METHOD FOR POWER SHIFTING IN HYBRID AUTOMATIC TRANSMISSIONS BY MEANS OF A DUAL-CLUTCH STRATEGY INVOLVING TRANSFORMATION

Abstract

A method for power shifting in hybrid automatic transmissions which can have any topology and are equipped with any number of additional drive units includes a generic transformation of the effective correlations between real transmission variables into virtual variables relating to a dual-clutch transmission such that a dual-clutch power shifting core having typical basic shifting modes can be used.

Claims

1. A method for output-neutral load switching of hybridized automatic transmissions with an arbitrary number of gears and a number n of clutches and with a first of p drive units and at least one further of p drive units on the basis of a transformation of real transmission variables of the hybridized automatic transmission to virtual variables of a dual-clutch transmission with associated dual-clutch-transmission-specific basic shifting modes comprising the following steps: initiation of a shifting process for a gear-change pair (i, j) from a gear i with an actual gear ratio (y.sub.i) to a gear j with a target gear ratio (y.sub.j) in dependence on a target gear preselection, sensing of actual variables of the hybridized automatic transmission and of the first and/or of the at least one further drive unit, wherein the actual variables comprise at least one of the following variables: a drive shaft rpm (.sub.in) of at least one drive shaft of the hybridized automatic transmission, an output shaft rpm (.sub.out) of an output shaft of the hybridized automatic transmission, a drive torque (T.sub.in) made available by the first and/or by the at least one further drive unit and present at the at least one drive shaft of the hybridized automatic transmission, currently set clutch capacities (T.sub.cap) of the n clutches and/or a minimally and/or maximally available drive torque (T.sub.in,min, T.sub.in,max) of the first and/or of the at least one further drive unit, selection of at least one transformation factor in dependence on at least one actual variable and on the gear-change pair (i, j) from tables of states, calculation of at least one transformation equivalent for the calculation of at least one dual-clutch-transmission-specific actuating quantity by the basic shifting mode of the dual-clutch transmission in dependence on at least one actual variable and/or on the at least one transformation factor, calculation of at least one dual-clutch-transmission-specific actuating quantity by a basic shifting mode in dependence on at least one actual variable and/or on the at least one transformation factor and/or on the at least one transformation equivalent, calculation of at least one automatic-transmission-specific actuating quantity in dependence on at least one actual variable and/or on the at least one transformation factor and/or on the at least one transformation equivalent and/or on the at least one dual-clutch-transmission-specific actuating quantity and implementation of the at least one automatic-transmission-specific actuating quantity by at least one actuator and by the at least one further drive unit.

