EP2976550A1

EP2976550A1 - Vehicle gearbox

Info

Publication number: EP2976550A1
Application number: EP14704810.2A
Authority: EP
Inventors: Johannes Kaltenbach; Peter Ziemer; Kai Borntraeger
Original assignee: ZF Friedrichshafen AG
Current assignee: ZF Friedrichshafen AG
Priority date: 2013-03-20
Filing date: 2014-02-18
Publication date: 2016-01-27
Also published as: WO2014146840A1; DE102013204918A1; US20160061304A1

Abstract

The invention relates to a vehicle gearbox, with a drive shaft (AW), with two gearbox input shafts (GE1, GE2), with at least one separator element (K2, X1), which is assigned to the second gearbox input shaft (GE2), with a main shaft (HW), with an output shaft (AB) and with at least three planetary gear sets (PG1, PG2, PG3), wherein one sub-gearbox (TG1, TG2) is associated with each of the gearbox input shafts (GE1, GE2), wherein one of the two sub-gearboxes (TG1, TG2) comprises at least the first planetary gear set (PG1) and the other of the two sub-gearboxes (TG1, TG2) comprises at least the second planetary gear set (PG2), wherein said shafts (AB, AW, GE1, GE2, HW) are actively connected, or can be actively connected to the planetary gear sets (PG1, PG2, PG3) in such a manner that at least seven forward gears ("1", "2", "3", "4", "5", "6", "7") can be selected sequentially under load by means of the two sub-gearboxes (TG1, TG2), wherein one of these gears may be selected as a direct gear or overdrive gear.

Description

Fahrzeuqqetriebe

The invention relates to a vehicle transmission according to the preamble of patent claim 1.

Increasing demands on the performance of vehicles, the highest possible efficiency and low fuel consumption as well as low emissions result in a relatively high number of gears in both the passenger car and commercial vehicle sector. At the same time, the available space is limited and the weight of the transmission should not or only slightly increased compared to existing designs as possible. In addition, gearboxes are sought that enable traction interruption-free gear changes, are inexpensive to produce and can be used with relatively little effort in various drive concepts.

Are known vehicle transmission with two partial transmissions, such as dual-clutch transmission. In these transmissions, two generally frictionally engaged input-side clutches together with one or more wheel planes or sets of wheels form a partial transmission, each with a power path, which are preselected and active alternately, so that in a sequential change by an overlapping opening and closing of the clutches power shift sequence arises.

Double clutch transmissions can also be designed as a group transmission. Such group transmissions have a multiple-speed main group, usually in countershaft design, and effective as a split gearbox Vorschaltgruppe and / or effective as a range gear rear-mounted in countershaft or planetary construction. As a result, a multiplication of the number of gears of the transmission can be achieved.

Already known are also dual-clutch planetary gear. DE 10 2004 014 081 A1 shows such a dual-clutch transmission with only one transmission input shaft, in which three planetary gear sets and two frictionally engaged and a plurality of form-locking switching elements are arranged, wherein the frictional switching elements for switching different power paths and the positive-locking switching elements for setting different translation stages in the power paths are effective, and in which a total of seven forward gears and one reverse gear are available. In a portion of the gears traction interruption-free gear changes by means of frictional switching elements are feasible.

From DE 10 2010 028 026 A1 discloses a hybrid powertrain for a vehicle with an internal combustion engine and one or more electric machines is known in which a gear in countershaft design has two partial transmissions. One or both partial transmissions one or each associated with an electric machine. At least one electric machine of a partial transmission is operatively connected via a positive switching element with the internal combustion engine.

The applicant's unpublished DE 10 2012 201 366 A1 shows a hybrid powertrain for a motor vehicle, having an internal combustion engine and at least one electric machine, in which a transmission has at least one transmission input shaft, one transmission output shaft and three planetary gear sets, with two between a drive and the second planetary gear set Power paths or partial transmission are each formed with a fixed input gear ratio, and in which the first planetary gear is assigned to the first or the second power path. The driveaway electric machine is associated with the first power path and can be brought into operative connection with the transmission input shaft or the internal combustion engine via a dog clutch or claw brake. Furthermore, the second planetary gear set is connectable to the first and second power paths. The third planetary gear set is in turn connectable to the second power path and the second planetary gear set as well as drivingly connected on the output side to the transmission output shaft.

For the realization of six to eight sequentially power shiftable forward gears seven to nine preferably form-fitting executed switching elements are arranged, wherein the switching elements predominantly in double-sided Schaltelemen- th or switching packages, each with two switching positions, which are mutually actuated by an actuator summarized. In the case of a possible, geometrically stepped, shift pattern of this transmission, the dog clutch or dog brake, which connects the driving-active electric machine to the transmission input shaft, is opened and closed sequentially. In a closed clutch state results in an internal combustion engine driving in the odd gears. In an open clutch state results in an electromotive driving operation in the odd gears and an internal combustion engine driving in the even gears. In the gear changes, a load circuit via the electric motor driven gears as support gears.

Against this background, the invention has the object to provide a vehicle transmission, which allows a comparatively large number of traction-free switchable gears, which is inexpensive to produce, and which can be used for conventional and hybrid powertrains.

The solution of this problem arises from the features of the main claim, while advantageous embodiments and further developments of the invention are the dependent claims can be removed.

The invention is based on the finding that a vehicle transmission consisting of several planetary gear sets, which can be coupled to one another, can be operated by a suitable connection to two transmission input shafts in an internal combustion engine drive train and in a hybrid drive train, wherein the transmission input shafts via separating clutches or separating brakes to the drive machine or the prime movers can be coupled. The planetary gear sets allow a large number of gears in a compact design that manages with relatively few wheel planes. Two transmission input shafts can be used in particular to form a dual-clutch transmission with two independent power paths, so that a power shiftable, sequential gear sequence can be realized. By connecting an electric machine to one of the two transmission input shafts, a hybrid drive can be realized. A switchable coupling The performance paths with each other can also extend the translation and drive options.

Accordingly, the invention is based on a vehicle transmission with a drive shaft, with a first and a second transmission input shaft, with at least one separating element which is assigned to the second transmission input shaft, with a main shaft, with an output shaft and with at least one first, second and third planetary gearset, which comprise at least in each case a ring gear, a sun gear and a planet carrier with planetary gears as elements, as well as with a plurality of switching elements for switching gear ratios or drive connections. The two transmission input shafts are each assigned a partial transmission, and one of the two partial transmissions has at least the first planetary gear set and the other of the two partial transmissions has at least the second planetary gear set.

To achieve the object is also provided that the first planetary gear is the two transmission input shafts drive technology upstream, wherein a first of the elements of the first planetary gear, which is effective as the drive element, the drive side directly or indirectly connected to the drive shaft or connectable and the transmission side means the at least one separating element with the second transmission input shaft is connectable, wherein a second of the elements of the first planetary gear, which is effective as its output element, the transmission side connected to the first transmission input shaft or connectable, wherein the first transmission input shaft with the second transmission input shaft or at least with the Main shaft is connectable, wherein the two transmission input shafts are each operatively connected to one or both of the second and third planetary gear sets, and in which the main shaft with the output shaft or at least mi t is connected to one of the elements of the third planetary gear set, and that at least seven sequentially power-shiftable forward gears are switchable by means of the two partial transmissions, wherein one of these forward gears is a direct gear or an overdrive gear. By this arrangement, a vehicle transmission is provided which can be used both in a drive train of a hybrid vehicle (hybrid transmission) and as a conventional power shift transmission, which has a relatively large number of gears and a relatively simple and compact design. The gears of the various embodiments of this vehicle transmission are completely or at least largely power-shiftable, resulting in a comfortable driving operation. The vehicle transmission has two input shafts, each of which forms two independent power paths or partial transmissions with one of the first two planetary gear sets, wherein one gear can be preselected in the respective load-free partial transmission, while the respective other partial transmission currently transmits the applied load. Thus, two independent power paths are switchable between the drive and the second planetary gear set. A third planetary gear set downstream of the power paths in terms of drive technology can be used flexibly with the two partial transmissions individually or together in operative connection.

The proposed vehicle transmission is operable as a two-input shaft transmission, for example, with two input friction clutches or an input clutch and an input brake for selectively connecting the partial transmissions, wherein the drive torque of an internal combustion engine is transmitted to the respective partial transmission. However, it is also possible that one of the two partial transmissions can be driven by an electric motor directly by an electric machine, wherein a positive coupling can fulfill the function of a separating or coupling element to the other partial transmission and / or to an internal combustion engine. By reversing the direction of rotation of the electric drive, reverse drive operation can be implemented in a hybrid drive train.

In a conventional, so purely internal combustion engine drive, a reversing gear for a reversal of rotation for the realization of reverse gear ratios can be provided and used at various points in the proposed transmission structure. Furthermore, the arrangement allows a demand coupling of the two partial transmission with each other, which is advantageously used in particular for the realization of direct gears and / or Overdrivegängen. The largest gear is preferably switchable as a direct gear or overdrive gear. In addition, a simple extension of the transmission with an additional transmission group, which is the main gearshift enabled, possible.

The vehicle transmission according to the invention is thus very flexibly insertable into a hybrid drive train, a dual-clutch transmission drive train, a group transmission drive train or combinations thereof both in the passenger and in the utility vehicle area.

According to a preferred embodiment of the invention can be provided that the vehicle transmission is designed as a dual-clutch transmission, with a first separating element and a second separating element, which are designed as friction clutches, that the first friction clutch is connected on the input side to the ring gear of the first planetary gear, which as the output element is effective that the first friction clutch is connected on the output side to the first transmission input shaft, that the second friction clutch is connected on the input side to the planet carrier of the first planetary gear, which is effective as its drive element, and that the second friction clutch is connected on the output side to the second transmission input shaft ,

Accordingly, the vehicle transmission can be designed as a dual-clutch transmission with two partial transmissions in planetary design. The drive shaft, the two friction clutches with the two transmission input shafts, the main shaft and the output shaft and the planetary gear sets may preferably be in a compact, coaxial arrangement in which a plurality of wave levels are superposed, on which switching elements are arranged, which for variable coupling of transmission elements or Shafts of the wheelsets of shift actuators are actuated.

