Abstract:
A shifting device for a dual clutch transmission as a speed-change transmission for motor vehicles, including two coaxially arranged input shafts, each being activatable via a clutch, an axially parallel output shaft and gearwheel sets that are mounted on the shafts and can be shifted by means of shifting clutches to allow a plurality of forward gears and a reverse gear, the gearwheel sets being subdivided into a sub-transmission having the one input shaft and a sub-transmission having the other input shaft. To achieve a wider gear-ratio spread in a structurally compact construction, a planetary gear train that can be shifted between two gear ratio stages is mounted upstream of the sub-transmission having the hollow input shaft.

Description:
CROSS-REFERENCES TO RELATED APPLICATIONS 
     This application is the U.S. National Stage of International Application No. PCT/EP2013/003187, filed Oct. 23, 2013, which designated the United States and has been published as International Publication No. WO 2014/067635 and which claims the priority of German Patent Application, Serial No. 10 2012 021 598.3, filed Oct. 30, 2012, pursuant to 35 U.S.C. 119(a)-(d). 
     BACKGROUND OF THE INVENTION 
     The present invention relates to a shifting device for a dual clutch transmission as change-speed transmission for motor vehicles. 
     Such dual clutch transmissions can be used as automated switching transmissions with good transmission efficiency, and due to the division into two subtransmissions and two separating clutches can be switched fast and without interruption of traction. In order to optimally adjust such change speed transmissions to the drive power of the drive aggregate or internal combustion engines, a wide transmission ratio spread is desired, which can be realized for example by providing a higher number of forward gears (in the absence of excessive transmission steps). 
     SUMMARY OF THE INVENTION 
     It is an object of the present invention to set forth a dual clutch transmission of the generic type, which enables wide transmission ratio spread and in an increased number of in particular forward gears, while at the same time being of compact construction and well manageable in terms of control. 
     The object is solved with the features of patent claim  1 . Advantageous embodiments and refinements of the invention are set forth in the dependent claims. 
     According to the invention it is proposed that a planetary transmission, which can be shifted between two transmission stages, is arranged upstream of the subtransmission A with the input hollow shaft. The planetary transmission results in a doubling of the number of gears in the subtransmission A, whereby in the subtransmission B in spite of the realizable wider transmission ratio spread the number of the gearwheel sets can be reduced compared to a conventional gearwheel set so that the overall length of the transmission is not increased and the additional costs can be kept low. The forward gears realized in the subtransmission A can be driven in two different transmission ratios per gearwheel set, wherein the shifting can also be accomplished relatively easily and automated. The gears in the subtransmission B are assigned so that the shifting of the gears without interruption of traction is retained. 
     As is known, the planetary transmission can have an input element, an output element and a transmission element, which can be fixed via a brake for shifting the lower (i.e., shorter) transmission stage, which leads to a higher output torque. Shifting into the 1:1 transmission stage can be accomplished via the clutch K 1 , wherein the clutch K 1  couples two elements of the planetary transmission (for example the input element and the transmission element) with each other, so that no transmission losses occur. 
     A separate clutch enables switching into a higher transmission ratio 1:1, wherein the clutch connects two elements of the planetary transmission with each other so that no transmission losses occur in the 1:1 transmission ratio. The brake and the clutch are preferably hydraulically actuatable friction based elements, for example of a multi-disc construction. 
     The brake and the clutch K 1  of the dual transmission are preferably hydraulically actuatable, friction-acting elements for example in the manner of a multi-disc construction. The dual transmission can be drivingly connected with a driving rive aggregate or an internal combustion engine. For this the dual clutch, which is constructed as multi-disc clutch, has a clutch housing, on which a force outputting shaft of the internal combustion engine is fixed in rotative fixed relationship. When the internal combustion engine is activated the clutch housing is therefore constantly rotating. The force flux is therefore transmitted via the clutch housing either to the first input shaft or to the second input shaft. 
     The planetary transmission can have a first sun gear as input element, a second sun gear as output element and a web as transmission element, on which multi-stage planet gears are supported, which are in engagement with the two sun gears. Hereby the input element (for example the driving sun gear) is constantly drivingly connected with the housing of the dual clutch transmission, while the transmission element or the web can be activated via the clutch K 1  of the dual clutch transmission. This has the advantage that a starting of the motor vehicle for example in the first forward gear or in the reverse gear and in the lower transmission stage of the planetary gear transmission can be controlled via the brake, which is connected to the web, wherein optionally the corresponding clutch K 1  can be configured with a lower clutch torque. This for example makes it possible to omit the separate clutch (i.e. K 3  in  FIG. 8 ) for shifting the planetary transmission. 
