Abstract:
A power-dividing gear train assembly for motor vehicles includes a differential that outputs torque to two output shafts. The output torque applied to the output shafts is changeable by using a superimposing gear train, which comprises a plurality of epicyclic gear trains and is drivingly connected indirectly or directly with the output shafts. A drive motor is drivingly coupled to the superimposing gear train and the transmission ratio of the superimposing gear train is set such that an output shaft of the drive motor is still when the output shafts are rotating synchronously. To achieve a more precise and quicker-reacting redistribution of the drive torques, at least one epicyclic gear train, which reduces torque and cooperates with the differential, is connected upstream of the superimposing gear train.

Description:
CROSS-REFERENCE 
     This application is the U.S. national stage filing of International Application No. PCT/EP2008/003148 filed Apr. 18, 2008, which claims priority to German patent application no. 10 2007 017 185.6 filed Apr. 12, 2007. 
     TECHNICAL FIELD 
     The invention relates to a power-dividing gear train assembly for motor vehicles. 
     RELATED ART 
     Such a power-dividing gear train assembly is described in DE 10 2005 021 023 A1. There, the differential, in this case an axle differential, drives two driven wheels of the motor vehicle via two output shafts. Further, a superimposing gear train formed by two drivingly-coupled planetary gear trains is provided with a drive motor, the output torques to the output shafts being redistributable via the differential by means of said superimposing gear train. The transmission ratios of the superimposing gear train are advantageously set such that the driving shaft of the drive motor is still when the output shafts are rotating synchronously (=straight-ahead driving of the vehicle with no rotational speed difference between the wheels). As this driving condition is prevalent with motor vehicles, an advantageous design is achieved with regard to efficiency and wear of drive parts; moreover, the drive motor can be actuated precisely and quickly from the still state in both rotational directions for redistribution of driving torques. 
     SUMMARY 
     It is an object of the invention to improve a power-dividing gear train assembly of the conventional type such that, for an unchanged high transmission efficiency, an even more precise and faster-reacting redistribution of output torques is achieved with a structurally advantageous design of the superimposing gear train and the drive motor. 
     According to the invention, it is proposed that at least one epicyclic gear train, which reduces torque and interacts with the differential, is connected upstream of the superimposing gear train. When the output shafts are rotating synchronously, the epicyclic gear train rotates as a block, i.e. only slightly diminishing the efficiency of the gear train assembly. The epicyclic gear train succeeds in transmitting, to a large extent, the redirected torque flow T directly via the differential, so that the load on the superimposing gear train is reduced. Further, the dimensions of the superimposing gear train and the drive motor, which may be a hydraulic motor or preferably an electric motor, can be decreased, as smaller control torques are required for the redistribution. 
     In principle, epicyclic gear trains can be connected to the three shafts of the differential, as proposed, in order to reduce the torques required for redistribution. A plurality of epicyclic gear trains can also be connected to one differential. Instead of, or in addition to, an axle differential, a longitudinal differential may be designed in an appropriate manner as the power-dividing gear train assembly in a four-wheel drive motor vehicle. The differential may be designed as a bevel gear differential, if necessary. 
     However, the differential is preferably designed in a known manner as a positive-ratio epicyclic gear train, wherein the outer gear is driven and the output shafts are drivingly connected to the planet carrier and the sun gear, respectively, wherein it is of particular structural advantage that at least one planetary gear set of the differential is connected to the superimposing gear train via at least one additional coaxial central gear. By accelerating or decelerating the planetary gear set, a redistribution of the output torque of the differential can be controlled in an advantageous manner. When the planetary gears are accelerated or decelerated by the drive motor, the resulting torque ratio deviates from the transmission ratios of the differential, whereby a torque-vectoring functionality can be produced. 
     Here, the epicyclic gear train with the differential can be designed as a reduction gear set having only one central gear meshing with a planetary gear set (=4-shaft gear train). The two other elements of the epicyclic gear train are therefore formed by the already-existing output elements of the differential, i.e. the planet carrier and the sun gear. Accordingly, redistributed output torques pass directly via the differential. 
     In another aspect of the present teachings, couplings of the central gear of the epicyclic gear train with the differential are described that are advantageous in view of structure and transmission ratio. 
     In an advantageous further development of the invention, two upstream-connected epicyclic gear trains can have two central gears interacting with the differential, which central gears are drivingly connected to the superimposing gear train, wherein the central gears drive the radial-inner or the radial-outer planetary gears as a positive-ratio gear train and as a negative-ratio gear train (=5-shaft gear train). 