The proposed transmission structure allows a gear sequence in which the gears can be, for example, geometrically stepped, that is, with an increasing difference in the shifting speed of the maximum speed in the gears. The load transition from the respective active gear to the following gear can be determined by an over- cutting opening and closing of the two friction clutches or separating elements take place, wherein traction interruption-free gear changes can be realized.

A judged as advantageous Basisradsatz arrangement for such a dual-clutch transmission can be realized in that the three planetary gear sets are switchable by means of a first, a second and a third switching element with two switching positions and by means of a fourth switching element with a switching position,

in the first planetary gear set, the ring gear can be connected to the first transmission input shaft by means of the first separating element, the sun gear is fixed or lockable on a non-rotatable component, and the planet carrier can be connected to the second transmission input shaft by means of the second separating element, that in the second planetary gear set the ring gear by means of the first switching element with the first transmission input shaft and by means of the second switching element with the second transmission input shaft is connected, the sun gear is fixed or detectable on a rotationally fixed component, and the planet carrier by means of the second switching element with the second transmission input shaft and by means of the third switching element with the planet carrier the third planetary gear set is connectable,

that in the third planetary gear set the ring gear is fixed or detectable on a non-rotatable component, the sun gear by means of the third switching element with the planet carrier of the second planetary gear set is connectable and by means of the first switching element with the first transmission input shaft is connected, and the planet carrier is connected to the output shaft .

that the main shaft is directly connected to the output shaft,

that eight forward gears are switchable, which are sequentially load-switchable by means of the first and second separating element,

wherein the seventh gear is switchable as a direct gear, which is switchable by means of the second separating element, the second switching element and the third switching element,

and that the eighth gear is an overdrive gear, which is switchable by means of the first separating element and the fourth switching element, wherein by the fourth switching element element a direct connection of the first transmission input shaft to the main shaft can be produced.

It should be noted at this point that a switching element may comprise both a single and a plurality of switching devices combined to form so-called switching packets. A switching position is understood to be a position of a switching element in which a non-positive connection of two components exists or is produced by the switching element. A switching element with, for example, two switch positions can therefore alternately produce or release a first or a second non-positive connection. A switching element also has a neutral position in which it is positioned connectionless. The switching elements can be designed as inexpensive form-locking claw switching elements.

In the disclosed base gearset, the first planetary gearset belongs to the first sub-transmission defined by the first friction clutch and the first transmission input shaft, and the second planetary gearset belongs to the second sub-transmission defined by the second friction clutch and the second transmission input shaft. Since in each case one of the elements of the planetary gear, namely the sun gear, in the first and the second planetary gear set to a non-rotatable component is detected or at least detectable, and in each case a second element, namely the ring gear, connected to the first or second transmission input shaft or connectable, the two first planetary gear sets act as input constants of their partial transmissions with a respective fixed transmission ratio.

A power-shiftable sequential gear sequence results from combining the Planetenradsatzübersetzungen, such that the respective sequence can be selected without load and the load transition by deactivating the respective one power path and activating the other power path via the separating elements or friction clutches.

The arrangement of this Basisradsatzes a compact dual clutch transmission is realized with a direct gear and an overdrive gear, wherein the Direct gear on the second clutch and the Overdrivegang are switchable via the first clutch. According to this arrangement, with three planetary gear sets and four shift elements, which have a total of seven shift positions, eight load-shiftable forward gears are possible. The gear ratios are preferably geometrically stepped.

Since the translation of the first planetary gearset of the first transmission input shaft is connected upstream of drive technology, is formed by the direct connection of the first transmission input shaft via the main shaft to the output shaft in this arrangement not the direct gear but the overdrive gear. For the production of the direct connection of the first transmission input shaft with the main shaft and with the, in this first embodiment with the main shaft fixedly connected output shaft, only the fourth switching element is required. This switching element is required in this embodiment, only for the preparation of this direct connection.

The overdrive gear has only low drag losses. This is achieved by the main shaft is guided directly to the output, the second planetary gear and its switchable input and output side coupling are realized on higher wave levels, ie on wave planes, which coaxially over a through the drive shaft, the first transmission input shaft, the main shaft and the output shaft defined wave plane lie.

The planetary gear sets can be designed as a simple Minusget ebe, ie as a planetary gear with a negative state translation, the state translation is given by the gear ratio of two planetary gear elements at festgehaltenem planet carrier and receive the numbers of teeth of internal gears ring gears or wheels according to the standard standard negative sign. The two rotating in the case of stationary translation elements, ie ring gear and sun gear, have opposite directions of rotation. If, however, the planet carrier used as input or output element and one of the other two elements, such as the sun, held, resulting in the same direction of rotation between the drive and output. In principle, plus planetary gear sets for the vehicle transmission are possible, in which case the planet carrier and Hohlradanbindungen are to be exchanged, since the ring gear and sun have this same direction due to double Planetenradreihen. The stand ratio then increases by the amount 1 compared to a corresponding negative gear.

In order to further simplify the base gear set, it can be provided that the three planetary gear sets can be switched by means of the first, second and third shift element with two shift positions each and by means of the fourth shift element with a shift position,

that the main shaft is connected to the sun gear of the third planetary gear set,

the seventh gear can be switched as a direct gear, which is switchable by means of the second disconnecting element and the second shifting element as well as the third shifting element, wherein the eighth gear is an overdrive gear, which is switchable by means of the first separating element and by means of the second, third and fourth switching element, wherein by means of the fourth switching element, a connection of the first transmission input shaft to the second transmission input shaft can be produced.

Accordingly, in this Basisradsatz an overdrive gear is generated by a coupling of the two partial transmission instead of a direct connection of the first transmission input shaft to the main shaft. As a result, a wave plane can be saved in the region of the main shaft between the second and third planetary gear set.

This can be realized, in particular, in that the torque-transmitting elements of all three planetary gear sets are connected in series for switching the overdrive gear, wherein on the drive side the first transmission input shaft is connectable to the output element of the first planetary gear set via the first separating element, and the output side is the third output element to the output shaft of the transmission is connected, so that the planetary gear sets are indeed coupled to each other, but only the translation of the first planetary gear is effective for the output. The actual partial transmission coupling takes place via the fourth

Switching element. Furthermore, the second and third switching element for realizing the direct gear and the overdrive gear are to be switched.

According to a further embodiment of the invention it can be provided that the three planetary gear sets are switchable by means of the first, second and third switching element, each with two switching positions,

in the first planetary gear set, the ring gear can be connected to the first transmission input shaft by means of the first separating element, the sun gear is fixed or lockable on a non-rotatable component, and the planet carrier can be connected to the second transmission input shaft by means of the second separating element, that in the second planetary gear set the ring gear by means of the first switching element with the first transmission input shaft and by means of the second switching element with the second transmission input shaft is connectable, the sun gear is fixed or detectable on a non-rotatable component, and the planet carrier by means of second switching element with the second transmission input shaft and by means of the third switching element with the planet carrier of the third planetary gear set is connectable,

that in the third planetary gear set the ring gear is fixed or detectable on a non-rotatable component, the sun gear by means of the third switching element with the planet carrier of the second planetary gear set is connectable and by means of the first switching element with the first transmission input shaft and the planet carrier is connected to the output shaft .

that seven forward gears are switchable, which are sequentially load-shiftable by means of the first and second separating element,

wherein the seventh gear is a direct gear, which is switchable by means of the second separating element and by means of the second and third switching element.

Accordingly, seven forward gears are switchable in this Basisradsatz with only three double shift elements. Since a fourth switching element is required only for the partial transmission coupling in eighth gear, can be constructed by simply omitting this switching element, a particularly compact and lightweight seven-speed dual-clutch transmission. The seventh gear is the highest gear in this transmission as a direct gear switchable.

In addition, it can be provided that a further switching element is arranged, by means of which in the third planetary gear, the ring gear is alternately releasably lockable to a rotationally fixed component or connectable to the planet carrier.

Accordingly, in the Basisradsatz the connection of the ring gear of the third planetary gear set can be performed by an additional switching element as a detachable connection. In those gears in which the translation of the third planetary gear is needed, the ring gear is detected. In those gears in which the translation of the third planetary gear set is not needed, with dissolved ring gear instead of a now free running of the planetary gears and / or the sun gear, the circuit of a block circulation of the third planetary gear set are made possible. As a result, unnecessary bearing losses of freely moving wheels can be avoided in the courses concerned. The block circulation can be achieved in that the additional switching element in addition to the switching position for detecting the ring gear, a second switching position for connecting two elements of the planetary, for example, the ring gear with the planet carrier having. The additional switching element ensures in this switching position defined by the block circulation speed ratios on the third planetary, without being self-load.

In principle, it is also possible, with dissolved ring gear to bring about the block circulation by suitable combinations of switching positions of other, already existing switching elements and to dispense with the second switching position, if this is a possible shift pattern of the transmission allowed and this is appropriate.

It is also possible, even with the first two planetary gear set element, ie in particular the sun, each releasably connectable by an additional switching element with the non-rotatable component or gear housing and to allow a block circulation to reduce bearing losses.

In order to save further installation space and weight, adjacent switching elements, which are never closed simultaneously in the possible or at least in the preferred switching schemes, can be combined to form switching elements with a plurality of switching positions, which are actuated alternately via a single actuator, as switching packets. As is known, double-sided or double-acting shift elements, each with two shift positions and an intermediate neutral position in various transmissions, are frequently used. The transmission structure of the invention also allows triple switching elements. Accordingly, it can be provided that, in particular in the second embodiment of the Basisradsatzes, for example, the respective first and fourth switching elements are summarized in a single switching element with three switching positions summarized. This is possible because the fourth shift element is needed only in the highest gear for the partial transmission clutch. This results in an additional space and weight advantage.

