DOUBLE GEARBOX, AND METHOD FOR ENGAGING AN OVERALL GEAR RATIO THEREIN

A double clutch transmission having a first and a second input shaft; a countershaft; a drive output shaft; a first sub-transmission having two gearsets, of which one gearset can be selectively connected to transmit torque between the first input shaft and the countershaft; a second sub-transmission having two gearsets, of which one gearset can be selectively connected to transmit torque between the second input shaft and the countershaft; and a third sub-transmission having one gearset which can be connected to transmit torque between the countershaft and the drive output shaft. In this case, one of the gearsets of the first sub-transmission coincides with one of the gearsets of the second sub-transmission.

FIELD OF THE INVENTION

The invention concerns a shiftable double transmission with intermediate gearing, preferably for a utility vehicle.

BACKGROUND OF THE INVENTION

A double transmission comprises a first and a second input shaft, each of which can be coupled to a drive input shaft for the transmission of torque by way of an associated friction clutch, the drive input shaft being connected for example to an internal combustion engine. With each input shaft is associated a sub-transmission which usually has a plurality of gearsets that can be engaged in alternation between the input shaft concerned and a countershaft. A further sub-transmission usually comprises a plurality of gearsets which can be engaged between the countershaft and the drive output shaft. By virtue of the shiftable gearsets and the clutches, various overall gear ratios can be engaged between the driveshaft and the drive output shaft.

To carry out a gearshift free from traction force interruption between a first and a second engaged overall gear ratio, between in each case one of the two respective input shafts and the output shaft various gearsets are engaged, wherein however only one of the friction clutches is closed. Then the closed friction clutch is opened and the open friction clutch is closed.

DE 10 2012 213 517 A1 relates to a dual clutch transmission with intermediate gearing for utility vehicles. In addition to the three above-mentioned sub-transmissions a shiftable connection between the first input shaft and the output shaft is possible. In combination with a downstream group transmission that can be shifted in two stages, twelve forward gears and four reversing gears can be obtained.

DE 2012 217 503 A1 shows a transmission of similar configuration, which enables eleven forward and four reversing gears to be engaged without traction force interruption.

SUMMARY OF THE INVENTION

The purpose of the present invention is to indicate an improved double transmission with intermediate gearing and a better control method for the double transmission. The invention achieves these objectives by virtue of the characteristics specified in the independent claims. Preferred embodiments are described in the subordinate claims.

A double transmission comprises a first and a second input shaft; a countershaft; a drive output shaft; a first sub-transmission having at least two gearsets, of which selectively one can be connected to transmit torque between the first input shaft and the countershaft; a second sub-transmission having at least two gearsets, of which selectively one can be connected to transmit torque between the second input shaft and the countershaft; and a third sub-transmission having at least one gearset, which can be connected to transmit torque between the countershaft and the drive output shaft. In this case, one of the gearsets of the first sub-transmission coincides with one of the gearsets of the second sub-transmission.

The double transmission can in particular be used for a utility vehicle, wherein a gearshift free from traction force interruption between various gear ratios can for example have consumption advantages for the drive motor connected to the transmission. Thanks to the double utilization of one of the gearsets for the first and for the second sub-transmission, the number of gearsets and thus also the number of wheel planes of the double transmission can be reduced. In that way the double transmission can be made more compact or lighter. If the first and second sub-transmissions together have three gearsets and the third sub-transmission has a further two gearsets, then with a total of five wheel planes six overall gear ratios can be obtained with the double transmission. The first and the second sub-transmission can also have more than two gearsets independently of one another.

In a particularly preferred embodiment, a shifting element is also provided in order to connect the first input shaft for the transmission of torque to a first intermediate shaft, which acts upon the drive output shaft. If the intermediate shaft is connected directly to the drive output shaft, then by actuating the shifting element a direct gear can be obtained, in which the rotational speed of the first input shaft is the same as the rotational speed of the drive output shaft. In addition, by means of the first two sub-transmissions a reduction ratio can be produced between the second input shaft and the first input shaft. For this, in the first two sub-transmissions different gear ratios are engaged, so enabling a torque flow from the second input shaft, via the second sub-transmission, to the countershaft and from there, via the first sub-transmission, to the first input shaft. In that way additional overall gear ratios of the double transmission can be obtained. An overall gear ratio in which the shifting element between the first input shaft and the first intermediate shaft is closed, is also known as a coupling gear.