2. The method for load switching of hybridized automatic transmissions according to claim 1, wherein the selection of the transformation factors comprises at least one of the following steps: selection of coefficients (a.sup.(i,j)) determining the automatic-transmission topology in dependence on the gear-change pair (i, j) from a table of states. selection of effective factors (b.sup.(i,j)) of clutch capacities (T.sub.cap) .sub.r to be set, of the n clutches and of a drive torque (T.sub.EM) delivered by one of the at least one further drive units in dependence on the gear-change pair (i, j) from a table of states, indexing (idx.sup.(i,j)) of the none or at least one on-coming (idx.sub.kom.sup.(i,j)) and of the none or at least one off-going (idx.sub.geh.sup.(i,j)) clutch and of the none or at least one clutch that remains closed (idx.sub.blb.sup.(i,j)) of then clutches and of the status (idx.sub.Em.sup.(i,j)) of the at least one further drive unit in dependence on the gear-change pair (i, j) and/or on the selected mode of operation from a table of states, selection of a gear-change-pair-dependent drive mass moment of inertia (J.sub.in.sup.(i,j)) of the hybridized automatic transmission and of a gear-change-pair-dependent output mass moment of inertia (J.sub.out.sup.(i,j)) of the hybridized automatic transmission in dependence on the gear-change pair (i, j) from a table of states, selection of coefficients (c.sup.i,j)) for determination of cutting torques (T.sub.cut,blb) for the m clutches that remain closed and for determination of a holding torque (T.sub.cut,EM) of the at least one further drive units in dependence on the gear-change pair (i, j) from a table of states and/or selection of maximally transmittable clutch capacities (T.sub.cap,.sub.max) of the n clutches in dependence on at least one actual variable, wherein the calculation of the at least one transformation equivalent comprises at least one of the following steps: calculation of an equivalent drive mass moment of inertia (J.sub.in.sup.(DCT)) in dependence on the gear-change-pair-dependent drive mass moment of inertia (J.sub.in.sup.(i,j)) and on the gear-change-pair-dependent output mass moment of inertia (J.sub.out.sup.(i,j)) and on an rpm ratio (.sub.out/.sub.in) of the output shaft rpm (.sub.out) and on the drive shaft rpm (.sub.in) and on the coefficients (a.sup.i,j)), calculation of dual-clutch-transmission-specific input-shaft rpms (.sub.in.sup.(i)) and (.sub.in.sup.(j)) in dependence on the gear-change pair (i, j) and on the output-shaft rpm (.sub.out) as well as on the actual gear ratio (y.sub.i) and on the target gear ratio (y.sub.j), calculation of effective directions of the cutting torques (T.sub.cut,blb) for the m clutches that remain closed in dependence on the gear-change pair (i, j) and on the clutch rpms (.sub.in.sup.(i,j)) and (.sub.out.sup.(i,j)) of the n clutches, calculation of effective-direction-adapted coefficients ({tilde over (c)}.sup.(i,j)) in dependence on the calculated effective directions of the cutting torques (T.sub.cut,blb) and on the coefficients (c.sup.(i, j)) for determination of the cutting torques (T.sub.cut,blb) for the m clutches that remain closed, calculation of the cutting torques (T.sub.cut,blb) on the m clutches that remain closed and of the holding torque (T.sub.cut,EM) of the at least one further drive unit independence on the effective-direction-adapted coefficients ({tilde over (c)}.sup.(i,j)) and on the drive torque (T.sub.in) and of the first and/or of the at least one further drive unit and on the currently set clutch capacities (T.sub.cap) of the n clutches and on the current output gradient ({dot over ()}.sub.out) and on the drive torque (T.sub.EM) currently made available by the at least one further drive unit and present at an element of the hybridized automatic transmission and/or calculation of a dual-clutch-transmission-specific extra-contact-pressure factor (k.sub.b,scale.sup.(DCT) and/or on a dual-clutch-transmission-specific extra-contact-pressure offset value (k.sub.b,offset.sup.(DCT)) in dependence on the gear-change pair (i, j) and on the effective factors (b.sup.(i,j)) and in dependence on global scaling factors or clutch-individual scaling factors and/or global offset values or clutch-individual offset values of the n clutches.

3. The method for load switching of hybridized automatic transmissions according to claim 1, wherein the at least one dual-clutch-transmission-specific actuating quantity comprises one of the following variables: a relative drive gradient ({dot over ()}.sub.VKM) and/or a relative drive torque (T.sub.VKM) of the first drive unit for rpm transfer, basic-clutch capacities (T.sub.cap,kom.sup.(DCT), (T.sub.cap,geh.sup.(DCT)) for load acceptance during the shifting process for the on-coming and off-going clutch, wherein the basic-clutch capacities (T.sub.cap,kom.sup.(DCT), (T.sub.cap,geh.sup.(DCT)) can be mathematically converted by evaluation with the respective effective direction to basic-clutch torques (T.sub.cl,kom,nom.sup.(DCT), (T.sub.cl,geh,nom.sup.(DCT)) and/or basic extra-contact-pressure clutch capacities (T.sub.b,kom.sup.(DCT), T.sub.b,geh.sup.(DCT)) for extra-contact-pressure control for the on-coming and off-going clutch in dependence on the dual-clutch-transmission-specific extra-contact-pressure factor (k.sub.b,scale.sup.(DCT)) and/or on the dual-clutch-transmission-specific extra-contact-pressure offset value (k.sub.b,offset .sup.(DCT)) and/or a dual-clutch-transmission-specific load-switching torque (T.sub.EM.sup.(DCT)) of at least one further dual-clutch-transmission-equivalent drive unit.

