Powershift transmission for a motor vehicle and method for controlling a powershift transmission

This application claims priority from German patent application serial no. 10 2013 208 201.0 filed May 6, 2013.

The invention concerns a powershift transmission of a motor vehicle, comprising two partial transmissions each with an input shaft, each of which can be brought into driving connection on the input side, via a main separator clutch, with the driveshaft of a drive engine and on the output side, by way of plurality of selectively engaged gear steps that have different gear ratios, with a common output shaft. The invention also concerns a method for controlling shifts in a powershift transmission of that type.

A powershift transmission is understood to be a variable-speed transmission in which a gearshift takes place under load, i.e. with no interruption of the force transfer between the driveshaft of the drive engine and the output shaft of the transmission. During this, the force flow in a gearshift is transmitted by the actuation of frictional shifting elements from a currently load-bearing force transmission branch with a current, actual gear ratio, by operating with slip for a short time, continuously to a parallel force transmission branch with a target gear ratio. Known structures for powershift transmissions are the planetary automatic transmission and the dual clutch transmission.

In a planetary automatic transmission the shift-related force transmission transfer takes place by the actuation of transmission-internal frictional shifting elements such as shifting clutches and shifting brakes, in that with time overlap, at least one friction shifting element that is load-carrying when the loaded gear is engaged, is opened, and at least one friction shifting element that carries the load when the target gear is engaged, is closed.

In a dual clutch transmission the shift-related force transmission transfer takes place by actuating the transmission-external friction clutches arranged between the driveshaft of the drive engine and, respectively, one of the two input shafts, in that first the target gear is engaged in the partial transmission that is not carrying the load and then, with time overlap, the friction clutch associated with the gear under load is opened and the friction clutch associated with the input shaft of the target gear is closed. The typical structure of a dual clutch transmission and a method for controlling shifts in a dual clutch transmission, in which a traction upshift takes place without any electronic action in the engine control system, can be seen in DE 10 2004 005 789 A1.

The disadvantages of shifting processes in the known powershift transmissions are, for one, the need to operate with frictional slip during the phase of force transfer from the force transmitting branch containing the initially loaded gear to the force transmitting branch containing the target gear. The frictional slip in the friction shifting elements or friction clutches dissipates drive energy, i.e. converts it into heat. Likewise, the frictional slip necessarily results in wear of the friction linings of the friction shifting elements and clutches concerned. The dissipation of drive energy increases the fuel consumption of the motor vehicle concerned, and wear on the friction lining limits the useful life of the powershift transmission.

To avoid these disadvantages, in DE 10 2006 036 758 A1 a dual clutch transmission is proposed, in which the input shafts of the two partial transmissions can in each case be connected in a rotationally fixed manner to the driveshaft of the drive engine in each case by means of a main separator clutch in the form of an unsynchronized claw clutch. In that case, as starting and synchronization means two electric machines are provided, each of them in driving connection with one of the two input shafts. Alternatively a single electric machine can also be provided, which can be brought into driving connection with one or other of the two input shafts in alternation by means of a shifting mechanism. It is true that this design of a dual clutch transmission does avoid slipping operation of the otherwise common friction clutches. However, the disadvantage of a dual clutch transmission of such a design is the high outfitting cost of the two electric machines or of the one electric machine together with the shifting mechanism, as well as at least one electrical energy accumulator and the associated control technology.

Accordingly, the purpose of the present invention is to propose a powershift transmission of the structural type mentioned earlier, in which during a gearshift between two gears of the two partial transmissions, slipping operation of the otherwise common friction clutches can be avoided with relatively little outfitting cost. In addition, a method for controlling shifts in such a powershift transmission is indicated.

Accordingly, the invention concerns a powershift transmission for a motor vehicle which comprises two partial transmissions, each of which can be brought into driving connection on the input side, via a main separator clutch, with the driveshaft of a drive engine and on the output side, by way of a plurality of selectively engaged gear steps with different gear ratios, with a common drive output shaft. In this powershift transmission it is provided that the two input shafts on the input side can be connected by the respective main separator clutch to an associated intermediate shaft and can be disconnected therefrom, and that the first intermediate shaft is connected directly in a rotationally fixed manner to the driveshaft of the engine, while in contrast the second intermediate shaft is in driving connection with the first intermediate shaft by way of a variator V with a continuously adjustable transmission ratio i V and a ratio at rest that has the value one (i V — 0 =1).

Thus, the invention starts from a powershift transmission comprising two partial transmissions TG 1 , TG 2 each with a respective input shaft GE 1 , GE 2 , which can in each case be brought into driving connection on the input side, via a main separator clutch CL 1 , CL 2 , with the driveshaft TW of a drive engine VM and on the output side, by way of a plurality of selectively engaged gear steps (G 1 to G 7 , R) that have different gear ratios, with a common output shaft GA.

Accordingly, the structure of this powershift transmission corresponds in large measure to that of a dual clutch transmission known per se. However, the difference from a dual clutch transmission is that instead of the therein customary friction clutches arranged between the driveshaft TW of the drive engine VM and the input shafts GE 1 , GE 2 , it is provided that on the input side the two input shafts GE 1 , GE 2 can be connected by their respective main separator clutches CL 1 , CL 2 to an associated intermediate shaft ZW 1 , ZW 2 and disconnected therefrom. The first intermediate shaft ZW 1 is connected directly and rotationally fixed to the driveshaft TW of the drive engine VM, whereas in contrast the second intermediate shaft ZW 2 is in driving connection with the first intermediate shaft ZW 1 , via a variator V, with a continuously adjustable transmission ratio i V and a ratio at rest with the value one (i V — 0 =1). A starting clutch K is preferably arranged on the driveshaft TW of the engine VM. In a known way, the variator V can for example be in the form of a conical wrap-round transmission or a toroidal transmission.

Thus, by means of the variator the rotational speeds n ZW1 , n ZW2 of the two intermediate shafts ZW 1 , ZW 2 can be changed relative to one another. This makes it possible, during a shifting process carried out as a powershift from an initially loaded gear (G 2 , G 4 , G 6 ) of the first partial transmission TG 1 to a target gear (G 1 , G 3 , G 5 , G 7 ) of the second partial transmission TG 2 , by appropriate adjustment of the variator V, to synchronize both the main separator clutch CL 2 arranged between the second intermediate shaft ZW 2 and the second input shaft GE 2 , and also the gear clutch (E, F, G, H) of the target gear. Likewise this makes it possible, during a shifting process carried out as a powershift from an initially loaded gear (G 1 , G 3 , G 5 , G 7 ) of the second partial transmission TG 2 to a target gear (G 2 , G 4 , G 6 ) of the first partial transmission TG 1 , by appropriate adjustment of the variator V as well as by simultaneous rotational speed adaptation of the drive engine VM, to synchronize both the main separator clutch CL 1 arranged between the first intermediate shaft ZW 1 and the first input shaft GE 1 and also the gear clutch (A, B, C) of the target gear. It is therefore possible for both of the main separator clutches CL 1 , CL 2 and also the gear clutches (A to H) to be designed inexpensively as unsynchronized claw clutches.

With the powershift transmission according to the invention, compared with a dual clutch transmission the shift-related friction slip operation of the friction clutches on the input side, and compared with a powershift transmission known from DE 10 2006 036 758 A1 the expenditure for additional electric machines and their supply and control systems, are avoided.

The two main separator clutches CL 1 , CL 2 , the two intermediate shafts ZW 1 , ZW 2 and the variator V are preferably combined in a preassembled powershift module LSM. On the one hand this simplifies the assembly of the powershift transmission. On the other hand it provides the possibility of using the same basic transmission in combination with the powershift module LSM, as a powershift transmission according to the invention, or in combination with two friction clutches on the input side, which can be combined in a dual clutch module, as a dual clutch transmission.

A second embodiment of the powershift transmission according to the invention is characterized in that on the input side the two input shafts GE 1 , GE 2 can be connected and disconnected by means of their respective main separator clutches CL 1 , CL 2 , to and from an associated intermediate shaft ZW 1 , ZW 2 , and that the first intermediate shaft ZW 1 is connected directly and in a rotationally fixed manner to the driveshaft TW of the engine VM, whereas the second intermediate shaft ZW 2 is in driving connection with the first intermediate shaft ZW 1 by way of a variator V having a continuously adjustable transmission ratio i V and a gear ratio at rest having the value one (i V — 0 =1), and by way of a superposition gearset SG with two input elements E 1 , E 2 and an output element A 0 , wherein the first input element E 1 is connected directly and rotationally fixed to the first intermediate shaft ZW 1 , the second input element E 2 is in driving connection via the variator V with the first intermediate shaft ZW 1 , and the output element A 0 is connected in a rotationally fixed manner to the second intermediate shaft ZW 2 .

The structure of this second embodiment of the powershift transmission differs from the first embodiment in that the two intermediate shafts ZW 1 , ZW 2 are not in driving connection with one another directly via the variator V. Rather, in this case the driving connection of the two intermediate shafts ZW 1 , ZW 2 has two parallel force transmission branches connected to the first intermediate shaft ZW 1 , one of which contains the variator V, and which are merged in the superposition gearset SG whose output element A 0 is connected to the second intermediate shaft ZW 2 in a rotationally fixed manner. This configuration has the advantage that force transmission when the second partial transmission TG 2 is carrying the load (loaded gear G 1 , G 3 , G 5 or G 7 ) takes place only partially by way of the variator V, which results in higher efficiency. However, a disadvantage of this is that in this case its transmission ratio i V must have a larger adjustment range in order to be able to produce a given rotational speed difference between the two intermediate shafts ZW 1 , ZW 2 . However, the principle of the mode of operation of this second embodiment of the powershift transmission according to the invention corresponds to that of the first embodiment.

To achieve a compact and at the same time robust structure, the superposition gearset SG is preferably made in the form of a simple planetary gearset PG with a sun gear S, a planetary carrier T carrying a plurality of planetary gearwheels and a ring gear U, in which the sun gear S is the first input element E 1 , the ring gear U is the second input element E 2 and the planetary carrier T is the output element A 0 .

Alternatively, the superposition gearset SG can also be in the form of a simple planetary gearset PG with a sun gear S, a planetary carrier T carrying a plurality of planetary gearwheels and a ring gear U, in which the ring gear U is the first input element E 1 , the sun gear S is the second input element E 2 and the planetary carrier T is the output element A 0 .

For the reasons mentioned earlier, with this second embodiment of the powershift transmission according to the invention it is also advantageous for the two main separator clutches CL 1 , CL 2 , the two intermediate shafts ZW 1 , ZW 2 , the variator V and the superposition gearset SG to be combined in a pre-assembled powershift module LSM′.

A third embodiment of the powershift transmission according to the invention is characterized in that on the input side, the two input shafts GE 1 , GE 2 can be connected to or disconnected from an associated intermediate shaft ZW 1 , ZW 2 by means of their respective main separator clutch CL 1 , CL 2 , and the two intermediate shafts ZW 1 , ZW 2 are directly in driving connection with one another via a variator V with a continuously adjustable transmission ratio i V and a ratio at rest which has the value one (i V — 0 =1), and in driving connection by way of a transfer box VG having an input element E 0 and two output elements A 1 , A 2 , with the driveshaft TW of the drive engine VM, wherein the input element E 0 is connected in a rotationally fixed manner to the driveshaft TW of the engine VM, the first output element A 1 is connected in a rotationally fixed manner to the first intermediate shaft ZW 1 and the second output element A 2 is connected in a rotationally fixed manner to the second intermediate shaft ZW 2 .

The structure of this third embodiment of the powershift transmission according to the invention differs from the first and second embodiments in that although the two intermediate shafts ZW 1 , ZW 2 are directly in driving connection with one another via the variator V as in the first embodiment, neither of the two intermediate shafts ZW 1 , ZW 2 is directly connected to the driveshaft TW of the engine VM. Rather, in this case the two intermediate shafts ZW 1 , ZW 2 are in driving connection with one another and with the driveshaft TW of the engine VM by way of the transfer box VG. As in the second embodiment, in relation to the second partial transmission TG 2 this has the advantage that when the first partial transmission TG 1 is carrying the load (loaded gear G 2 , G 4 , G 6 or R) and when the second partial transmission TG 2 is carrying the load (loaded gear G 1 , G 3 , G 5 or G 7 ), in each case the force transmission takes place only partially by way of the variator V, which results in higher efficiency. On the other hand, to set a given rotational speed difference between the two intermediate shafts ZW 1 , ZW 2 the variator V does not need a larger adjustment range, since it is arranged immediately between the two intermediate shafts ZW 1 , ZW 2 .