In a further preferred embodiment, the third sub-transmission comprises at least two gearsets, of which one can optionally be connected to transmit torque between the countershaft and the drive output shaft and wherein one of the gearsets is preferably non-reversing (i.e. it maintains its rotational direction), whereas it is also preferable for at least one of the other gearsets to be reversing (i.e. it can reverse its rotational direction). By virtue of the gearsets associated with it, the third sub-transmission produces various drive output constants by way of which the countershaft can be coupled to the drive output shaft. In the torque flow between one of the input shafts and one of the intermediate shafts there are usually no reversing gearsets or two reversing gearsets, so that the rotational directions of the input shaft and the intermediate shaft are the same. If a reversing and a non-reversing gearset are in the torque flow, then a negative overall gear ratio is produced so that a reversing gear can be obtained.

Building on the above-described double transmission, various variants can be formed. In a first variant a group transmission with two gearsets is provided in addition, of which either one or the other can be connected to transmit torque between the first intermediate shaft and the drive output shaft. In that way the number of overall gear ratios of the double transmission can be doubled. It is particularly preferred that the group transmission can be powershifted in order to make it possible for an overall gearshift free from traction force interruption to be carried out between any consecutive overall gear ratios of the double transmission. In total the double transmission can thereby produce twice-seven overall gear ratios, not counting the reversing gears.

In a preferred embodiment the group transmission comprises a planetary gearset with a sun gear, a planetary gear and a ring gear. In this case the sun gear is permanently connected to the first intermediate shaft and the planetary gear to the drive output shaft. The ring gear can be shifted either to idle or to be connected to the first intermediate shaft in a torque-transmitting manner. If both the sun gear and the ring gear are connected to the first intermediate shaft, then the step-down ratio of the planetary gearset is equal to 1. In contrast, if the ring gear is idling, then there is a positive reduction ratio between the first intermediate shaft and the drive output shaft.

In another variant a further shifting element is provided, in order to connect a gearset of the third sub-transmission to transmit torque to a second intermediate shaft. In addition a group transmission is provided, which comprises two gearsets, of which, either one can be connected to transmit torque between the first intermediate shaft and the drive output shaft, or the other can be connected to transmit torque between the second intermediate shaft and the drive output shaft. Thus, torque can be transmitted between the main transmission with the three sub-transmissions and the group transmission by way of two different intermediate shafts, whereby advantageous shifting combinations can be made possible.

Preferably, this group transmission also comprises a planetary gearset with a sun gear, a planetary gear and a ring gear. In particular the sun gear can be permanently connected to the first intermediate shaft and the planetary gear to the drive output shaft. In this case the ring gear can either idle or be connected to transmit torque to the drive output shaft. If both the ring gear and the planetary gear or a planetary carrier are connected to the drive output shaft, then the step-down ratio of the group transmission is equal to one. If the ring gear is idling, then the gear ratio of the group transmission is positive.

In general a gearset of one of the sub-transmissions preferably comprises a spur gear system with at least two gearwheels. The gearwheels mesh with one another and usually have different radii in order to produce a step-up or a step-down ratio. Other embodiments are also possible, for example by means of an epicyclic gearset.

The two gearwheels are usually mounted on different shafts in such manner that an engagement between a shaft and an associated gearwheel can be permanent or releasable. For a releasable engagement a shifting element is usually used, which forms or separates an interlocked connection between one of the gearwheels and a shaft. The shifting element can in particular comprise a sliding sleeve which, for example, is connected to the shaft with interlock in the circumferential direction by gearteeth but can be displaced axially. In a first position the sliding sleeve can engage with interlock in a gearwheel on the same shaft and in a second position it can leave the gearwheel free. In relation to the shaft the gearwheel is a loose wheel so that it can rotate freely about the shaft when not engaged by the shifting element. In relation to its associated shaft, a gearwheel with which the loose wheel meshes is usually a fixed wheel and is therefore permanently in torque-transmitting connection with its shaft. In this case the connection can be optionally interlocking, frictional, or material-merged. In general, none of the gearwheels can be displaced axially relative to its associated shaft.

The shifting element described can also be used in a third axial position to form a torque-transmitting connection between a further loose wheel and the shaft. The position of the sliding sleeve in which there is no interlock with either of the loose wheels on the shaft is usually between the positions in which there is engagement with a respective gearwheel. Thus, two shifting elements can be made integrally with one another. In a corresponding manner a shifting element can also be used to form a separate torque-transmitting connection between two shafts. For this the shafts are preferably arranged coaxially and particularly preferably end to end opposite one another. A sliding sleeve connected to one of the shafts in a torque-transmitting manner can be brought axially into interlocking engagement with the other shaft, in order to close the shifting element. In another embodiment one of the shafts can also be axially movable in order to form or separate a shiftable engagement with the other shaft.

