Transmission, drive train and vehicle having transmission

Abstract

The invention relates to a transmission device (G), comprising an input shaft (10), a first output shaft (11), a second output shaft (12), a first planetary gear set (P1) and a second planetary gear set (P2) connected to the first planetary gear set, wherein the planetary gear sets (P1, P2) respectively comprise a plurality of elements (E11, E21, E31, E12, E22, E32). The input shaft (10), the two output shafts (11, 12), the planetary gear sets (P1, P2) and the elements of the planetary gear sets are arranged and constructed so that the torque introduced via the input shaft (10) is converted and distributed to the two output shafts (11, 12) in a defined ratio and so that the generation of accumulated torque is prevented, wherein at least one element (E31) of the first planetary gear set (P1) is connected to another element (E12) of the second planetary gear set in a rotationally fixed manner, and another element (E22) of the second planetary gear set (P2) is fixed to a rotationally fixed structural element (GG). A connecting device (40) is provided, which is arranged and constructed to connect the first output shaft (11) and the second output shaft (12) passively and thus without a controller and without an actuator, so that torque can be transmitted.

Description

Technical Field

[0001] The invention relates to a transmission, in particular a transmission for a motor vehicle. The invention also relates to a drive train and a vehicle. Background Art

[0002] Transmissions known from the prior art, for example DE 10 2011 079 975 A1, provide a torque conversion as a ratio of an output torque to an input torque and a speed ratio as a ratio of an input rotational speed to an output rotational speed.

[0003] DE 10 2018 112 880 A1 discloses an electric vehicle with an integrated differential. The electric vehicle drive train comprises an electric motor, a first and a second planetary transmission including a sun gear, a planet carrier and a ring gear element, a first and a second output shaft and a housing. Elements of the first planetary transmission are connected to the electric motor, the first driven shaft and an element of the second planetary transmission. Elements of the second planetary transmission are connected to the first planetary transmission, the housing and the second driven shaft. The first planetary transmission provides a differential reduction device, and the second planetary transmission provides a reversing and reduction device. An optional clutch can provide the function of a limited slip differential and distribute the torque to one driven shaft or the other driven shaft.

[0004] Such a transmission, which is also called a rolling differential, does not generate cumulative torques as is common in the prior art (e.g. on the differential cage). Preventing the generation of cumulative torques means that the sum of the individual torques applied to the two output shafts is not applied to rotating structural elements, such as input shafts, output shafts, elements of planetary gear sets.

[0005] In the case of bevel gear differentials, the expansion effect of the bevel gears under torque is used to generate friction between the corresponding bevel gears and the differential cage and thus achieve the desired locking effect. This effect can be intensified by using additional friction plates. According to the meaning, the friction force that depends on the tooth force is also used in spur gear differentials to achieve a torque-dependent locking effect. In addition, the principle of the torque-dependent locking effect is known from DE 10 2008 000444 A1 of the applicant.

[0006] The locking effect of the differential can also be realized in a speed-dependent manner. Thus, for example, a locking effect that is dependent on the speed difference can be realized using a fluid friction clutch, a so-called viscous clutch. For this purpose, the speed difference between the output shaft and the differential cage is usually used. The principle of a viscous drive is known, for example, from DE 39 12 304 A1.

[0007] In the case of bevel gear differentials and spur gear differentials, it is known to use preloaded bearing surfaces or sliding surfaces for producing a locking effect. This is known, for example, from DE 10 2011 085 119 B3 and DE 10 2015 223126 A1. Here, the locking effect is independent of the speed and torque.

[0008] Combinations of the aforementioned principles of locking effect are also often used. Summary of the invention

[0009] The object of the present invention is in particular to provide a transmission, in particular in the form of an integrated differential, which enables improved traction and improved driving characteristics. In addition, the object of the present invention is to provide a drive train and a motor vehicle having such a transmission.

[0010] This object is achieved by a transmission according to the invention, a drive train according to the invention and a vehicle according to the invention.

[0011] The transmission device on which the present invention is based comprises an input shaft, a first output shaft, a second output shaft, a first planetary gear set and a second planetary gear set connected to the first planetary gear set, wherein the planetary gear sets each comprise a plurality of elements, wherein the torque introduced on the input shaft is converted and distributed to the two output shafts in a defined ratio and the generation of accumulated torque is prevented, wherein at least one element of the first planetary gear set is connected to another element of the second planetary gear set in a rotationally non-relatively manner by means of a shaft, and another element of the second planetary gear set is fixed to a rotationally non-relatively structural element.

[0012] The transmission is characterized in that a connecting device is provided, which is arranged and configured to connect the first output shaft and the second output shaft passively and thus without a controller and without an actuator, so that torque can be transmitted. In other words, the locking torque can be provided passively by means of the present invention.

[0013] Therefore, the connecting means is configured to provide a torque-transmitting connection between the first and second output shafts without an actuator. That is, a locking torque can also be brought about without a controller and an actuator.

[0014] The connection can be frictional and therefore in particular implemented in dependence on a pressing force. The connection can be damped and therefore implemented in dependence on a differential speed. In the case of a frictional connection associated with a pressing force, the frictional connection can be implemented with a constant pressing force (for example achieved by means of preload) or with a pressing force associated with a torque. In the case of a pressing force that is dependent on a torque, in particular an axial force of the helical toothing or an axial force proportional to the torque from a curved ramp system can be used.

[0015] The locking effect of the transmission is produced by a locking torque, which can

[0016] - is load dependent, and / or

[0017] - is related to the differential speed, or

[0018] - is independent of speed and torque, i.e. is pre-set.

[0019] The lock value is understood to be the quotient of the difference between the two output torques and the sum of the two output torques. That is, at a lock value of 0%, both wheels have exactly the same torque. At a lock value of 100%, one of the outputs transmits 100% of the torque, while the other output transmits no torque.

[0020] The lock value S can be defined according to the formula as follows:

[0021] (or: =0%...100%)

[0022] Here, M(L) corresponds to the torque of the first output shaft, and M(R) corresponds to the torque of the second output shaft.

[0023] In the sense of the present invention, a "shaft" is to be understood as a rotatable component of a transmission, via which the respective associated parts of the transmission are connected to one another in a rotationally fixed manner, or via which such a connection is established when a corresponding shifting element is actuated. The respective shafts can be connected to one another either axially, radially, or both axially and radially. Thus, the respective shafts can also be present as intermediate parts, via which the respective parts are connected, for example radially.

[0024] These elements are present, in particular, in the form of a sun gear, a planet carrier and a ring gear.

[0025] In the sense of the invention, "axial" refers to the orientation in the direction of the longitudinal center axis along which the planetary gear sets are arranged coaxially with respect to one another. "Radial" refers to the orientation of the shafts lying on the longitudinal center axis in the diametrical direction.

[0026] If the element is fixed, it is blocked from rotational movement. The rotationally fixed component of the transmission can preferably be a permanently stationary component, preferably a housing of the transmission, a part of such a housing or a component connected thereto in a rotationally fixed manner.

[0027] The transmission device can be implemented, for example, as follows:

[0028] The input shaft is connected to the first element of the first planetary gear set in a rotationally fixed manner;

[0029] The first output shaft is connected to the second element of the first planetary gear set in a rotationally fixed manner;

[0030] wherein the third element of the first planetary gear set is connected to the first element of the second planetary gear set in a rotationally fixed manner;

[0031] wherein the second element of the second planetary gear set is fixed to a structural element of the transmission device that is fixed against relative rotation;

[0032] The second output shaft is connected to the third element of the second planetary gear set in a rotationally fixed manner.

[0033] Thus, a transmission is provided which enables torque conversion, torque distribution and a locking effect by means of a single integrated component. The invention therefore relates to a combined ratio and differential transmission which enables torque conversion, torque distribution to an output shaft and generation of a locking torque with the aid of a housing support. In addition, the shaft connecting the first and second planetary gear sets can be mounted in an inexpensive, space-optimized and low-loss manner.

[0034] The description of torque conversion is understood as follows:

[0035] The transmission has two output shafts, which describe the transformation of the transmission with respect to the sum of the torques of the input torque. The speed ratio of the individual output shafts is not initially defined. Only when the two output shafts are coupled, for example via the wheels of a vehicle on the road, does a defined speed occur. If the two output shafts rotate at the same speed, for example when driving in a straight line, the speed ratio can be formed as the speed ratio between the input speed and one of the two identical output speeds, as in the prior art. In all other cases, the speed ratio of the transmission cannot be named using the usual definition of a speed ratio.

[0036] The two planetary gear sets can be arranged axially adjacent to each other. However, the first planetary gear set can also be arranged radially inside the second planetary gear set. The latter embodiment is also called a nested arrangement of planetary gear sets.

[0037] Preferably, the toothing of two elements of the first and second planetary gear sets that are connected to one another, namely the toothing of the third element of the first planetary gear set and the toothing of the first element of the second planetary gear set, are formed on the same component.