A disadvantage of this third embodiment can be considered to be that due to the coupling of the intermediate shafts ZW 1 , ZW 2 to the driveshaft TW of the drive engine VM by way of the transfer box VG, any transmission ratio adjustment of the variator V is necessarily accompanied by a change of the engine rotational speed n VM , i.e. it necessitates an adaptation of the rotational speed of the drive engine VM. In other respects the mode of operation of this third embodiment of the powershift transmission according to the invention corresponds in large measure to those of the first and second embodiments.

To achieve a compact and at the same time robust structure, the transfer box VG is preferably in the form of a simple planetary gearset PG with a sun gear S, a planetary carrier T that carries a plurality of planetary gearwheels and a ring gear U, in which the planetary carrier T is the input element E 0 , the sun gear S is the first output element A 1 and the ring gear U is the second output element A 2 .

Alternatively, the transfer box VG can also be made as a simple planetary gearset PG with a sun gear S, a planetary carrier T carrying a plurality of planetary gearwheels and a ring gear U, in which the planetary carrier T is the input element E 0 , the ring gear U is the first output element A 1 and the sun gear S is the second output element A 2 .

For the reasons mentioned earlier, it is also advantageous in this third embodiment of the powershift transmission according to the invention for the two main separator clutches CL 1 , CL 2 , the two intermediate shafts ZW 1 , ZW 2 , the variator V and the transfer box to be structurally combined in a pre-assembled powershift module LSM*.

Since in the first and third embodiments of the powershift transmission according to the invention the two intermediate shafts ZW 1 , ZW 2 are directly in driving connection with one another by way of the variator V, in this case the variator V should have an adjustment range whose highest ratio i V — max corresponds at least to the maximum gear interval φ max =(i G — n /i G — n+1 ) max of the gears (G 1 to G 7 ), and whose lowest ratio i V — min corresponds at least to the reciprocal of the maximum gear interval φ max of the gears (G 1 to G 7 ) (i V — max ≧φ max , i V — min ≦1/φ max ).

Since in the second embodiment of the powershift transmission according to the invention the two intermediate shafts ZW 1 , ZW 2 are only indirectly in driving connection with one another, namely via the superposition gearset and the variator, then assuming a symmetrical structure of the superposition gearset SG the variator V should have an adjustment range whose highest transmission ratio i V — max corresponds to at least twice the maximum gear interval 2*φ max of the gears (G 1 to G 7 ) and whose lowest gear interval corresponds to at least the reciprocal of twice the maximum gear interval 2*φ max of the gears (i V — max ≧2*φ max , i V — min ≦½*φ max ). If an adjustment range as large as that cannot be realized with a single variator V, then to do this a serial arrangement of more than one variator can be provided.

In order, when no shifting processes are taking place, to avoid a permanent force flow via the variator V and the associated loss of efficiency, it can advantageously be provided that the variator V can be connected to and disconnected from the first intermediate shaft ZW 1 by means of a separator clutch and that the two intermediate shafts ZW 1 , ZW 2 can be directly or indirectly connected to and disconnected from one another by means of a bridging clutch J.

Since the two clutches z, J are engaged and disengaged in alternation, the separator clutch I and the bridging clutch J can be combined in a dual shifting element so that compared with a separate arrangement of the clutches I, J one clutch control element can be saved.

The main separator clutches CL 1 , CL 2 are preferably in the form of unsynchronized claw clutches, since they can be synchronized by adjusting the variator V and this type of structure is particularly inexpensive, space-saving and robust.

It is also possible, however, for the main separator clutches to be frictionally synchronized synchronous clutches CL 1 ′, CL 2 ′ or disk clutches CL 1 *, CL 2 *. In that case the main separator clutches (CL 1 ′, CL 2 ′; CL 1 *, CL 2 *) can also be synchronized as necessary by compressing their frictional elements.

The gear clutches (A to H) for engaging the gears (G 1 to G 7 , R) are also preferably unsynchronized claw clutches, since these too can be synchronized by adjusting the variator V.

Alternatively however, the gear clutches (A′ to H′) for engaging the gears (G 1 to G 7 , R) can also be made as frictionally synchronized synchronous clutches, so that these too can be synchronized as necessary by compressing their friction elements.

Furthermore, it can be provided that one of the two main separator clutches XΛ 1 ′, CL 2 ′; CL 1 *, CL 2 * is in the form of a friction clutch and the other main separator clutch is a claw clutch. This design can exclude any sticking during a variator adjustment for load takeover if the main separator clutch made as a friction clutch is operated with slight slipping.

Finally, as already indicated earlier, in all the aforesaid embodiments it is preferably provided that the driveshaft TW of the drive engine VM is connected to the first intermediate shaft ZW 1 by way of a starting clutch K.

A method according to the invention for controlling shifts in a powershift transmission that comprises two partial transmissions TG 1 , TG 2 , each with a respective input shaft GE 1 , GE 2 , which can be brought into driving connection on the input side with the driveshaft TW of a drive engine VM by means of a main separator clutch (CL 1 , CL 2 ; CL 1 ′, CL 2 ′; CL 1 *, CL 2 *) and on the output side, by way of a plurality of selectively engaged gear steps (G 1 to G 7 , R) having different gear ratios, with a common output shaft GA, wherein the two input shafts GE 1 , GE 2 can be connected on the input side by way of the respective main separator clutch (CL 1 , CL 2 ; CL 2 ′; CL 1 *, CL 2 *) to a respective associated intermediate shaft ZW 1 , ZW 2 or disconnected therefrom, wherein the first intermediate shaft ZW 1 is connected directly and rotationally fixed to the driveshaft TW of the drive engine VM and the second intermediate shaft ZW 2 (in the first embodiment of the powershift transmission) is in driving connection with the first intermediate shaft ZW 1 directly, or (in the second embodiment of the powershift transmission) indirectly via a variator V with a continuously adjustable transmission ratio i V , provides that a gearshift carried out as a powershift from an initially loaded gear (for example the fourth gear G 4 ) of the first partial transmission TG 1 to a target gear (for example the fifth gear G 5 ) of the second partial transmission TG 2 , takes place with the following steps:

a) synchronization of the second main separator clutch (CL 2 , CL 2 ′, CL 2 *) by adjusting the transmission ratio of the variator V,

b) engagement of the second main separator clutch (CL 2 , CL 2 ′, CL 2 *),

c) synchronization of the gear clutch (F, F′) of the target gear (G 5 ) by adjusting the transmission ratio of the variator V,

d) engagement of the gear clutch (F, F′) of the target gear (G 5 ),

e) load uptake by the second partial transmission TG 2 by a slight transmission ratio adjustment of the variator V,

f) disengagement of the first main separator clutch (CL 1 , CL 1 ′, CL 1 *),

g) disengagement of the gear clutch (B, B′) of the loaded gear (G 4 ),

h) rotational speed adaptation of the drive engine VM with simultaneous readjustment of the variator V to its rest transmission ratio (i V — 0 =1).

In this method, regardless of the structure of the main separator clutches (CL 1 , CL 2 ; XΛ 1 ′, CL 2 ′; CL 1 *, CL 2 *) and of the gear clutches (A to H; A′ to H′) it is provided that during a shift from an initially loaded gear of the first partial transmission TG 1 to a target gear of the second partial transmission TG 2 , here for example an upshift from the fourth gear G 4 to the fifth gear G 5 , both the second main separator clutch (CL 2 , CL 2 ′, CL 2 *) and the gear clutch (F, F′) of the target gear (G 5 ) are in each case synchronized by appropriate adjustment of the variator V (process steps a and c). For the load uptake by the second partial transmission TG 2 the two intermediate shafts ZW 1 , ZW 2 are then braced relative to one another by a slight ratio adjustment of the variator V in the direction toward a higher transmission ratio i V (process step e), whereby the first main separator clutch (CL 1 , CL 1 ′, CL 1 *) and the first input shaft GE 1 are relieved of the load from the transmitted engine torque M VM of the drive engine VM and the second main separator clutch (CL 2 , CL 2 ′, CL 2 *) and the second input shaft GE 2 take up the bad of the transmitted engine torque M VM of the engine VM.

When at least the second main separator clutch (CL 2 ′, CL 2 *) is a frictionally synchronized synchronous clutch or a disk clutch, the process step a) can be replaced by the following step:

a*) synchronization of the second main separator clutch (CL 2 ′, CL 2 *) by means of the friction elements of the second main separator clutch (CL 2 ′, CL 2 *).

In that case the second input shaft GE 2 , which before the shifting process is not mechanically connected to the second intermediate shaft ZW 2 and the output shaft GA and which rotates relatively slowly under the action of the active drag torques, is accelerated by the compression of the friction elements of the second main separator clutch (CL 2 ′, CL 2 *) until it reaches the synchronous rotational speed. For this, no ratio adjustment of the variator V is needed. Thereafter, the second main separator clutch (CL 2 ′, CL 2 *) is engaged if in the form of a synchronous clutch (CL 2 ′) or closed by compression if in the form of a disk clutch (CL 2 *) (process step b).

If the gear clutches (A′ to H′) for engaging the gear steps (G 1 to G 7 , R) are made as frictionally synchronized synchronous clutches, the process steps a) to d) can also be replaced by the following steps:

a′) synchronization of the gear clutch (F′) of the target gear (G 5 ) by the friction elements of the gear clutch (F′),

b′) engagement of the gear clutch (F′) of the target gear (G 5 ),

c′) synchronization of the second main separator clutch (CL 2 , CL 2 ′, CL 2 *) by adjusting the transmission ratio of the variator V,

d′) engagement of the second main separator clutch (CL 2 , CL 2 ′, CL 2 *).

In that case the second input shaft GE 2 is first accelerated due to the pressure of the friction elements of the gear clutch (F′) of the target gear (G 5 ) until it reaches the synchronous rotational speed and the gear clutch (F′) is then engaged (process steps a′ and b′). For this, no ratio adjustment of the variator V is needed. Thereafter, the second main separator clutch (CL 2 , CL 2 ′, CL 2 *) is synchronized by an appropriate transmission ratio adjustment of the variator V and then engaged (process steps c′ and d′). Compared with the first two method variants, the sequence of synchronizing the second main separator clutch (CL 2 , CL 2 ′, CL 2 *) and the gear clutch (F′) of the target gear (G 5 ) is therefore reversed.

Since in the first and second embodiments of the powershift transmission according to the invention the first intermediate shaft ZW 1 is in each case indirectly connected in a rotationally fixed manner to the driveshaft TW of the drive engine VIM, the gearshift from an initially loaded gear of the second partial transmission TG 2 to a target gear of the first partial transmission TG 1 cannot take place as described previously. Consequently, it is provided for this that a gearshift carried out as a powershift, from an initially loaded gear (for example gear G 5 ) of the first partial transmission TG 1 to a target gear (for example G 6 ) of the second partial transmission TG 2 takes place by the following steps:

i) synchronization of the first main separator clutch (CL 1 , CL 1 ′, CL 1 *) by a ratio adjustment of the variator V at the same time as a rotational speed adaptation of the drive engine VM,

j) engagement of the first main separator clutch (CL 1 , CL 1 ′, CL 1 *),

k) synchronization of the gear clutch (A, A′) of the target gear (G 6 ) by adjusting the transmission ratio of the variator V and at the same time adapting the rotational speed of the drive engine VM,

l) engagement of the gear clutch (A, A′) of the target gear (G 6 ),

m) load uptake by the first partial transmission TG 1 by a slight ratio adjustment of the variator V,

n) disengagement of the second main separator clutch (CL 2 , CL 2 ′, CL 2 *),

o) disengagement of the gear clutch (F, F′) of the loaded gear (G 5 ),

p) resetting of the variator V to its rest transmission ratio (i V — 0 =1).