A further space advantage can be achieved in that the second planetary gear set is arranged radially above the third planetary gear, wherein these two planetary gear sets are constructed axially nested one inside the other. This can save a wheel plane and thus the transmission structure can be shortened axially.

In addition, it can be provided that in the vehicle transmission, a first separating element and a second separating element are arranged, wherein the first separating element is designed as a brake, by means of which the sun gear of the first planetary gear on a non-rotatable member (GH) is braked or solvable by this that the second separating element is designed as a friction clutch, which is connected on the input side to the planet carrier of the first planetary gear set, which is effective as the drive element, and the output side is connected to the second transmission input shaft, and wherein the ring gear of the first planetary gear set, which as its output element is effective, is connected to the first transmission input shaft.

Accordingly, instead of two friction clutches, a friction clutch and a brake may alternatively be used. This is possible because the first planetary gear set is effective as an input constant of the first subtransmission. Accordingly, instead of a first clutch for activating the gears of the first partial transmission each brake is closed and thus the sun gear braked and opened the second clutch while activating the gears of the second partial transmission in each case the second clutch closed and the brake for load-free preselection of the respective subsequent gear is solved. The brake thus assumes the function of the first separating element. A circuit diagram of the gearbox bes may be the same in both embodiments, ie with two friction clutches or with a friction clutch and a brake.

In order to realize at least one reverse gear in the vehicle transmission in the case of a purely internal combustion engine driving operation, another simple planetary gear set may be arranged in the same, which is effective as a reversing gear between the drive and output direction reversal. The reversing wheel set may be integrated at different locations in the transmission structure.

According to a further embodiment of the invention can be provided in this regard, that for the realization of up to eight reverse gears, a fourth planetary, which is effective as a Wenderadsatz, and a fifth switching element, which has two switching positions, arranged axially in front of the first planetary gear and upstream of this drive technology are, wherein in the fourth planetary gear, the ring gear is connected to the planet carrier of the first planetary gear, the sun gear is connected to the drive shaft, and the planet carrier by means of the fifth switching element alternately lockable on a non-rotatable component or with the sun gear of the fourth planetary gear set is connectable.

The Wenderadsatz is therefore at the transmission input, in the power flow before the first planetary gear and the separating elements, integrated in the transmission. The eight reverse gears are sequentially power-shiftable via the two disconnect couplings. The fifth shift element is for switching between the reverse ratios and the forward gear ratios.

This arrangement allows eight reverse gears, which can be translated relatively short. For example, the reverse gears can have about 1.5 times as high a ratio as the corresponding forward gears. Accordingly, in particular reverse gears can be realized, which generate a very low driving speed when idling of an engine designed as an internal combustion engine, so that with completely closed friction clutch and without accelerator pedal operation a comfortable, sensitive rear maneuvering al- lein by pressing the brake pedal is possible. Because of the short translation of the drive torque is in the reverse gears, a torque limit of the engine useful to limit the load on the transmission.

In another preferred embodiment of the invention can be provided that for the realization of up to four reverse gears, a fourth planetary gear, which is effective as a Wenderadsatz, and a fifth switching element, which has two switching positions, axially disposed between the first planetary gear set and the second planetary gear set In the fourth planetary gear set, the ring gear by means of the second switching element is alternately connected to the ring gear or the planet carrier of the second planetary gear set, the sun gear of the fourth planetary gear set is connected to the second transmission input shaft, and the planet carrier by means of Fifth switching element alternately lockable on a non-rotatable component or connectable to the sun gear of the fourth planetary gear set.

This arrangement allows four reverse gears, the larger one

Gear jump compared to the forward gears may have. The reverse gears are realized in this arrangement, all over the same friction clutch or the same partial transmission and thus not power shiftable. However, by changing the load-carrying friction clutch shifting under load between a reverse gear and a forward gear is possible.

In a further embodiment of the invention can be provided that for the realization of up to four reverse gears a fourth planetary gear, which is effective as a Wenderadsatz, and a fifth switching element, which has two switching positions, axially disposed between the first planetary gear set and the second planetary gear set and are arranged upstream of the second planetary gear set, wherein in the fourth planetary gear, the ring gear by means of the second switching element is alternately connectable to the ring gear or the planet carrier of the second planetary gear set, the sun rad of the fourth planetary gear netenradsatzes is connected to the second transmission input shaft, and the planet carrier by means of the fifth switching element alternately lockable on a non-rotatable component or with the sun gear of the fourth planetary gear set is connectable.

According to another variant of the invention it is provided that for the realization of one or two reverse gears, a fourth planetary, which is effective as a Wenderadsatz, and a fifth switching element, which has a switching position, arranged axially between the second planetary gear set and the third planetary gear and the second planetary gear set are arranged downstream of the drive, that the fifth switching element and the second switching element are combined into a single switching element with three switching positions, that in the fourth planetary gear, the ring gear is connected to the planet carrier of the second planetary gear set, the sun gear by means of the fifth switching element with the second transmission input shaft is connectable, and wherein the planet carrier is detected or detectable together with the sun gear of the second planetary gear set on a non-rotatable component.

Although this gear arrangement allows only two reverse gears. For this purpose, however, a separate fifth switching element can be saved, since the function for switching from the forward gears in the reverse gears in the existing second switching element can be integrated as a third switching position, which reduces the cost and space requirements. In addition, in this arrangement, the first and fourth switching elements can be combined into a single switching element with three switching positions.

The transmission structure according to the invention with two partial transmissions or with two power paths via two transmission input shafts also allows a simple implementation in a hybrid powertrain.

In a further preferred embodiment of the invention, accordingly, the vehicle transmission may be formed as a so-called hybrid transmission, wherein it is provided that the second transmission input shaft to the rotor an electric machine is drivingly connected, that a first separating element and a second separating element are arranged, wherein the first separating element is designed as a friction clutch, by means of which the drive shaft with the planet carrier of the first planetary gear set, which is effective as the drive element, connectable, and wherein the second separating element is designed as a form-locking coupling, by means of which the planet carrier of the first planetary gear set can be connected to the second transmission input shaft in a controlled manner.

Accordingly, the second transmission input shaft may be drivingly connected to the rotor of an electric machine. By the electric motor then a purely electric motor driven driving is possible in the second part of the transmission. The first partial transmission can be operated by internal combustion engine. Instead of a friction clutch can be arranged on the second transmission input shaft, a positive coupling, via which the electric machine with the planet carrier, so the drive element of the first planetary gear set is connectable. This allows an electric motor and combustion engine combined driving operation. Due to the switchable connection between the electric machine and the internal combustion engine, the known hybrid functions are also possible, such as charging a battery, boosting and starting the internal combustion engine by the electric machine. In order to decouple the planet carrier of the first planetary gear set and thus the transmission completely from the internal combustion engine, or to connect the internal combustion engine as needed, a friction clutch is arranged at the transmission input in this embodiment.

A possible shift pattern for this hybrid transmission, with a load-shiftable gear sequence, may correspond to a shift pattern of a transmission according to the embodiments with two friction clutches or with a friction clutch and a brake.

Due to the possibility of reversing the direction of rotation of the electromotive drive reverse gears can be realized without an additional Wenderadsatz in the hybrid powertrain, in particular, the translation of the smallest forward gear for the electromotive reversing operation is available. In another embodiment of the vehicle transmission for a hybrid drive train is provided that the vehicle transmission is designed as a so-called hybrid transmission, in which the second transmission input shaft is operatively connected to the rotor of the electric machine, in which a separating element is arranged, which is designed as a form-locking coupling , By means of which the planet carrier of the first planetary gear train side is connectable to the second transmission input shaft, and in which the drive shaft with the planet carrier of the first planetary gear set, which is effective as the drive element is connected.

Accordingly, it is also possible to dispense with the hybrid design of a vehicle transmission according to the invention entirely on an input-side friction clutch and provide only a positive coupling for switchable connection of the electric machine with the drive shaft via the first planetary gear set. The starting operation of the vehicle then takes place exclusively via the electric machine.

Furthermore, it can be provided that the form of a positive coupling formed separating element and the fourth switching element are combined as a single switching element with two switching positions in which alternately the planet carrier of the first planetary gear set side is connectable to the second transmission input shaft, or the first connectable to the second transmission input shaft is. This ensures that in the hybrid transmission all the switching elements are designed as double-sided switching packages, resulting in further cost and space advantages.

The embodiments of the vehicle transmission according to the invention described so far allow with three planetary gear sets eight powershift forward gears, including a direct gear and an overdrive gear, or seven powershifted forward gears in the absence of an overdrive gear. With an additional reversing wheel, up to eight reverse gears are possible. When using a generator and electric motor operable electric machine can alternatively To be an electric motor driven reversing function without additional Wenderadsatz be realized.

In addition, the vehicle transmission can be expanded by an extension with a front group and / or a Nachschaltgruppe to a group transmission, whereby the number of revolutions of a main transmission in the construction of the embodiments described so far can be doubled. This can be useful in particular for applications in commercial vehicles.

According to a further preferred embodiment of the invention can therefore be provided that the vehicle transmission is designed as a dual-clutch group gearbox, in which the first, second and third planetary gear set are switchable at least by means of a first, second and third switching element with two switching positions,

in that a range group is arranged downstream of the third planetary gear set,

the range group comprising a fourth planetary gear set formed as a turning wheel, to which a fifth shift element is associated with a shift position for shifting a reverse gear group, and a fifth planetary gear set, which is associated with a sixth shift element with two shift positions for switching between a slow and a fast forward gear group the fourth planetary gear set, the ring gear is connected to the sun gear of the fifth planetary gear set, the sun gear is connected to the planet carrier of the third planetary gear set, and the planet carrier is connected to the ring gear of the fifth planetary gear and is detectable by means of the fifth switching element on a rotationally fixed component,

that in the fifth planetary gear set, the sun gear by means of the sixth switching element alternately lockable on a non-rotatable component or is connectable to the planet carrier, and the planet carrier is connected to the output shaft,

that at least fourteen forward gears and at least seven reverse gears are switchable by means of five shift elements with a total of nine shift positions, that at least thirteen of the at least fourteen forward gears are power-shiftable and the fourteenth forward gear is a direct gear, and in which the at least seven reverse gears are all power-shiftable.