[0038] Preferably, the toothing inclination of the third element of the first planetary gear set and the toothing inclination of the first element of the second planetary gear set have at least similar magnitudes, preferably the same magnitudes, and preferably the same signs. The same magnitudes and the same signs, i.e. the same inclinations, enable the connection member or the coupling shaft to be free of axial forces, so that a costly axial bearing can be omitted. It is also achieved thereby that the forces from the helical toothing are supported on the connection element according to the invention, without the toothing forces being guided lossily via the housing bearing. This support on the connection element in turn facilitates the provision of torque information and activation energy in terms of magnitude and direction on the connection element.

[0039] The pitch or lead of the helical teeth The axial distance measured along the associated axis of rotation is understood to be the distance that would be required to cause the tooth to extend 360° around the axis if the tooth were not actually extended beyond the actual width of the gear. In the case of threads, the term pitch is used in a similar way. Thus, a helical gear with multiple teeth is similar to a multi-start thread. In the case of spindles, the term lead is also often used for the corresponding dimensions.

[0040] Preferably, the input shaft for introducing torque into the transmission is connected to a drive machine, in particular an electric machine or an internal combustion engine. In the case of an electric machine, preferably, the rotor of the electric machine is connected to the input shaft in a rotationally fixed manner. Preferably, the rotor is connected to the input shaft via at least one speed ratio stage.

[0041] The electric machine can be arranged either coaxially with the planetary gear set or parallel to their axes. In the former case, the rotor of the electric machine can either be directly connected to the input shaft in a rotationally fixed manner or can be coupled to the input shaft via one or more intermediate gear ratio stages, wherein the latter can more advantageously design the electric machine with a higher speed and lower torque. At least one gear ratio stage can be designed as a spur gear stage and / or a planetary stage.

[0042] If the electric machine is arranged offset from the planetary gear set axis, the coupling takes place via one or more intermediate gear ratio stages and / or traction mechanism transmissions. One or more gear ratio stages can also be implemented individually here either as a spur gear stage or as a planetary stage. The traction mechanism transmission can be either a belt drive or a chain drive.

[0043] In the case of a coaxial arrangement of the electric motor, it is preferred that the first output shaft is guided through the rotor of the electric motor. This makes the transmission device with the electric motor particularly compact.

[0044] Preferably, the fixed-shaft transmission ratio of the second planetary group is at least approximately calculated by subtracting 1 from the reciprocal value of the fixed-shaft transmission ratio of the first planetary group, that is:

[0045] For two planetary groups, it is implemented as a negative planetary group (for example according to Figure 2 or Figure 3 ), this calculation rule causes the power output torque to be divided into two output shafts in half when ignoring transmission losses. This is particularly advantageous when the present invention is used to divide the torque into two wheels on the same axle.

[0046] If a different torque split is desired or the planetary gear set is implemented differently (e.g. Figures 4 to 9 ), the calculation rules can be limited in a meaningful way ( Fig.19 ). Since in operation under real conditions the asymmetry of the transmission losses to the two output shafts can lead to slight deviations from the calculation rule in favor of obtaining the same power output torque on both shafts, the wording "at least approximately" is used. In addition, this expression is used because, for example, due to acoustic requirements, it is sometimes not possible to fully comply with the calculation rule while maintaining an integer number of teeth and a favorable combination of teeth.

[0047] Preferably, the number of planets of the second planetary gear set is greater than the number of planets of the first planetary gear set. With this configuration, despite the application of the above-mentioned calculation rules, large transmission ratios can still be achieved, which in turn can achieve a particularly compact and inexpensive electric machine.

[0048] Preferably, the second planetary gear set has five, six, seven or eight planets. Preferably, the first planetary gear set has three or four planets.

[0049] Another positive effect of a large or greater number of planets on the second planetary gear set is that the sun gear and / or the ring gear of the second planetary gear set can thereby be designed very thin-walled and thus very light, inexpensive and space-saving.

[0050] In addition, a large or greater number of planets can enable a more uniform introduction of the toothing forces into the sun gear and / or the ring gear, thereby resulting in a smaller elastic deformation of the sun gear or the ring gear.

[0051] Furthermore, the support of the shaft connecting the first and second planetary gear sets can be improved by a high number of planets on the second planetary gear set. This is possible because the planetary wheels of the second planetary gear set are fixedly mounted relative to the housing and the mentioned shaft is centered therein.

[0052] Of course, it is also possible to make the number of planetary gears of the second planetary gear set greater than the number of planetary gears of the first planetary gear set without applying the calculation rule.

[0053] Preferably, the drive engine is mounted transversely to the direction of travel. Preferably, the two output shafts are connected to the wheels of the vehicle in a rotationally fixed manner.

[0054] Preferably, the two output shafts divide the introduced torque onto different axles of the vehicle. Thus, it is possible to arrange a longitudinal transfer case transmission (also referred to as a longitudinal transfer case), i.e. a transmission which divides the introduced torque onto, for example, multiple axles, in particular the front and rear axles of the vehicle.

[0055] The torque distribution of the transmission does not necessarily have to be distributed evenly to the output shafts. In particular, in the case of an embodiment as a longitudinal transfer gearbox, an uneven distribution between one axle and the other axles is possible. For example, the torque provided by the input shaft can be distributed in such a way that 60% is directed to the rear axle and 40% to the front axle.

[0056] The two planetary gear sets can be implemented as negative planetary gear sets or as positive planetary gear sets. A combination of negative and positive planetary gear sets is also possible.

[0057] The negative planetary gear set consists of the following elements in a manner and method known in principle to those skilled in the art, namely a sun gear, a planet carrier and a ring gear, wherein the planet carrier guides at least one, but preferably a plurality of planet gears in a rotatable manner and the planet carrier meshes in detail with the sun gear and the surrounding ring gear, respectively.

[0058] In a spurious planetary gear set, the following elements are also present: a sun gear, a ring gear and a planet carrier, wherein the latter guides at least one planetary gear set, wherein one planet gear meshes with an inner sun gear and another planet gear meshes with a surrounding ring gear, and the planet gears mesh with one another.

[0059] In the case of allowing the connection of each element, the negative planetary set can be transformed into a positive planetary set, wherein, compared with the implementation scheme as a negative planetary set, the ring gear connection mode and the planet carrier connection mode are interchanged, and the fixed shaft transmission ratio is increased by one. Conversely, the positive planetary set can also be replaced by a negative planetary set, as long as the connection of the elements of the transmission device allows this. Here, compared with the positive planetary set, the ring gear connection mode and the planet carrier connection mode are also interchanged, and the fixed shaft transmission ratio is reduced by one, and the sign is also changed. However, within the scope of the present invention, the two planetary gear sets are preferably implemented as negative planetary sets respectively.

[0060] Preferably, both planetary gear sets are implemented as negative planetary gear sets. Negative planetary gear sets have good efficiency and can be arranged axially side by side and radially nested.

[0061] In the case of a combination of a negative planetary gear set and a positive planetary gear set in a nested arrangement, it is preferred that the radially inner planetary gear set is a negative planetary gear set and the radially outer planetary gear set is a positive planetary gear set. Here, on the one hand, easy nestability is maintained. In addition, in this regard, a fixed ring gear offers the advantage that the (usually) poor efficiency caused by the positive planetary gear set only affects a single output shaft.

[0062] A ratio transmission or a multi-speed transmission, preferably a 2-speed transmission, can additionally be connected upstream of the transmission. The ratio transmission or the multi-speed transmission can also be a component of the transmission and can be used to design an additional speed ratio, for example, by converting the speed of the drive engine into a speed ratio and driving the input shaft with this converted speed ratio. The multi-speed transmission or the ratio transmission is present in particular in the form of a planetary transmission.

[0063] The elements of the transmission can preferably be implemented as follows:

[0064] a) A transmission with two negative planetary gear sets, wherein:

[0065] - the first element of the first planetary gear set is a sun gear,

[0066] - the second element of the first planetary gear set is a planet carrier, and

[0067] - the third element of the first planetary gear set is a ring gear, and wherein

[0068] - the first element of the second planetary gear set is the sun gear,

[0069] - the second element of the second planetary gear set is a planet carrier, and

[0070] - The third element of the second planetary gear set is the ring gear.

[0071] This transmission can be referred to as a first concept with two negative planetary gear sets.

[0072] b) A transmission with two negative planetary gear sets, wherein

[0073] - the first element of the first planetary gear set is a sun gear,

[0074] - the second element of the first planetary gear set is the ring gear, and

[0075] - the third element of the first planetary gear set is a planet carrier, and wherein

[0076] - the first element of the second planetary gear set is the ring gear,

[0077] - the second element of the second planetary gear set is a planet carrier, and

[0078] - The third element of the second planetary gear set is the sun gear.

[0079] This transmission can be referred to as the second concept with two negative planetary gear sets.