In a particularly preferred embodiment, the double transmission also comprises a first friction clutch for connecting a drive input shaft to the first input shaft and a second friction clutch for connecting the drive input shaft to the second input shaft. In this case the double transmission can also be called a dual-clutch transmission. The double transmission or its shifting elements and the friction clutches can be controlled by a common control unit. In that way sequential control when shifting the double transmission between different overall gear ratios, in particular without interrupting the traction force between the drive input shaft and the drive output shaft, can be improved. The entire double transmission with the dual clutch can be designed as a separately handled unit, for example to be used in a motor vehicle, in particular a utility vehicle.

A method for engaging an overall gear ratio in the above-described double transmission with a dual clutch comprises the steps of engaging a first overall gear ratio between one of the input shafts and the drive output shaft and a second overall gear ratio between the other input shaft and the drive output shaft, and controlling a shift from the first to the second overall gear ratio by means of the friction clutches. In this, in each case the overall gear ratios are engaged in one of two ways. In the first way just one of the gearsets of the first or second sub-transmission is connected to transmit torque between one of the input shafts and the countershaft and just one gearset of the third sub-transmission is connected to transmit torque between the countershaft and the first intermediate shaft. In the second way just one gearset of the second sub-transmission is connected to transmit torque between the second input shaft and the countershaft. In addition just one gearset of the first sub-transmission is connected to transmit torque between the countershaft and the second input shaft, and the first input shaft is connected to transmit torque to the first intermediate shaft. By means of this procedure many of the physically possible overall gear ratios of the double transmission can be engaged without traction force interruption. Combinations of overall gear ratios with which this is not possible can nevertheless be obtained without traction force interruption by briefly engaging another overall gear ratio (“supporting gear”), as explained in greater detail below.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

FIG. 1shows a dual-clutch transmission100, in a first embodiment. The dual-clutch transmission100comprises a dual clutch105, a double transmission110and an optional group transmission115. The dual clutch105and the double transmission110can be made and used separately from one another, or made integrally with one another. The double transmission110, also called the main transmission, comprises a first sub-transmission120, a second sub-transmission125and a third sub-transmission130. The dual clutch105comprises a first friction clutch135and a second friction clutch140. A drive input shaft145of the dual clutch105can be continuously variably, frictionally connected by means of the first friction clutch135to a first input shaft150of the double transmission110and, independently of that, continuously variably, frictionally connected by means of the second friction clutch140to a second input shaft155of the double transmission110. Usually the input shafts150,155are made coaxial, such that a solid shaft runs inside a hollow shaft. It is preferable for whichever input shaft150,155is transmitting torque in gear step1(seeFIG. 2) to be a solid shaft. When the double transmission110is used, for example in a truck, this gear step can be used for starting off from rest.

An output side of the double transmission110is formed by an intermediate shaft160, which acts upon an input side of the group transmission115. A second intermediate shaft165, which can provide an alternative power transfer between the double transmission110and the group transmission115(seeFIGS. 4 and 7), is not present in the embodiment shown. On its output side, the group transmission115acts upon a drive output shaft170, which can be the same as the first intermediate shaft160if the group transmission115is omitted.

The first sub-transmission120is associated with the first input shaft150and the second sub-transmission125with the second input shaft155. In the embodiment illustrated, the sub-transmissions120,125each comprise two gearsets175, each of them comprising the two gearwheels that mesh with one another, one of these gearwheels in each case acting upon a countershaft180. As shown inFIG. 1a plurality of countershafts180each with associated gearsets175can be used, in order to increase the load-bearing capacity of the double transmission110.

To selectively form or break torque flow between the countershaft180and one of the shafts150,155or165, shifting elements A to G, H and L are provided. Each shifting element A to L is designed to either make or break a torque-transmitting connection. In this case an only partial transfer of torque, in particular by means of friction, is not provided, but rather, an interlocked engagement is made or broken by axial displacement of a shifting or coupling element. In the embodiment shown the shifting elements A, B, C, D, F and G act in each case between a gearwheel of a gearset175and one of the shafts150,155and165. The respective corresponding gearwheels of the gearsets175are connected permanently to the countershaft180. A converse arrangement, in which a shifting element makes or breaks the torque flow through the countershaft180, is also possible. The shifting element E is designed to make or break a torque flow between the first input shaft150and the first intermediate shaft160.