Accordingly, the number of revolutions of a seven-speed main transmission with the features of the invention by means of a range group can be doubled by this gear arrangement. The changeover of the area group is interrupted without further action. However, the interruption of tractive effort can be minimized by designing a smaller gear jump in the range changeover and thus a rather low loss of speed compared to the other gears. All other gears, including the seven reverse gears, are power shiftable. In particular, the downstream translation of the fifth planetary gear set can provide short gears, as are generally required for commercial vehicle applications. The ratios of the reverse gears may be comparable to the ratios of the corresponding forward gears.

According to a further preferred embodiment of the invention, it can be provided that the vehicle transmission is designed as a dual-clutch group transmission in which the first, second and third planetary gear set is switchable, each with at least two switching positions, by means of a first, second and third shifting element,

a fourth planetary gear set, which is effective as a range group, and a sixth shift element, which has two shift positions for switching between a slow and a fast forward gear group, are arranged downstream of the third planetary gear set,

that in the fourth planetary gear set the ring gear is alternately lockable on a rotationally fixed component or connectable to the planet carrier, the sun gear is connected to the planet carrier of the third planetary gear set, and the planet carrier is connected to the output shaft,

a fifth planetary gear set, which is effective as a turning wheel set, and a fifth switching element, which has two switching positions, are arranged upstream of the first planetary gear set, that in the fifth planetary gear, the ring gear is connected to the planet carrier of the first planetary gear, the sun gear is connected to the drive shaft, and the planet carrier by means of the fifth switching element alternately lockable on a non-rotatable component or connectable to the sun, that by means of five switching elements with a total ten shift positions at least fourteen forward gears and at least seven reverse gears are switchable that of the at least fourteen forward gears at least thirteen are power shiftable and the fourteenth forward gear is a direct gear, and that the at least seven reverse gears are all power shiftable.

This last arrangement therefore has a fourth planetary gear set, which is arranged downstream of the main gear as a range group, and a fifth planetary gear set, which is connected upstream of the main gear as a reversing gear set. In one possible shift scheme, a lower and an upper, respectively seven-speed gear group can be realized, wherein the range change is traction-interrupted. The gear jump of the area change is expediently designed to be relatively small. As a result, the range change can take place at a minimized speed adaptation of the internal combustion engine, which favors a particularly short shift time. The reverse gears can be translated relatively short and have, for example, the 1, 8-fold translation of the corresponding forward gears, which is advantageous for a shunting operation. In order to limit the load on the transmission, a torque limitation of the internal combustion engine in the reverse drive mode makes sense.

To further illustrate the invention, the description is accompanied by a drawing with several embodiments. In this shows

1 shows a transmission diagram of a first embodiment of a vehicle transmission according to the invention, with two input-side clutches and three planetary gear sets, with a switchable direct connection between a first transmission input shaft and an output shaft, 2 is a transmission diagram of planetary gear sets for a vehicle transmission according to FIG. 1, FIG.

3 is a circuit diagram for the 8-speed vehicle transmission according to FIG. 1, with a direct gear and an overdrive gear, FIG.

4 shows a transmission diagram of a second embodiment of a vehicle transmission according to the invention, with a shiftable partial transmission clutch via a transmission input shaft connection,

5 is a circuit diagram for the 8-speed vehicle transmission according to FIG. 4, with a direct gear and an overdrive gear, FIG.

6 is a transmission diagram of a third embodiment of a vehicle transmission, with a triple switching element,

7 shows a transmission diagram of a fourth embodiment of a vehicle transmission, exclusively with double shift elements,

8 is a circuit diagram for the 7-speed vehicle transmission of FIG. 7 with a direct gear,

9 is a transmission diagram of a fifth embodiment of a vehicle transmission, with radially nested planetary gear sets,

10 is a transmission diagram of a sixth embodiment of a vehicle transmission, with an input-side clutch and an input-side brake,

11 is a circuit diagram for the 8-speed vehicle transmission according to FIG. 10, with a direct gear and an overdrive gear, FIG.

12 is a transmission diagram of a seventh embodiment of a vehicle transmission, with a first hybrid drive assembly, 13 is a circuit diagram of the 8-speed vehicle transmission according to FIG. 12, with a direct gear and an overdrive gear, FIG.

14 is a transmission diagram of an eighth embodiment of a vehicle transmission, with a second hybrid drive arrangement,

15 is a circuit diagram for the 8-speed vehicle transmission according to FIG. 14, with a direct gear and an overdrive gear, FIG.

6 is a transmission diagram of a ninth embodiment of a vehicle transmission, with a first arrangement of a reverse planetary gear set,

17 is a transmission diagram of planetary gear sets for a vehicle transmission according to FIG. 1 6,

18 is a circuit diagram for the 8-speed vehicle transmission according to FIG. 1 6, with a direct gear and an overdrive gear and with eight reverse gears,

19 is a transmission diagram of a tenth embodiment of a vehicle transmission, with a second arrangement of a reverse planetary gear set,

20 is a circuit diagram for the 8-speed vehicle transmission according to FIG. 19, with a direct gear and an overdrive gear and with four reverse gears,

21 is a transmission diagram of an eleventh embodiment of a vehicle transmission, with a third arrangement of a reverse planetary gear,

22 is a transmission diagram of planetary gear sets for a vehicle transmission according to FIG. 21, FIG.

23 is a circuit diagram for the 8-speed vehicle transmission according to FIG. 21, with a direct gear and an overdrive gear and with two reverse gears, 24 is a transmission diagram of a twelfth embodiment of a vehicle transmission, with an additional switching element,

25 is a circuit diagram for the 8-speed vehicle transmission of FIG. 24, with a direct gear and an overdrive gear,

26 shows a transmission diagram of a 14-speed embodiment of a vehicle transmission, with a range group with an integrated reverse gear planetary gear set,

27 is a translation scheme with state translations of planetary gear sets for a vehicle transmission according to FIG. 26,

28 is a circuit diagram for the 14-speed vehicle transmission of FIG. 26, with a direct gear and seven reverse gears,

29 is a transmission diagram of a fourteenth embodiment of a vehicle transmission, with a range group and with a reverse planetary gear as a Vorschaltgruppe,

Fig. 30 is a translation scheme with state translations of planetary gear sets for a vehicle transmission according to FIG. 29, and

Fig. 31 is a circuit diagram for the 14-speed vehicle transmission of FIG. 29, with a direct gear and seven reverse gears.

As an introduction it should be mentioned that in the figures all components of the same or functionally identical are provided with the same reference numerals for the sake of simplicity.

Accordingly, the vehicle transmission schematically illustrated in FIG. 1 essentially comprises three planetary gear sets PG1, PG2, PG3, a drive shaft AW, two transmission input shafts GE1, GE2, two frictional clutches. te separating elements K1, K2, a main shaft HW and an output shaft AB, which are located in a mutually coaxial arrangement.

The three planetary gear sets PG1, PG2, PG3 are designed as simple minus gears, each comprising a radially outer ring gear HR1, HR2, HR3, an inner sun gear SR1, SR2, SR3 and a planet carrier PT1, PT2, PT3, wherein the planet carrier PT1, PT2 , PT3 each have a plurality of planet gears PR1, PR2, PR3, which mesh with the sun gear SR1, SR2, SR3 and the ring gear HR1, HR2, HR3.

The first planetary gear set PG1 is arranged at the transmission input. His planet carrier PT1 is the input side rotatably connected to the drive shaft AW, which is drivingly connected to a drive machine, not shown, for example, designed as an internal combustion engine. The planet carrier PT1 is thus effective as a drive element of the first planetary gear PG1. The planet carrier PG1 of the first planetary gear set PG1 is the transmission side connected to the second friction clutch K2. The second friction clutch K2 is drive-connected on the output side or on the transmission side to the second transmission input shaft GE2, which is designed as a hollow shaft. The sun gear SR1 of the first planetary gear set PG1 is fixed to a non-rotatable component GH, for example, a transmission housing. The ring gear HR1 of the first planetary gear set PG1 is drive-connected to the first friction clutch K1 and thus effective as an output element of the first planetary gear PG1. The first friction clutch K1 is drive-connected to the first transmission input shaft GE1 on the transmission side, which is formed with respect to the second transmission input shaft GE2 as a radially inner shaft, the transmission side emerges from the radially outer second transmission input shaft GE2. The first planetary gear set PG1 forms with the first friction clutch K1 and the first transmission input shaft GE1 a first partial transmission TG1 with a first fixed input ratio. To produce a drive connection between the drive shaft AW and the second friction clutch K2, the first planetary gear set PG1 is only bridged via the planet carrier PT1. The second planetary gear set PG2 following in the axial direction and in terms of drive technology forms with the second friction clutch K2 and the second transmission input shaft GE2 a second partial transmission TG 2 with a second fixed input ratio, the ring gear HR2 of which being the drive element and the planet carrier PT2 acting as the output element. The sun gear SR2 of the second planetary gear set PG2 is in turn fixed to the rotationally fixed component GH.

Coaxially and axially adjacent to the two transmission input shafts GE1, GE2, the main shaft HW is arranged. The output end of the main shaft HW is guided axially through the third planetary gear PG3 and rotatably connected to the output shaft AB.

The three planetary gear sets PG1, PG2, PG3 are by means of a first, a second and a third switching element S1, S2, S3, each with two switching positions A / B, C / D, E / F, which are mutually switchable, and by means of a fourth switching element S4 with only one switch position G switchable.

The first transmission input shaft GE1 or the first friction clutch K1 can be connected by means of the first shifting element S1 in its first shifting position A to the ring gear HR2 of the second planetary gear set PG2. By the second switching position B of the first switching element S1, the first transmission input shaft GE1 and the second friction clutch K2 with the sun gear SR3 of the third planetary gear set PG3 can be connected. Furthermore, in the second planetary gear PG2 of the planet carrier PT2 via the third switching element S3 in its second switching position F with the planet carrier PT3 of the third planetary gear PG3 connectable. In addition, the first transmission input shaft GE1 via the fourth switching element S4 in its switching position G directly to the main shaft HW and thus to the output shaft AB connectable.