     Alternatively, the central gear may be combined with an epicyclic gear train, which is composed of a sun gear having planetary gears supported on carriers of the sun gear of the differential, which sun gear is connected to the superimposing gear train, and an outer gear forming a structural unit with the central gear. Hence, for still higher transmission ratios, the epicyclic gear train is integrated into the differential as a negative-ratio gear train in a structurally advantageous manner. 
     Here, it may be particularly advantageous with regard to the structure and manufacturing, if the central gear is designed as an internal-toothed and external-toothed ring gear that meshes in an overhung manner with the outer planet gears of the differential and with the planet gears of the integrated epicyclic gear train. 
     Alternatively, the central gear can be designed as an internal-toothed ring gear that meshes with the inner planet gears of the differential and with the planet gears of the integrated epicyclic gear train. 
     Further, the epicyclic gear train can be produced from a combination of two elementary gear trains, two of its three shafts being connected with two of the three shafts of the differential. Here, the two elementary gear trains may be designed as a reduced gear set having one negative-ratio gear train and one positive-ratio gear train or as a negative-ratio gear train. In this way, it is achieved that especially high transmission ratios (=especially low torque flows in the superimposing gear train and the drive motor) can be provided. 
     In a further embodiment of the invention, the superimposing gear train may be formed of two interconnected planetary gear trains, which are drivingly connected to the drive motor via an element and are supported in a fixed manner relative to the housing via a further element, and which effect the torque distribution via the epicyclic gear train and the differential. 
     In particular, the epicyclic gear train and the superimposing gear train can be designed in a manner, which is especially advantageous with regard to the structure and manufacturing, as one structural unit with the differential, to which structural unit the drive motor is attached. In this way, drive couplings of the sub-gear trains can be implemented especially easily. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWING 
       Several exemplary embodiments of the invention are described below with further details and advantages. The schematic drawings show in: 
         FIG. 1  a power-dividing gear train assembly for motor vehicles having a possible arrangement of its superimposing gear train with an epicyclic gear train connected upstream of the drive motor, utilizing Wolf symbolic representations; 
         FIG. 2  a power-dividing gear train assembly for motor vehicles having an axle differential designed as a positive-ratio gear train, an epicyclic gear train with an external-toothed central gear depicted as a reduced coupled gear set and a superimposing gear train having an electric motor; 
         FIG. 3  the power-dividing gear train assembly according to  FIG. 2 , however with an internal-toothed annulus gear as a central gear; 
         FIG. 4  the power-dividing gear train assembly according to  FIGS. 2 and 3  and the superimposing gear train as Wolf symbolic representations; 
         FIG. 5  the power-dividing gear train assembly according to  FIG. 2 , wherein the central gear meshes with additional planet gears that are integrally formed with a planetary gear set of the differential; 
         FIG. 6  the power-dividing gear train assembly according to  FIG. 2 , wherein the external-toothed central gear meshes with planet gears that are separately supported on the planet carrier of the differential; 
         FIG. 7  the power-dividing gear train assembly according to  FIGS. 5 and 6  as Wolf symbolic representations; 
         FIG. 8  a power-dividing gear train assembly having a differential and an epicyclic gear train with two central gears that are drivingly connected with the superimposing gear train; 
         FIG. 9  the resulting Wolf symbolic representations of the power-dividing gear train assembly according to  FIG. 8 ; 
         FIG. 10  a power-dividing gear train assembly having a differential and an epicyclic gear train, wherein a further epicyclic gear train designed as a negative-ratio gear train is integrated into the central gear designed as a ring gear, and which has a downstream superimposing gear train; 
         FIG. 11  the power-dividing gear train assembly according to  FIG. 10 , wherein the central gear is formed, however, as an internal-toothed annulus gear, which meshes with the inner planetary gears of the differential and the planet gears of the integrated epicyclic gear train; 
         FIG. 12  the power-dividing gear train assembly according to  FIGS. 10 and 11  as Wolf symbolic representations; 
         FIG. 13  a further power-dividing gear train assembly having a differential, wherein the planetary gear train is composed of the combination of two elementary gear train assemblies, two of its three shafts being respectively connected with two of the three shafts of the differential; 
         FIG. 14  the power-dividing gear train assembly according to  FIGS. 13 and 15  as Wolf symbolic representations; and 
         FIG. 15  the power-dividing gear train assembly according to  FIG. 13 , wherein the epicyclic gear train is again produced from a combination of two elementary gear trains. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     In  FIG. 1 , a power-dividing gear train assembly for motor vehicles is illustrated utilizing Wolf symbolic representations, which power-dividing gear train assembly can be used as a longitudinal differential (in four-wheel-drive motor vehicles) and/or as a transverse differential and an axle differential, respectively. 