[0080] c) The transmission has two negative planetary gear sets, where

[0081] - the first element of the first planetary gear set is the ring gear,

[0082] - the second element of the first planetary gear set is a planet carrier, and

[0083] - the third element of the first planetary gear set is a sun gear, and wherein

[0084] - the first element of the second planetary gear set is the sun gear,

[0085] - the second element of the second planetary gear set is a planet carrier, and

[0086] - The third element of the second planetary gear set is the ring gear.

[0087] This transmission can be referred to as a fifth concept with two negative planetary gear sets.

[0088] d) A transmission having a positive planetary gear set and a negative planetary gear set, wherein the second planetary gear set is a negative planetary gear set, wherein

[0089] - the first element of the first planetary gear set is a sun gear,

[0090] - the second element of the first planetary gear set is the ring gear, and

[0091] - the third element of the first planetary gear set is a planet carrier, and wherein

[0092] - the first element of the second planetary gear set is the sun gear,

[0093] - the second element of the second planetary gear set is a planet carrier, and

[0094] - The third element of the second planetary gear set is the ring gear.

[0095] This transmission is somewhat the first concept with a spur planetary gear set.

[0096] e) A transmission having a positive planetary gear set and a negative planetary gear set, wherein the first planetary gear set is a negative planetary gear set, wherein

[0097] - the first element of the first planetary gear set is a sun gear,

[0098] - the second element of the first planetary gear set is a planet carrier, and

[0099] - the third element of the first planetary gear set is a ring gear, and wherein

[0100] - the first element of the second planetary gear set is the sun gear,

[0101] - the second element of the second planetary gear set (P2) is the ring gear, and

[0102] - The third element of the second planetary gear set is the planet carrier.

[0103] This transmission is somewhat the first concept with a spur planetary gear set.

[0104] f) The transmission has two spur planetary gear sets, where

[0105] - the first element of the first planetary gear set is a sun gear,

[0106] - the second element of the first planetary gear set is the ring gear, and

[0107] - the third element of the first planetary gear set is a planet carrier, and wherein

[0108] - the first element of the second planetary gear set is the sun gear,

[0109] - the second element of the second planetary gear set is the ring gear, and

[0110] - The third element of the second planetary gear set is the planet carrier.

[0111] This transmission is in a way the first concept with two spur planetary gear sets.

[0112] The following transmission is preferred, wherein the connecting element has a friction surface so that the connection can be established by means of friction force. That is, at least two corresponding friction surfaces are arranged between the two output shafts. The connecting element is in particular a diaphragm clutch.

[0113] The friction surfaces of the multi-disk clutch can be designed to be flat or conical, wherein the conical embodiment provides a higher friction torque with the same number of friction surfaces.

[0114] A transmission is preferred in which the friction element implemented as a friction plate or a friction cone or a component adjacent to the friction plate or the friction cone is arranged and configured so that the axial play of the second element of the first planetary gear set and / or the axial play of the third element of the second planetary gear set and / or the axial play of the input shaft can be set. In other words, one of the three elements, two of the three elements or all three elements can be set in terms of their axial play.

[0115] A transmission is preferred in which at least one friction surface is a component of the second element of the first planetary gear set or a component of the third element of the second planetary gear set. This reduces the number of components and thus the costs. These elements are preferably relatively large, heavy and well connected thermally, so that in this case, the heat generated in the friction element can also be well stored and / or discharged.

[0116] The following transmission is preferred, wherein at least one friction surface is arranged on the following component, which guides the output torque to one of the two output shafts. This also reduces the number of components, thereby also reducing costs. In addition, the same thermal advantages can be achieved.

[0117] The following transmission is preferred, in which

[0118] - Input shaft and / or

[0119] - a first element of a first planetary gear set and / or

[0120] - the second element of the first planetary gear set and / or

[0121] - The third element of the second planetary gear set

[0122] Floating support.

[0123] It is particularly preferred that the input shaft and the first element of the first planetary gear set are supported in a floating manner. It is also preferred that the second element of the first planetary gear set and the third element of the second planetary gear set are supported in a floating manner.

[0124] The floating bearing has a play of a few tenths of a millimeter in the axial direction, that is, the corresponding shaft has no significant fixation or tightening in the axial direction. This makes it possible to easily guide the toothing forces in the direction of the friction element. In addition, the floating bearing is inexpensive, has only a small design complexity, and allows for expansion of the shaft in the play area, for example due to temperature.

[0125] A transmission is preferred in which the contact pressure of the friction surface is brought about by means of a constant contact pressure.

[0126] The following transmission is preferred, wherein a constant pressing force is achieved by means of a preloaded spring. That is, the torque that can be transmitted between the two output shafts is achieved by means of the preload of the friction element. The preloaded spring can be implemented, for example, as a slotted disc spring or as a corrugated spring.

[0127] A transmission is preferred in which the friction surfaces preloaded by the spring are additionally loaded with an axial force from the helical toothing of the first element of the first planetary gear set and / or an axial force from the helical toothing of the third element of the second planetary gear set. This makes it possible to achieve a torque-dependent locking effect in addition to the constant locking effect caused by the spring while using the same friction elements.

[0128] Preferably, the friction element implemented as a friction plate or a friction cone is implemented in a corrugated shape. In the corrugated embodiment, a preload is generated that acts through the friction surface. As a result, an initial locking torque can be provided.

[0129] An initial locking torque is understood to be a locking torque which is present substantially independently of the transmitted torque and independently of the level of the existing differential speed.

[0130] The following transmission is preferred, wherein the pressing force of the friction surface is brought about by means of a torque-dependent pressing force.The pressing force for transmitting the torque between the two output shafts is provided by a torque acting in the drive shaft or in one of the two output shafts.

[0131] A transmission is preferred in which the contact force is brought about by means of a helical toothing, wherein the sign of the angle of inclination is selected such that in forward motion in traction operation (and thus also in reverse motion in freewheeling operation) the first element of the first planetary set and the third element of the second planetary set interact with each other. Thus, the input shaft and the second output shaft interact with each other. As a result, a contact force proportional to the torque is obtained at the bearing point, for example at the friction element between the two output shafts.

[0132] The disadvantage of the advantageous solutions mentioned so far is that the possible axial forces are limited to the forces of the toothing. Although the inclination angle can be increased, there are other criteria that must be taken into account when designing the toothing. In addition, the forces are usually introduced into the thrust bearing, which is often lossy. In addition, the use of the axial forces from the toothing only works in the case of a single torque sign, that is, in the present case in traction operation.

[0133] This shortcoming is solved by the fact that the pressing force is promoted by means of the curve ramp. That is, the torque is guided by the curve ramp and thus the pressing force for the friction element is generated.

[0134] The curved ramp can be arranged in particular at

[0135] - between the third element of the second planetary gear set and the second output shaft or

[0136] - between the second element of the first planetary gear set and the first output shaft or

[0137] - between the input shaft and the first element of the first planetary gear set.

[0138] The curve ramp has a torque input and a coaxial torque output. The transmission therebetween provides a curve geometry which causes the input and output to be subjected to a torque-proportional expansion force in the axial direction during the torque transmission. Such a curve ramp is known in particular as a torque-dependent contact device in the form of a continuously variable transmission. The curve ramp can be implemented in a rolling manner (for example with balls, rollers or tapered rollers) or in a sliding manner. The expansion effect of the curve ramp is effective in traction operation and in freewheeling operation relative to the spindle drive.

[0139] Preferably, the pressing force of the friction surface is brought about by means of a pressing force dependent on the differential speed. The differential speed between the two output shafts is used, for example, to operate a hydraulic pump which actuates a friction clutch via a hydraulic cylinder.

[0140] A transmission is preferred in which the connecting element is a fluid friction clutch. A damping-acting fluid friction clutch, also known as a viscous clutch, acts between the two output shafts. In this variant, the locking torque increases with increasing differential speed. The differential speed is therefore also the reason for the torque transfer between the two output shafts.

[0141] In summary, it can thus be established that the connecting means can be passively actuated in different ways, namely,

[0142] - by the axial force of the helical teeth or

[0143] - With the help of preload or

[0144] - Torque with ramp or

[0145] - by means of the differential speed between the two output shafts.

[0146] In particular, the transmission is part of a motor vehicle drive train for a hybrid vehicle or an electric vehicle and is arranged between a drive engine of the motor vehicle, which is designed as an internal combustion engine or an electric machine, and further components of the drive train that follow in the direction of the force flow to the drive wheels of the motor vehicle. The input shaft of the transmission is preferably coupled to the crankshaft of the internal combustion engine or the rotor shaft of the electric machine. The transmission can also be part of a drive train for a conventional motor vehicle, i.e. a vehicle that is driven only by an internal combustion engine.