The second transmission input shaft GE2 or the second friction clutch K2 is connected via the second switching element S2 alternately in its first switching position C with the ring gear HR2 or in its second switching position D with the planet carrier PT2 of the second planetary PG2 connectable. In the third planetary gear set PG3, the ring gear HR3 is fixed to the non-rotatable component GH, the sun gear SR3 is connected via the third switching element S3 in its first switching position E with the planet carrier PT2 of the second planetary gear set PG2, and the planet carrier PT3 is the output side fixed to the output shaft AB connected.

FIG. 2 shows as a numerical example a possible translation of the three planetary gear sets PG1, PG2, PG3 designed as negative gear units, whereby the positive planetary gear set transmission i_PG effective in the gear structure according to FIG. It can be seen that the drive and output elements have the same direction of rotation.

3 shows those shift positions of the transmission which are activated to set a respective gear, are indicated in the shift pattern with the lower case letter "x." Accordingly, in the transmission according to FIG The gears are activated in the switching sequence in a sequential change by closing the two clutches K1, K2, wherein by overlapping opening and closing of the clutches K1, K2 of the load transition between the two partial transmissions TG1, The operation of the transmission according to Fig. 1 is thus that of a dual-clutch transmission.

For example, the gear shift between the first gear "1" and the second gear "2" is as follows:

In the first gear "1", the second clutch K2 is closed, the second gear TG2 is in the switching position C, wherein the ring side of the ring gear HR2 of the second planetary gear PG2 with the second transmission input shaft GE2 and on the drive side the second friction clutch K2 and the planet carrier PT1 of the first planetary gearset PG1 are connected to the drive shaft AW. The third shifting element S3 is located in its first shift position E, in which the planet carrier PT2 of the second planetary gearset PG2 acts as an output element and is connected to the sun gear SR3 of the third planetary gear PG3, so that the translation of the third planetary PG3 acts on the output shaft AB. According to the example of FIGS. 2 and 3, a gear ratio of i = 4.99 results for the first gear "1".

In the second gear "2", the shift position E of the third shift element S3 remains in. In addition, the first shift element S1 is moved into its first shift position A, in which the ring gear HR2 of the second planetary gear set PG2 is connected to the first transmission input shaft GE1. because when switched first gear "1", the first clutch K1 is still open and thus the first part of the transmission TG1 is still inactive.

To carry out the gear change first gear "1" in the second gear "2" now the second clutch K2 is opened and the first clutch K1 closed, wherein the frictional engagement in the second clutch K2 degraded and built up in the first clutch K1. As a result, the load transfer from the second partial transmission TG2 to the first partial transmission TG1 takes place without loss of traction in the drive train. Subsequently, the second switching element S2 can be opened without load to disengage their previously switched in the first gear "1" switching position C.

In the circuit diagram of Fig. 3, for each of the eight gears "1" to "8", the gear ratio i is indicated. The gears "1" to "8" have a constant gear jump phi = 1, 31, ie a geometric gradation on. The seventh gear "7" is designed as a direct gear. In this gear "7", the drive shaft AW is drive-connected to the output shaft AB via the three planet carriers PT1, PT2, PT3 of the three planetary gear sets PG1, PG2, PG3 when the second clutch K2 is closed.

The eighth gear "8", which can be activated by closing the fourth shift element S4, is designed as an overdrive gear or overdrive gearshift in the relevant shift position G of the fourth shift element S4 and when the fourth shift element S4 is engaged. sener first clutch K1 is the translation of the first planetary PG1 directly on the output effective.

In the transmission structure according to FIG. 1, no reverse gear is included to clarify the basic structure of the transmission. The transmission structure therefore forms a Basisradsatz, which can be extended by a Wenderadsatz to realize at least one reverse gear.

FIG. 4 shows a base gearset modified in relation to FIG. 1, in which the output-side end of the main shaft HW is not connected directly to the output shaft AB but to the sun gear SR3 of the third planetary gear set PG3. In addition, via the fourth switching element S4 not a direct connection between the first transmission input shaft GE1 and the main shaft HW but a connection between the first transmission input shaft GE1 and the second transmission input shaft GE2 produced, whereby a coupling of the two partial transmission TG1, TG2 can be realized. This eliminates a wave plane between the main shaft plane HW and the planetary carrier shaft plane PT2 of the second planetary gear set PG2. Otherwise, this transmission structure corresponds to the transmission structure according to FIG. 1.

An associated circuit diagram is shown in Fig. 5. The eighth gear "8" is designed as Overdrivegang and is realized in this transmission structure by coupling the two partial transmission TG1, TG2 by means of the fourth switching element S4 and by switching the planet carrier PT2, PT3 of the second and third planetary PG2, PG3 on the output shaft AB For this purpose, the second and the third shift element S2, S4 are switched to their shift positions D and F. When shifting from the seventh gear "7", which is designed as a direct gear, into the overdrive gear (eighth gear "8"), the Switching positions D and F of the second and third switching elements S2, S4 in the region of the second and third planetary gear set PG2, PG3 otherwise the circuit diagram of Fig. 5 corresponds to the circuit diagram of Fig. 3 of the Basisradsatzes of FIG. The transmission structure according to FIG. 4 can be expanded by a turning wheel set in order to realize reverse gears. Different reverse gear variants will be explained later.

FIG. 6 shows a largely identical arrangement to the transmission structure according to FIG. 4, in which, however, the first and fourth switching elements S1, S4 are combined to form a triple switching element S4 / S1 with a total of three switching positions A, B, G. The three switch positions A, B, G are sequentially and mutually actuated with a common actuator. As the circuit diagram of FIG. 5 shows, these switch positions A, B, G are never closed at the same time, since they are assigned to the same partial transmission TG1. Therefore, this triple switching element S1 / S4 can be used with its three switch positions A / B / G.

Fig. 7 shows a further variant of the Basisradsatzes according to FIG. 4, in which, however, the fourth switching element S4 is omitted. As a result, a transmission structure is formed, which has only the first, the second and the third double switching element S1, S2, S3. Since the omitted fourth shift element S4 was required for the overdrive gear "8", only one shift pattern with seven gears "1" to "7" can be realized in the gear arrangement according to Fig. 7. This shift scheme is shown in Fig. 8. It corresponds to Figs to the omitted Overdrivegang the circuit diagram of FIG. 5 of the transmission of FIG. 4th

FIG. 9 shows a further variant of the base gearset according to FIG. 4, in which, however, the second and third planetary gear sets PG2, PG3 are arranged coaxially and radially one above the other in a common gear plane. The ring gear HR3 of the third planetary gearset PG3 is rotatably connected to the sun gear SR2 of the second planetary gear set PG2. The connection of the individual wheelset elements as well as the circuit diagram otherwise correspond to the transmission structure according to FIG. 4 or the circuit diagram according to FIG. 5.

10 shows a transmission structure in which a brake B1 is arranged instead of the first friction clutch K1. This is possible because the first planetary gearset PG1 is effective as an input constant of the first subtransmission TG1. By the brake B1, the sun gear SR1 of the first planetary gear set PG1 is festbremsbar to the stationary component GH and detachable from this. The first transmission input shaft GE 2, however, is rotatably connected to the ring gear HR1 of the first planetary gear set PG1.

A circuit diagram shown in Fig. 1 1 for the transmission structure of FIG. 10 largely corresponds to the circuit diagram of FIG. 5, wherein instead of the first clutch K1 now the brake B1 is actuated. The ratios of the eight forward gears "1" to "8" and the stationary ratios i_0 and the effective ratios i_PG of the planetary gear sets PG1, PG2, PG3 are identical to the transmission according to FIG.

Fig. 12 shows an embodiment, which is referred to as a hybrid transmission, because this transmission is advantageously used in the drive train of a hybrid vehicle with internal combustion engine and electric motor drive. In this arrangement, instead of a second separating element or a second friction clutch K2, the rotor EMR of an electric machine EM is connected to the second transmission input shaft GE2. Furthermore, a positive separating clutch X1 is present, by means of which the second transmission input shaft GE2 can be connected to the planet carrier PT1 of the first planetary gear set PG1 at its closed switching position X. The first separating element K1 is designed as a transmission input-side friction clutch, by means of which the planet carrier PT1 of the first planetary gear set PG1 on the drive side with the drive shaft AW connectable or detachable from this. As a result, an internal combustion engine-electric motor combined driving operation and a decoupling of the internal combustion engine from the drive train are possible. The gear structure otherwise corresponds to the transmission according to FIG. 4.

It can be seen that in the odd gears "1", "3 ^" , "5", "7", which are associated with the second partial transmission TG2, the drive means of the electric machine EM and means of the internal combustion engine or takes place. The first friction clutch K1 can remain closed in all gears. Basically, however, the odd the gears "1", "3", "5", "7", with open friction clutch K1, even a purely electric motor driving possible. In the even gears "2 ^" , "4", "6", "8", which are assigned to the first partial transmission TG1, the drive is effected only by the internal combustion engine or via the closed friction clutch K1. When switching from the seventh gear "7" in the eighth gear "8", ie from the direct gear in the overdrive gear, the switch positions D, F remain the second and third planetary gear sets PG2, PG3. However, the overdrive gear "8" can only be driven by an internal combustion engine since the clutch X1 must be opened in the overdrive gear "8" due to the partial gear coupling. In the gear changes, a load circuit via the electric motor driven gears can be done as support gears.

Fig. 14 shows a second embodiment of a hybrid transmission. In this transmission, the input-side first friction clutch K1 is omitted in comparison to the transmission of FIG. 12 and the fourth shift element S4 and the positive-locking disconnect clutch X1 are combined as a single, double-sided operable switching element X / S4. The hybrid transmission of FIG. 14 thus comes completely without friction clutches. The startup operations in this transmission therefore take place only by electric motor with the aid of an electric machine EM whose rotor EMR is drive-connected to the second transmission input shaft GE2. Characterized in that the fourth switching element S4 is combined with the positive separating clutch X1 to a double switching element X1 / S4, only double switching elements S1, S2, S3, X1 / S4 are present in the transmission. This is also possible in the hybrid transmission according to FIG. 12.