4. The method for load switching of hybridized automatic transmissions according to claim 1, wherein the calculation of the automatic-transmission-specific actuating quantities comprises at least one of the following steps: calculation of load-switching clutch capacities (T.sub.cap,kom.sup.(AT), (T.sub.cap,geh.sup.(AT)) for the on-coming and off-going clutch and calculation of a load-switching torque (T.sub.EM.sup.(AT)) of the at least one further drive unit in dependence on the basic clutch capacities (T.sub.cap,kom.sup.(DCT), T.sub.cap,geh.sup.(DCT)) and on the effective factors (b.sup.i,j)) and on the dual-clutch-transmission-specific load-switching torque (T.sub.EM.sup.(DCT)) of the at least one further dual-clutch-transmission-equivalent drive unit for load acceptance, calculation of an engagement torque (T.sub.in) of the first and/or of the at least one further drive unit and/or at least one engagement torque (T.sub.cl) of the none or at least one on-coming and/or of the none or at least one off-going clutch and/or of the none or at least one clutch that remains closed in dependence on the gear-change pair (i, j) and on the relative drive gradients ({dot over ()}.sub.VKM) and/or on the relative drive torque (T.sub.VKM) of the first drive for rpm transfer, calculation at least of a compensating torque (T.sub.cl,komp) of the none or at least one on-coming and/or of the none or at least one off-going and/or of the none or at least one open clutch and/or on a compensating torque (T.sub.EM,komp) of the at least one further drive unit in dependence on the engagement torque (T.sub.in) of the first and/or of the at least one further drive unit and/or on the at least one engagement torque (T.sub.cl) of the none or at least one on-coming and/or of the none or at least one off-going and/or of the none or at least one closed clutch and/or on the output gradient ({dot over ()}.sub.out ) and/or on the coefficients (a.sub.out.sup.(i,j), a.sub.in.sup.(i,j)) and/or on the gear-change-pair-dependent output mass moment of inertia (J.sub.out.sup.(i,j)), calculation of extra-contact-pressure clutch capacities (T.sub.b,blb) of the none or at least one clutch that remains closed in dependence on the cutting torques (T.sub.cut,blb) for the m clutches that remain closed and on the global scaling factor or clutch-individual scaling factors and/or on the global offset values or clutch-individual offset values of the n clutches, calculation of extra-contact-pressure clutch capacities (T.sub.b,kom.sup.(AT), (T.sub.b,geh.sup.(AT)) of the at least one on-coming and of the at least one off-going clutch in dependence on the basic extra-contact-pressure clutch capacities (T.sub.b,kom.sup.(DCT), (T.sub.b,geh.sup.(DCT)) and on the effective factors (b.sup.(i,j)). calculation of the clutch capacities (T.sub.cap) to be set for the n clutches and of the drive torque (T.sub.EM) to be set for the at least one further drive unit in dependence on the load-switching clutch capacities (T.sub.cap,kom.sup.(AT), (T.sub.cap,geh.sup.(AT)) and/or on the load-switching torque (T.sub.EM.sup.(AT)) of the at least one further drive unit and/or on the extra-contact-pressure clutch capacities (T.sub.b,kom.sup.(AT), (T.sub.b,geh.sup.(AT)) for the none or at least one on-coming and the none or at least one off-going clutch and/or on the extra-contact-pressure clutch capacities (T.sub.b,blb) of the none or at least one clutch that remains closed and/or on the cutting torques (T .sub.cut,blb) of the m clutches that remain closed and/or on the engagement torque (A.sub.in) of the first and/or of the at least one further drive unit and/or on the at least one engagement torque (T,.sub.cl) of the none or at least one on-coming and/or of the none or at least one off-going clutch and/or of the none or at least one clutch that remains closed and/or on the at least one compensating torque (T.sub.cl,komp) of the none or at least one on-coming and/or of the none or at least one off-going and/or of the none or at least one open clutch and/or on the compensation torque (T.sub.EM,komp) of the at least one further drive unit.