In this method, regardless of the structure of the main separator clutches (CL 1 , CL 2 ; CL 1 ′, CL 2 ′; CL 1 *, CL 2 *) and of the gear clutches (A to H; A′ to H′) it is provided that during a shift from an initially loaded gear of the second partial transmission TG 2 to a target gear of the first partial transmission TG 1 , here for example an upshift from the fifth gear G 5 to the sixth gear G 6 , both the first main separator clutch (CL 1 , CL 1 ′, CL 1 *) and the gear clutch (A, A′) of the target gear (G 6 ) are each synchronized by an appropriate transmission ratio adjustment of the variator V, and at the same time the rotational speed of the drive engine VM is adapted (process steps i and k). In the context of the invention it goes without saying that the rotational speed adaptation of the drive engine VM is carried out as much as possible so as to avoid a collapse of the torque at the output shaft GA, i.e. the engine power, or product of the engine torque M VM and engine speed n VM , is kept constant (M VM *n VM =const.) during the adaptation.

For the first partial transmission TG 1 to take up the load, the two intermediate shafts ZW 1 , ZW 2 are then braced relative to one another by a slight ratio adjustment of the variator V in the direction toward a lower transmission ratio i V (process step m), whereby the second main separator clutch (CL 2 , CL 2 ′, CL 2 *) and the second input shaft GE 2 are relieved of the transmitted engine torque M MV of the drive engine VM, whereas the first main separator clutch (CL 1 , CL 1 ′, CL 1 *) and the first input shaft GE 1 are loaded by uptake of the transmitted engine torque M VM of the drive engine VM.

When at least the first main separator clutch (CL 1 ′, CL 1 *) is in the form of a frictionally synchronized synchronous clutch or disk clutch, the process step i) can also be replaced by the following step:

i*) synchronization of the first main separator clutch (CL 1 ′, CL 1 *) by means of the friction elements of the first main separator clutch (CL 1 ′, CL 1 *).

In that case the first input shaft GE 1 , which before the shifting operation is not mechanically connected to the first intermediate shaft ZW 1 and to the output shaft GA and which rotates relatively slowly under the action of active drag torques, is accelerated by the compression of the friction elements of the first separator clutch (CL′, CL 1 *) until it reaches the synchronous rotational speed. For this, no ratio adjustment of the variator V is needed. Then the first main separator clutch (CL 1 ′, CL 1 *), if it is in the form of a synchronous clutch (CL′), is engaged, and if it is a disk clutch (CL 1 *) it is closed by compression (process step j).

If the gear clutches (A′ to H′) for engaging the gear steps (G 1 to G 7 , R) are frictionally synchronized synchronization clutches, then process steps i) to l) can also be replaced by the following steps:

i′) synchronization of the gear clutch (A′) of the target gear (G 6 ) by means of the friction elements of the gear clutch (A′) of the target gear (G 6 ),

j′) engagement of the gear clutch (A′) of the target gear (G 6 ),

k′) synchronization of the first main separator clutch (CL 1 , CL 1 ′, CL 1 *) by adjusting the transmission ratio of the variator V and at the same time adapting the rotational speed of the drive engine VM,

l′) engagement of the first main separator clutch (CL 1 , CL 1 ′, CL 1 *).

In that case the first input shaft GE 1 is first accelerated by compressing the friction elements of the gear clutch (A′) of the target gear (G 6 ) until it reaches the synchronous speed and the gear clutch (A′) is then engaged (process steps i′) and j′)). For this, no ratio adjustment of the variator V is needed. Thereafter the first main separator clutch (CL 1 , CL 1 ′, CL 1 *) is synchronized by an appropriate transmission ratio adjustment of the variator V in combination with a corresponding rotational speed adaptation of the drive engine VM, and is then engaged (process steps k′ and l′). Thus, compared with the two previously described method variants the synchronization sequence of the first main separator clutch (CL 1 , CL 1 ′, CL 1 *) and the gear clutch (A′) of the target gear (G 6 ) is reversed.

Since in the third embodiment of the powershift transmission according to the invention the two intermediate shafts ZW 1 , ZW 2 are respectively in driving connection with the driveshaft TW of the drive engine VM by way of the transfer box VG and with one another by way of the variator V in a symmetrical arrangement, a gearshift from an initially loaded gear of one partial transmission (TG 1 or TG 2 ) to a target gear of the other partial transmission (TG 2 or TG 1 ) can in each case take place in the same way. Consequently a method according to the invention is provided for controlling shifts in a powershift transmission which comprises two partial transmissions TG 1 , TG 2 each with a respective input shaft GE 1 , GE 2 , which on the input side can be brought into driving connection, by means of a main separator clutch (CL 1 , CL 2 ; CL 1 ′, CL 2 ′; CL 1 *, CL 2 *), with the driveshaft TW of a drive engine VM and on the output side, by way of a plurality of selectively engaged gear steps (G 1 to G 7 , R) having different gear ratios, with a common output shaft GA, wherein on the input side the two input shafts GE 1 , GE 2 can be connected to and disconnected from an associated intermediate shaft ZW 1 , ZW 2 by means of the respective main separator clutch (CL 1 , CL 2 ; CL 1 ′, C: 2 ′; CL 1 *, CL 2 *), and wherein the two intermediate shafts ZW 1 , ZW 2 are directly in driving connection with one another via a variator V having a continuously adjustable transmission ratio i V and, via a transfer box GV having one input element E 0 and two output elements A 1 , A 2 , with the driveshaft TW of the drive engine VM, in which method a gearshift carried out as a powershift from an initially loaded gear (for example the third gear G 3 ) of one partial transmission (for example TG 2 ) to a target gear (for example the second gear G 2 ) of the other partial transmission (for example TG 1 ) takes place by the following steps:

q) synchronization of the main separator clutch (CL 1 , CL 1 ′, CL 1 *) of the other partial transmission (TG 1 ) by adjusting the transmission ratio of the variator V and at the same time adapting the rotational speed of the drive engine VM,

r) engaging the main separator clutch (CL 1 , CL 1 ′, CL 1 *) of the other partial transmission (TG 1 ),

s) synchronizing the gear clutch (C, C′) of the target gear (G 2 ) by adjusting the transmission ratio of the variator V and at the same time adapting the rotational speed of the engine (VM),

t) engaging the gear clutch (C, C′) of the target gear (G 2 ),

u) load uptake by the other partial transmission (TG 1 ) by a slight ratio adjustment of the variator V,

v) disengaging the main separator clutch (CL 2 , CL 2 ′, CL 2 *) of the one partial transmission (TG 2 ),

w) disengaging the gear clutch (G, G′) of the loaded gear (G 3 ),

x) adapting the rotational speed of the drive engine VM and at the same time resetting the variator V to its rest transmission ratio (i V — 0 =1).

In this method, regardless of the structure of the main separator clutches (CL 1 , CL 2 ; CL 1 ′, CL 2 ′; CL 1 *, CL 2 *) and of the gear clutches (A to H; A′ to H′), it is provided that during a shift from an initially loaded gear of one partial transmission (TG 1 or TG 2 ) to a target gear of the other partial transmission (TG 2 or TG 1 ), here for example during a downshift from the third gear G 3 to the second gear G 2 , both the first main separator clutch (CL 1 , CL 1 ′, CL 1 *) and also the gear clutch (C, C′) of the target gear (G 2 ) are each synchronized by an appropriate adjustment of the transmission ratio of the variator V. Since the intermediate shaft (ZW 1 ) associated with the target gear (G 2 ) is in driving connection via the transfer box VG with the driveshaft TW of the drive engine VM, in each case a rotational speed adaptation of the drive engine VM is carried out, so avoiding an abrupt rotational speed change at the output shaft GA (process steps q and s). In the context of the invention it goes without saying that to avoid a torque collapse at the output shaft GA, the rotational speed adaptation of the drive engine VM is carried out as much as possible in such manner that the engine power, i.e. the product of the engine torque M VM and the engine rotational speed n VM , is kept constant during this (M VM *n VM const.).

For the load uptake by the other partial transmission (TG 1 ) the two intermediate shafts ZW 1 , ZW 2 are then braced relative to one another by a slight transmission ratio adjustment of the variator V in the direction toward a higher ratio i V (process step u), whereby the second main separator clutch (CL 2 , CL 2 ′, CL 2 *) and the input shaft (GE 2 ) of the one partial transmission (TG 2 ) are relieved from the transmitted engine torque M VM of the drive engine VM whereas the first main separator clutch (CL 1 , CL 1 ′, CL 1 *) and the input shaft (GE 1 ) of the other partial transmission (TG 1 ) are loaded by the transmitted engine torque M VM of the drive engine VM.

If at least the main separator clutch (for example CL′, CL*) of the other partial transmission (for example TG 1 ) is in the form of a frictionally synchronized synchronous clutch or a disk clutch, then process step q) can also be replaced by the following step:

q*) synchronization of the main separator clutch (CL 1 ′, CL 1 *) of the other partial transmission (TG 1 ) by means of the friction elements of the main separator clutch (CL 1 ′, CL 1 *) of the other partial transmission (TG 1 ).

In that case, the input shaft (GE 1 ) of the other partial transmission (TG 1 ), which before the shifting process has no mechanical connection to the associated intermediate shaft (ZW 1 ) and to the output shaft GA and which is rotating relatively slowly under the action of active drag torques, is accelerated due to compression of the friction elements of the associated main separator clutch (CL′, CL 1 *) until it reaches the synchronous rotational speed. For this, no ratio adjustment of the variator V is needed. Thereafter, the main separator clutch (CL 1 ′, CL 1 *) concerned is engaged if in the form of a synchronous clutch (CL 1 ′) or closed by compression if in the form of a disk clutch (CL 1 *) (process step r).

If the gear clutches (A′ to H′) for engaging the gear steps (G 1 to G 7 , R) are made as frictionally synchronized synchronous clutches, the process steps q) to t) can also be replaced by the following steps:

q′) synchronization of the gear clutch (C′) of the target gear (G 2 ) by means of its friction elements,

r′) engagement of the gear clutch (C′) of the target gear (G 2 ),

s′) synchronization of the main separator clutch (CL 1 , CL 1 ′, CL 1 *) of the other partial transmission (TG 1 ) by adjusting the transmission ratio of the variator V and at the same time adapting the rotational speed of the drive engine VM,

t′) engagement of the main separator clutch (CL 1 , CL 1 ′, CL 1 *) of the other partial transmission (TG 1 ).

In that case the input shaft (GE 1 ) of the other partial transmission (TG 1 ) is first accelerated by compressing the friction elements of the gear clutch (C′) of the target gear (G 2 ) until it reaches the synchronous speed and the gear clutch (C′) is engaged (process steps q′ and r′). For this, no ratio adjustment of the variator V is needed. Thereafter, the associated main separator clutch (CL 1 , CL 1 ′, CL 1 *) is synchronized by an appropriate adjustment of the transmission ratio of the variator V in combination with a corresponding rotational speed adaptation of the engine VM, and is then engaged (process steps s′ and t′). Compared with the two previously described method variants, therefore, the sequence in which the main separator clutch (CL 1 , CL 1 ′, CL 1 *) of the other partial transmission (TG 1 ) and the gear clutch (C′) of the target gear (G 2 ) are synchronized, is reversed.

When no shifting processes are taking place, in order to avoid a permanent force flow by way of the variator V and the associated loss of efficiency, in all the method variants described above it is preferably provided that in the absence of shifting processes the variator V is decoupled by disengaging a separator clutch arranged between the variator V and the first intermediate shaft ZW 1 , and bridged by engaging a bridging clutch J arranged directly or indirectly between the two intermediate shafts ZW 1 , ZW 2 .

To clarify the invention further, the description of a drawing with a number of example embodiments is given below.

