9of the pair of bevel gear wheels8being connected with a carrier bearing10of a gear device6of a known differential gearing device11.

The carrier bearing10thus represents an input shaft of the gearing device6that is positively connected, via further pairs of bevel gear wheels12,13, with the output shafts14,15of the gearing device6, which in turn are positively connected with the drive wheels5A,5B.

A planetary gear device16,17is arranged between the carrier bearing10and the each of the output shafts14,15in the power path26,27, which shows a multiplication factor depending on the direction of a rotational speed difference between the drive wheels of the drive axle5of the vehicle. The multiplication factor of the planetary gear devices16,17depends on the friction forces appearing during operation of the gear device6as well as their gear transmission ratios, the planetary gear device16or17assigned to the faster turning drive wheel5A or5B showing a lower multiplication factor than the planetary gear device17or16, which is assigned to the slower drive wheel5B or5A. Thus, the degree of distribution of the torque between the two drive wheels5A and5B transmitted to the vehicle drive axle5varies as a function of the rotational speed-dependent multiplication factors of the planetary gear devices16and17.

In addition, a frictionally engaged switch element18,19designed as a brake in this description is assigned to the planetary gear devices16and17, whereby the multiplication factors of the planetary gear devices16and17can be varied as a function of the transmission capacity of the switch elements18,19.

The planetary gear devices16and17respectively show several double planet gears20A,20B and21A,21B, only two double planet gears of the planetary gear devices16and17being illustrated inFIG. 2. The double planet gears20A to21B are constructed as continuous helical gear wheels that mesh respectively with a first gear teeth area20A1,20B1and21A1,21B1respectively with a first sun gear22or23. In addition, the double planet gears20A to21B mesh with the second gear areas20A2and20B2and21B1,21B2with the second sun gears24,25respectively. In this case, the sun gears22and23are respectively connected in a rotationally fixed manner with the output shafts14and15and the sun gears24and25in a rotationally fixed manner with the carrier bearing10, while planet carriers28,29of the planetary gear devices16and17are, positively connected rotationally fixed manner with the disks30,31of the brakes18and19. Thus the planet carriers28,29of the double planet gears20A to21B are positively connected with the brakes18and19, so that the rotational speeds of the planet carriers28,29vary as a function of the respectively adjusted transmission capacities of the brakes18,19.

With completely disengaged brakes18and19, a driving torque of the drive unit3adjacent to the gear device6is transmitted equally, via the output shafts14and15, to both drive wheels5A and5B along a longitudinal shaft7, as in conventional differential gears.

Upon the appearance of a rotational speed difference between the drive wheels5A and5B or the probability of such a rotational speed difference, the transmission capacities of both switch elements18and19during the start up procedure of the vehicle1are set to values, so that most of the driving torque of the drive unit3transmitted to the vehicle drive axle5is conveyed, to an increased extent, to the slower rotating drive wheel, to the traction in the area of the drive wheels5A and5B of the drive axis of the vehicle.

The sun gears22and23as well as the sun gears24and25are designed in the same way in the present description, the sun gears22and24as well as the sun gears23and25being constructed with different diameters.

As embodiment of the double planet gears20A to21B is basically the same, only the double planet gear20A will be addressed in the following description ofFIG. 3.

InFIG. 3, a tooth32of the double planet gear20A is shown alone, schematically across its entire tooth width b. In this case, the width of the first gear teeth area20A1is designated more precisely with b1and the width of the second gear teeth area20A2with b2.

As is apparent fromFIG. 3, the tooth32, which is designed in the same manner as the other teeth of the double planet gear20A, shows the same tooth shape in the first gear teeth area20A1as in the second gear teeth area20A2, a crown circle diameter da1in the first gear teeth area20A1being smaller than the crown circle diameter da2in the second gear teeth area20A2.

Thus, the double planet gear20A is designed in a stepped manner, the tooth profiles of the sun gears22and24meshing with the double planets20A in the area of the crown circle diameters being adapted to the crown circle diameters da1and da2in the gear teeth areas20A1and20A2of the double planet gear20A, so that the pitch circle diameters dw1and dw2between the double planet gear20A and the sun gears22and24are arranged in the center between the crown circle diameters da1and da2and the effective root diameters dNf1and dNf2of the double planet gear20A and the sun gears22and24. Thus, a profile of the slide velocity vgillustrated inFIG. 3develops in the area between the effective root diameters dNf1and dNf2in the gear teeth areas20A1and20A2and the crown circle diameters da1and da2, the slide velocity, vg, in the area of the pitch circle diameters dw1and dw2being equal to zero and the slide velocities in the area of the tooth base and the tooth crown being of the same size.

A total contact ratio between the teeth of the sun gears22and23and tooth32and the teeth of the double planet gear20A respectively is larger than 1 to minimize friction, while a profile contact ratio that is formed by the quotient of the normal base pitch and the transverse normal base pitch between the gear teeth areas20A1,20A2of the double planet gear20A and the sun gears22,24is smaller than 1, preferably equal to 0.7. Thus, less than one tooth of the double planet gear20A meshes with the sun gears22and24in one transverse section. Since the double planet gear20A is designed with a helical gearing and the total contact ratio is larger than 1, uniform power transmission is achieved in a simple manner. In addition, an overlap ratio, which corresponds to the number of tooth pairs meshing across the gear teeth areas b1and b2and the gear teeth areas20A1and20A2is at least 1 in order to guarantee a faultless kinematic transmission system.

The helical gearings of the planetary gear devices16and17show better synchronization properties than straight gearings and thus are characterized by more uniform breakaway torque and starting efficiency, the planetary gear devices16and17being characterized by a meshing sequence to reduce the breakaway torque that is described in more detail inFIG. 4.

FIG. 4shows a strongly schematic side view of the planetary gear device16with three double planets20A,20B and20C with their first gear teeth areas20A1,20B1and20C1meshing with sun gear22. In this case, the meshing between the double planet gears20A to20C and the sun gear22is provided so that torque can only be transmitted between one of the double planet gears20A to20C and the sun gear22, the double planet gears20A to20C successively meshing in a torque transmitting manner in a rotary sequence with sun gear22. This results in that the stiffness gradients of the individual gear teeth meshes are superimposed between the double planet gears20A to20C and the sun gear22and a minimal stiffness variation is achieved, the planetary gear device16being characterized by low breakaway torque.

In general, the planetary gear device according to the present invention shows double planet gears or stepped planets with an equal number of teeth for both the steps and gear teeth areas. To guarantee a high manufacturing quality, and thus a high load capacity, the double planet gears are designed with continuous gearings, which can be manufactured at low production costs. In order to implement the gearings between the double planet gears and the sun gears of the planetary gear device according to the present invention with a small gear teeth module, with a profile contact ratio smaller than 1, a total contact ratio larger than 1, as well as an overlap ratio which is at least 1 the crown circle diameters of the gear teeth areas of the double planet gears are designed offset and are optimized for meshing with both sun gears.

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