5. The method for load switching of hybridized automatic transmissions according to claim 1, wherein the calculation of the transformation equivalents comprises, alternatively or additionally, the calculation of dual-clutch-transmission-specific maximally settable clutch capacities (T.sub.cap,geh,max .sup.(DCT), (T.sub.cap,kom,max.sup.(DCT)) in dependence on the maximally transmittable clutch capacities (T.sub.cap,max) of the n clutches and/or on the minimally and/or maximally available drive torques (T.sub.in,min, T.sub.in,max) of the first and/or of the at least one further drive unit, wherein the basic clutch capacities (T.sub.cap,kom.sup.(DCT), (T.sub.cap,geh.sup.(DCT)) for load acceptance during the shifting process for the on-coming and the off-going clutch are additionally determined in dependence on the dual-clutch-transmission-specific maximally settable clutch capacities (T.sub.cap,geh,max.sup.(DCT), (T.sub.cap,kom,max.sup.(DCT)).

6. The method for load switching of hybridized automatic transmissions according to claim 1, wherein the selection of the transformation factors comprises, alternatively or additionally, the selection of a dual-clutch-transmission-specific minimally and/or maximally realizable drive gradient ({dot over ()}.sub.min.sup.(DCT), {dot over ()}.sub.max.sup.(DCT)) or a dual-clutch-transmission-specific minimally and/or maximally realizable drive-gradient change ({dot over ()}.sub.min.sup.(DCT), {dot over ()}max.sup.(DCT)) in dependence on at least one actual variable and/or on the maximally transmittable clutch capacities (T.sub.cap,max) of the n clutches and/or on the minimally and/or maximally available drive torque (T.sub.in,min, T.sub.in,max) of the first and/or of the at least one further drive unit.

Description

Exemplary Embodiment

[0087] Further features, application possibilities and advantages of the invention will become apparent from the following description of exemplary embodiments of the invention, which are illustrated schematically in the figures. In this connection, all described or illustrated features individually or in arbitrary combination form the subject matter of the invention, regardless of their association in the claims or their cross-referencing as well as regardless of their wording or illustration in the description and in the figures.

[0088] Herein,

[0089] FIGS. 1a-b show the diagram of an automatic transmission (AT) hybridized by means of an E-machine on the basis of a transmission topology and a table of clutch states,

[0090] FIGS. 2a-c show a schematic comparison of clutch capacities of a dual-clutch transmission (DCT), clutch capacities and the drive torque of an automatic transmission hybridized with an E-machine as well as an rpm transfer illustrated in simplified manner by an eCVT gear,

[0091] FIGS. 3a-c show the comparison of a load switching compensated with an E-machine and one not compensated,

[0092] FIG. 4 shows the schematic representation of a dual-clutch-transmission-equivalent view of an arbitrary hybridized automatic transmission with an E-machine,

[0093] FIG. 5 shows the schematic representation of a dual-clutch-transmission-equivalent view of an arbitrary hybridized automatic transmission including a shiftable eCVT transmission, with two E-machines and output-power branching (output split) and

[0094] FIG. 6 shows the schematic representation of a dual-clutch-transmission-equivalent view of an arbitrary hybridized automatic transmission including a shiftable eCVT transmission, with two E-machines and input-power branching (input split) or double power branching (compound split).

[0095] The method for load switching of hybridized automatic transmissions is based on parts of the method, disclosed in the non-prepublished German Patent Application DE 10 2015 120 599.8, for load switching of automatic transmissions, the entire content of which is herewith explicitly incorporated in the present disclosure. In comparison therewith, at least one further drive machine in the form of an E-machine is additionally involved on the gear change, which also comprises the mode-of-operation changeover of the hybridized automatic transmission. From the viewpoint of the transmission control or transmission regulation, the E-machine corresponds to a torque source analogous to a clutch, except only that the effective direction can be bilateral.