     In a further advantageous embodiment of the invention the planetary transmission can be constructed as minus transmission (standard transmission ratio for example i o =−3), which allows further improving the proportion of clutch power or the transmission efficiency in the speed-reduction stage. 
     In a further preferred embodiment of the invention, the change-speed transmission can have eight forward gears, of which the first and the second forward gear, the fourth and the fifth forward gear and the seventh and eighth forward gear are assigned to the subtransmission A with the upstream arranged planetary transmission, and the third and sixth gear are assigned to the subtransmission B. The eight forward gears can thus be realized by five gearwheel sets, wherein the number of forward gears of the three gearwheel sets of the subtransmission A is doubled via the shiftable planetary transmission. 
     In addition, in the case of a change-speed transmission with more than eight forward gears the ninth forward gear can be assigned to the subtransmission B via a further gearwheel set and optionally the forward gears  10  and  11  can be assigned to the subtransmsision A also via a further gearwheel set. 
     Finally, a gearwheel set with an intermediate gear wheel can be assigned to the subtransmission A with the upstream arranged planetary gear transmission for providing two reverse gears; as an alternative the gearwheel set with intermediate gear wheel can be arranged in the subtransmission B for forming only one reverse gear. The latter has the advantage that the starting gears (forward-reverse) are not controlled via the same starting element and with this an uneven clutch wear can be avoided. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWING 
       In the following, multiple exemplary embodiments of the invention are explained in more detail by way of the included schematic drawings. It is shown in: 
         FIG. 1  a block diagram of a twin clutch transmission for an all-wheel drive motor vehicle, with two subtransmissions A and B, wherein a planetary transmission is arranged upstream of the subtransmission A; 
         FIG. 2  a further planetary transmission alternative to the one of  FIG. 1 , configured as minus transmission; 
         FIG. 3  a shifting matrix illustrating the shifting sequences for a dual clutch transmission having eight forward gears; 
         FIGS. 4 to 7  respectively partial views of further exemplary embodiments. 
         FIG. 8  a comparative example not according to the invention; and 
         FIG. 9  a further dual clutch transmission with a total of ten forward gears and a reverse gear R and four shifting groups. 
     
    
    
     DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS 
       FIG. 1  shows very, schematically a dual clutch transmission  12  as change-speed transmission for motor vehicles, with two coaxial transmission input shafts  14 ,  16 , which are drivingly connectable with, a driving drive aggregate or an internal combustion engine via two separating clutches K 1 , K 2 . For this the dual clutch transmission, which is for example configured as multi-disc clutch, has a clutch housing  44 , on which an output shaft  11  of the internal combustion engine is fixed for co-rotation with the housing. When the internal combustion engine is active, the clutch housing  44  therefore permanently rotates. The force flux is therefore conducted via the clutch housing  44  either to the first input shaft  14  or to the second input shaft  16 . The input shaft  14  is constructed as a hollow shaft. 
     An output shaft  18  is provided axially parallel to the input shafts  14 ,  16 , and in the exemplary embodiment is formed by a first output shaft  18   a  and a second coaxial output shaft  18   b , which is configured as a hollow shaft. 
     Via an output stage  25  consisting of spur gears  24 ,  26 , the first output shaft  18   a  outputs to a front axle differential  22  (only indicated schematically) attached to the transmission housing  20 , while the second output shaft  18   b  outputs to an interaxle differential  30 , whose output elements or bevel gears  32 ,  34  are connected on one side with the output shaft  36  for driving the front axle differential  22  and on the other end with an output shaft  36  for driving a rear axle differential (not shown) of the motor vehicle. 
     The shown layout of the output shaft  18  with the interaxle differential  30  makes it possible for individual forward gears to output directly to the front axle differential, while other forward gears and the reverse gear form an all-wheel drive via the interaxle differential  30 . 
     When outputting only to an axle differential  22  (front wheel drive) or via the output shaft  36  to a rear axle differential (rear wheel drive), the output shaft  18  can also be configured one-piece and without interaxle differential  30 . 
     The mentioned shafts  14 ,  16 ,  18  etc. are rotatably supported via only schematically indicated rolling bearings in the transmission housing  20 . 
     The twin clutch transmission  12  is divided into a subtransmission A and a subtransmission B, wherein the input hollow shaft  14  only extends within the subtransmission A, while the input shaft  16  is guided through the input shaft  14  into the subtransmission B up to its end wall  20   a.    