[0147] In the sense of the invention, two components of a transmission are "connected" or "coupled" or "connected to one another" in a rotationally fixed manner, meaning that these components are permanently coupled, so that they cannot rotate independently of one another. In this respect, no switching elements are provided between these components, which can be elements of a planetary gear set and / or shafts and / or rotationally fixed components of the transmission, but rather the corresponding components are firmly coupled to one another. A rotationally elastic connection between two components is also understood to be rotationally fixed. In particular, a rotationally fixed connection can also include a universal joint, for example in order to enable a steering movement or a bouncing of a wheel.

[0148] According to another aspect, a drive train for a vehicle is provided, which has a transmission with the aforementioned features. The advantages of the transmission also apply to a drive train having such a transmission.

[0149] According to a further aspect, a vehicle is provided, which has a drive train with a transmission, the transmission having the aforementioned features. The advantages of the transmission also apply to a vehicle having such a transmission.

[0150] In summary, the invention provides a transmission and a vehicle having such a transmission, which has an integrated design, i.e. a torque conversion and torque distribution and a compact and axially short design, especially in the case of a nested arrangement. The transmission is also characterized by good efficiency and low costs due to low complexity. Significantly lower tooth forces occur. Furthermore, erosion problems can be reduced. Furthermore, freely selectable locking values ​​can be achieved, without requiring actuators for this purpose.

[0151] The invention is not limited to the feature combinations specified in the independent claims or in the dependent claims. In addition, the possibility of combining individual features with one another is given, which can also be seen from the following description of preferred embodiments of the invention or directly from the drawings. The reference of the claims to the drawings by the use of reference numerals shall not limit the scope of protection of the claims. BRIEF DESCRIPTION OF THE DRAWINGS

[0152] The advantageous embodiments of the present invention described below are shown in the accompanying drawings.

[0153] Figures 1a to 1e A schematic view showing a motor vehicle drive train;

[0154] Figures 2 to 5 A schematic illustration of a preferred embodiment of a transmission device that can be used in the drive train of the motor vehicle of FIG. 1 is shown;

[0155] Figure 6A schematic view showing another preferred embodiment of a transmission device that can be used in the motor vehicle drive train of FIG. 1 ;

[0156] Figures 7 to 9 A schematic view of another preferred embodiment of a transmission device that can be used in the drive train of the motor vehicle of FIG. 1 is shown;

[0157] Figure 9a Shown for Figure 1c A specific implementation form of the transmission device of the drive system;

[0158] Figures 10 to 13 A schematic view of another preferred embodiment of a transmission device that can be used in the drive train of the motor vehicle of FIG. 1 is shown;

[0159] Fig.14 Show according to Figure 3 A cross-sectional view of an embodiment of the present invention;

[0160] Figures 15 to 18 A schematic diagram showing the working principle of the present invention;

[0161] Fig.19 An overview of the fixed-axis transmission ratios of the various embodiments is shown;

[0162] Fig. 20 Show Figures 2 to 19 A schematic diagram of the working mode of the connecting device according to the invention between the first and second output shafts of the transmission; and

[0163] Figure 21 to Figure 25 Show Fig. 20 A preferred embodiment of the connecting device in . DETAILED DESCRIPTION

[0164] Figures 1a to 1e In each case a schematic view of a transmission G of a motor vehicle drive train 100 of a vehicle 1000 in the form of a passenger vehicle is shown.

[0165] according to Figure 1a The drive train 100 of FIG. 1 shows an electric drive that drives the rear axle A of the vehicle 1000. The drive train includes a transmission G that divides the drive torque of the electric motor EM to two output shafts 11 and 12. The transmission G and the electric motor are arranged in a common housing. The forward driving direction is indicated by an arrow 99. In addition, Figure 1a It can be seen that the transmission G and the electric machine EM are oriented transversely to the direction of travel of the vehicle.

[0166] according to Figure 1bThe drive train 100 of the vehicle 1000 shows an internal combustion engine drive that drives the rear axle A of the vehicle 1000. The drive train includes a transmission G that divides the drive torque of the internal combustion engine VM onto two output shafts 11 and 12, wherein a further transmission is arranged between the transmission G and the internal combustion engine VM, for example an automatic transmission of the vehicle. The forward driving direction is indicated by an arrow 99. In addition, from Figure 1b It can be seen in FIG. 2 that the transmission G and the internal combustion engine VM are oriented longitudinally with respect to the direction of travel of the vehicle.

[0167] according to Figure 1c The drive train 100 of the vehicle 1000 shows an internal combustion engine drive that drives the rear axle A and the front axle B. The drive train includes a transmission G that divides the drive torque of the internal combustion engine VM to the axles A and B, wherein a further transmission is arranged between the transmission G and the internal combustion engine VM, for example an automatic transmission of the vehicle. The transmission G can then be connected to the axle differential of the rear wheel axle A via the output shaft 11 and to the axle differential of the front axle B via the output shaft 12. The forward driving direction is indicated by the arrow 99. In addition, from Figure 1c It can be seen in FIG. 2 that the transmission G and the internal combustion engine VM are oriented longitudinally with respect to the direction of travel of the vehicle.

[0168] according to Figure 1d The drive train 100 of the vehicle 1000 shows an electric drive that drives the front axle B, i.e., a front transverse electric drive. The drive train includes a transmission G that divides the drive torque of the electric motor EM to two output shafts 11 and 12. The transmission G and the electric motor are arranged in a common housing. The forward driving direction is indicated by an arrow 99. In addition, the Figure 1d It can be seen that the transmission G and the electric machine EM are oriented transversely to the direction of travel of the vehicle.

[0169] according to Figure 1e The drive train 100 shows an all-wheel electric drive that drives the rear axle A and the front axle B of the vehicle 1000. It is a transmission device implemented as a longitudinal transfer case. The drive train includes a transmission device G, which divides the driving torque of the electric machine EM to two output shafts 11 and 12. The output shaft 11 transmits the torque to the front axle B, while the output shaft 12 transmits the torque to the rear axle A, and then the respective torque is introduced into the respective axle differential. The transmission device G and the electric machine are arranged in a common housing. The forward driving direction is indicated by the arrow 99. In addition, from Figure 1e It can be seen that the transmission G and the electric machine EM are oriented transversely to the direction of travel of the vehicle.

[0170] The following Figures 2 to 14The transmission is shown in a number of preferred embodiments. In these figures, there is a connecting element 40 according to the invention, but it is not shown, which can connect the two output shafts 11, 12 to each other in a rotationally fixed manner. The connecting element 40 will be referred to in the following. Figure 20 to Figure 25 Elaborate in detail.

[0171] Figure 2 The transmission device G of a first preferred embodiment is shown. The transmission device G comprises an input shaft 10, a first output shaft 11, a second output shaft 12, a first planetary gear set P1 and a second planetary gear set P2 connected to the first planetary gear set P1. The planetary gear sets P1 and P2 are respectively designed as negative planetary gear sets in the present case. The planetary gear sets P1 and P2 respectively comprise a plurality of elements E11, E21, E31, E12, E22, E32, wherein the first element E11 is a sun gear SO1, the second element E21 is a planet carrier PT1, and the third element E31 of the first planetary gear set P1 is a ring gear HO1. In the second planetary gear set P2, the first element E12 is a sun gear SO2, the second element E22 is a planetary gear carrier PT2, and the third element E32 is a ring gear HO2. The planetary gear carriers PT1 and PT2 respectively carry a plurality of planetary gears, which are shown but not labeled. The planetary gears mesh with the respective radially inner sun gears on the one hand and with the respective surrounding ring gears on the other hand.

[0172] The input shaft 10 is currently connected to the first element E11 in a rotationally fixed manner. The first output shaft 11 is connected to the second element E21 of the first planetary gear set in a rotationally fixed manner. The second output shaft 12 is connected to the third element E32 of the second planetary gear set in a rotationally fixed manner. The third element E31 of the first planetary gear set P1 is connected to the first element E12 of the second planetary gear set P2 in a rotationally fixed manner, while the second element E22 of the second planetary gear set P2 is fixed to the rotationally fixed structural element GG. The rotationally fixed structural element GG is the transmission housing of the transmission G.

[0173] The third element E31 , ie the ring gear HO1 of the first planetary gear set P1 , and the first element E12 , ie the sun gear SO2 of the second planetary gear set, form a common component which in the present case is present as shaft 3 .

[0174] from Figure 2 As can be seen from FIG. 1 , the input shaft 10 , the first output shaft 11 and the second output shaft 12 are arranged coaxially with each other. The two planetary gear sets P1 and P2 are also arranged coaxially with each other. According to this embodiment, the two planetary gear sets P1 and P2 are arranged spaced apart from each other in the axial direction.