[0096] With a hybridized automatic transmission such as illustrated in FIG. 1a, several modes of operation can be represented, e.g. electric driving, internal-combustion-engine driving and hybridized driving. These modes of operation correspond to those of a so-called parallel hybrid vehicle. FIG. 1a shows the diagram of the transmission topology of an automatic transmission according to Lepelletier with a simple planet gearset, a Ravigneaux gearset, three clutches C1, C2 and C3, two brakes C4 and C5, which in the following will likewise be described and referred to as clutches, and an E-machine EM.

[0097] FIG. 1b shows a table of clutch states, which determines, for each fixed (conventional), each hybridized-operated (parallel), each purely electrically actuated (electric) and each continuously variably transmitted gear (eCVT), whether the status (clutch states) of the respective clutch (FR1 to FR5) is to be open or closed (marked by x). During a gear change, an eCVT mode takes place, in which the gear ratio can be adjusted continuously variably between primary drive source, here internal combustion engine, and output by the E-machine and the internal combustion engine is supported via the E-machine. According to FIG. 1b, an eCVT mode can be realized in the present example for all gear changes in which either the first (FR1 ) or the second clutch (FR2) remains closed.

[0098] FIG. 2 shows two gear-change processes, from the 1st gear into the 2nd gear and from the 2nd gear into the 3rd gear. As can be seen in FIG. 2b, during rpm transfer the E-machine (T.sub.EM) takes over the function of the load-bearing clutch in the shifting process of conventional automatic transmissions. Depending on shifting type (pull or push), this corresponds to the on-coming or off-going clutch. During the load acceptance, the off-going clutch, still bearing the load, is relieved by the imposition of a drive torque with the E-machine. As soon as the off-going clutch is load-free, it is opened. The state achieved at this instant corresponds to an eCVT gear. If the gear-change process were ended at this instant, the eCVT1 gear would be engaged, in which the gear ratio of the automatic transmission could be adjusted infinitely variably with change of the applied drive torque of the internal combustion machine VKM and of the supporting drive torque of the E-machine EM. The rpm transfer or adjustment of the rpm is realized by the eCVT gear in dependence on the rpm gradient instructed by the load-switching core. This phase lasts between seconds 2 and 4 (gear-change process from 1st gear into 2nd gear). During this time, as already mentioned, the rpm transfer takes place, which is implemented by a reduction of the drive torque (T.sub.VKM) of the first drive unit, i.e. of the internal combustion engine. At the level of the dual-clutch transmission, this can be seen in FIG. 2a. Finally, FIG. 2c shows the rpms of the off-going gear (.sub.ist), of the on-coming gear (.sub.Ziel), of the internal combustion engine (.sub.VKM) and of the E-machine (.sub.EM).

[0099] The comparison of a compensated and non-compensated gear change can be seen in FIG. 3. The compensation takes place during the rpm transfer. In the process, the rpm of the internal combustion engine (VKM) is transferred in eCVT mode to the rpm of the target gear by a combined VKM-EM engagement, whereby the drive gradient {dot over ()}.sub.out can be smoothly maintained. The compensation of a disturbance torque is illustrated in the non-prepublished German Patent Application DE 10 2015 120 601.3, the entire content of which is herewith explicitly incorporated in the present disclosure, and where the specific determination of the individual drive, engagement and compensation torques is explained by way of examples.

[0100] Specifically, the rpm ratios of the current gear .sub.ist, of the on-coming gear .sub.Ziel, of the internal combustion engine .sub.VKM and of the E-machine .sub.EM can be seen in FIG. 3a. These are identical both during a compensated and during a non-compensated gear-change process.

