Planetary gear device with two sun gears having different diameters

This application claims priority from German Application Serial No. 10 2006 054 404.8 filed Nov. 18, 2006.

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

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.

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.

Further advantages and advantageous further developments of the present invention result from the patent claims and the exemplary embodiment described below, in principle, with reference to the figures. The figures show:

FIG. 1 is a schematic illustration of a vehicle;

FIG. 2 is a gear diagram of a differential transmission which is designed with two planetary gear devices according to the present invention;

FIG. 3 is a schematic illustration of part of a double planet gear; and

FIG. 4 is a side view of a schematic partial illustration of a planetary gear device according to the present invention.

FIG. 1 shows a vehicle 1 with a power train 2 . In the present description, the power train 2 includes a drive unit 3 and a main transmission 4 which can be a transmission of the type known from practice. In the exemplary embodiment shown in FIG. 1 , the power engine 3 is 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 device 6 that is shown in more detail in FIG. 2 is arranged between the main transmission 4 that is provided for illustrating various gear transmission ratios and a vehicle drive axle 5 , which is connected in a known manner with at least one drive wheel 5 A, 5 B on each side of the vehicle.

A driving torque of the drive unit 3 is transmitted via a longitudinal shaft 7 that is connected with a non-illustrated transmission output shaft of the main transmission 4 and can be introduced via a pair of bevel gear wheels 8 into the gearing device 6 , one bevel gear wheel 9 of the pair of bevel gear wheels 8 being connected with a carrier bearing 10 of a gear device 6 of a known differential gearing device 11 .

The carrier bearing 10 thus represents an input shaft of the gearing device 6 that is positively connected, via further pairs of bevel gear wheels 12 , 13 , with the output shafts 14 , 15 of the gearing device 6 , which in turn are positively connected with the drive wheels 5 A, 5 B.

A planetary gear device 16 , 17 is arranged between the carrier bearing 10 and the each of the output shafts 14 , 15 in the power path 26 , 27 , which shows a multiplication factor depending on the direction of a rotational speed difference between the drive wheels of the drive axle 5 of the vehicle. The multiplication factor of the planetary gear devices 16 , 17 depends on the friction forces appearing during operation of the gear device 6 as well as their gear transmission ratios, the planetary gear device 16 or 17 assigned to the faster turning drive wheel 5 A or 5 B showing a lower multiplication factor than the planetary gear device 17 or 16 , which is assigned to the slower drive wheel 5 B or 5 A. Thus, the degree of distribution of the torque between the two drive wheels 5 A and 5 B transmitted to the vehicle drive axle 5 varies as a function of the rotational speed-dependent multiplication factors of the planetary gear devices 16 and 17 .

In addition, a frictionally engaged switch element 18 , 19 designed as a brake in this description is assigned to the planetary gear devices 16 and 17 , whereby the multiplication factors of the planetary gear devices 16 and 17 can be varied as a function of the transmission capacity of the switch elements 18 , 19 .

The planetary gear devices 16 and 17 respectively show several double planet gears 20 A, 20 B and 21 A, 21 B, only two double planet gears of the planetary gear devices 16 and 17 being illustrated in FIG. 2 . The double planet gears 20 A to 21 B are constructed as continuous helical gear wheels that mesh respectively with a first gear teeth area 20 A 1 , 20 B 1 and 21 A 1 , 21 B 1 respectively with a first sun gear 22 or 23 . In addition, the double planet gears 20 A to 21 B mesh with the second gear areas 20 A 2 and 20 B 2 and 21 B 1 , 21 B 2 with the second sun gears 24 , 25 respectively. In this case, the sun gears 22 and 23 are respectively connected in a rotationally fixed manner with the output shafts 14 and 15 and the sun gears 24 and 25 in a rotationally fixed manner with the carrier bearing 10 , while planet carriers 28 , 29 of the planetary gear devices 16 and 17 are, positively connected rotationally fixed manner with the disks 30 , 31 of the brakes 18 and 19 . Thus the planet carriers 28 , 29 of the double planet gears 20 A to 21 B are positively connected with the brakes 18 and 19 , so that the rotational speeds of the planet carriers 28 , 29 vary as a function of the respectively adjusted transmission capacities of the brakes 18 , 19 .

With completely disengaged brakes 18 and 19 , a driving torque of the drive unit 3 adjacent to the gear device 6 is transmitted equally, via the output shafts 14 and 15 , to both drive wheels 5 A and 5 B along a longitudinal shaft 7 , as in conventional differential gears.