The drawing shows:

FIG. 1 a : Schematic view of a first embodiment of the powershift transmission according to the invention,

FIG. 1 b : Schematic view of a first further development of the powershift transmission shown in FIG. 1 a,

FIG. 1 c : Schematic view of a second further development of the powershift transmission shown in FIG. 1 a,

FIG. 1 d : Schematic view of a further development of the powershift transmission shown in FIG. 1 c,

FIG. 2 a : Schematic view of a second embodiment of the powershift transmission according to the invention,

FIG. 2 b : Schematic view of a first further development of the powershift transmission shown in FIG. 2 a,

FIG. 2 c : Schematic view of a second further development of the powershift transmission shown in FIG. 2 a,

FIG. 2 d : Schematic view of a third further development of the powershift transmission shown in FIG. 2 a,

FIG. 2 e : Schematic view of a further development of the powershift transmission shown in FIG. 2 d,

FIG. 3 a : Schematic view of a third embodiment of the powershift transmission according to the invention,

FIG. 3 b : Schematic view of a first further development of the powershift transmission shown in FIG. 3 a,

FIG. 3 c : Schematic view of a second further development of the powershift transmission shown in FIG. 3 a,

FIG. 3 d : Schematic view of a third further development of the powershift transmission shown in FIG. 3 a,

FIG. 3 e : Schematic view of a further development of the powershift transmission shown in FIG. 3 d,

FIGS. 4 a - 4 e : A shifting process of the powershift transmission shown in FIGS. 1 a and 2 a , controlled in accordance with the invention, represented in a time diagram,

FIGS. 5 a - 5 e : A shifting process of the powershift transmission shown in FIGS. 1 d and 2 e , controlled in accordance with the invention, represented in a time diagram,

FIGS. 6 a - 6 e : A controlled shifting process of the powershift transmission shown in FIGS. 1 b and 2 b , represented in a time diagram,

FIGS. 7 a - 7 e : A further controlled shifting process of the powershift transmission shown in FIGS. 1 a and 2 a , represented in a time diagram,

FIGS. 8 a - 8 e : A further controlled shifting process of the powershift transmission shown in FIGS. 1 d and 2 e , represented in a time diagram,

FIGS. 9 a - 9 e : A further controlled shifting process of the powershift transmission shown in FIGS. 1 b and 2 b , represented in a time diagram,

FIG. 10 a - 10 e : A controlled shifting process of the powershift transmission shown in FIG. 3 a , represented in a time diagram,

FIG. 11 a - 11 e : A controlled shifting process of the powershift transmission shown in FIG. 3 e , represented in a time diagram,

FIG. 12 a - 12 e : A controlled shifting process of the powershift transmission shown in FIG. 3 b , represented in a time diagram.

Thus, FIG. 1 a shows a schematic view of a first embodiment variant of a powershift transmission designed according to the invention. The powershift transmission consists of a basic transmission comprising two partial transmissions TG 1 , TG 2 each with a respective input shaft GE 1 , GE 2 , each of which can be brought into driving connection with a common output shaft GA by way of a plurality of selectively engaged gear steps G 1 to G 7 , R having different gear ratios, as well as components arranged between the two input shafts GE 1 , GE 2 and the driveshaft TW of a drive engine VM. In this case the drive engine VM is an internal combustion engine and the driveshaft TW has a starting clutch K. According to the definition used here, the driveshaft TW thus also includes the output side of the starting clutch K.

The basic transmission is derived from a dual clutch transmission and in this case, for example, is designed such that the two input shafts GE 1 , GE 2 are arranged coaxially one over another and the output shaft GA is arranged coaxially with and axially close to the input shafts GE 1 , GE 2 .

The first input shaft GE 1 is arranged centrally and is in driving connection with a first countershaft VG 1 by way of a first input constant EK 1 . The first countershaft VG 1 is in the form of a hollow shaft arranged coaxially over a second countershaft VG 2 . Between the first countershaft VG 1 and the output shaft GA are arranged three spur gear stages Z 2 , Z 4 , ZR with different gear ratios. The two spur gear stages Z 2 and Z 4 , by way of which the force flow takes place in the second gear G 2 and in the fourth gear G 4 , each consist of a fixed wheel connected in a rotationally fixed manner to the first, hollow countershaft VG 1 and a loose wheel mounted to rotate on the output shaft GA but that can be connected in a rotationally fixed manner thereto by a respective, associated gear clutch B, C. The spur gear stage ZR is designed as a reversing spur gear stage by way of which the force flow takes place in the reversing gear R. This reversing spur gear stage ZR comprises a fixed wheel connected in a rotationally fixed manner to the first countershaft VG 1 , a loose wheel mounted to rotate on the output shaft GA but that can be connected to it in a rotationally fixed manner by means of an associated gear clutch D, and arranged between these two gearwheels, an intermediate wheel indicated by a dotted line.

By means of a gear clutch A the first input shaft GE 1 can be connected directly and in a rotationally fixed manner to the output shaft GA, whereby a sixth gear G 6 acting as a direct-drive gear can be engaged. The first input shaft GE 1 , the first countershaft VG 1 and a first section of the output shaft GA and the spur gear stages EK 1 Z 2 , Z 4 , ZR and gear clutches A, B, C, D arranged between those transmission shafts, form the first partial transmission TG 1 , with which therefore the even-numbered forward gears G 2 , G 4 , G 6 and the reversing gear R are associated.

The second input shaft GE 2 is in the form of a hollow shaft arranged over the first input shaft GE 1 . The second input shaft GE 2 is in driving connection with the second countershaft VG 2 by way of a second input constant EK 2 . Between the second countershaft VG 2 and the output shaft GA are arranged four spur gear stages Z 1 , Z 3 , Z 5 , Z 7 with different gear ratios. These spur gear stages Z 1 , Z 3 , Z 5 , Z 7 , by way of which the force flow takes place in the first G 1 , the third G 3 , the fifth G 5 and the seventh G 7 gears respectively, in each case consist of a fixed wheel connected in a rotationally fixed manner to the second countershaft VG 2 and a loose wheel mounted to rotate on the output shaft GA but which can be connected in a rotationally fixed manner thereto by means of a respectively associated gear clutch E, F, G, H. The second input shaft GE 2 , the second countershaft VG 2 and a second section of the output shaft GA, as well as the spur gear stages EK 2 , Z 1 , Z 3 , Z 5 , Z 7 and gear clutches E, F, G, H arranged between these transmission shafts, form the second partial transmission TG 2 , with which therefore the odd-numbered forward gears G 1 , G 3 , G 5 , G 7 are associated.

In the present case the gear clutches A, B, C, D, E, F, G, H are for example in the form of unsynchronized claw clutches and are combined in pairs in double shifting elements S 1 , S 2 , S 3 , S 4 .

On the input side the two input shafts GE 1 , GE 2 can in each case be connected to and disconnected from two associated intermediate shafts ZW 1 , ZW 2 by means of a respective main separator clutch CL 1 , CL 2 . In this case the main separator clutches CL 1 , CL 2 are for example in the form of unsynchronized claws clutches. The first intermediate shaft ZW 1 is connected directly to the driveshaft TW of the drive engine VM in a rotationally fixed manner by means of the starting clutch K. The second intermediate shaft ZW 2 is in driving connection with the first intermediate shaft ZW 1 by way of a variator V with continuously adjustable transmission ratio i V and a rest transmission ratio equal to one i V — 0 =1). The two main separator clutches CL 1 , CL 2 , the two intermediate shafts ZW 1 , ZW 2 and the variator V are combined in a pre-assembled powershift module LSM.

Thus, in a load-carrying gear (G 2 , G 4 , G 6 or R) of the first partial transmission TG 1 the force flow takes place from the driveshaft TW of the drive engine VM, via the first intermediate shaft ZW 1 and the first input shaft GE 1 (the main separator clutch CL 1 being engaged), directly or via the first countershaft VG 1 and one of the associated spur gear stages (Z 2 , Z 4 or ZR), to the output shaft GA (gear clutch A, B, C or D engaged). In a load-carrying gear (G 1 , G 3 , G 5 or G 7 ) of the second partial transmission TG 2 the force flow, in contrast, takes place from the driveshaft TW of the drive engine VM, via the first intermediate shaft ZW 1 , the variator V, the second intermediate shaft ZW 2 , the second input shaft GE 2 (main separator clutch CL 2 engaged), the second countershaft VG 2 and one of the associated spur gear stages (Z 1 , Z 3 , Z 5 or Z 7 ), to the output shaft GA (gear clutch E, F, G or H engaged), this path being characterized by a lower transmission efficiency due to the force transmission by way of the variator V.

However, by a transmission ratio adjustment of the variator V the rotational speeds n ZW1 , n ZW2 of the two intermediate shafts ZW 1 , ZW 2 can advantageously be changed relative to one another, which in a gearshift between an initially loaded gear and a target gear in different partial transmissions TG 1 , TG 2 , can be used for synchronizing the main separator clutch (CL 1 or CL 2 ) of the partial transmission (TG 1 or TG 2 ) that comprises the target gear and the gear clutch (A, B, C or D, or else E, F, G or H) of the target gear (G 2 , G 4 , G 6 or R, or else G 1 , G 3 , G 5 , or G 7 ). Thus, the powershift module LSM can replace the friction clutches usually present in a dual clutch transmission, whereby during powershifts the friction slipping that usually takes place therein is eliminated so that a higher efficiency is achieved. Such process sequences for the control of gearshifts carried out as powershifts will be described in detail later.

FIG. 1 b shows schematically a first further development of the first embodiment variant of the powershift transmission shown in FIG. 1 a , which consists in that the gear clutches A′ to H′ of the two partial transmissions TG 1 , TG 2 are now in the form of frictionally synchronized, synchronous clutches combined in pairs in dual shifting elements S 1 ′, S 2 ′, S 3 ′, S 4 ′. This makes it possible, during a gearshift between a loaded gear and a target gear in different partial transmissions TG 1 , TG 2 , to reverse the sequence in which the main separator clutch (CL 1 or CL 2 ) of the partial transmission (TG 1 or TG 2 ) that comprises the target gear and the gear clutch (A′, B′, C′ or D′, or else E′, F′, G′ or H′) of the target gear (G 2 , G 4 , G 5 or R, or else G 1 , G 3 , G 5 , G 7 ) are synchronized, i.e. first by pressing together the relevant friction elements to synchronize the gear clutch (A′, B′, C′ or D′, or else E′, F′, G′ or H′) of the target gear (G 2 , G 4 , G 6 or R, or else G 1 , G 3 , G 5 , or G 7 ), and thereafter, by adjusting the transmission ratio of the variator V, to synchronize the main separator clutch (CL 1 or CL 2 ) of the partial transmission (TG 1 or TG 2 ) that comprises the target gear. Such process sequences for the control of gearshifts carried out as powershifts will be described in detail later.

FIG. 1 c shows schematically a second further development of the first embodiment of the powershift transmission shown in FIG. 1 a , which consists in that the two main separator clutches CL 1 ′, CL 2 ′ are now in the form of frictionally synchronized synchronous clutches. This makes it possible, during a gearshift between an initially loaded gear and a target gear in different partial transmissions TG 1 , TG 2 , to carry out the synchronization of the main separator clutch (CL 1 ′ or CL 2 ′) of the partial transmission (TG 1 or TG 2 ) comprising the target gear, which is done at the beginning, by compressing the friction elements of the main separator clutch (CL 1 ′ or CL 2 ′) concerned instead of adjusting the transmission ratio of the variator V.

A further deviation compared with the first embodiment of the powershift transmission shown in FIG. 1 a consists in that the variator V can now be connected to and disconnected from the first intermediate shaft ZW 1 by means of a separator clutch I, and the two intermediate shafts ZW 1 , ZW 2 can be connected to by a bridging clutch J, directly and in a rotationally fixed manner, and disconnected from one another thereby. By disengaging the separator clutch I and engaging the bridging clutch J, in a load-carrying gear (G 1 , G 3 , G 5 or G 7 ) of the second partial transmission TG 2 , outside shifting processes a permanent force flow via the variator and the associated loss of efficiency can be avoided.

FIG. 1 d shows schematically a further development of the embodiment of the powershift transmission shown in FIG. 1 c , which consists in that the two main separator clutches CL 1 *, CL 2 * are now in the form of disk clutches. As in the previously described variant this makes it possible, during a gearshift between an initially loaded gear and a target gear in different partial transmissions TG 1 , TG 2 , to carry out the synchronization of the main separator clutch (CL 1 *, CL 2 *) of the partial transmission (TG 1 or TG 2 ) that comprises the target gear, which is done at the beginning, by pressing together the friction elements of the main separator clutch (CL 1 *, CL 2 *) concerned instead of adjusting the transmission ratio of the variator V. Such process sequences for the control of gearshifts carried out as powershifts will be described in detail later on.

A further change compared with the previously described variant of the powershift transmission shown in FIG. 1 c , consists in that the separator clutch I and the bridging clutch J are now combined in a dual shifting element S 5 , whereby compared with when the clutches I, J are arranged separately, one clutch control element can be saved.