[0175] The input shaft 10 can be connected to the drive engine and thus introduce the input torque into the transmission G. That is, the input shaft and the output shaft rotate in the same direction. By connecting the two planetary gear sets P1 and P2 to each other and supporting the second element E22 on the housing GG, the introduced input torque can be divided to the two output shafts 11 and 12. Here, the transmission not only assumes the function of a speed ratio transmission, but also assumes the function of a differential transmission. This means that the introduced torque is not only transformed, but also divided to different output shafts. In this embodiment, no change of rotation direction occurs.

[0176] Figure 3 Another preferred embodiment of the transmission device G is shown. Figure 2 Compared with the implementation plan, according to Figure 3 The embodiment of FIG. 1 shows a radially nested arrangement of two planetary gear sets P1 and P2. Figure 2 The embodiment of the invention provides a radially extremely compact solution, and according to Figure 3 The embodiment of can realize a transmission device G which is extremely compact in the axial direction. The first planetary gear set P1 forms a radially inner planetary gear set. The second planetary gear set P2 forms a radially outer planetary gear set. Therefore, the first planetary gear set P1 is located radially inside the second planetary gear set P2. In this embodiment, the connection between the first ring gear HO1 of the first planetary gear set P1 and the sun gear SO2 of the second planetary gear set is also designed as a single component, which is also present as shaft 3 in the present case. In this embodiment, there is also no change in the direction of rotation.

[0177] Figure 4 A further preferred embodiment of a transmission device G is shown. Figure 2 In contrast, the first planetary gear set P1 is now designed as a spur gear set. This means that the third element E31 of the first planetary gear set is designed as a planetary carrier, which is non-rotatably connected to the first element E12 of the second planetary gear set, i.e. the sun gear SO2. The second element E21 is now designed as a ring gear HO1 and is non-rotatably connected to the first output shaft 11. The third element E31 of the first planetary gear set and the first element E12 of the second planetary gear set are again designed as the same component, which is currently present as shaft 3. The rest of the reference Figure 2 statement.

[0178] Figure 5 Another preferred embodiment of the transmission device G is shown. Figure 2Compared with the embodiment of the present invention, both planetary gear sets P1 and P2 are now constructed as positive planetary gear sets. Therefore, the second element E21 is constructed as the ring gear HO1 and is connected to the first output shaft 11 in a rotationally fixed manner. The third element E31 is now constructed as the planet carrier PT1 and is connected to the first element E12 of the second planetary gear set P2, that is, the sun gear SO2, in a rotationally fixed manner. The second element E22 of the second planetary gear set P2 is now constructed as the ring gear HO2 and is fixed to the rotationally fixed structural element GG. The third element E32 of the second planetary gear set P2 is constructed as the planet carrier PT2 and is connected to the second output shaft 12 in a rotationally fixed manner.

[0179] Therefore, the planet carrier connection mode and the ring gear connection mode in the two planetary gear sets P1 and P2 have been interchanged. Figure 2 statement.

[0180] Figure 6 Another preferred embodiment of the transmission is shown. Figure 2 Compared with the embodiment of the present invention, the second planetary gear set P2 is now constructed as a positive planetary gear set, while the first planetary gear set P1 remains unchanged. Therefore, the ring gear HO2 of the second planetary gear set P2 is fixed to the housing GG. In addition, the planet carrier PT2 is connected to the second output shaft 12 in a rotationally fixed manner. Therefore, the planet carrier connection mode and the ring gear connection mode of the second planetary gear set have been interchanged. The rest of the reference Figure 2 statement.

[0181] Figure 7 Another preferred embodiment of the transmission device G is shown. Figure 6 The implementation form is different in that, according to Figure 7 The embodiment of the invention is provided with radially nested planetary gear sets P1 and P2. The radially inner planetary gear set is the first planetary gear set P1. The radially outer planetary gear set is the second planetary gear set P2. Figure 6 or Figure 2 statement.

[0182] Figure 8 A transmission device G in another preferred embodiment is shown. Figure 2Compared with the embodiment of the present invention, the present invention has the following differences. First, a drive machine in the form of an electric motor EM is provided. The electric motor EM includes a stator S fixed relative to the housing and a rotor R. The rotor R of the electric motor EM is connected to the first element E11 of the first planetary gear set, namely the sun gear SO1, in a rotationally fixed manner. Another difference is that the second element E21 of the first planetary gear set is configured as a ring gear HO1 and is connected to the first output shaft 11 in a rotationally fixed manner. In addition, the third element E31 of the first planetary gear set P1 is configured as a planet carrier PT1 and is connected to the first element E12 of the second planetary gear set P2 in a rotationally fixed manner, which is currently configured as a ring gear HO2. The second element E22 of the second planetary gear set is also configured as a planet carrier PT2 and is fixed to the housing GG. Therefore, the third element E32 is configured as a sun gear SO2 and is connected to the second output shaft in a rotationally fixed manner. In this preferred embodiment, a change in the direction of rotation of the input speed occurs. In this embodiment, the nesting of the planetary gear sets P1 and P2 cannot be achieved.

[0183] In other words, the torque is still introduced through the sun gear SO1 of the first planetary gear set P1, while the power output is guaranteed through the ring gear HO1. Figure 2 Unlike the embodiment of the invention, the planet carrier of the first planetary gear set P1 is now connected to the ring gear HO2 of the second planetary gear set in a rotationally fixed manner. Figure 2 Compared to the implementation scheme of , the power output of the second planetary gear set is therefore realized via the sun gear SO2.

[0184] Fig. 9 Another preferred embodiment of the transmission device G is shown. Figure 2 The embodiment of the present invention differs as follows. Firstly, a drive machine is provided in the form of an electric motor EM, which has a stator S fixed relative to the housing and a rotor R, which is connected to the input shaft 10 in a rotationally fixed manner, and the input shaft 10 is connected to the first element E11, which is designed as the ring gear HO1 of the first planetary gear set P1 in the present case. The first output shaft 11 is connected to the second element E21 of the first planetary gear set P1 (in the present case, the planet carrier PT2). The third element E31 of the first planetary gear set P1 (in the present case, it is designed as the sun gear SO1) is connected to the first element E12 of the second planetary gear set P2, namely the sun gear SO2, in a rotationally fixed manner. The remaining elements of the second planetary gear set remain unchanged.

[0185] According to Figure 2 The implementation forms are different. Fig. 9 In the embodiment of , the torque is introduced via the ring gear HO1 of the first planetary gear set P1, while the power output of the first planetary gear set P1 is still achieved via the planet carrier PT1. Figure 2In contrast, the two planetary gear sets P1 , P2 are connected via a common sun gear, which is present in the present case as shaft 3 .

[0186] Figure 9a Shown for Figure 1c The specific implementation form of the transmission device G of the drive system. The power output end 12 transmits the torque to the rear axle A. The power output end 11 transmits the torque to the front axle B. It can be clearly seen that the output shafts 11 and 12 are arranged parallel to each other (not coaxially). The second output shaft 12 of the second planetary gear set P2 is meshed with the intermediate gear ZZ, which is in turn connected to the following shaft, which in turn transmits the torque to the rear axle differential (not shown).

[0187] Fig.10 A drive train 100 of a vehicle is shown having a transmission of a preferred embodiment, wherein a ratio transmission in the form of a planetary gear set P3 is additionally connected upstream of the transmission G.

[0188] Transmission G is referred to herein according to Figure 3 The planetary gear set P3 is configured as a negative planetary gear set and has a first element E13 configured as a sun gear, a second element E23 configured as a planet carrier, and a third element E33 configured as a ring gear HO3. The second element E23 of the third planetary gear set is connected to the input shaft 10 of the transmission G in a rotationally fixed manner.

[0189] In addition, the planetary gear P3 is equipped with a shifting element SE. The shifting element SE is set up to fix the third element E33 on the anti-rotational structural element GG. In addition, the shifting element SE is set up to connect the third element E33 with the first element E13 of the third planetary gear set in the second shifting position, that is, to interlock. If the planetary gear set is interlocked, the speed ratio is always 1 regardless of the number of teeth. In other words, the planetary gear set rotates as a whole. In the third shifting position, the third element E33 is neither fixed to the housing nor interlocks the planetary gear set P3. In this case, the shifting element SE is in a neutral shifting position. The first shifting position of the shifting element SE is marked with the reference numeral G1, which also represents the first gear stage. The second shifting position is marked with the reference numeral G2, which also represents the second gear stage. The first element E13 of the planetary gear set P3 is connected to a not shown drive machine via the input shaft 14. If the shifting element SE is in its neutral position, the driving torque introduced into the speed ratio transmission device P3 is not transmitted to the input shaft 10 of the transmission device G.

[0190] In addition, from Fig.10It can be easily seen that the ratio transmission P3 is arranged coaxially with the input shaft 10 and the output shafts 11, 12. In addition, it can be clearly seen how the first output shaft 11 is guided through the input shaft 10, which is implemented as a hollow shaft, and further guided through the other shaft 14, which is implemented as a hollow shaft. The two output shafts 11, 12 are connected to the drive wheels 20 respectively. A constant velocity joint 15 is provided to enable wheel movements such as steering movements and / or bouncing. The switching element SE is represented here as a double switching element with a form lock. Single switching elements, in particular load switching elements, are also conceivable.