Upon the appearance of a rotational speed difference between the drive wheels 5 A and 5 B or the probability of such a rotational speed difference, the transmission capacities of both switch elements 18 and 19 during the start up procedure of the vehicle 1 are set to values, so that most of the driving torque of the drive unit 3 transmitted to the vehicle drive axle 5 is conveyed, to an increased extent, to the slower rotating drive wheel, to the traction in the area of the drive wheels 5 A and 5 B of the drive axis of the vehicle.

The sun gears 22 and 23 as well as the sun gears 24 and 25 are designed in the same way in the present description, the sun gears 22 and 24 as well as the sun gears 23 and 25 being constructed with different diameters.

As embodiment of the double planet gears 20 A to 21 B is basically the same, only the double planet gear 20 A will be addressed in the following description of FIG. 3 .

In FIG. 3 , a tooth 32 of the double planet gear 20 A is shown alone, schematically across its entire tooth width b. In this case, the width of the first gear teeth area 20 A 1 is designated more precisely with b 1 and the width of the second gear teeth area 20 A 2 with b 2 .

As is apparent from FIG. 3 , the tooth 32 , which is designed in the same manner as the other teeth of the double planet gear 20 A, shows the same tooth shape in the first gear teeth area 20 A 1 as in the second gear teeth area 20 A 2 , a crown circle diameter d a1 in the first gear teeth area 20 A 1 being smaller than the crown circle diameter d a2 in the second gear teeth area 20 A 2 .

Thus, the double planet gear 20 A is designed in a stepped manner, the tooth profiles of the sun gears 22 and 24 meshing with the double planets 20 A in the area of the crown circle diameters being adapted to the crown circle diameters d a1 and d a2 in the gear teeth areas 20 A 1 and 20 A 2 of the double planet gear 20 A, so that the pitch circle diameters d w1 and d w2 between the double planet gear 20 A and the sun gears 22 and 24 are arranged in the center between the crown circle diameters d a1 and d a2 and the effective root diameters d Nf1 and d Nf2 of the double planet gear 20 A and the sun gears 22 and 24 . Thus, a profile of the slide velocity v g illustrated in FIG. 3 develops in the area between the effective root diameters d Nf1 and d Nf2 in the gear teeth areas 20 A 1 and 20 A 2 and the crown circle diameters d a1 and d a2 , the slide velocity, v g , in the area of the pitch circle diameters d w1 and d w2 being 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 gears 22 and 23 and tooth 32 and the teeth of the double planet gear 20 A 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 areas 20 A 1 , 20 A 2 of the double planet gear 20 A and the sun gears 22 , 24 is smaller than 1, preferably equal to 0.7. Thus, less than one tooth of the double planet gear 20 A meshes with the sun gears 22 and 24 in one transverse section. Since the double planet gear 20 A 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 b 1 and b 2 and the gear teeth areas 20 A 1 and 20 A 2 is at least 1 in order to guarantee a faultless kinematic transmission system.

The helical gearings of the planetary gear devices 16 and 17 show better synchronization properties than straight gearings and thus are characterized by more uniform breakaway torque and starting efficiency, the planetary gear devices 16 and 17 being characterized by a meshing sequence to reduce the breakaway torque that is described in more detail in FIG. 4 .

FIG. 4 shows a strongly schematic side view of the planetary gear device 16 with three double planets 20 A, 20 B and 20 C with their first gear teeth areas 20 A 1 , 20 B 1 and 20 C 1 meshing with sun gear 22 . In this case, the meshing between the double planet gears 20 A to 20 C and the sun gear 22 is provided so that torque can only be transmitted between one of the double planet gears 20 A to 20 C and the sun gear 22 , the double planet gears 20 A to 20 C successively meshing in a torque transmitting manner in a rotary sequence with sun gear 22 . This results in that the stiffness gradients of the individual gear teeth meshes are superimposed between the double planet gears 20 A to 20 C and the sun gear 22 and a minimal stiffness variation is achieved, the planetary gear device 16 being 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.

1 Vehicle

2 Power train

3 Power unit

4 Main transmission

5 Vehicle drive axle

5 A, B Drive wheels

6 Gear device

7 Longitudinal shaft

8 Pair of bevel gear wheels

9 Bevel gear wheel

10 Carrier bearing

11 Differential gear device

12 , 13 Pair of bevel gear wheels

14 , 15 Output shaft

16 , 17 Planetary gear device

18 , 19 Switch element

20 A to 20 C Double planet gear

20 A 1 to 20 C 1 First gear teeth area

20 A 2 to 21 B 2 Second gear teeth area

22 to 25 Sun gear

26 , 27 Power path

28 , 29 Planet carrier

30 , 31 Disks

32 Tooth

b Tooth width

b 1 Tooth width of first gear teeth area

b 2 Tooth width of second gear teeth area

d a Crown circle diameter

d Nf1 Effective root diameter

d w2 Pitch circle diameter

v g Slide speed