FIG. 2 a shows schematically a second embodiment of the powershift transmission. This powershift transmission comprises the same basic transmission as the first embodiment in FIG. 1 a . Likewise, on the input side the two input shafts GE 1 , GE 2 can be connected to and disconnected from an associated intermediate shaft ZW 1 , ZW 2 by means of the respective main separator clutch CL 1 , CL 2 . In addition, the first intermediate shaft ZW 1 is again connected directly and in a rotationally fixed manner to the driveshaft TW of the drive engine VM. By contrast with the first embodiment in FIG. 1 a , however, this time the second intermediate shaft ZW 2 is in driving connection with the first intermediate shaft DTI via a variator V with a continuously adjustable transmission ratio i V and a rest ratio equal to one (i V — 0 =1) and also via a superposition gearset SG with two input elements E 1 , E 2 and one output element A 0 . The first input element E 1 is connected directly and rotationally fixed to the first intermediate shaft ZW 1 , the second input element E 2 is in driving connection with the first intermediate shaft ZW 1 via a variator V, and the output element A 0 is connected to the second intermediate shaft ZW 2 in a rotationally fixed manner.

In this case the superposition gearset SG is in the form of a simple planetary gearset PG with a sun gear S, a planetary carrier T carrying a plurality of planetary gearwheels, and a ring gear U, in which the sun gear S is the first input element E 1 , the ring gear U is the second input element E 2 and the planetary carrier T is the output element A 0 .

In a load-carrying gear (G 2 , G 4 , G 6 or R) of the first partial transmission TG 1 , the force flow again takes place from the driveshaft TW of the drive engine VM, via the first intermediate shaft ZW 1 and the first input shaft GE 1 (with the main separator clutch CL 1 engaged), directly or via the first countershaft VG 1 and one of the associated spur gear stages (Z 2 , Z 4 or ZR), to the output shaft GA (gear clutch A, B, C or D engaged). In a load-carrying gear (G 1 , G 3 , G 5 or G 7 ) of the second partial transmission TG 2 , however, the force flow now takes place from the driveshaft TW of the engine VM, via the first intermediate shaft ZW 1 and then via two parallel transmission branches, namely on the one hand directly to the first input element E 1 (the sun gear S) of the superposition gearset SG (planetary transmission PG), and on the other hand via the variator V to the second input element (ring gear U) of the superposition gearset SG (planetary gearset PG), where the two transmission branches merge and the corresponding rotational speeds and torques are superposed. The force flow then continues from the output element A 0 (planetary carrier T) of the superposition gearset SG (planetary gearset PG), via the second intermediate shaft ZW 2 , the second input shaft GE 2 (with main separator clutch CL 2 engaged), the second countershaft VG 2 and one of the associated spur gear stages (Z 1 , Z 3 , Z 5 or Z 7 ), to the output shaft GA (with E, F, G or H engaged).

By virtue of the drive-technical connection of the variator V, therefore, in this case only part of the engine torque M VM is transmitted via the variator V, so that the transmission efficiency is less impaired than in the first embodiment of the powershift transmission shown in FIG. 1 a . However it can be a disadvantage of this arrangement that to produce a given rotational speed difference (Δn ZW =n ZW2 −n ZW1 ) between the two intermediate shafts ZW 1 , ZW 2 , a larger transmission ratio adjustment of the variator V is needed, since this is reduced by the superposition gearset SG and indeed halved if the configuration of the superposition gearset SG is symmetrical, as preferred. In addition to the two main separator clutches CL 1 , CL 2 , the two intermediate shafts ZW 1 , ZW 2 and the variator V, the pre-assembled powershift module LSM′ now also contains the superposition gearset SG.

In principle, however, as with the first embodiment of the powershift transmission shown in FIG. 1 a , in this case too ( FIG. 2 a ) the rotational speeds n ZW1 , n ZW2 of the two intermediate shafts ZW 1 , ZW 2 can be changed relative to one another by adjusting the transmission ratio of the variator V, which can be using during a gearshift between an initially loaded gear and a target gear in different partial transmissions TG 1 , TG 2 for synchronizing the main separator clutch (CL 1 or CL 2 ) of the partial transmission (TG 1 or TG 2 ) comprising the target gear and the gear clutch (A, B, C or D, or else E, F, G or H) of the target gear (G 2 , G 4 , G 6 or R, or else G 1 , G 3 , G 5 or G 7 ). Such process sequences for the control of gearshifts carried out as powershifts will be described in detail farther on.

FIG. 2 b shows schematically a first further development of the second embodiment of the powershift transmission in FIG. 2 a . This further development consists in that the gear clutches A′ to H′ of both partial transmissions TG 1 , TG 2 are this time in the form of frictionally synchronized synchronous clutches combined in pairs in dual shifting elements S 1 ′, S 2 ′, S 3 ′, S 4 ′. This makes it possible, during a gearshift between an initially loaded gear and a target gear in different partial transmissions TG 1 , TG 2 , to reverse the synchronization sequence of the main separator clutch (CL 1 or CL 2 ) of the partial transmission (TG 1 or TG 2 ) that comprises the target gear and the gear clutch (A′, B′, C′ or D′, or else E′, F′, G′ or H′) of the target gear (G 2 , G 4 , G 6 or R, or else G 1 , G 3 , G 5 or G 7 ), i.e. first, by pressing together the relevant friction elements of the gear clutch (A′, B′, C′ or D′, or else E′, F′, G′ or H′) of the target gear (G 2 , G 4 , G 6 or R, or else G 1 , G 3 , G 5 or G 7 ), to synchronize that clutch, and thereafter, by adjusting the transmission ratio of the variator V, to synchronize the main separator clutch (CL 1 or CL 2 ) of the partial transmission (TG 1 or TG 2 ) that contains the target gear. Such process sequences for controlling gearshifts carried out as powershifts will be described in detail later.

FIG. 2 c shows schematically a second further development of the second embodiment of the powershift transmission in FIG. 2 a , which consists in a modified drive-technical connection mode of superposition gearset SG. In this case too the superposition gearset SG is in the form of a simple planetary gearset PG with a sun gear S, a planetary carrier T carrying a plurality of planetary gearwheels, and a ring gear U. However, now the ring gear U forms the first input element E 1 of the superposition gearset SG, which is connected directly and rotationally fixed to the first intermediate shaft ZW 1 , and the sun gear S forms the second input element E 2 of the superposition gearset SG, which is in driving connection with the first intermediate shaft ZW 1 via the variator V. On the other hand, as before the planetary carrier T forms the output element A 0 of the superposition gearset SG, which is connected directly and in a rotationally fixed manner to the second intermediate shaft ZW 2 . The operational mode of this variant, however, corresponds exactly to that of the second embodiment of the powershift transmission shown in FIG. 2 a.

FIG. 2 d shows schematically a third further development of the second embodiment of the powershift transmission in FIG. 2 a , which consists in that the two main separator clutches CL 1 ′, CL 2 ′ are now both in the form of frictionally synchronized synchronous clutches. This makes it possible, during a gearshift between an initially loaded gear and a target gear in different partial transmissions TG 1 , TG 2 , to carry out the synchronization of the main separator clutch (CL 1 ′ or CL 2 ′) of the partial transmission (TG 1 or TG 2 ) that comprises the target gear, which synchronization is done at the beginning, by pressing together the friction elements of the main separator clutch (CL 1 ′ or CL 2 ′) concerned instead of by adjusting the transmission ratio of the variator V.

A further change compared with the second embodiment of the powershift transmission in FIG. 2 a consists in that the variator V can now be connected to and disconnected from the first intermediate shaft ZW 1 by means of a separator clutch I, and the two intermediate shafts ZW 1 , ZW 2 can indirectly be connected to and disconnected from one another in a rotationally fixed manner by means of a bridging clutch J, i.e. by a rotationally fixed connection of the first intermediate shaft ZW 1 to the output element A 0 (planetary carrier T) of the superposition gearset SG (planetary gearset PG). By disengaging the separator clutch I and engaging the bridging clutch J, in a load-carrying gear (G 1 , G 3 , G 5 or G 7 ) of the second partial transmission TG 2 , outside of shifting processes a permanent force flow via the variator V and the associated loss of efficiency can be avoided.

FIG. 2 e shows schematically a further development of the embodiment variant of the powershift transmission in FIG. 2 d , which consists in that the two main separator clutches CL 1 *, CL 2 * are now in the form of disk clutches. As with the variant described previously, this makes it possible during a gearshift between an initially loaded gear and a target gear in different partial transmissions TG 1 , TG 2 , to carry out the synchronization of the main separator clutch (CL 1 *, CL 2 *) of the partial transmission (TG 1 or TG 2 ) containing the target gear, which is done at the beginning, by compressing the friction elements of the main separator clutch (CL 1 *, CL 2 *) concerned, instead of by adjusting the transmission ratio of the variator. Such process sequences for the control of gearshifts carried out as powershifts will be described in detail later.

A further change compared with the previously described variant of the powershift transmission shown in FIG. 2 d consists in that the separator clutch I and the bridging clutch J are now combined in a dual shifting element S 5 , whereby compared with the separate arrangement of the two clutches I, J, one clutch control element is saved.

FIG. 3 a shows schematically a third embodiment of the powershift transmission according to the invention. This powershift transmission comprises the same basic transmission as the first embodiment in FIG. 1 a . Likewise, on the input side the two input shafts GE 1 , GE 2 can be connected to and disconnected from a respective associated intermediate shaft ZW 1 , ZW 2 by means of a respective main separator clutch CL 1 , CL 2 . As in the powershift transmission according to FIG. 1 a , the two intermediate shafts ZW 1 , ZW 2 are directly in driving connection with one another via a variator V with continuously adjustable transmission ratio i V and a rest transmission ratio of one (i V — 0 =1). In contrast to the embodiment of the powershift transmission shown in FIG. 1 a , however, the two intermediate shafts ZW 1 , ZW 2 are in this case in driving connection with the driveshaft TW of the drive engine VM, via a transfer box VG with one input element E 0 and two output elements A 1 , A 2 , wherein the input element E 0 is connected in a rotationally fixed manner to the driveshaft TW of the engine VM, the first output element A 1 is connected in a rotationally fixed manner to the first intermediate shaft ZW 1 , and the second output element A 2 is connected in a rotationally fixed manner to the second intermediate shaft ZW 2 . In FIG. 3 a the transfer box VG is in the form of a simple planetary gearset PG with a sun gear S, a planetary carrier T a plurality of planetary gearwheels, and a ring gear U, in which the planetary carrier T is the input element E 0 , the sun gear S is the first output element A 1 , and the ring gear U is the second output element A 2 .

In a load-carrying gear (G 2 , G 4 , G 6 or R) of the first partial transmission TG 1 , the force flow now takes place starting from the driveshaft TW of the drive engine VM, into the input element E 0 (planetary carrier T) of the transfer box VG (planetary gearset PG), where the engine torque M VM and the engine rotational speed n VM are distributed to the two output elements A 1 , A 2 (sun gear S and ring gear U of the transfer box VG (planetary gearset PG). The force flow then continues along two parallel transmission branches, namely on the one hand from the first output element A 1 (sun gear S) of the transfer box VG (planetary gearset PG) directly to the first intermediate shaft ZW 1 and on the other hand (shortly afterward) from the second output element A 2 (ring gear U) of the transfer box VG (planetary gearset PG), via the second intermediate shaft ZW 2 and the variator V, also to the first intermediate shaft ZW 1 . From the first intermediate shaft ZW 1 , in which the two transmission branches merge, the force flow then continues by way of the first input shaft GE 1 (main separator clutch CU engaged) directly or via the first countershaft VG 1 and one of the associated spur gears stages (Z 2 , Z 4 or R) to the output shaft GA (gear clutch A, B, C or D engaged).

In a load-carrying gear (G 1 , G 3 , G 5 or G 7 ) of the second partial transmission TG 2 , the force flow takes place from the driveshaft TW of the engine VM to the output shaft GA (main separator clutch CL 2 and gear clutch E, F, G or H engaged), with a correspondingly interchanged allocation of the output elements A 1 , A 2 (sun gear S, ring gear U) of the transfer box VG (planetary gearset PG) and the intermediate shafts ZW 1 , ZW 2 , by analogy identically by way of the second input shaft GE 2 (main separator clutch CL 2 engaged), the second countershaft VG 2 and one of the spur gear stages (Z 1 , Z 3 , Z 5 or Z 7 ) of the second partial transmission (gear clutch E, F, G or H engaged).