     In the subtransmission A three forward-gears gearwheel sets I, II, V and a reverse-gear gearwheel set R with an integrated reverse gearwheel are arranged, which are composed in a known manner of fixed gearwheels and idler gears, wherein the idler gears can be switched via synchronizing clutches (generally designated  38 ). The gearwheel sets I, II, V, R form the forward gears  1 / 2 ,  4 / 5 ,  7 / 8  and two reverse gears R  1 / 2  (as will be explained below). 
     In the subtransmission B two forward-gear gearwheel sets II and IV are provided, which are also formed by fixed gearwheels and idler gears, which can be switched via a synchronizing clutch  38 , and which in the corresponding transmission configuration form the forward gears  3  and  6 . 
     A planetary transmission  40  is arranged upstream of the subtransmission A adjacent the dual clutch K 1 , K 2 , which planetary transmission  40  is configured coaxial to the output hollow shaft  14  and which can be shifted between two transmission stages or between a lower/higher transmission stage and a 1:1 transmission stage. 
     The input element of the planetary transmission  40  is formed by a first sun gear  42 , which is connected to the housing  44  of the dual clutch K 1 , K 2  and is thus constantly driven. 
     The web  46  as transmission element of the planetary transmission  40  carries multiple rotatably supported multi-stage planet gears  48 , of which the one, greater gearwheel meshes with the sun gear  42  and can be drivingly connected with the clutch K 1 . 
     The smaller gearwheel of the multi-stage planet gears  48  is connected with a second sun gear  50  as output element of the planetary transmission  40  with the input hollow shaft  14  of the subtransmission A. 
     Further the hollow shaft  52  carrying the web  46  is connected to a brake B which is constructed similar to a multi-disc clutch, and via which the web  46  can be fixed by braking through open or closed-loop control. 
     In the following description, as an example, a lower transmission stage is shifted at disengaged clutch K 1  and at actuated brake B, while at engaged clutch K 1  and at non-actuated brake B the transmission stage is shifted which is higher than the lower transmission stage, so that the invention can be described as follows: 
     When the clutch K 1  is disengaged and the brake B is actuated, the sun gear  42  drives the sun gear  50  via the multi-stage planet gears  48 , and the sun gear  50  drives the input shaft  14  in the lower transmission stage; at correspondingly shifted gearwheel sets of the subtransmission A this corresponds to the forward gears  1  (here the starting of the motor vehicle is controlled by the brake B via the exerted braking torque),  4  and  7 , and when the gearwheel set R is shifted to the reverse gear R 1 . When the brake B is disengaged the drive power is interrupted. 
     For shifting the planetary transmission  40  to the 1:1 transmission stage, the clutch K 1  is engaged, wherein the sun gear  42  and the web  46  are now coupled together and the planetary transmission  40  is blocked in itself. The drive torque is correspondingly divided via the clutch K 1  and the housing  44 . In the 1:1 transmission stage of the planetary transmission  40  the forward gears  2 ,  5 ,  8  and the reverse gear R 2  can be shifted. 
     For shifting the forward gears  3  and  6  of the subtransmission B the gearwheel sets II or IV are shifted in a known manner via the synchronizing clutch  38  and are activated via the separating clutch K 2 . Hereby the clutch K 1  and the brake B of the planetary transmission  40  are disengaged. 
     The shifting matrix according to  FIG. 3  shows the shifting sequence for example when passing through the forward gears  1  to  8  and the two possible reverse gears R 1  and R 2 . 
     When the motor vehicle is to be started in the 1 st  gear, the drive torque is introduced via the brake B after coupling the gearwheel set I by means of the synchronizing clutch  38 , wherein the planetary transmission  40  as described above is in the lower transmission stage (this is respectively indicated for example with IP; P stands for planetary transmission active). 
     Subsequently the 2 nd  gear is shifted by disengaging the brake B and engaging the clutch K 1 . This can be controlled without any interruption of traction. In the shifting processes described so far the 3 rd  gear can already be engaged beforehand in the subtransmission B via the synchronizing clutch  38 . 
     When shifting into the 3 rd  gear the clutch K 1  is disengaged and without interruption of traction the clutch K 2  is simultaneously engaged. 
     When driving tin the 3 rd  gear the gearwheel set III for the 4 th  and 5 th  gear can be shifted subtransmission A via the synchronizing clutch  38 . The gears  4  and  5  are then activated in analogy to the gears  1  and  2  as described above via the brake B (IIIP) and subsequently via the clutch K 1  (III). 
     The same applies to the 6 th  gear in the subtransmission B and the further gears  7  and  8  via the subtransmission A. 
     For shifting the reverse gears via the reverse-gear gearwheel set R, the gearwheel set R is shifted via the synchronizing clutch  38  and then either the brake B is actuated (R 1  or RP) or at disengaged brake B the clutch K 1  is engaged (R 2  or R). 