An associated circuit diagram is shown in FIG. 15. From this it can be seen that the even gears "2", "4", "6", "8" of the first partial transmission TG1 are switched without an input-side isolating element.

Figures 1 6 to 23 show various embodiments for installing a Wenderadsatzes in the transmission structure of FIG. 4 for the realization of reverse gears. Accordingly, according to FIG. 1 6, a fourth planetary gear set PG 4, which is effective as a turning wheel set, is arranged. The fourth planetary gear PG4 is the first planetary gear set PG1 axially and drive technology upstream and thus associated with the first partial transmission TG1. In addition, there is a fifth shift element S5 with two shift positions V, R for switching between a forward drive operation and a reverse drive operation. The ring gear HR4 of the fourth planetary gear set PG4 is connected to the planet carrier PT1, that is to say to the drive element of the first planetary gear set PG1. The sun gear SR4 of the fourth planetary gear set PG4 is connected to the drive shaft AW. The planet carrier PT4 of the fourth planetary gearset PG4 is mutually connectable by means of the fifth switching element S5 in its first switching position V to the drive shaft AW or detectable in its second switching position R on the non-rotatable component GH. By connecting the planet carrier PT4 of the fourth planetary gear set PG4 with the drive shaft AW, this is simultaneously connected to the sun gear SR4 of the fourth planetary gear set PG4, so that in the forward drive mode the turning gear set PG4 rotates in the block. By fixing the planet carrier PT4 of the fourth planetary gear set PG4 on the non-rotatable component GH, the negative stationary gear ratio of the fourth planetary gearset PG4 designed as a negative gear is active, so that the direction of rotation between the driving sun gear SR4 and the output ring gear HR4 of the fourth planetary gear set PG4 for a reverse drive operation turns.

Fig. 17 shows in a numerical example a translation table containing the additional planetary gear PG4, from which it can be seen that its effective translation i_PG = -1, 5 of the state ratio iO = -1, 5 corresponds.

For the eight forward gears "1" to "8" corresponds to the circuit diagram of the circuit diagram of Fig. 5 of the transmission structure of FIG. 4, wherein additionally the fifth switching element S5 always in the forward gear shift position V is located. In addition, eight reverse gears R1, R2, R3, R4, R5, R6, R7, R8 are realized and sequentially load switchable, wherein the fifth switching element S5 is always in the reverse gear position R. The ratio of the eight reverse gears R1 to R8 corresponds approximately to 1.5 times the eight forward gears "1" to "8". 19 shows a transmission structure with an alternative connection of a fourth planetary gear set PG4 between the first planetary gear set PG1 and the second planetary gear set PG2. In this case, the ring gear HR4 of the fourth planetary gear set PG4 can be connected to the ring gear HR2 or to the planet carrier PT2 of the second planetary gear set PG2 by means of the second switching element S2. The sun gear SR4 of the fourth planetary gear set PG4 is connected to the second transmission input shaft GE2. The planet carrier PT4 of the fourth planetary gear set PG4 is mutually connectable in its first switching position V with the second transmission input shaft GE2 or detectable in its second switching position R on the non-rotatable component GH by the fifth switching element S5. By connecting the planet carrier PT4 of the fourth planetary gear set PG4 with the second transmission input shaft GE2, this is simultaneously connected to the sun gear SR4 of the fourth planetary gear set PG4, so that in the forward drive mode, the fourth planetary gear set PG4 rotates in the block.

19 shows that in this transmission four non-power-shift reverse gears R1 to R4 are realized, in which in each case the second separating element K2 is closed and the first separating element K1 is opened. Their translations correspond approximately to 1, 5 times the corresponding forward gears.

Fig. 21 shows another arrangement of a fourth planetary gear set PG4, which is effective as a Wenderadsatz for the transmission. In this transmission structure, the fourth planetary gear set PG4 is disposed axially between the second planetary gear set PG2 and the third planetary gear set PG3. The fifth shift element S5 requires only one shift position R for activating the reverse drive function and is combined with the second shift element S2 to form a triple shift element S2 / S5 with three shift positions C, D, R. Furthermore, the first switching element S1 and the fourth switching element S4 are combined to form a further triple switching element S4 / S1 with three switching positions G, B, A. The planet carrier PT4 of the fourth planetary gear PG4 is detected on the rotationally fixed component GH. The sun gear SR4 of the fourth planetary gear set PG4 is for switching the Reversing operation with the second transmission input shaft GE2 connectable. The ring gear HR4 of the fourth planetary gear set PG4 is connected to the planet carrier PT2 of the second planetary gear set PG2.

A possible transmission scheme and a circuit diagram of this transmission structure according to FIG. 21 are shown in FIGS. 22 and 23. Accordingly, the fourth planetary gear set PG4 is shorter in length with i_PG = -1.8 compared to the transmission diagram according to FIG. 17 of the transmission structures according to FIG and Fig. 19. It is apparent from the circuit diagram of Fig. 23 that two reverse gears R1, R2 are shiftable, whose gear ratios approximately correspond to the gear ratios of the associated forward gears (first gear "1" and fifth gear "5", respectively).

Fig. 24 shows a gear similar to that of Fig. 4, but with an additional switching element S7 with two switching positions H, I is arranged to alternately in the third planetary gear PG3 whose ring gear HR3 with the non-rotatable member GH or with the planet PT3 connect to. As a result, this third planetary gear set PG3 can optionally be blocked.

In an associated circuit diagram shown in Fig. 25, which largely corresponds to the circuit diagram of FIG. 5, Accordingly, in the lower four forward gears "1" to "4", the effective ratio of the third planetary gear set PG3 is switched by the ring gear HR3 of the third planetary gear set PG3 is detected. In the upper four forward gears "5" to "8" is the third planetary gear PG3 in the block circulation.

Fig. 26 shows an extension of the previously proposed transmission structure to a group transmission. The third planetary gear set PG3 for this purpose a range group GP driving technology and axially downstream. This range group GP has a fourth planetary gear set PG4 designed as a reversing gear set, to which a fifth shift element S5 with a single shift position R for shifting a reverse gear group is assigned, and a fifth planetary gear set PG5 which has a sixth shift element S6 with two shift positions L, H for switching between a slow and a fast forward gear group is assigned. The ring gear HR4 of the fourth planetary gear set PG4 is connected to the sun gear SR5 of the fifth planetary gear set PG5. The planet carrier PT4 of the fourth planetary gear set PG4 is connected to the ring gear HR5 of the fifth planetary gear set PG5 and together with this by the fifth switching element S5 for switching the reverse drive function on the non-rotatable component GH detectable. The sun wheel SR4 of the fourth planetary gear set PG4 is connected to the planet carrier PT3 of the third planetary gear set PG3. Furthermore, the sun gear SR5 of the fifth planetary gearset PG5 connected to the ring gear HR4 of the fourth planetary gear set PG5 is lockable by the sixth shifting element S6 for shifting a lower gear group on the non-rotatable member GH and connectable with the planet carrier PT5 of the fifth planetary gear set PG5 for shifting an upper gear group the fifth planetary gear set PG5 is blocked. A fourth switching element S4 for switching a partial transmission coupling is omitted in this transmission structure. Accordingly, the main transmission of the transmission structure according to FIG. 26 corresponds to that of the 7-speed transmission according to FIG. 7.

A possible translation table with state ratios i_0 of the five planetary gear sets PG1, PG2, PG3, PG4, PG5 is shown in FIG. 27. A possible possible shift scheme is shown in FIG. 28. Accordingly, a doubling of the number of revolutions of the main transmission is achieved, so that a total of fourteen forward gears "1" to "14" and seven reverse gears R1 to R7 are switchable. The largest forward gear "14" is designed as a direct gear.

The switching of the area group GP between the seventh gear "7" and the eighth gear "8" is traction interrupted because in these two courses, the second separating element K2 is closed. The gear jump phi between these two gears "7", "8" is therefore designed to be slightly smaller. All other gear changes are power shiftable. Due to the switched gear ratio of the fifth planetary gear PG5 the seven forward gears "1" to "7" in the lower gear group and the seven reverse gears R1 to R7 are very short translated. The transmission according to FIG. 26 therefore permits very low maneuvering speeds and is therefore particularly suitable for commercial vehicles. Fig. 29 shows a second embodiment of a group transmission. In this transmission, a range group GP includes a fourth planetary gear set PG4 and a sixth shift element S6 with two shift positions L, H for switching between a slow and a fast forward gear group. A fifth planetary gear set PG5 is arranged as a front-end group. In addition, a fifth shift element S5 is arranged with two shift positions V, R for switching between a forward drive operation and a reverse drive operation.

The fifth planetary gear set PG5 and the Vorschaltgruppe PG5 is comparable to the upstream planetary gear set PG4 of the transmission of FIG. The ring gear HR5 of the fifth planetary gearset PG5 is connected to the planet carrier PT1 of the first planetary gearset PG1. The sun gear SR5 of the fifth planetary gear set PG5 is connected to the drive shaft AW. The planet carrier PT5 of the fifth planetary gear set PG5 is mutually connectable in its first switching position V with the drive shaft AW by the fifth switching element S5 or detectable in its second switching position R on the rotationally fixed component GH. By connecting the planet carrier PT5 of the fifth planetary gear set PG5 with the drive shaft AW, this is simultaneously connected to the sun gear SR5 of the fifth planetary gear set PG5, so that in the forward drive the fifth planetary gear set PG5 rotates in the block. By fixing the planetary carrier PT5 of the fifth planetary gear set PG5, the negative stationary gear ratio of the fifth planetary gear set PG5 is effective, so that the direction of rotation between the driving sun gear SR5 and the driven side ring gear HR5 reverses for a reverse drive operation.