By virtue of the drive-technical connection of the variator V, in this case only part of the engine torque M VM is transmitted via the variator V, so that the transmission efficiency is less impaired than in the first embodiment of the powershift transmission according to FIG. 1 a . Compared with the second embodiment of the powershift transmission shown in FIG. 2 a , the direct driving connection of the two intermediate shafts ZW 1 , ZW 2 via the variator V is advantageous, since to produce a given rotational speed difference (Δn ZW =n ZW2 −n ZW1 ) between the intermediate shafts ZW 1 , ZW 2 there is now (as in the first embodiment of the powershift transmission according to FIG. 1 a ) no need for a larger transmission ratio adjustment of the variator V. However, it can be a disadvantage of this arrangement that every change of the transmission ratio i V of the variator V requires a change of the engine rotational speed n VM , i.e. a rotational speed adaptation of the engine VM is necessary. In addition to the two main separator clutches CL 1 , CL 2 , the two intermediate shafts ZW 1 , ZW 2 and the variator V, the pre-assembled powershift module LSM* also comprises the transfer box VG.

In principle, however, as with the first and second embodiments of the powershift transmission shown respectively in FIGS. 1 a and 2 a , here too by means of a transmission ratio adjustment of the variator V the rotational speeds n ZW1 , n ZW2 of the two intermediate shafts ZW 1 , ZW 2 can be changed relative to one another, and during a gearshift between a loaded gear and a target gear in different partial transmissions TG 1 , TG 2 this can be used to synchronize the main separator clutch (CL 1 or CL 2 ) of the partial transmission that comprises the target gear (TG 1 or TG 2 ) and the gear clutch (A, B, C or D, or else E, F, G or H) of the target gear (G 2 , G 4 , G 6 or R, or else G 1 , G 3 , G 5 or G 7 ). Such a process sequence for the control of a gearshift carried out as a powershift will be described in detail later.

FIG. 3 b shows schematically a first further development of the third embodiment of the powershift transmission shown in FIG. 3 a , which consists in that the gear clutches A′ to H′ in both partial transmissions TG 1 , TG 2 are now in the form of frictionally synchronized synchronous clutches, which are combined in pairs in dual shifting elements S 1 ′, S 2 ′, S 3 ′, S 4 ′. This makes it possible, during a gearshift between an initially loaded gear and a target gear in different partial transmissions TG 1 , TG 2 , to reverse the synchronization sequence of the main separator clutch (CL 1 or CL 2 ) of the partial transmission (TG 1 or TG 2 ) comprising the target gear and of the gear clutch (A′AP, B′, C′ or D′, or else E′, F′, G′ or H′) of the target gear (G 2 , G 4 , G 6 or R, or else G 1 , G 3 , G 5 or G 7 ), i.e. to synchronize first the gear clutch (A′, B′, C′ or D′, or else E′, F′, G′ or H′) of the target gear (G 2 , G 4 , G 6 or R, or else G 1 , G 3 , G 5 or G 7 ) by pressing together the relevant friction elements, and thereafter to synchronize the main separator clutch (CL 1 or CL 2 ) of the partial transmission (TG 1 or TG 2 ) that comprises the target gear, by adjusting the transmission ratio of the variator V. Such a process sequence for the control of a gearshift carried out as a powershift will be described in detail later.

FIG. 3 c shows schematically a second further development of the third embodiment of the powershift transmission according to FIG. 3 a , which consists in a modified drive-technical connection mode of the transfer box VG. In this case too the transfer box VG is a simple planetary gearset PG with a sun gear S, a planetary carrier T carrying a plurality of planetary gearwheels, and a ring gear U. However, the ring gear U now forms the first output element A 1 of the transfer box VG and is connected in a rotationally fixed manner to the first intermediate shaft ZW 1 and the sun gear S forms the second output element A 2 of the transfer box VG and is connected in a rotationally fixed manner to the second intermediate shaft ZW 2 . On the other hand, the planetary carrier T, as before, forms the input element E 0 of the transfer box VG and is connected directly and in a rotationally fixed manner to the driveshaft TW of the drive engine VM. The mode of action of this variant, however, corresponds exactly to that of the third embodiment of the powershift transmission shown in FIG. 3 a.

FIG. 3 d shows schematically a third further development of the third embodiment of the powershift transmission according to FIG. 3 a , which consists in that the two main separator clutches CL 1 ′, CL 2 ′ are this time in the form of frictionally synchronized synchronous clutches. This makes it possible, during a gearshift between an initially loaded gear and a target gear in different partial transmissions TG 1 , TG 2 to carry out the synchronization of the main separator clutch (CL 1 ′ or CL 2 ′) of the powershift transmission (TG 1 or TG 2 ) comprising the target gear, which is done at the beginning, by compressing the friction elements of the main separator clutch (CL 1 ′ or CL 2 ′) concerned instead of by adjusting the transmission ratio of the variator V.

A further change compared with the third embodiment of the powershift transmission according to FIG. 3 a , consists in that the variator V can be connected to and disconnected from the first intermediate shaft ZW 1 by means of a separator clutch I, and the two intermediate shafts ZW 1 , ZW 2 can be indirectly connected in a rotationally fixed manner to and disconnected from one another by means of a bridging clutch J, i.e. by a rotationally fixed connection of the input element E 0 (planetary carrier T) of the transfer box VG (planetary gearset PG) to the second intermediate shaft ZW 2 . By disengaging the separator clutch I and engaging the bridging clutch J, in a load-carrying gear (G 2 , G 4 , G 6 or R or else G 1 , G 3 , G 5 or G 7 ) of either partial transmission (TG 1 or TG 2 ), outside of shifting processes a permanent force flow via the variator V and an associated loss of efficiency can be avoided.

FIG. 3 e shows schematically a further development of the embodiment of the powershift transmission according to FIG. 3 d , which consists in that the two main separator clutches CL 1 *, CL 2 * are now in the form of disk clutches. As with the previously described variant, this makes it possible during a gearshift between an initially loaded gear and a target gear in different partial transmissions TG 1 , TG 2 , to carry out the synchronization of the main separator clutch (CL 1 * or CL 2 *) of the partial transmission (TG 1 or TG 2 ) that comprises the target gear, which is done at the beginning, by compressing together the friction elements of the main separator clutch (CL 1 * or CL 2 *) concerned, instead of by adjusting the transmission ratio of the variator V. Such a process sequence for the control of a gearshift carried out as a powershift, is described in detail below.

A further change compared with the previously described variant of the powershift transmission shown in FIG. 3 d , consists in that the separator clutch I and the bridging clutch J are now combined in a dual shifting element S 5 , whereby compared with a separate arrangement of the clutches I, J, one clutch control element is saved. In contrast to the separate arrangement in FIG. 3 d , however, the bridging clutch J is now arranged between the input element E 0 (planetary carrier T) of the transfer box VG (planetary gearset PG) and the first intermediate shaft ZW 1 , but the bridging clutch J still performs the same function, namely to connect the two intermediate shafts ZW 1 , ZW 2 in a rotationally fixed manner when it is engaged, by blocking the transfer box VG.

Below, with reference to FIGS. 4 to 12 typical gearshifts, each carried out as powershifts between an initially loaded gear and a target gear in different partial transmissions TG 1 , TG 2 , will be explained for all of the above-described embodiments and further developments of the powershift transmission designed according to the invention. For that purpose FIGS. 4 to 12 are each divided into five part-figures (for example FIGS. 4 a to 4 e ), each being a time diagram, in which in the first part-figure (such as FIG. 4 a ) in each case the rotational speed variation n VM (t) of the drive engine VM or its driveshaft TW, the rotational speed variations n ZW1 (t) and n ZW2 (t) of the two intermediate shafts ZW 1 , ZW 2 , the rotational speed variations n GE1 (t) and n GE2 (t) of the two input shafts GE 1 , GE 2 , and the rotational speed variation n GA (t) of the output shaft GA during the gearshift process concerned are shown. The next two part-figures (such as FIGS. 4 b and 4 c ) show, respectively, the shift condition of the two main separator clutches (CL 1 , CL 2 or CL 1 *, CL 2 *) during the gearshift process concerned, wherein the disengaged or open condition is always denoted by a 0 and the engaged or closed condition always by a 1. The last two part-figures (such as FIGS. 4 d and 4 e ) show the respective shift condition of the two gear clutches (B, F or B′, F′; A, F or A′, F′; C, G or C′, G′) involved in the gearshift during the gearshift process concerned, wherein once again the disengaged or open condition is always denoted by a 0 and the engaged or closed condition always by a 1.

FIGS. 4 a to 4 e show the sequence of a gearshift carried out as a powershift, which can be used in the powershift transmission according to FIG. 1 a and the powershift transmission according to FIG. 2 a , during a shift form an initially loaded gear in the first partial transmission TG 1 to a target gear in the second partial transmission TG 2 . As an example, the gearshift portrayed is an upshift from the fourth gear G 4 to the fifth gear G 5 .

Between times t 1 and t 2 the second intermediate shaft ZW 2 is slowed, by an appropriate transmission ratio adjustment of the variator V, i.e. by increasing the transmission ratio i V of the variator V defined between the first intermediate shaft ZW 1 and the second intermediate shaft ZW 2 , down to the rotational speed n GE2 of the second input shaft GE 2 whereby the second main separator clutch CL 2 is synchronized, which is then engaged at time t 2 ( FIGS. 4 a , 4 c ). Until then the second input shaft GE 2 rotated at a relatively low rotational speed under the action of drag torques, since until then it was not in driving connection either with the second intermediate shaft ZW 2 or with the output shaft GA.

After coupling the second input shaft GE 2 to the second intermediate shaft ZW 2 , the latter is accelerated by an appropriate transmission ratio adjustment of the variator V, i.e. by reducing the transmission ratio i V of the variator V, to the synchronous speed at the gear clutch F of the target gear G 5 . When the synchronous speed is reached at time t 3 the gear clutch F of the target gear G 5 is engaged ( FIG. 4 e ), whereby the target gear G 5 is engaged within the transmission. Thereafter, between times t 3 and t 4 the transmission ratio of the variator V is adjusted in the direction toward a higher ratio i V , by virtue of which the first intermediate shaft ZW 1 is relieved of the transmitted engine torque M VM and the second intermediate shaft ZW 2 thereby takes up the load, i.e. the load is taken up by the second partial transmission TG 2 or the target gear G 5 from the first partial transmission TG 1 or the originally loaded gear G 4 .

When freedom from load is reached at time t 4 , the first main separator clutch CL 1 ( FIG. 4 b ) and the gear clutch B of the originally loaded gear G 4 ( FIG. 4 d ) are disengaged, whereby the loaded gear G 4 is disengaged within the transmission.

Thereafter, between times t 4 and t 5 the rotational speed is adapted, i.e. the drive engine UM is slowed down to the rotational speed of the second intermediate shaft ZW 2 or the second input shaft GE 2 coupled to it (n ZW2 =n GE2 ), and at the same time the variator V is reset to its rest ratio (i V — 0 =1). During this, in order to avoid a torque collapse at the output shaft GA, which would be perceptible as a jerk, the rotational speed of the drive engine VM is adapted as much as possible so that the engine power, i.e. the product of the engine torque M VM and the engine rotational speed n VM , is kept constant (M VM *n VM =const.). When the driveshaft TW of the drive engine VM and the two intermediate shafts ZW 1 , ZW 2 are running synchronously, namely at time t 5 , the shifting process has ended. Meanwhile, the first input shaft GE 1 which, since time t 4 , has no longer been coupled to the first intermediate shaft ZW 1 and is no longer in driving connection with the output shaft GA, approaches a relatively low rotational speed where it is maintained under the action of drag torques.

The shifting process described can also be carried out in this form with the other variants of the powershift transmission concerned according to FIGS. 1 b to 1 d and 2 b to 2 e , even though the characteristic of friction-synchronization of the main separator clutches (CL 1 ′, CL 2 ′ or CL 1 *, CL 2 *) made as synchronous clutches or disk clutches in the variants shown in FIGS. 1 c and 1 d and in FIGS. 2 d and 2 e , and the corresponding characteristic of the gear clutches (A′ to H′) made as synchronous clutches in the variants shown in FIGS. 1 b and 2 b , are then not utilized.