[0191] Fig.11 A further preferred embodiment of a vehicle drive train with a transmission according to the invention is shown. Figure 2 A preferred embodiment of the present invention. Fig.10 Different, according to Fig.11 The embodiment has no gear ratio transmission upstream. The drive is constructed as an electric motor EM. The electric motor EM has a stator S fixed relative to the housing and a rotor R. The rotor R is connected to the input shaft 10 in a rotationally fixed manner. It can be clearly seen that the electric motor EM is arranged coaxially with the input shaft 10 and the output shafts 11, 12. In addition, the electric motor is also arranged coaxially with the planetary gear sets P1, P2. The input shaft 10 is designed as a hollow shaft, through which the first output shaft 11 is guided. The rest of the reference Fig.10 statement.

[0192] Fig.12 A further drive train 100 with a transmission G is shown in a preferred embodiment. Fig.11 Compared with the embodiment of the invention, the planetary gear sets P1 and P2 are not arranged axially side by side, but are radially overlapped, i.e., nested. Therefore, the transmission device G is Figure 3 The preferred implementation form of. Fig.11 and Figure 3 statement.

[0193] Fig.13 A further preferred embodiment of a drive train 100 is shown. Fig.11 , wherein, in contrast to the latter, the electric motor EM is not arranged coaxially, but parallel to the axis of the transmission G. Here, the connection is realized via a spur gear stage SRS, which consists of a first spur gear SR1 and a second spur gear SR2. The first spur gear SR1 is connected to the input shaft 10 in a rotationally fixed manner. The spur gear SR1 then meshes with a spur gear SR2 which is arranged in a rotationally fixed manner on the input shaft EW of the electric motor EM and which establishes a connection with the rotor of the electric motor EM (not shown further at this moment) inside the electric motor EM.

[0194] In other respects, according to Fig.13 Implementation plan and basis Fig.11 The embodiments of the invention correspond to the above, so reference is made to the description in this respect.

[0195] Fig.14 Shown according to Figure 3 sectional view of a preferred embodiment of a transmission G. The central shaft is the output shaft 11. In this figure, the input shaft 10 coincides with the sun gear of P1, i.e. in other words, the input shaft 10 is connected to the sun gear of the first planetary gear set P1. The sun gear of the first planetary gear set P1 in turn meshes with the planet gears of the first planetary gear set P1. The planet gears of the first planetary gear set P1 in turn mesh with the surrounding ring gear of the first planetary gear set P1, wherein the ring gear simultaneously forms the sun gear of the second planetary gear set P2. The sun gear of the second planetary gear set P2 in turn meshes with the planet gears of the second planetary gear set P2. The planet gears of the second planetary gear set P2 in turn mesh with the ring gear surrounding the planet gears of the second planetary gear set P2.

[0196] It can be clearly seen that the number of planets of the second planetary gear set is greater than the number of planets of the first planetary gear set. According to this embodiment, the second planetary gear set P2 has six planets, while the first planetary gear set P1 has four planets.

[0197] This arrangement makes it possible to achieve large transmission ratios, which in turn allows for a particularly compact and inexpensive electric motor.

[0198] However, according to the calculation rules

[0199]

[0200] A large transmission ratio leads to a smaller fixed-shaft transmission ratio on the second planetary gear set P2. A smaller fixed-shaft transmission ratio leads to a smaller planet diameter. A small planet diameter worsens the meshing and reduces the installation space for the planet bearings.

[0201] It has been found that having a greater number of planets in the second planetary gear set compared to the first planetary gear set can counteract this effect.

[0202] The following Figures 15 to 17 The invention shows the force introduction and force support compared with the prior art (for example DE 10 2011 079 975 A1). In contrast to the prior art, in particular Figure 2 and Figure 3 The preferred embodiment with two negative planetary gears has already been described in . However, this consideration also applies to the remaining embodiments in a meaningful manner.

[0203] for Figures 15 to 17 The generally applicable ones are:

[0204] On the first planetary gear set P1, the torque of the input shaft 10 is converted into a power output torque for the first power output end 11. The third element E31 of the first planetary gear set P1 (which is also the first element E12 of the second planetary gear set P2) is driven in the opposite direction due to its reaction torque. The reverse movement of the third element E31 is allowed, so that a part of the mechanical drive power (preferably 50% in the case of a transverse differential and straight-line driving) is guided into the second planetary gear set via the first planetary gear set P1.

[0205] Furthermore, the reverse rotation increases the speed ratio to the first power output end 11 (fixed shaft transmission ratio i0=-3, and only a speed ratio of i=4 can be achieved when the ring gear is fixed).

[0206] In the second planetary group P2, the direction of rotation introduced at the first element E12 (reverse) is transformed into the power output movement of the second power output end 12 (forward) by means of the housing support E22. Here, the torque introduced into the second planetary group P2 and the torque led to the second power output end 12 are accumulated into the housing support torque. The second planetary group P2 only transmits the part of the mechanical power (typically 50%) that is directed to the second power output end 12. The second planetary group P2 is only loaded with a part of the power, which has a positive effect on the overall efficiency.

[0207] In the prior art, the torque conversion is usually carried out by means of a housing support. The reaction torque of the ratio transmission is hereby directed directly into the housing and is not used to generate the second power output torque. As a result, a transmission must first be designed for the cumulative torque of the two output shafts (usually double the torque). A separate differential transmission is then required in order to divide this type of cumulative torque, which is not needed anywhere, into two power output torques again.

[0208] each Figures 15 to 18 Specifically shown:

[0209] Fig.15 The first planetary gear set P1 of the transmission G (right) and the first stage of the spur gear differential of the prior art (left) are shown schematically. The force input from the planetary gears to the sun gear takes place in parallel via three stationary, i.e. fixed, meshing parts. The power output to the first output shaft takes place via the sun gear.

[0210] In contrast, the force introduction according to the preferred embodiment takes place in parallel via eight moving, i.e. rotating, meshings. Four meshings are located between the sun gear SO1 and the four planetary gears. Four further meshings act between the respective planetary gears and the not shown ring gear HO1. The power output to the first output shaft 11 is achieved via the planetary carrier PT1. The technical effect of this is that the tooth forces acting on the first planetary gear set are significantly lower.

[0211] Fig.16 The second planetary gear set P2 of the transmission G (right) and the second stage of the stepped planetary element of the prior art (left) are schematically shown. The force input from the planetary gears to the sun gear is realized via three stationary, that is, fixed meshing parts. The power output to the second output shaft is carried out via the sun gear.

[0212] In contrast, the introduction of force into the second planetary gear set P2 is carried out in parallel via six moving, that is, rotating meshing parts according to the preferred embodiment. The six meshing parts act between one of the six planetary gears and the ring gear HO2. The fixed planet carrier PT2 carrying the six planetary gears and the sun gear SO2 are not shown. The power output to the second output shaft 12 is realized via the ring gear HO2. The technical effect is that the tooth forces acting on the second planetary gear set are significantly less due to the larger effective diameter and the larger number of possible planetary parts.

[0213] Fig.17 The introduction of the supporting torque into the housing is schematically shown. According to the prior art, in the case of stepped planets (left) the force is introduced into the fixed ring gear via three parallel teeth.

[0214] According to the preferred embodiment, the forces are introduced into the fixed planet carrier PT2 via 12 parallel meshings. Six meshings act between the sun gear SO2 of the second planetary gear set and the six planet gears. Six further meshings act between each planet gear of the second planetary gear set and the ring gear HO2. The technical effect is that the tooth forces acting on the second planet carrier PT2 are significantly lower.

[0215] Fig.18 Shows Figures 15 to 17 Another view of the principle shown in detail in FIG.

[0216] According to the invention, the maximum torque in the gear set (right) corresponds to the power output torque of a single wheel. According to the laws of physics, only the housing support has a high torque coefficient.

[0217] According to the prior art, the stepped planetary set (left) is obtained from the input torque M 输入The differential divides this high torque into two halves of the wheel torque M 输入1 and M 输入2 .

[0218] In the figure, the magnitude of the torque is shown symbolically as it passes through the transmission. The direction of rotation is not shown.

[0219] Fig.19 An overview of the calculation rules for the fixed-axis transmission ratios of the various embodiments is given. These fixed-axis transmission ratios produce power output torques of equal magnitude and sign on the two output shafts 11, 12, ignoring transmission losses. 01 represents the fixed shaft transmission ratio of the first planetary gear set P1, i 02 Denotes the fixed axis transmission ratio of the second planetary gear set P2. According to the application of the transmission device, one of the planetary gear set configurations with a corresponding fixed axis transmission ratio can be selected.

[0220] Fig. 20 A schematic diagram is shown, in which the following should be explained Figures 2 to 13 The basic working mode of the connecting device according to the invention in the transmission device.