In the fourth planetary gear set PG4 of the transmission according to FIG. 29, which acts as a range group GP, the sun gear SR4 of the fourth planetary gear set PG4 is connected to the planet carrier PT3 of the third planetary gear set PG3. The ring gear HR4 of the fourth planetary gearset PG4 can be detected alternately on the non-rotatable component GH by means of the sixth shifting element S6, whereby the gear ratio of the fourth planetary gearset PG4 is activated, or with the Planetenträ- eng PT4 of the fourth planetary gear PG4 connectable, whereby the block circulation of the fourth planetary gear set PG4 is activated.

A possible translation table with state translations i_0 of the five planetary gear sets PG1, PG2, PG3, PG4, PG5 as a numerical example is shown in FIG. 30. A possible circuit diagram resulting therefrom is shown in FIG. 31. Thus, fourteen forward speeds "1" to "14" and seven reverse speeds R1 to R7 are realized. The ratios of the forward gears "1" to "14" are similar to those of the circuit diagram of FIG. 28 of the group transmission of FIG. 26. The reverse gears R1 to R7, however, are again shorter translated. The translations correspond approximately to 1, 8 times the corresponding forward gears. This transmission is therefore particularly suitable for a very sensitive shunting operation.

LIST OF REFERENCES

A, B, C, D, E, F switch positions

G, H, I, L, R, V switch positions

AB output shaft

B1 separating element, brake

AW drive shaft

EM electric machine

EMR rotor of the electric machine EM

GE1 First transmission input shaft

GE2 second transmission input shaft

GH Non-rotating component, housing

GP area group

HR1, HR2, HR3 ring gears

HR4, HR5 ring gears

HW main shaft

K1, K2 separating elements, friction clutches

PG1, PG2, PG3 planetary gear sets

PG4, PG5 planetary gear sets

PR1, PR2, PR3 planetary gears

PR4, PR5 planet gears

PT1, PT2, PT3 planet carrier

PT4, PT5 planet carrier

R1, R2, R3, R4 reverse gears

R5, R6, R7, R8 reverse gears

S1. S2, S3, S4 switching elements

S5, S6, S7 switching elements

SR1, SR2, SR3 sun gears

SR4, SR5 sun wheels

TG1, TG2 partial transmission

X1 separating element, positive coupling i Gear ratio i0 Stator gear ratio translation i_PG Planetary gear ratio

phi gear jump

"1" to "14" forward

Claims

claims

1 . Vehicle transmission having a drive shaft (AW), with a first and second transmission input shaft (GE1, GE2), with at least one separating element (K2, X1), which is the second transmission input shaft (GE2) associated with a main shaft (HW), with a Output shaft (AB) and at least one first, second and third planetary gear set (PG1, PG2, PG3), which as elements at least one ring gear (HR1, HR2, HR3), a sun gear (SR1, SR2, SR3) and a planet carrier ( PT1, PT2, PT3) with planetary gears (PR1, PR2, PR3) and with a plurality of switching elements (S1, S2, S3, S4, S5, S6, S7) for switching gear ratios or drive connections, wherein the two transmission input shafts (GE1, GE2) in each case a partial transmission (TG1, TG2) is assigned, wherein one of the two partial transmissions (TG1, TG2) at least the first planetary gear set (PG1) and the other of the two partial transmissions (TG1, TG2) at least the second planetary gear set (PG2), characterized et,

the first planetary gear set (PG1) is arranged upstream of the two transmission input shafts (GE1, GE2),

a first of the elements (HR1, PT1, SR1) of the first planetary gear set (PG1), which is effective as its drive element, is connected or connectable on the drive side directly or indirectly with the drive shaft (AW) and on the transmission side by means of the at least one separating element (K2, X1) is connectable to the second transmission input shaft (GE2),

a second of the elements (HR1, PT1, SR1) of the first planetary gear set (PG1), which is effective as its output element, is connected or connectable to the first transmission input shaft (GE1) on the transmission side,

in which the first transmission input shaft (GE1) can be connected to the second transmission input shaft (GE2) or at least to the main shaft (HW), wherein the two transmission input shafts (GE1, GE2) are each connected to one or both of the second and third planetary gear sets (PG2, PG3). are operatively connected,

the main shaft (HW) is connected to the output shaft (AB) or at least to one of the elements (HR3, PT3, SR3) of the third planetary gear set (PG3), and that by means of the two partial transmissions (TG1, TG2) at least seven sequentially power-shiftable forward gears ("1", "2", "3", "4", "5", "6", "7") are switchable, wherein one This forward gears is a direct gear or overdrive gear.

2. Vehicle transmission according to claim 1, characterized in that the vehicle transmission is designed as a dual-clutch transmission, with a first separating element (K1) and a second separating element (K2), which are designed as friction clutches (K1, K2) that the first friction clutch ( K1) is connected on the input side to the ring gear (HR1) of the first planetary gearset (PG1), which is effective as its output element, that the first friction clutch (K1) is connected on the output side to the first transmission input shaft (GE1) that the second friction clutch ( K2) on the input side to the planet carrier (PT1) of the first planetary gear set (PG1) is connected, which is effective as the drive element, and that the second friction clutch (K2) on the output side to the second transmission input shaft (GE2) is connected.

3. Vehicle transmission according to claim 1 or 2, characterized in that the three planetary gear sets (PG1, PG2, PG3) by means of a first, a second and a third switching element (S1, S2, S3) with two switching positions (A, B, C , D; E, F) and by means of a fourth switching element (S4) with a switching position (G) are switchable,

in that in the first planetary gear set (PG1) the ring gear (HR1) can be connected to the first transmission input shaft (GE1) by means of the first separating element (K1), the sun gear (SR1) is fixed or lockable on a non-rotatable component (GH), and the planet carrier (PT1) can be connected to the second transmission input shaft (GE2) by means of the second separating element (K2),

in the case of the second planetary gear set (PG2), the ring gear (HR2) can be connected to the first transmission input shaft (GE1) by means of the first shift element (S1) and the second transmission input shaft (GE2) by means of the second shift element (S2), the sun gear (SR2) on a non-rotatable component (GH) is detected or ascertainable, and the planet carrier (PT2) by means of the second switching element (S2) with the second transmission input shaft (GE2) and by means of the third Schaltele- ment (S3) with the planet carrier (PT3) of the third planetary gear set (PG3) is connectable,

in that in the third planetary gear set (PG3) the ring gear (HR3) is fixed or lockable on a non-rotatable component (GH), the sun gear (SR3) can be connected to the planet carrier (PT2) of the second planetary gear set (PG2) by means of the third shifting element (S3) is and by means of the first switching element (S1) with the first transmission input shaft (GE1) is connectable, and the planet carrier (PT3) with the output shaft (AB) is connected,

in that the main shaft (HW) is directly connected to the output shaft (AB), that eight forward gears ("1", "2", "3", "4", "5", "6", "7", "8 ") are switchable by means of the first and second separating element (K1, K2) sequentially, wherein the seventh gear (" 7 ") is switchable as a direct gear, by means of the second separating element (K2), the second switching element (S2 ) and the third switching element (S3) is switchable,

and that the eighth gear ("8") is an overdrive gear, which is switchable by means of the first separating element (K1) and the fourth switching element (S4), wherein by the fourth switching element (S4), a direct connection of the first transmission input shaft (GE1) with the Main shaft (HW) can be produced.

4. Vehicle transmission according to claim 1 or 2, characterized in that the three planetary gear sets (PG1, PG2, PG3) by means of the first, second and third switching element (S1, S2, S3) with two switching positions (A, B, C, D , E, F) and by means of the fourth switching element (S4) with a switching position (G) are switchable,

in the case of the second planetary gear set (PG2), the ring gear (HR2) can be connected to the first transmission input shaft (GE1) by means of the first shift element (S1) and the second transmission input shaft (GE2) by means of the second shift element (S2), the sun gear (SR2) detected on a non-rotatable component (GH) or can be detected, and the planet carrier (PT2) by means of the second switching element (S2) with the second transmission input shaft (GE2) and by means of the third switching element (S3) with the planet carrier (PT3) of the third planetary gear set (PG3) is connectable,

the main shaft (HW) is connected to the sun gear (SR3) of the third planetary gear set (PG3),

in that eight forward gears ("1", "2", "3", "4", "5 ^" , "6", "7", "8") are switchable, which can be shifted by means of the first and second separating element (K1, K2 ) are sequentially load switchable that the seventh gear ("7") is switchable as a direct gear, which is switchable by means of the second separating element (K2) and the second switching element (S2) and the third switching element (S3),

wherein the eighth gear ("8") is an overdrive gear, which is switchable by means of the first separating element (K1) and the second, third and fourth switching element (S2, S3, S4),

wherein by the fourth switching element (S4), a compound of the first transmission input shaft (GE1) with the second transmission input shaft (GE2) can be produced.

5. Vehicle transmission according to claim 1 or 2, characterized in that the three planetary gear sets (PG1, PG2, PG3) by means of the first, second and third switching element (S1, S2, S3) with two switching positions (A, B, C, D ; E, F) are switchable,

in that in the first planetary gear set (PG1) the ring gear (HR1) can be connected to the first transmission input shaft (GE1) by means of the first separating element (K1), the sun gear (SR1) is fixed or lockable on a non-rotatable component (GH), and the planet carrier (PT1) can be connected to the second transmission input shaft (GE2) by means of the second separating element (K2), in the case of the second planetary gear set (PG2), the ring gear (HR2) can be connected to the first transmission input shaft (GE1) by means of the first shift element (S1) and the second transmission input shaft (GE2) by means of the second shift element (S2), the sun gear (SR2) on a non-rotatable component (GH) is detected or detectable, and the Pianetenträger (PT2) by means of the second switching element (S2) with the second transmission input shaft (GE2) and by means of the third switching element (S3) with the planet carrier (PT3) of the third planetary gear set connectable is

in that in the third planetary gear set (PG3) the ring gear (HR3) is fixed or lockable on a non-rotatable component (GH), the sun gear (SR3) can be connected to the planet carrier (PT2) of the second planetary gear set (PG2) by means of the third shifting element (S3) is, and by means of the first switching element (S1) with the first transmission input shaft (GE1) is connectable and the planet carrier (PT3) is connected to the output shaft (AB),

in that seven forward gears ("1", "2", "3", "4", "5", "6", "7") can be shifted, which are sequentially load-switchable by means of the first and second separating elements (K1, K2) wherein the seventh gear ("7") is a direct gear, which is switchable by means of the second separating element (K2) and the second and third switching element (S2, S3).