FIGS. 5 a to 5 e now show the sequence of the same gearshift carried out as a powershift from the fourth gear G 4 to the fifth gear G 5 , which can be used with the embodiment variant of the powershift transmission according to FIG. 1 d and with the embodiment variant of the powershift transmission according to FIG. 2 e , i.e. when the two main separator clutches CL 1 *, CL 2 * are in the form of disk clutches, during a shift from an initially loaded gear of the first partial transmission TG 1 to a target gear of the second partial transmission TG 2 .

Between times t 1 * and t 2 * the second input shaft GE 2 , which to begin with rotates freely at a relatively low rotational speed, is accelerated by the continuous closing of the second main separator clutch CL 2 * ( FIG. 5 c ) up to the rotational speed n ZW2 of the second intermediate shaft ZW 2 ( FIG. 5 a ) and, when the synchronous rotational speed is reached at the second main separator clutch CL 2 *, namely at time t 2 *, the clutch is fully closed by over-compression. Thereafter, the second input shaft GE 2 together with the second intermediate shaft ZW 2 are slowed to the synchronous rotational speed at the gear clutch F of the target gear G 5 by an appropriate transmission ratio adjustment of the variator V, i.e. by increasing the ratio i V of the variator. When the synchronous rotational speed is reached at time t 3 the gear clutch F of the target gear G 5 is engaged ( FIG. 5 e ), whereby the target gear G 5 is engaged within the transmission. The remainder of the shifting sequence ( FIGS. 5 a , 5 d ) corresponds to the sequence described earlier with reference to part- FIGS. 4 a to 4 e . The advantage of this process variant compared with the shifting sequence according to FIGS. 4 a to 4 e is that the relatively large transmission ratio adjustment required at the beginning of the shifting process in the latter case for synchronizing the second main separator clutch CL 2 , can in this case be avoided.

When, as in the embodiment variants of the powershift transmission shown in FIGS. 1 c and 2 d , the main separator clutch CL 1 ′ and CL 2 ′ are in the form of synchronous clutches, the above-described process variant can be carried out to a large extent identically. In such a case the second input shaft GE 2 is accelerated between times t 1 * and t 2 * to the synchronous rotational speed by pressing together the friction elements of the second main separator clutch CL 2 ′, and the second main separator clutch CL 2 ′ is then engaged when the synchronous rotational speed is reached at time t 2 *.

Finally, FIGS. 6 a to 6 e show the sequence of the same gearshift carried out as a powershift, which can be used with the powershift transmission according to FIG. 1 b and with the powershift transmission according to FIG. 2 b , i.e. when the gear clutches A′ to H′ are made as frictionally synchronized synchronous clutches, for a gearshift from an initially loaded gear of the first partial transmission TG 1 to a target gear of the second partial transmission TG 2 .

Between times t 1 ′ and t 2 ′ the second input shaft GE 2 , at first rotating freely at a relatively slow rotational speed, is now accelerated to the synchronous rotational speed of the gear clutch F′ of the target gear G 5 by pressing together the friction elements of the gear clutch F′ ( FIGS. 6 a , 6 e ). When the synchronous rotational speed is reached at time t 2 ′, the gear clutch F′ of the target gear G 5 is engaged, whereby the target gear G 5 is engaged within the transmission. Thereafter, by an appropriate transmission ratio adjustment of the variator V, i.e. by increasing the ratio i V of the variator, the second intermediate shaft ZW 2 is slowed to the rotational speed of the second input shaft GE 2 and thereby the second main separator clutch CL 2 is synchronized, which is then engaged at time t 3 ( FIG. 6 c ). The remainder of the shifting sequence corresponds to the sequence described earlier with reference to part- FIGS. 4 a to 4 e and 5 a to 5 e (see here FIGS. 6 a to 6 e ). Thus, in this process variant the synchronization sequence of the second main separator clutch CL 2 and the gear clutch F′ of the target gear G 5 is exchanged compared with the two previously described process variants. Compared with the shifting sequence according to FIGS. 4 a to 4 e , the present process variant also has the advantage that the relatively large transmission ratio adjustment of the variator V required at the beginning of the shifting process in the variant according to FIGS. 4 a to 4 e in order to synchronize the second main separator clutch CL 2 , can be avoided.

Since in the first and second embodiments of the powershift transmission shown, respectively, in FIGS. 1 a and 2 a , in contrast to the second intermediate shaft ZW 2 it is the first intermediate shaft ZW 1 which is in each case connected directly and rotationally fixed to the driveshaft TW of the drive engine VM, a gearshift carried out as a powershift from an initially loaded gear of the second partial transmission TG 2 to a target gear of the first partial transmission TG 1 necessarily takes place otherwise than in the reverse shifting direction.

Accordingly, FIGS. 7 a to 7 e show the sequence of a gearshift carried out as a powershift, which can be used in the powershift transmission shown in FIG. 1 a and in the powershift transmission shown in FIG. 2 a during a gearshift from an initially loaded gear of the second partial transmission TG 2 to a target gear of the first partial transmission TG 1 . The shift described is, for example, an upshift from the fifth gear G 5 to the sixth gear G 6 .

Between times t 1 and t 2 the first intermediate shaft ZW 1 , which is connected in a rotationally fixed manner to the driveshaft TW of the drive engine VM, in combination with a simultaneous rotational speed adaptation of the engine VM by an appropriate transmission ratio adjustment of the variator V, i.e. by reducing the ratio i V of the variator V defined between the first intermediate shaft ZW 1 and the second intermediate shaft ZW 2 , is slowed to the rotational speed n GE1 of the first input shaft GE 1 , and thus the first main separator clutch CL 1 is synchronized and is engaged at time t 2 ( FIG. 7 b ). Until then the first input shaft GE 1 was rotating under the action of drag torques at a relatively slow rotational speed since until then it was neither coupled to the first intermediate shaft ZW 1 , nor was it in driving connection with the output shaft GA. In order during this to avoid a torque collapse at the output shaft GA that would be perceptible as a jerk, the rotational speed adaptation of the drive engine VM is carried out in such manner that the engine power, i.e. the product of the engine torque M VM and the engine rotational speed n VM , is kept constant (M VM *n VM =const.).

When the first input shaft GE 1 has been coupled to the first intermediate shaft ZW 1 , in combination with a simultaneous rotational speed adaptation of the drive engine VM by an appropriate transmission ratio adjustment of the variator V, i.e. by increasing the transmission ratio i V of the variator V, the first intermediate shaft ZW 1 is accelerated to the synchronous rotational speed at the gear clutch A of the target gear G 6 . When the synchronous rotational speed is reached at time t 3 the gear clutch A of the target gear G 6 is engaged, whereby the target gear G 6 is engaged within the transmission ( FIG. 7 d ). Thereafter, between times t 3 and t 4 , by means of a slight transmission ratio adjustment of the variator V in the direction toward a lower ratio i V , by virtue of which the second intermediate shaft ZW 2 is relieved of the transmitted engine torque M VM and the first intermediate shaft ZW 1 correspondingly takes up the load, the load is taken up by the first partial transmission TG 1 or the target gear G 6 from the second partial transmission TG 2 or the originally loaded gear G 5 .

When the second intermediate shaft ZW 2 becomes free from load, which happens at time t 4 , the second main separator clutch CL 2 and the gear clutch F of the originally loaded gear G 5 are disengaged, whereby the loaded gear G 5 is disengaged within the transmission ( FIGS. 7 c , 7 e ).

Thereafter, between times t 4 and t 5 the variator V is reset to its rest transmission ratio (i V — 0 =1), whereby the load-free second intermediate shaft ZW 2 is slowed to the rotational speed n ZW1 =n VM of the first intermediate shaft ZW 1 and the drive engine VM and its driveshaft TW. A rotational speed adaptation of the drive engine VM is now no longer needed, since this has already taken place completely. When the variator V reaches its rest ratio (i V — 0 =1) at time t 5 , the shifting process is completed ( FIG. 7 a ). Meanwhile the second input shaft GE 2 , which since time t 4 is no longer coupled to the second intermediate shaft ZW 2 and is no longer in driving connection with the output shaft GA, approaches a relatively low rotational speed at which it is maintained under the action of drag torques.

The shifting process described can also be carried out in this form with the other variants of the powershift transmission concerned and shown in FIGS. 1 b to 1 d , 2 b to 2 e , even if the characteristic of frictional synchronization of the main separator clutches (CL′, CL 2 ′ or CL 1 *, CL 2 *) made as synchronous clutches or disk clutches, respectively, in the variants according to FIGS. 1 c and 1 d and FIGS. 2 d and 2 e , and the corresponding characteristics of the gear clutches (A′ to H′) made as synchronous clutches in the variants shown in FIGS. 1 b and 2 b , are not utilized.

FIGS. 8 a to 8 e now show the sequence of the same gearshift carried out as a powershift, during a shift from an initially loaded gear in the second partial transmission TG 2 to a target gear in the first partial transmission TG 1 , for the case of the powershift transmission according to FIG. 1 d and the powershift transmission according to FIG. 2 e , i.e. when the two main separator clutches CL 1 *, CL 2 * are in the form of disk clutches.

Between times t 1 * and t 2 * the first input shaft GE 1 , which to begin with rotates freely at a relatively low rotational speed, is accelerated by the continuous closing of the first main separator clutch CL 1 * to the rotational speed n ZW1 of the first intermediate shaft ZW 1 and when the synchronous rotational speed is reached at the first main separator clutch CL 1 * at time t 2 *, the clutch is fully closed by compression ( FIGS. 8 a , 8 b ). Thereafter the first input shaft GE 1 together with the first intermediate shaft ZW 1 , in combination with a simultaneous rotational speed adaptation of the drive engine VM by an appropriate transmission ratio adjustment of the variator V, i.e. by reducing the ratio i V of the variator V, are slowed down to the synchronous rotational speed of the gear clutch A of the target gear G 6 . When the synchronous rotational speed is reached at time t 3 the gear clutch A of the target gear G 6 is engaged, whereby the target gear G 6 is engaged within the transmission ( FIGS. 8 a , 8 d ). The remainder of the shifting sequence corresponds to the sequence described previously with reference to part- FIGS. 7 a to 7 e (see FIGS. 8 a to 8 e ). The advantage of this process variant compared with the shifting sequence shown in FIGS. 7 a to 7 e is that the relatively large transmission ratio adjustment of the variator V required at the beginning of the shifting process in the latter case for the synchronization of the main separator clutch CL 1 , can be avoided.

When in accordance with the variants of the powershift transmission shown in FIGS. 1 c and 2 d the two main separator clutches CL 1 ′, CL 2 ′ are in the form of synchronous clutches, this process variant can be carried out in a largely identical manner. In that case between times t 1 * and t 2 * the first input shaft GE 1 is accelerated to the synchronous rotational speed of the first main separator clutch CL 1 ′ by pressing the friction elements of the clutch together, and when the synchronous rotational speed has been reached at time t 2 * the main separator clutch CL 1 ′ is engaged.

Finally, FIGS. 9 a to 9 e show the sequence of the same gearshift carried out as a powershift, that can be used with the powershift transmission of FIG. 1 b and the powershift transmission of FIG. 2 b , i.e. when the gear clutches A′ to H′ are in the form of frictionally synchronized synchronous clutches, for a shift from an initially loaded gear of the second partial transmission TG 2 to a target gear of the first partial transmission TG 1 .

Between times t 1 ′ and t 2 ′ the first input shaft GE 1 , initially rotating at a relatively low rotational speed, is now accelerated, by pressing together the friction elements of the gear clutch A′ of the target gear G 6 , to the synchronous rotational speed of the gear clutch A′ ( FIGS. 9 a , 9 d ). When the synchronous rotational speed is reached at time t 2 ′ the gear clutch A′ of the target gear G 6 is engaged, whereby the target gear G 6 is engaged within the transmission ( FIG. 9 d ). Thereafter, in combination with a simultaneous rotational speed adaptation of the drive engine VM by an appropriate transmission ratio adjustment of the variator V, i.e. a reduction of the ratio i V of the variator V, the first intermediate shaft ZW 1 is slowed down to the rotational speed n GE1 of the first input shaft GE 1 and the first main separator clutch CL 1 is thus synchronized and engaged at time t 3 ( FIG. 9 b ). The remainder of the shifting sequence corresponds to the sequence described earlier with reference to part- FIGS. 7 a to 7 e and 8 a to 8 e (see FIGS. 9 a to 9 e ). Thus, in the present process variant, compared with the two process variants described above the synchronization sequence of the first main separator clutch CL 1 and the gear clutch A′ of the target gear G 6 is reversed. Compared with the shifting sequence according to FIG. 7 , the present process variant also has the advantage that the relatively large transmission ratio adjustment of the variator V required at the beginning of the shifting process for synchronizing the first main separator clutch CL 1 , can be avoided.