[0221] Figure 20 to Figure 25 The transmission gear set used in Figure 3 and Fig.12 The transmission gear set corresponds to the one known in the art. That is, the embodiment in which the first and second planetary gear sets P1, P2 are arranged radially one above the other. At this point, it should be pointed out that the connecting element 40 to be described below between the output shafts 11, 12 can be used in Figures 2 to 13 in each of the implementation forms.

[0222] about Fig. 20 The structural embodiment of the transmission device G shown in FIG. Figure 3 For the connection between the motor EM and the input shaft 10, refer to Fig.12 statement.

[0223] From the basis Figure 3 or Fig.12 Based on the embodiment of , a connecting element 40 is additionally provided, which can connect the first output shaft 11 and the second output shaft 12 in a rotationally fixed manner in the actuated state. In the present case, the connecting element is embodied as a multi-disk clutch by way of example.

[0224] The schematic diagram shows the axial forces generated in the helical gearing. The axial force F_E32 acts on the third element E32 of the second planetary gear set P2, namely the ring gear HO2. The axial force F_E11 acts on the first element E11 of the first planetary gear set P1, namely the sun gear SO1.

[0225] The respective second elements E21, E22, i.e. the two planet carriers PT1, PT2, are axially balanced. Shaft 3 is also axially balanced by means of a suitable toothing inclination on shaft 3. Therefore, the axial forces F_E32 and F_E11 are the only axial forces that can be felt on the planetary gear sets P1 and P2. The inclination direction is selected such that the arrow direction corresponds to "forward" traction operation.

[0226] The axial force F_E11 present in the helical toothing of the sun gear SO1 presses via the bearing 31 onto the planet carrier PT1 of the first planetary gear set P1 (note: in Fig. 20 E21 and PT1 are not shown in the figure). The axial force F_E32 present in the helical toothing in the ring gear HO2 pulls the second output shaft 12. The second output shaft 12 is supported on the planet carrier PT2 of the second planetary gear set P2 via the friction element 40. In other words, the forces F_E11 and F_E32 support each other. This means that in the more frequent forward traction operation, the force from the helical toothing is not guided via the transmission housing.

[0227] The axial force of the helical gearing is thus transmitted via the multi-disk clutch 40. As a result, a load-dependent, torque-proportional pressing force of the friction element is generated during forward travel in traction operation (and during reverse travel in freewheeling operation), and thus a load-dependent locking torque is generated. The force flow of the axial force present due to the helical gearing is closed via the friction element of the multi-disk clutch 40.

[0228] The advantage of this principle is that in addition to the torque-proportional locking effect, it also has lower bearing losses. In addition, the advantage is that the acoustic properties are improved, because the axial forces of the toothing and thus their superimposed axial vibrations are not introduced into the housing.

[0229] If the output shafts are assumed to be of equal speed, the thrust bearing 31 is the only bearing that transmits the axial forces in the case of differential rotational speeds. In the case of alternative fixed support of the rotor R and the second output shaft 12, the two fixed bearings transmit the axial forces to the housing in the case of rotational speeds. The losses are high.

[0230] The bearings 41 and 42 shown toward the housing GG are loaded with axial forces from the toothing during forward movement in freewheeling operation (and during reverse movement in traction operation). They are preferably mounted with axial play in order to achieve a floating bearing.

[0231] Fig.21 Show Fig. 20 The first embodiment of the principle of , in particular, has a friction element 40 designed as an axial plain bearing. An axial plain bearing can also be understood as a diaphragm clutch having only one friction surface pair.

[0232] The rotor R and the electric motor EM are supported in a floating manner. The sun gear SO1 is directly embedded in the rotor shaft. Therefore, the rotor shaft also constitutes the input shaft 10. In the forward traction operation, the rotor shaft is pressed onto the first planetary carrier PT1 of the first planetary gear set P1 via the axial adjustment disk 31 and the axial needle bearing 32. The axial force is transmitted to the ring gear HO2 via the adjacent axial sliding bearing 40 in order to close the force flow. Because the sliding bearing 40 is arranged on the shaft, that is, the second output shaft 12, with a relatively small diameter and has only a non-conical friction surface pair, the locking torque that can be generated is relatively low. The advantage of this solution lies in the relatively simple structure and the low thermal load of the friction element 40.

[0233] Fig. 22 Shows Fig. 20 Another embodiment of the principle of the present invention, in particular, has a friction element 40 implemented as a diaphragm clutch. The advantage of a diaphragm clutch is that, compared to a sliding bearing ( Fig.21 ) can produce a larger locking torque.

[0234] According to Fig.21 Compared to the embodiment of the invention, a diaphragm clutch 40 with three friction surface pairs is provided instead of a sliding bearing. Other numbers of friction surfaces are also conceivable. However, an odd number of friction surface pairs, such as one pair, three pairs, five pairs or seven pairs, is particularly advantageous because the bearing position can be eliminated. The increase in the number of friction surface pairs acting in parallel contributes to an increase in the locking value. The diaphragm clutch 40 is particularly advantageous for the movement requirements of the transmission or the vehicle. In the case of a diaphragm clutch, the locking torque can be increased by increasing the friction diameter, that is, the diameter of the friction element.

[0235] Fig.23 An embodiment of a transmission having a cone clutch as a connecting means 40 between the first output shaft 11 and the second output shaft 12 is shown. This embodiment has a locking value according to Fig. 22 Due to the smaller number of friction surfaces and components, this embodiment can be more cost-effective than an embodiment with friction linings. However, due to the smaller friction surfaces, the cone clutch can be less thermally loaded at the same time with higher surface contact. Fig.23The friction cone is implemented in one piece with the output shaft 11 or the first planetary carrier PT1. However, the friction cone can also be implemented in multiple pieces, that is to say, it is manufactured separately and connected to the first output shaft 11 or the first planetary carrier PT1 via a driving gear, for example. A one-piece friction cone is cheaper. A multi-piece friction cone is usually more expensive. However, a multi-piece friction cone achieves greater flexibility in material selection and is more suitable for preventing undesirable centering effects in the friction cone.

[0236] Fig.24 An embodiment of a transmission device with a preloaded spring 45 is shown. Fig. 22 The embodiment of also enables a further, fixedly predefined locking torque in addition to the torque-proportional locking torque. Fig.24 The connecting device 40 is as follows Fig. 22 The spring preload 45 loads two of the three friction surface pairs with a superimposed constant axial force. Preferably, an even number of spring-loaded friction surface pairs is selected so that a thrust bearing can be omitted. In this regard, it should be pointed out that this should not be confused with Fig. 22 confused, in Fig. 22 An odd number of friction surface pairs are provided in the invention so that the thrust bearing can be eliminated.

[0237] Therefore, in Fig.24 In the embodiment, the two friction surface pairs on the left are spring-loaded and additionally loaded in a torque-dependent manner. The friction surface pair on the right is only loaded in a torque-dependent manner. The spring element 45 is implemented as a slotted disc spring and is fixed with a retaining ring.

[0238] The advantage of a torque-independent locking element is that a certain torque can be introduced to the wheel that is gripping the road even when the wheel has no grip, that is, when the grip is equal to zero. In the case of a purely torque-proportional locking action, the locking force is equal to zero when one of the wheels is in the air (ignoring mass effects and friction). The principle of preload can also be used without introducing the axial force from the helical toothing via the friction surface.

[0239] Fig.25 An embodiment of a transmission having a curved ramp 50 concept is shown. Figure 20 to Figure 24 The embodiments of each use the axial force caused by the helical toothing in order to contact the friction element and thus generate friction. This has the disadvantage that the possible axial force is limited to the force of the toothing. Although the angle of inclination can be increased. However, there are other criteria that must be taken into account when designing the toothing. In addition, the force is introduced into the thrust bearing, which is usually lossy. In addition, the use of the axial force from the toothing only works in the case of a single torque sign, that is, in the current traction operation.

[0240] These shortcomings are reflected in Fig.25 The embodiment of the present invention solves the problem by guiding the torque via the curve ramp 50 and generating a pressing force for the friction element 40. In this case, the output torque of the ring gear HO2 is guided on the path to the second output shaft 12 via the curve ramp 50. Here, the curve ramp 50 generates an axial force that acts in proportion to the guided torque. This axial force in turn loads the friction element 40, which is a diaphragm-type switching element, and generates a locking torque proportional to the torque. The advantage of the curve ramp is that it also works in the coasting operation and the axial force can be set arbitrarily by selecting the slope of the curve ramp. Here, in the case of the curve ramp 50, for the same locking force in the traction operation and in the coasting operation, the traction side and the coasting side can be implemented in the same way. However, for different locking forces, the traction side and the coasting side can also be implemented differently. However, they can also be implemented with tapered roller rolling bodies or cylindrical roller rolling bodies or in a sliding manner.