The group transmission115is preferably in the form of an epicyclic transmission shown as an example inFIG. 1as a planetary gearset with a sun gear185, a planetary gear190and a ring gear195. In other embodiments, however, another type of gearing can be used for the group transmission115. In the embodiment of the group transmission115shown, torque can be coupled by means of the first intermediate shaft160to the sun gear185and decoupled by means of the planetary gear or a planetary carrier at the drive output shaft170. In a first operating condition the ring gear195can be connected by means of the shifting element H to the first intermediate shaft160or the sun gear185, so that the gear ratio of the group transmission115is equal to 1 (one). In a second operating condition the ring gear195can be braked by means of the shifting element L, so that it is at rest for example relative to a housing of the group transmission115and/or the double transmission110. In that case a step-down ratio is obtained between the first intermediate shaft160and the drive output shaft170. It is particularly preferable for the group transmission115to be designed so that it can be powershifted, i.e. the shifting element H can be released and the shifting element L closed (or conversely) while torque is transmitted continuously between the first intermediate shaft160and the drive output shaft170by the group transmission115.

In the embodiment illustrated the gearsets175each comprise gearwheels with spur teeth, so that each gearset175lies in a rotational plane about a rotational axis, in order that the input shafts150,155, the first intermediate shaft160and if appropriate the second intermediate shaft165, and usually also the drive input shaft145and/or the drive output shaft170, are mounted in a rotatable manner. These rotational planes are also called wheel planes and in the embodiment shown are numbered from1to5from the left toward the right.

It is proposed to associate one gearset175, in the embodiment shown the gearset175of the wheel plane2, if necessary selectively with the first sub-transmission120or the second sub-transmission125. In this case the two sub-transmissions150,155comprise together only three gearsets175of the wheel planes1,2and3. If the shifting element B is closed, then the gearset175of wheel plane2is connected between the second input shaft155and the countershaft180, but instead, if the shifting element C is closed, then the gearset175of wheel plane2is connected between the first input shaft150and the countershaft180.

The shifting element E can on the one hand be used, when the first friction clutch135is closed, to engage a direct gear so that the drive input shaft145rotates at the same speed as the first intermediate shaft160. There should be no permanent connection of the direct gear to the countershaft180, so as to design the direct gear in a pre-selectable manner. If instead the second friction clutch140is closed, then instead of the third sub-transmission130, the first sub-transmission120can be used to transmit torque from the countershaft180to the first intermediate shaft160. For this, one of the shifting elements C or D and one of the shifting elements A or B is closed, but the shifting elements B and C may not be closed at the same time. Torque from the drive input shaft145is then transmitted via the second friction clutch140to the second intermediate shaft165, from there via a gearset175of one of the wheel planes1or2to the countershaft180, onward via one of the gearsets175of the wheel planes2or3to the first input shaft150and farther onward via the shifting element E to the first intermediate shaft160.

In the preferred embodiment illustrated, the third sub-transmission130has two gearsets175of the wheel planes4and5. One of the gearsets175, in wheel plane5in the embodiment shown, has three instead of two gearwheels that mesh with one another so that the gearset175has the same rotational direction on its input side and on its output side. In contrast to the other, reversing gearsets175, whose gearwheels rotate in pairs in different directions, in this case the gearset175of wheel plane5is of non-reversing design in order to produce one or more reversing gears of the double transmission110.

FIG. 2shows a shifting matrix200for the dual-clutch transmission100ofFIG. 1In columns from left to right a gear step Gg and shift conditions of the first friction clutch135(K1), the second friction clutch140(K2) and the shifting elements A to G and L and H are shown. On the right next to these, in further columns, an overall gear ratio i and a gear interval φ are entered. Examples of the overall gear ratios that can be obtained with appropriately sized gearsets175are indicated. The overall gear ratio i shows the ratio of the rotational speed of the drive input shaft145relative to the rotational speed of the drive output shaft170of the dual-clutch transmission100. The gear interval φ denotes the ratio of the overall gear ratios i of neighboring gear steps Gg.

In the shifting matrix200, in columns A-G, L, H a dot denotes respectively a closed connection of the respective switching element, otherwise the respective connection is open. In the case of the friction clutches135,140when a dot is shown it is assumed that a torque flow is enabled, usually by static friction without slip. A transition between the indicated gear ratios, in particular without interruption of traction, may require a slipping friction clutch135,140.