[0101] The torque ratios, illustrated in FIG. 3b, during a gear-change process from gear 1 to gear 2 differ depending on whether a compensated gear-change process (right) or a non-compensated gear-change process (left) is taking place. At second 2, the load T.sub.KS of the off-going clutch C5 is transmitted to the E-machine T.sub.EM. After the load acceptance by the E-machine has taken place, a reduction of the drive torque T.sub.in of the internal combustion engine begins (approx. second 2.3), in order to transfer the rpm of the internal combustion engine .sub.VKM as seen in FIG. 3a to the new target rpm .sub.Ziel. Because of the coupling between drive and output via the power path, which remains closed, of clutch 1, the torque T.sub.K1 transmitted via this power path to the output (compare with approx. second 2.3 in FIG. 3b left) is reduced, which is perceptible by a gradient collapse {dot over ()}.sub.out at the output. In order to compensate this collapse, a combined engagement takes place, comprising a smaller drive-torque reduction T.sub.in and an increased drive torque T.sub.EM of the E-machine (see FIG. 3b right), respectively in comparison with the non-compensated gear-change process. Thereby the torque transmitted via the power path that remains closed is held almost constant and, after a necessary first reduction has occurred, caused by the changing gear ratio during the load acceptance, the output gradient {dot over ()}.sub.out likewise remains approximately constant (see FIG. 3c right) and a smoother and more comfortable gear-change process takes place without traction-force interruption. The reduction of the output gradient may be prevented by the application of the wheel-torque concept. In the process, the drive torque is correspondingly increased during the load acceptance.

[0102] As soon as the rpm of the internal combustion engine .sub.VKM has adapted to the rpm of the new on-coming gear .sub.Ziel, the load is transmitted from the E-machine to the on-coming clutch C4 (see time interval from second 3.5 to approx. 3.8 in FIG. 3b right).

[0103] What is important for the method according to the invention is the transformation of the measured or present actual variables into dual-clutch-transmission-specific variables, more precisely into the transformation equivalents. From these variables, the DCT load-switching core then determines the DCT-specific actuating variables, which in turn are back-transformed into automatic-transmission-specific actuating variables. In order that this transformation can also be applied for hybridized automatic transmissions, the further drive unit, here the E-machine EM, must be transformed into a dual-clutch-transmission-equivalent view. Such a view is represented as a dual-clutch-transmission diagram in FIG. 4. In this connection, it is apparent that the E-machine EM is assumed to be a further parallel-shifted sub-transmission. The power flow takes place from the internal combustion engine VKM via one of the three sub-transmissions to the output. Via the clutches K1 and K2, the two conventional sub-transmissions connect the internal combustion engine alternatingly with the output. Each of these two sub-transmissions has a fixed transmission ratio i.sub.1 or i.sub.2, which differs depending on engaged gear stage. The transmission ratio is designated as fixed because the clutch in the closed state relays the VKM rpm unchangeably to the transmission element forming the gear ratio, e.g. gearwheels. In contrast to this, the E-machine EM is able in the third sub-transmission to map the function of a clutch, i.e. the approaching and the matching of the VKM rpm to the rpm of the sub-transmission input shaft, and the function of a CVT transmission. In the process, the rpm of the E-machine and thus the relative rpm between VKM output shaft and sub-transmission input shaft are infinitely varied. The application of this simple transmission scheme for automatic transmissions is particularly advantageous because, independently thereof, where the E-machine EM is disposed in the automatic transmission, the application of the transmission scheme is ensured via the transformation factors.

[0104] The calculation of the actuating variables takes place on the basis of a wheel-torque or requested-torque concept, which interprets the drive torque to be transmitted via the drive wheels to the roadway as a requested torque on the basis of the driver's request and/or of a crankshaft-torque concept. In conventional transmissions, the crankshaft-torque concept is adopted for the most part. In this case, a drive torque is predetermined that is applied by the internal combustion engine VKM directly on the crankshaft. In the process, however, the crankshaft torque can be mathematically converted into the wheel torque via the gear ratio, supplied by the transmission control unit, between drive and output. The crankshaft torque can also be mathematically converted into the wheel torque during the change from a fixed gear into an eCVT gear, because the internal combustion engine VKM in the eCVT gear also has a fixed torque-transmission ratio relative to the output. Thus the application of the invention is always possible as a wheel-torque-based control unit.