     The dual clutch transmission can have further gears, for example forward gears  9  to  11 . For this an additional gearwheel set VII would have to be provided according in the subtransmission A according to  FIG. 1 , which preferably can be arranged instead of the shown reverse gear gearwheel set RW and which can form the gears  10  and  11  in the corresponding transmission ratio configuration. 
     Further an additional gearwheel set VI for a 9 th  gear and the reverse gear gearwheel set RW can be arranged in eh subtransmission B, which can be shifted via a common synchronizing clutch  38 . 
       FIG. 2  shows a further alternative configuration of the upstream arranged planetary gear; functionally same parts are again provided with the same reference signs. 
     The planetary transmission according to  FIG. 2  is configured as minus transmission  80 , i.e., it has a negative standard transmission ratio (in the exemplary embodiment of i 0 =−3). This means the ratio between the angular velocities or the rotational speeds of the center gear shafts in a planetary transmission when the web stands still or is imagined to stand still. 
     The minus transmission  80  has as input element a first sun gear  82  which is connected to the housing  44  of the dual clutches, K 2  and is therefore constantly driven. 
     Further a web  84  is provided as output element which is drivingly connected with the hollow shaft  14  of the subtransmission A. 
     A second sun gear  86  forms the transmission element of the minus transmission  80  which is connected to the clutch K 1  via the hollow shaft  52 . The clutch. K 2  is connected to the input shaft  16  of the subtransmission B as described above. 
     The web  84  carries first multi-stage planet gears  88  which are rotatably supported, and of which the smaller gearwheel meshes with the sun gear  82 , while the greater gearwheel is in engagement with radially inwardly situated second planetary gears  90  and these with the second sun gear  86 . 
     The housing-fixed brake B is in turn connected to the hollow shaft  52  of the second sun gear  86 . 
     The shifting function of the minus transmission  80  is, aside from the different torque flux (indicated by arrows), equal to the planetary transmission  40 . In the lower transmission stage the brake B is actuated or the sun gear  86  is fixed. The sun gear  82  drives the web  84  via the multi-stage planet gears  88  and the planet gears  90 , and the web drives the input hollow shaft  14  of the subtransmission B. 
     When the brake B and the clutch K 1  are disengaged, there is no transmission of drive power. 
     For shifting the minus transmission  80  to the 1:1 transmission ratio the clutch K 1  is engaged, whereby the two sun gears  82 ,  86  are coupled with each other via the web  84  or the minus transmission  80  is blocked. 
     The shifting sequence of the forward gears  1  to  8  or optionally up to  11  and the revere gears R corresponds to the one of  FIG. 1  or to the shifting matrix according to  FIG. 3 . 
     The following exemplary embodiments of  FIGS. 4 to 7  also respectively relate to a dual clutch transmission, of which however only the dual clutch and the planetary transmission  40  is shown. The function and the construction of the dual clutch transmission of  FIGS. 34 to 7  is comparable to the dual clutch transmission shown in  FIG. 1 , so that a detailed description is not required. 
     As in  FIG. 1 , in  FIG. 4  a planetary transmission  40  is also arranged upstream of the input shaft  14  of the not shown subtransmission A. In  FIG. 4  the input element of the planetary transmission  40  is also the sun gear  42 , which is connected to the housing  44  of the dual clutch K 1 , K 2  and thus constantly rotates. In contrast to  FIG. 1  the planetary transmission  40  in  FIG. 4  has a dual planetary gearwheel set of which the radially outer planet gears  49  are carried by the web  46 . The web  46  in  FIG. 4  is the transmission element. The output element in  FIG. 4  is formed by the outer hollow shaft  51 , which is connected to the input shaft  14  in rotative fixed relationship with the input shaft  14 . 
       FIGS. 5 to 7  respectively show further variations of the dual clutch transmission of  FIG. 1 . Thus in  FIG. 5  the clutch K 1  is arranged downstream of the clutch K 2  and the brake B in the force flux direction via a free space  51 . Constructively advantageously, the spur gear  24  of the output stage  25  protrudes into the free space  51 . In addition, the clutch K 2  is arranged in the force flux direction upstream of the brake B. 
     In  FIG. 6  the brake B is arranged downstream to the clutch K 2  in the force flux direction. also in this case the spur gear  24  of the output stage protrudes into the free space  51 . The brake B is arranged in force flux direction upstream of the clutch K 1 . 