6. Vehicle transmission according to one of claims 1 to 5, characterized in that a further switching element (S7) is arranged, by means of the third planetary gear set (PG3), the ring gear (HR3) alternately releasably lockable on a non-rotatable component (GH) or with the planet carrier (PT3) is connectable.

7. Vehicle transmission according to one of claims 1 to 6, characterized in that the first switching element (S1) and the fourth switching element (S4) to a single switching element (S1 / S4) summarized with three switch positions (A, B, G) are.

8. Vehicle transmission according to one of claims 1 to 7, characterized in that the second planetary gear set (PG2) is arranged radially above the third planetary gear set (PG3), and that these two planetary gear sets (PG2, PG3) are constructed axially nested.

9. Vehicle transmission according to one of claims 1 to 8, characterized in that a first separating element (B1) and a second separating element (K2) are arranged, wherein the first separating element (B1) is designed as a brake, by means of which the sun gear (SR1 ) of the first planetary gear set (PG1) can be braked or released from a non-rotatable component (GH),

the second separating element (K2) is designed as a friction clutch, which is connected on the input side to the planet carrier (PT1) of the first planetary gear set (PG1), which is effective as its driving element, and is connected on the output side to the second transmission input shaft (GE2) .

and wherein the ring gear (HR1) of the first planetary gear set, which is effective as its output element, is connected to the first transmission input shaft (GE1).

10. Vehicle transmission according to one of claims 1 to 9, characterized in that for the realization of up to eight reverse gears (R1, R2, R3, R4, R5, R6, R7, R8), a fourth planetary gear set (PG4), which as a Wenderadsatz is effective, as well as a fifth switching element (S5), which has two switching positions (R, V), arranged axially in front of the first planetary gear set (PG1) and upstream of this drive technology,

wherein in the fourth planetary gear set (PG4), the ring gear (HR4) is connected to the planet carrier (PT1) of the first planetary gear set (PG1), the sun gear (SR4) is connected to the drive shaft (AW), and the planetary carrier (PT4) by means of fifth switching element (S5) alternately lockable on a non-rotatable component (GH) or with the sun gear (SR4) of the fourth planetary gear set (PG4) is connectable.

1 1. Vehicle transmission according to one of claims 1 to 9, characterized in that for the realization of up to four reverse gears (R1, R2, R3, R4) a fourth planetary gear set (PG4) effective as a turning wheel set; and a fifth switching element (S5) having two shift positions (R, V) disposed axially between the first planetary gear set (PG1) and the second planetary gear set (PG2), and second planetary gear set (PG2) are upstream in terms of drive technology,

wherein in the fourth planetary gear set (PG4), the ring gear (HR4) by means of the second switching element (S2) alternately with the ring gear (HR2) or the planet carrier (PT2) of the second planetary gear set (PG2) is connectable, the sun gear (SR4) of the fourth planetary gear set (PG4) is connected to the second transmission input shaft (GE2), and the planet carrier (PT4) by means of the fifth switching element (S5) alternately lockable on a non-rotatable component (GH) or with the sun gear (SR4) of the fourth planetary gear set (PG4) is connectable ,

12. Vehicle transmission according to one of claims 1 to 9, characterized in that for the realization of one or two reverse gears (R1, R2), a fourth planetary gear set (PG4), which is effective as a Wenderadsatz, and a fifth switching element (S5), which a shift position (R), arranged axially between the second planetary gear set (PG2) and the third planetary gear set (PG3) and the second planetary gear set (PG2) are arranged downstream of the drive, that the fifth switching element (S5) and the second switching element (S2) to a single switching element (S2 / S5) are combined with three switching positions (C, D, R),

that in the fourth planetary gear set (PG4) the ring gear (HR4) is connected to the planet carrier (PT2) of the second planetary gear set (PG2), the sun gear (SR4) is connectable to the second transmission input shaft (GE2) by means of the fifth shift element (S5), and wherein the planet carrier (PT4) is detected or detectable together with the sun gear (SR2) of the second planetary gear set (PG2) on a non-rotatable component (GH).

13. Vehicle transmission according to one of claims 1 to 12, characterized in that the vehicle transmission is designed as a hybrid transmission, in which the second transmission input shaft (GE2) with the rotor (EMR) of an electric machine (EM) is drive-connected, in that a first separating element (K1) and a second separating element (X1) are arranged, wherein the first separating element (K1) is designed as a friction clutch, by means of which the drive shaft (AW) is connected to the planet carrier (PT1) of the first planetary gear set (PG1), which is effective as its drive element, is connectable, and in which the second separating element (X1) is designed as a positive coupling, by means of which the planet carrier (PT1) of the first planetary gear set (PG1) is the transmission side with the second transmission input shaft (GE2) connectable.

14. Vehicle transmission according to one of claims 1 to 12, characterized in that the vehicle transmission is designed as a hybrid transmission, in which the second transmission input shaft (GE2) with the rotor (EMR) an electric machine (EM) is operatively connected, in which a Separating element (X1) is arranged, which is designed as a positive coupling, by means of which the planet carrier (PT1) of the first planetary gear set (PG1) is the transmission side connected to the second transmission input shaft (GE2), and in which the drive shaft (AW) with the planet carrier (PT1) of the first planetary gear set (PG1), which is effective as its driving element, is connected.

1 5. A vehicle transmission according to claim 13 or 14, characterized in that the form of a positive coupling formed separating element (X1) and the fourth switching element (S4) as a single switching element (S4 / X1) with two switching positions (G, X) are summarized, in which alternately the planet carrier (PT1) of the first planetary gear set (PG1) the transmission side with the second transmission input shaft (GE2) is connectable or the first with the second transmission input shaft (GE1, GE2) is connectable.

1 6. The vehicle transmission according to one of claims 1 to 1 5, characterized in that the vehicle transmission is designed as a dual clutch group transmission, in which the first, second and third planetary gear set (PG1, PG2, PG3) at least by means of a first, second and third switching element (S1, S2, S3) with two switching positions (A, B, C, D, E, F) is switchable, in that a range group (GP) is subordinate to the third planetary gear set (PG3),

comprising a fourth planetary gear set (PG4) designed as a turning wheel set, to which a fifth shift element (S5) having a shift position (R) for shifting a reverse gear group is assigned, and a fifth planetary gear set (PG5) having a sixth shift element (S6) with two shift positions ( L, H) is assigned for switching between a slow and a fast forward gear group,

in the fourth planetary gear set (PG4) the ring gear (HR4) is connected to the sun gear (SR5) of the fifth planetary gear set (PG5), the sun gear (SR4) is connected to the planet carrier (PT3) of the third planetary gear set (PG3), and Planet carrier (PT4) with the ring gear (HR5) of the fifth planetary gear set (PG5) is connected and by means of the fifth switching element (S5) on a non-rotatable component (GH) is determined,

in the case of the fifth planetary gear set (PG5), the sun gear (SR5) can be detected alternately on a non-rotatable component (GH) by means of the sixth shifting element (S6) or can be connected to the planet carrier (PT5), and the planet carrier (PT5) is connected to the output shaft (PT5). AB),

that by means of five switching elements (S1, S2, S3, S5, S6) with a total of nine switch positions (A, B, C, D, E, F, H, L, R) at least fourteen forward gears ("1" to "14") and at least seven reverse gears (R1 to R7) are switchable,

at least thirteen are shiftable from the at least fourteen forward gears ("1" to "14") and the fourteenth forward gear ("14") is a direct gear, and wherein the at least seven reverse gears (R1 to R7) are all power shiftable.

17. Vehicle transmission according to one of claims 1 to 1 5, characterized in that the vehicle transmission is designed as a dual clutch group transmission, in which the first, second and third planetary gear set (PG1, PG2, PG3) at least by means of a first, second and third Switching element (S1, S2, S3) with two switch positions (A, B, C, D, E, F) is switchable,

in that the third planetary gear set (PG3) has a fourth planetary gear set (PG4), which acts as a range group (GP), and a sixth switching element (S6), which has two switch positions (L, H) for switching between a slow and a fast forward gear group, are arranged downstream of drive technology,

that in the fourth planetary gear set (PG4) the ring gear (HR4) is alternately lockable on a non-rotatable component (GH) or connectable to the planet carrier (PT4), the sun gear (SR4) is connected to the planet carrier (PT3) of the third planetary gear set (PG3) is connected, and the planet carrier (PT4) is connected to the output shaft (AB),

in that the first planetary gear set (PG1) has a fifth planetary gearset (PG5), which is effective as a turning wheel set, and a fifth switching element (S5), which has two switching positions (R, V), in terms of drive technology,

in the fifth planetary gear set (PG5), the ring gear (HR5) is connected to the planet carrier (PT1) of the first planetary gear set (PG1), the sun gear (SR5) is connected to the drive shaft (AW), and the planet carrier (PT5) is connected by means of fifth switching element (S5) alternately on a non-rotatable component (GH) detectable or with the sun gear (SR5) is connectable,

that by means of five switching elements (S1, S2, S3, S5, S6) with a total of ten switch positions (A, B, C, D, E, F, H, L, R, V) at least fourteen forward gears ("1" to "14 ") and at least seven reverse gears (R1 to R7) are switchable, that of the at least fourteen forward gears (" 1 "to" 14 ") are at least thirteen power shiftable and the fourteenth forward gear (" 14 ") is a direct gear, and that the at least seven reverse gears (R1 to R7) are all power shiftable.