[0105] Hybridized automatic transmissions may also have two or more E-machines. In FIGS. 5 and 6, variants with respectively two E-machines EM1, EM2 are represented. FIG. 5 shows the dual-clutch-transmission-equivalents, i.e. the transformed view of an output-power-branched automatic transmission, also known as output split. Therein one E-machine EM2 is connected directly with the crankshaft. The effective crankshaft torque, previously the drive torque of the internal combustion machine VKM, is then the sum of the drive torques of the VKM and of the E-machine EM2. The other E-machine EM1 supports this summation torque, whereby a torque is transmitted to the output. The changeover between a fixed gear with constant gear ratio and an eCVT mode with variable gear ratio then takes place as described above, with the difference that the drive-shaft torque is formed from two source torques. By analogy with a vehicle having parallel hybrid drive, a further degree of freedom is obtained in the engine and transmission control unit by the distribution of the summation torque to the internal combustion engine VKM and the E-machine EM2.

[0106] FIG. 6 shows the dual-clutch-transmission-equivalent view of an input-power-branched and a doubly power-branched automatic transmission. In an input-power-branched automatic transmission known in itself, also referred to as input split, the internal combustion engine VKM is supported by an E-machine EM1. This means that the E-machine EM1, for example, engages with and applies a torque on the sun wheel of a planetary transmission, so that the internal combustion engine VKM is able to apply a torque on, for example, the planet carrier, whereby the ring gear is able to deliver a desired and selectively transmitted drive torque to the output shaft. In the conventional automatic transmission without further drive units, a clutch or a brake would support the sun wheel against the housing. In contrast, the example described here has a further degree of freedom, which is characterized by an eCVT mode on the basis of variable rpms of the E-machine EM1. The second E-machine EM2 is coupled directly with the output. Due to the E-machine EM1, the wheel nominal torque can be distributed to the E-machine EM2 on the output and the internal combustion engine VKM. The changeover between a fixed gear with constant gear ratio and an eCVT mode with variable gear ratio then takes place as described above, with the difference that the magnitude of the drive torque of the internal combustion engine VKM can be varied, for constant wheel nominal torque, by application of a drive torque by the E-machine EM2 directly on the output. Thereby load-point displacements of the VKM and thus better efficiencies are possible. In addition, it is possible to use the E-machine EM2 on the output for compensation of disturbance influences in rpm transfers during changeover or gear-change processes.

[0107] A doubly power-branched automatic transmission, also referred to as compound split, has at least two sub-transmissions, preferably two planet-gear transmissions, and in terms of the transmission structure and the transmission control is much more complex than an input-power-branched automatic transmission. Nevertheless, it can be represented in the same dual-clutch-transmission-equivalent view as the latter, even on the basis of the equivalent treatment of clutches, brakes and, depending on the gear-change combination, E-machines, as equivalent to a clutch with additional degrees of freedom in magnitude and effective direction. This further illustrates the great advantageousness of the method according to the invention for output-neutral load switching. In the compound split or doubly branched automatic transmissions, both E-machines EM1 and EM2 may be used to support the internal combustion engine VKM. Thereby the torque applied by the further drive units can be distributed to the two E-machines in dependence on the nominal wheel torque and on the drive torque of the internal combustion engine VKM. In addition, a further degree of freedom is obtained by the fact that the torque for supporting the internal combustion engine VKM can be distributed to both E-machines. These additional degrees of freedom permit both a more efficient constructive design of the entire drive train and a load-point displacement of the individual drive units in the direction of (global) operating optimum. The changeover process between individual fixed gears or between a fixed gear and an eCVT mode takes place as described above, with the difference that both E-machines can be used for transfer of the load. Depending on E-machine used for the changeover to eCVT mode or depending on combination of two E-machines used, a different transmission ratio is obtained between internal combustion engine and output in dependence on support factors of the E-machines.