     In  FIG. 7  the clutch K 2  is arranged in the force flux direction upstream to the brake B and the clutch K 1  via the free space  51 , in which the output stage  25  is arranged. The clutch K 1  is arranged in the force flux direction upstream of the brake B. 
     In  FIG. 8  an alternative embodiment to that of  FIG. 1  of the upstream arranged planetary transmission is shown, which however is not included in the invention; functionally same parts are provided with the same reference signs. 
     The planetary transmission  60 , which is attached to the dual clutch K 1 , K 2 , has a ring gear  62  as input element, which is drivingly connected to the clutch K 1  via the hollow shaft  52 , a web  64  with rotatably supported planet gears  66  as output element and a sun gear  68  as transmission element. 
     The web  64  is directly drivingly connected with the input hollow shaft  14  of the subtransmission A, while the sun gear  68  is connected to the housing-fixed brake B via a further hollow shaft  70 . 
     Further a separate multi-disc clutch K 3  is arranged between the ring gear  62  and the hollow shaft  70  or the sun gear  68 , which separate clutch under hydraulic action connects the ring gear  62  with the sun gear  68  and thus forms the 1:1 transmission ratio of the planetary transmission  60 . The clutch K 1  then drives the hollow shaft  14  of the subtransmission (A) via the blocked planetary transmission  60 . 
     The planetary transmission can be shifted into the lower gears in that the clutch K 3  is disengaged and the brake B is actuated. Then the engaged clutch K 1  drives the input hollow shaft  14  via the ring gear  62 , the planet gears  66  and the web  64 , while the sun gear  68  as support element is braked fixed. 
     In contrast to the exemplary embodiments of  FIGS. 1 to 7  the input element (here the outer ring gear  62 ) in  FIG. 8  is not connected in rotative fixed relationship on the housing  44  of the dual clutch. For shifting between the transmission stages therefore not only the brake B has to be provided but in addition also the separate clutch K 3 . Correspondingly, in the construction with the separate clutch K 3  shown in  FIG. 8  an additional shifting element is required. 
       FIG. 9  shows a further dual clutch transmission which has a total of ten forward gears and a reverse gear R and for shifting groups  38 . In the subtransmission A four forward gear gearwheel sets II, IV, VI and VII are arranged, which are formed in a known manner by fixed gears and idler gears. The idler gears can be shifted via synchronizing clutches  38 . The gearwheel sets II, IV, VI, VII form the forward gears  2 / 3 ,  5 / 6 ,  8 / 9  and  10 . In addition the gearwheel set II forms a reverse gear R, as will be explained below. In the subtransmission B the forward gear gearwheel sets I, II, V are provided. These are also formed by fixed gears and idler gears which can be shifted via synchronizing clutches  38  and form the forward gears  1 ,  4  and  7 . 
     In  FIG. 9  the subtransmissions A and B have a common synchronizing clutch  38 . The common synchronizing clutch  38  is mounted on the output hollow shaft  18   b  and can be shifted with the gearwheel set I or with the gearwheel set IV. 
     The arrangement of the dual clutches K 1  and K 2  shown in  FIG. 9  essentially corresponds to the arrangement shown in  FIG. 5 . In contrast to  FIG. 5 , the planetary transmission  40  in  FIG. 9  further has an outer ring gear  71 , which meshes with the smaller gearwheels of the multi-stage planet gears  48 . The outer ring gear  71  is connected with an intermediate hollow shaft  73  which is supported coaxial to the input hollow shaft  14 . 
     Via a synchronizing clutch  38 , the intermediate hollow shaft  73  is connectable with the gearwheel set VII (for shifting the 10 th  forward gear) or with the gearwheel set II (for shifting the reverse gear R). In addition the gearwheel set II is connectable with the input hollow shaft  14  via a further synchronizing clutch  38 . 
     For shifting the reverse gear R, the gearwheel set II is shifted via the synchronizing clutch  38  arranged on the intermediate hollow shaft  73 . In addition the clutch K 1  is disengaged and the brake B is actuated. The force flux is therefore conducted to the gearwheel set II via the clutch housing  44 , the sun gear  42 , the multi-stage planet gears  48  and the outer ring gear  73  associated with reversing the direction of rotation and at shifted synchronizing clutches  38 . 
     For shifting the 10 th  forward gear the gearwheel set VII is shifted via the synchronizing clutch  38  arranged on the intermediate hollow shaft  73 . In addition the brake B is disengaged and the clutch K 1  is engaged. As a result the planetary transmission  40  is blocked, i.e., blocked in itself. The force flux is therefore conducted to the outer ring gear  71  via the clutch housing  44 , the blocked planetary transmission  40  without reversal of the rotation direction and further to the shifted gearwheel set VII.