[0241] Alternatively, the output torque of the planet carrier PT1 on the way to the first output shaft 11 can also be used in order to obtain the aforementioned torque-derived axial force with a curve ramp.

[0242] Additionally, similar to Fig.24 The friction element 40 can be preloaded with a spring in order to obtain an initial locking torque. This embodiment has the advantage that a relatively high locking value can be produced. The locking value can be set by the curve ramp geometry, the friction diameter and / or the number of friction surfaces.

[0243] In addition, due to the parallel friction surfaces, the surface-dependent friction loads are higher in the case of differential speeds than in the case of Fig.23 The cone variation is relatively low. Fig.24 The embodiment of the present invention is very suitable for sports needs. Of course, the curved slope 50 can also be used according to Fig.23 Cone clutch combination.

[0244] The present invention is fully described and illustrated with reference to the drawings and the specification. The description and illustration should be understood as examples and should not be interpreted restrictively. The present invention is not limited to the disclosed embodiments. Other embodiments or variations will be apparent to those skilled in the art when using the present invention and when carefully analyzing the drawings, the disclosure, and the following claims.

[0245] In the claims, the words "comprising" and "having" do not exclude the presence of further elements or steps. The indefinite article "a" or "an" does not exclude the presence of a plurality. A single element or a single unit may perform the functions of several units mentioned in the claims. The mere mention of some measures in several different dependent claims should not be understood to mean that a combination of these measures cannot also be used to advantage.

[0246] Reference numerals list

[0247] G-Transmission

[0248] GG anti-relative rotation structural elements and housings

[0249] E11 The first element of the first planetary gear set

[0250] E21 The second element of the first planetary gear set

[0251] E31 The third element of the first planetary gear set

[0252] E12 The first element of the second planetary gear set

[0253] E22 The second element of the second planetary gear set

[0254] E32 The third element of the second planetary gear set

[0255] E13 The first element of the third planetary gear set

[0256] E23 The second element of the third planetary gear set

[0257] E33 The third element of the third planetary gear set

[0258] P1 First planetary gear set

[0259] P2 Second planetary gear set

[0260] P3 Third planetary gear set

[0261] SO sun gear

[0262] PT planetary carrier

[0263] HO ring gear

[0264] EM Motor

[0265] S Stator

[0266] R Rotor

[0267] Input shaft of EW motor

[0268] SRS cylindrical gear stage

[0269] SR1 first cylindrical gear

[0270] SR2 second cylindrical gear

[0271] SE switching element

[0272] G1 first switching position, first gear

[0273] G2 second switching position, second gear

[0274] N Neutral Positioning

[0275] VM Internal Combustion Engine

[0276] A The rear axle of the vehicle

[0277] B Front axle of the vehicle

[0278] 3 Axis

[0279] 10 Input shaft

[0280] 11 First output shaft

[0281] 12 Second output shaft

[0282] 15 Universal joint

[0283] 20 wheels

[0284] 31 Bearings

[0285] 40 Connectors, friction clutches, diaphragm clutches, cone clutches,

[0286] Sliding bearings

[0287] 41 Bearings

[0288] 42 Bearings

[0289] 45 Spring

[0290] 50 Curve Slope

[0291] 99 Forward driving direction

[0292] 100 drive system

[0293] 1000 vehicles

[0294] i 01 Fixed shaft transmission ratio of the first planetary gear set

[0295] i 02 Fixed axis transmission ratio of the second planetary gear set

Claims

1. A transmission device (G), comprising an input shaft (10), a first output shaft (11), a second output shaft (12), a first planetary gear set (P1), and a second planetary gear set (P2) connected to the first planetary gear set, wherein: The planetary gear sets (P1, P2) respectively include a plurality of elements (E11, E21, E31, E12, E22, E32), wherein the input shaft (10), the two output shafts (11, 12), the planetary gear sets (P1, P2) and the elements of the planetary gear sets are arranged and configured such that: - so that the torque introduced via the input shaft (10) is converted and distributed to the two output shafts (11, 12) in a defined relationship, and - so as to prevent the generation of accumulated torque, - wherein at least one element (E31) of the first planetary gear set (P1) is connected to the other element (E12) of the second planetary gear set in a rotationally fixed manner by means of a shaft (3), and - the other element (E22) of the second planetary gear set (P2) is fixed to a rotationally fixed structural element (GG), A connecting device (40) is provided, which is arranged and constructed to connect the first output shaft (11) and the second output shaft (12) passively and thus without a controller and without an actuator, so that torque can be transmitted. wherein the input shaft (10) is connected to the first element (E11) of the first planetary gear set (P1) in a rotationally fixed manner; The first output shaft (11) is connected to the second element (E21) of the first planetary gear set (P1) in a rotationally fixed manner; wherein the third element (E31) of the first planetary gear set (P1) is connected to the first element (E12) of the second planetary gear set (P2) in a rotationally fixed manner; wherein the second element (E22) of the second planetary gear set (P2) is fixed to a structural element (GG) of the transmission device (G) that is prevented from relative rotation; The second output shaft (12) is connected to the third element (E32) of the second planetary gear set (P2) in a rotationally fixed manner.

2. The transmission device (G) according to claim 1, wherein: The tooth pitch of the third element (E31) of the first planetary gear set (P1) and the tooth pitch of the first element (E12) of the second planetary gear set (P2) have the same magnitude and have the same sign.

3. The transmission device (G) according to claim 1, wherein: The connecting element (40) has a friction surface, so that the connection can be established by means of frictional forces.

4. The transmission device (G) according to claim 3, wherein: The connecting device (40) is a diaphragm clutch.

5. The transmission device (G) according to claim 3, wherein: The friction surface is designed to be flat.

6. The transmission device (G) according to claim 3, wherein: The friction surface is conical in design.

7. The transmission device (G) according to claim 3, wherein: A friction element implemented as a friction plate or a friction cone or a component adjacent to the friction plate or the friction cone is arranged and constructed so that the axial clearance of the second element (E21) of the first planetary gear set (P1) and / or the third element (E32) of the second planetary gear set (P2) and / or the input shaft (10) can be set.

8. The transmission (G) according to claim 3, wherein at least one friction surface is a component of the second element (E21) of the first planetary gear set (P1) or a component of the third element (E32) of the second planetary gear set (P2).

9. The transmission device (G) according to claim 3, wherein at least one friction surface is arranged on a component which directs the output torque to one of the two output shafts (11, 12).

10. The transmission device (G) according to claim 1, wherein: The input shaft (10) and / or the first element (E11) of the first planetary gear set (P1) are supported in a floating manner.

11. The transmission device (G) according to claim 1, wherein: The second element (E21) of the first planetary gear set and / or the third element (E32) of the second planetary gear set are supported in a floating manner.

12. The transmission device (G) according to claim 3, wherein: The contact pressure of the friction surface is brought about by means of a constant contact pressure.

13. The transmission device according to claim 12, wherein: The constant contact pressure is brought about by means of a preloaded spring.

14. The transmission device (G) according to claim 13, wherein: The prestressed spring is designed as a disc spring.

15. The transmission device according to claim 13, wherein: The prestressed spring is designed as a corrugated spring.

16. The transmission device (G) according to claim 13, wherein: The spring-preloaded friction surface is additionally loaded by an axial force from the helical toothing of the first element (E11) of the first planetary gear set (P1) and / or an axial force from the helical toothing of the third element (E32) of the second planetary gear set (P2).

17. The transmission device according to claim 13, wherein: The friction element, which is designed as a friction plate or a friction cone, is designed to be corrugated.

18. The transmission device (G) according to claim 3, wherein: The contact pressure of the friction surface is brought about by means of a torque-dependent contact pressure.

19. The transmission device according to claim 18, wherein: The contact force is brought about by means of helical toothing, wherein the sign of the angle of inclination is selected such that in forward traction operation the first element (E11) of the first planetary gear set (P1) and the third element (E32) of the second planetary gear set (P2) interact with each other.

20. The transmission device (G) according to claim 12, wherein: The contact pressure is brought about by means of the curved ramp.

21. The transmission device (G) according to claim 20, wherein: The curve ramp is arranged between the third element (E32) of the second planetary gear set (P2) and the second output shaft (12).

22. The transmission device (G) according to claim 20, wherein: The curve ramp is arranged between the second element (E21) of the first planetary gear set (P1) and the first output shaft (11).

23. The transmission device (G) according to claim 3, wherein: The contact pressure of the friction surfaces is brought about by means of a contact pressure that is dependent on the differential speed.

24. The transmission device (G) according to claim 1, wherein: The connection device (40) is a fluid friction clutch.

25. The transmission device (G) according to claim 13, wherein: The prestressed spring is designed as a slotted disc spring.

26. Drive train having a transmission according to any one of claims 1 to 25.

27. A vehicle having a drive train according to claim 26 or a transmission according to any one of claims 1 to 25.

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