In the powershift transmission according to FIG. 3 a the two intermediate shafts ZW 1 , ZW 2 are in driving connection with the driveshaft TW of the drive engine VM via the transfer box VG in a symmetrical arrangement. Thus, a gearshift carried out as a powershift from an initially loaded gear of the first partial transmission TG 1 to a target gear in the second partial transmission TG 2 takes place, in principle, in the same manner as a gearshift carried out as a powershift from an initially loaded gear of the second partial transmission TG 2 to a target gear of the first partial transmission TG 1 . On the other hand, in this embodiment of the powershift transmission a transmission ratio adjustment of the variator V is always connected with a rotational speed change of the drive engine VM and its driveshaft TW when, during this, the rotational speed (n ZW1 or n ZW2 ) of one of the two intermediate shafts (ZW 1 or ZW 2 ) is kept substantially constant.

Accordingly, FIGS. 10 a to 10 e shows the sequence of a gearshift carried out as a powershift, which can be used with the powershift transmission shown in FIG. 3 a , for a shift from an initially loaded gear in one of the partial transmissions (TG 1 or TG 2 ) to a target gear in the other partial transmission (TG 2 or TG 1 ), i.e. in both shifting directions. The shift described is, as an example, a downshift from the third gear G 3 to the second gear G 2 , i.e. a shift from an initially loaded gear (G 3 ) of the second partial transmission TG 2 to a target gear (G 2 ) of the first partial transmission TG 1 .

Between times t 1 and t 2 the first intermediate shaft ZW 1 , which is in driving connection via the transfer box VG with the driveshaft TW of the drive engine VM, in combination with a simultaneous rotational speed adaptation of the engine VM by means of an appropriate transmission ratio adjustment of the variator V, i.e. by reducing the ratio i V of the variator V defined between the first and second intermediate shafts ZW 1 , ZW 2 , is slowed down to the rotational speed n GE1 of the first input shaft GE 1 and the first main separator clutch CL 1 is thus synchronized and engaged at time t 2 ( FIG. 10 b ). Until then the first input shaft GE 1 was rotating under the action of drag torques at a relatively low rotational speed, since until then it was neither coupled to the first intermediate shaft ZW 1 nor was it in driving connection with the output shaft GA. In order during this to avoid a torque collapse at the output shaft GA which would be perceived as a jerk, the rotational speed adaptation of the engine VM is carried out as much as possible in such manner that the engine power, i.e. the product of the engine torque M VM and the engine rotational speed n VM , is kept constant (M VM *n VM =const.).

After coupling the first input shaft GE 1 to the first intermediate shaft ZW 1 , in combination with a simultaneous rotational speed adaptation of the engine VM by means of an appropriate transmission ratio adjustment of the variator V, i.e. by increasing the ratio i V of the variator, the first intermediate shaft ZW 1 is accelerated to the synchronous rotational speed of the gear clutch C of the target gear G 2 . When the synchronous rotational speed is reached at time t 3 the gear clutch C of the target gear G 2 is engaged, whereby the target gear G 2 is engaged within the transmission ( FIGS. 10 , 10 d ). Thereafter, between times t 3 and t 4 a slight transmission ratio adjustment of the variator V in the direction toward a lower ratio i V relieves the second intermediate shaft ZW 2 from the transmitted engine torque M VM , whereas the first intermediate shaft ZW 1 takes up the load, completing the load uptake by the first partial transmission TG 1 or its target gear G 2 from the second partial transmission TG 2 or its originally loaded gear G 3 ( FIG. 10 a ).

When the second intermediate shaft ZW 2 has become free from load at time t 4 , the second main separator clutch CL 2 and the gear clutch G of the initially loaded gear G 2 are disengaged, whereby the initially loaded gear G 2 is disengaged within the transmission ( FIGS. 10 c , 10 e ).

Thereafter, between times t 4 and t 5 , in combination with a simultaneous rotational speed adaptation of the drive engine VM the variator V is reset to its rest transmission ratio (i V — 0 =1), whereby the load-free second intermediate shaft ZW 2 is accelerated to the rotational speed n ZW1 of the first intermediate shaft ZW 1 ( FIG. 10 a ). When the variator V reaches its rest ratio (i V — 0 =1) at time t 5 , the shifting process has ended. Meanwhile the second input shaft GE 2 , which since time t 4 has no longer been coupled to the second intermediate shaft ZW 2 and is no longer in driving connection with the output shaft GA, approaches a relatively low rotational speed at which it is kept by the action of drag torques.

The shifting process just described can also be carried out in this form with the other variants of the powershift transmission concerned according to FIGS. 3 b to 3 e , even if then the characteristic of frictional synchronization of the main separator clutches (CL 1 ′, CL 2 ′ or CL 1 *, CL 2 *) made as synchronous clutches or disk clutches, respectively, in the variants shown in FIGS. 3 d and 3 e , and the corresponding characteristic of the gear clutches (A′ to H′) made as synchronous clutches in the variant shown in FIG. 3 b , are not utilized.

FIGS. 11 a to 11 e show the sequence for the same gearshift carried out as a powershift, which can be used with the powershift transmission according to FIG. 3 a , i.e. when the two main separator clutches CL 1 *, CL 2 * are in the form of disk clutches, during a shift from an initially loaded gear of one partial transmission (TG 1 or TG 2 ) to a target gear of the other partial transmission (TG 2 or TG 1 ).

Between times t 1 and t 2 * the first input shaft GE 1 , which initially rotates freely at a relatively low rotational speed, is now accelerated by the continuous closing of the first main separator clutch CL 1 * to the rotational speed n ZW1 of the first intermediate shaft ZW 1 and when the synchronous rotational speed is reached at the first main separator clutch CL 1 * at time t 2 *, the clutch is fully closed by compression ( FIGS. 11 a , 11 b ). Thereafter the first input shaft GE 1 together with the first intermediate shaft ZW 1 , in combination with a simultaneous rotational speed adaptation of the drive engine VM by means of an appropriate transmission ratio adjustment of the variator V, i.e. by increasing the ratio i V of the variator V, are accelerated to the synchronous rotational speed at the gear clutch C of the target gear G 2 . When the synchronous rotational speed is reached at time t 3 , the gear clutch C of the target gear G 2 is engaged, whereby the target gear G 2 is engaged within the transmission ( FIG. 11 c ). The remainder of the shifting sequence corresponds to the sequence described above with reference to part- FIGS. 10 a to 10 e (see FIGS. 11 a to 11 e ). Compared with the shifting sequence according to FIG. 10 , the advantage of the present process variant is that the relatively large transmission ratio adjustment of the variator V required at the beginning in that case for synchronizing the first main separator clutch CL 1 , can be avoided.

If, as in the variant of the powershift transmission according to FIG. 3 d , the main separator clutches CL 1 ′, CL 2 ′ are in the form of synchronous clutches, the present process variant can be carried out in a largely identical manner. Between times t 1 * and t 2 * the first input shaft GE 1 is accelerated, by compressing the friction elements of the first main separator clutch CL 1 ′, to the synchronous rotational speed and when this is reached at time t 2 *, the first separator clutch CL 1 ′ is engaged.

Finally, FIGS. 12 a to 12 e show the sequence for the same gearshift carried out as a powershift, which can be used with the powershift transmission according to FIG. 3 b , i.e. when the gear clutches A′ to H′ are in the form of frictionally synchronized synchronous clutches, for shifting from an initially loaded gear of one partial transmission (TG 1 or TG 2 ) to a target gear of the other partial transmission (TG 2 or TG 1 ).

Between times t 1 ′ and t 2 ′ the first input shaft GE 1 , initially rotating freely at a relatively low rotational speed, is accelerated by pressing together the friction elements of the gear clutch C′ of the target gear G 2 to the synchronous rotational speed of the gear clutch C′ ( FIGS. 12 a , 12 d ). When the synchronous rotational speed is reached at time t 2 ′ the gear clutch C′ of the target gear G 2 is engaged, whereby the target gear G 2 is engaged within the transmission ( FIG. 12 d ). Thereafter, in combination with a simultaneous rotational speed adaptation of the drive engine VM by an appropriate transmission ratio adjustment of the variator, i.e. by increasing the ratio i V of the variator, the first intermediate shaft ZW 1 is accelerated to the rotational speed n GE1 of the first input shaft GE 1 and thereby the first main separator clutch CL 1 is synchronized, and then engaged at time t 3 ( FIG. 12 b ). The remainder of the shifting sequence corresponds to the sequence described earlier with reference to part- FIGS. 10 a to 10 e and 11 a to 11 e (see FIGS. 12 a to 12 e ). Thus, in the present process variant, compared with the two previously described process variants the synchronization sequence of the main separator clutch CL 1 and the gear clutch C′ of the target gear G 2 is reversed. Compared with the shifting sequence according to FIG. 10 , the present process variant has the advantage that the relatively large transmission ratio adjustment of the variator V required at the beginning of the shifting process in that case for synchronizing the first main separator clutch CL 1 , can be avoided.

A-H Gear clutches, claw clutches

A′-H′ Gear clutches, synchronous clutches

A 0 Output element of the superposition gearset SG

A 1 First output element of the transfer box VG

A 2 Second output element of the transfer box VG

CL 1 First main separator clutch, claw clutch

CL 1 ′ First main separator clutch, synchronous clutch

CL 1 * First main separator clutch, disk clutch

CL 2 Second main separator clutch, claw clutch

CL 2 ′ Second main separator clutch, synchronous clutch

CL 2 * Second main separator clutch, disk clutch

E 0 Input element of the transfer box VG

E 1 First output element of the superposition gearset SG

E 2 Second output element of the superposition gearset SG

EK 1 First input constant

EK 2 Second input constant

G 1 -G 7 Gear steps, forward gears

GA Output shaft

GE 1 First input shaft

GE 2 Second input shaft

I Separator clutch at the variator V

i G — n Gear ratio of the n-th gear

i G — n+1 Gear ratio of the (n+1)-th gear

i V Transmission ratio of the variator V

i V — 0 Rest transmission ratio of the variator V

i V — max Highest transmission ratio of the variator V

i V — min Lowest transmission ratio of the variator V

J Bridging clutch at the variator V

K Starting clutch

LSM Powershift module

LSM′ Powershift module

LSM* Powershift module

M VM Engine torque, torque from the internal combustion engine VM

n Rotational speed

n GA Rotational speed of the output shaft GA

n GE1 Rotational speed of input shaft GE 1

n GE2 Rotational speed of input shaft GE 2

n ZW1 Rotational speed of intermediate shaft ZW 1

n ZW2 Rotational speed of intermediate shaft ZW 2

n VM Engine rotational speed, rotational speed of the driveshaft TW

PG Planetary gearset

R Gear step, reversing gear

S Sun gear of the planetary gearset PG

S 1 -S 5 Coupling shifting elements

SG Superposition gearset, summation gearset

t Time

t 1 -t 5 Time points

t 1 ′, t 2 ′ Time points

t 1 *, t 2 * Time points

T Planetary carrier of the planetary gearset PG

TG 1 First partial transmission

TG 2 Second partial transmission

TW Driveshaft of the internal combustion engine VM

U Ring gear of the planetary gearset PG

V Variator

VG Transfer box, differential transmission

VG 1 First countershaft

VG 2 Second countershaft

VM Drive engine, internal combustion engine

Z 1 -Z 5 Spur gear stages

Z 7 Seventh spur gear stage

ZR Reversing spur gear stage

ZW 1 First intermediate shaft

ZW 2 Second intermediate shaft

Δn ZW Rotational speed difference between ZW 1 and ZW 2

φ max Maximum gear interval of G 1 -G 7