The gears Gg are numbered from 1 to 14 for forward gears and from R1to R3for reversing gears. In this case the gears3,10and R3are double-engaged. The overall gear ratios i of the double-engaged gears Gg are equal in pairs and differ on the one hand in which of the friction clutches135,140and on the other hand in which of the shifting elements B and C is closed. If the shifting element B is closed, then the number of the gear step is marked with b, while in contrast, if the shifting element C is closed the number of the gear step has c attached. A transition between the double-engaged gear steps3,10and R3from one variant to the respective other variant is possible at any time without traction force interruption. Thus, the gear steps3,10and R3can each be operated alternatively by way of the first friction clutch135or by way of the second friction clutch140.

The gear steps13and14have an overall gear ratio i smaller than 1 and are therefore also called overdrive gears (OD). Fourteen gear steps1-14are obtained since, by means of the double transmission110, seven different gears for each of the two conditions of the group transmission115can be engaged. Thus, the dual-clutch transmission100ofFIG. 1is also termed a 7×2 2OD configuration. In this, reversing gears are not counted. To carry out a shift without traction force interruption from one gear step Gg to another, both of the gears Gg are engaged by means of the shifting elements A to L, while only one of the friction clutches135,140is closed so that only one of the gears Gg is effective. Then the closed friction clutch135,140is opened and the open friction clutch135,140is closed. The one friction clutch135,140is opened and the other friction clutch135,140is dosed preferably simultaneously, so that a shift between the engaged gears Gg takes place without traction force interruption.

In the embodiment illustrated such a shift between gears7and8or between gears13and14is not directly possible. A shift between gears3and4or10and11can be made without traction force interruption by temporarily engaging gear5for a short time and partially closing the associated friction clutch135. Here, gear5is used as a so-termed supporting gear. Between gear3and gear4a powershift can be carried out as follows: (3a↔) 3b↔5↔4. In a corresponding manner a shift can be carried out between gears10and11: (10a↔) 10b↔12↔11. To enable the engagement of the supporting gear it is preferable to be able to make the countershaft180completely free from torque by means of the shifting elements A-G.

FIG. 3shows a shifting matrix200for the dual-clutch transmission100ofFIG. 1, which can be used instead of the shifting matrix200ofFIG. 2. By means of supporting gear shifts, the following transitions can be carried out without traction force interruption: (3a↔) 3b↔6↔4; 4↔6↔5; (10a↔) 10b↔13↔11 and 11↔13↔12. A transition between the double-engaged gear steps3,10and R3from one variant to the respective other variant is possible at any time without traction force interruption.

FIG. 4shows a further embodiment of a dual-clutch transmission100according toFIG. 1, wherein a powershiftable group transmission115is provided. An appropriate shifting matrix200for the configuration 12×1 2OD is shown inFIG. 5. The overall gear ratios i shown again relate to gearsets175whose dimensions are taken as examples.

Compared with the embodiment according toFIG. 1, the gearsets175of wheel planes4and5have been interchanged and the second intermediate shaft165can be connected in a shiftable manner by means of an additional shifting element J to the gearset175of wheel plane5of the third sub-transmission130. The second intermediate shaft165acts by way of a planetary carrier on the planetary gear190of the group transmission115and the planetary carrier acts directly on the drive output shaft170. By means of the shifting element L the ring gear195can be braked for example relative to a housing or some other static device, so that it is at rest, or by means of the shifting element H it can be connected to the drive output shaft170and to the planetary carrier of the planetary gear190.

FIG. 6shows a shifting matrix200for the dual-clutch transmission100in the embodiment ofFIG. 4, such that the shifting matrix200shown can be used as an alternative to that shown inFIG. 5. In this case the wheel planes1and3are interchanged and the dual-clutch transmission100is operated in a 13×1 2OD configuration. The shifting element J can be used in combination with a powershiftable group transmission115. A transition between the double-engaged gear steps3,9and R3from the variant concerned to the respective other variant can be carried out at any time without traction force interruption.

FIG. 7shows still another embodiment of the dual-clutch transmission100ofFIG. 1. In contrast to the embodiment shown inFIG. 4, in this case the shifting elements E, F, G and J are combined in a different way. Again, a powershiftable group transmission115is provided. Whereas in the embodiment ofFIG. 4the shifting elements E and J are single shifting elements and the shifting elements F and G are integrated with one another, in this case the shifting elements E and F respectively and G and J respectively are integrated. In that way the overall number of elements that have to be moved in order to engage a gear in the double transmission110is reduced.

The shifting matrix200for the embodiment shown corresponds to that ofFIG. 5, so that the configuration is 12×1 2OD, or to that ofFIG. 6with the configuration 13×1 1OD,

INDEXES