     In the exemplary embodiment, the power-dividing gear train assembly has an axle differential U, which is designed as a still-to-be-described epicyclic gear train or rather a planetary gear train, whose input shaft  12  distributes the driving torque generated by a drive source of the motor vehicle to the two output shafts  14 ,  16  and from these via universal joint shafts  18 ,  20  to the driven wheels  22 ,  24  of the motor vehicle. 
     A torque-reducing epicyclic gear train U z  is drivingly connected between the two output shafts  14 ,  16  and is connected to an attached drive motor via at least one input shaft  28  with a downstream, still-to-be-described superimposing gear train  26 . The superimposing gear train  26  is coupled back to one of the output shafts  14 ,  16  for redistribution of the output torque. Due to the epicyclic gear train U z , the to-be-redistributed output torque T 1 , T 2  is reduced to T St , wherein T St &lt;&lt;T 1 , T 2 . 
     Instead of the epicyclic gear train U z  between the output shafts  14 ,  16 , shown by solid lines in  FIG. 1 , the epicyclic gear train U z  (according to the broken lines) having the downstream-connected superimposing gear train  26  can also be drivingly arranged between the input shaft  12  and the output shafts  14  or  16 . Moreover, multiple epicyclic gear trains U z  can also be used simultaneously, if required. 
     It should be noted that the terms “ . . . shafts” and “drivingly connected” also are understood to include the epicyclic gear train elements of the differential U connected with said shafts, the additional epicyclic gear trains U z  and the superimposing gear train  26 , as is also readily derivable from the following figures. 
     In the structural embodiment,  FIG. 2  shows the axle differential U as a known positive-ratio planetary gear train having a driven outer gear  32  (cf. input shaft  12  of  FIG. 1 ), an output planet carrier  34  (cf. output shaft  14  of  FIG. 1 ) having radial-outer planet gears  36  and radial-inner planet gears  38  and a central sun gear  40  (cf. output shaft  16  in  FIG. 1 ). The planet gears  36  mesh with the outer gear  32  and the planet gears  38 ; additionally, the planet gears  38  are in engagement with the sun gear  40 . 
     Moreover, an external-toothed central gear  42  is provided, which is connected with the superimposing gear train  26  that is integrated into the power-dividing gear train assembly as presented below, and which is in engagement with the planet gears  36  of the differential U. 
     Consequently, the central gear  42 , together with the already-existing elements of the planet carrier  34  and sun gear  40  of the differential U, forms a reduced coupled gear set or the epicyclic gear train assembly U z , which is drivingly connected with the superimposing gear train  26 . Due to the integration of the epicyclic gear train U z , the power flow can be changed already in the differential U, without having to use additional power-guiding components. This results in a substantial reduction of costs and weight and increases the operational reliability of the power-dividing gear train assembly. 
     According to  FIG. 2 , the superimposing gear train  26  is composed of two drivingly-coupled planetary gear trains  44 ,  46 , which interact with an attached drive motor or rather an electric motor  48 . The transmission ratio of the planetary gear trains  44 ,  46  or the superimposing gear train  26  is set so that the drive shaft  50  of the electric motor  48  does not rotate during synchronous rotation of the output shafts  14 ,  16 . 
     Here, the central gear  42  of the epicyclic gear train U z  also forms the planet carrier  52  of the planetary gear train  46  having the planet gears  54 , while the planet gears  56  of the planetary gear train  44  are rotatably supported on a planet carrier  58  that is drivingly connected with the output shaft  16 . 
     The planet gears  54 ,  56  of the planetary gear trains  44 ,  46  mesh with a common outer gear  60  and each one meshes with a respective sun gear  62 ,  64 . Here, the sun gear  64  of the planetary gear train  46  is fixedly supported relative to the housing at  66 , whereas the sun gear  62  is connected with the drive shaft  50  of the electric motor  48 . 
       FIG. 3  shows a power-dividing gear train assembly corresponding to  FIG. 2 , which is described only as far as it differs from  FIG. 2 . Here, the central gear  42 ′ is formed as an inner-toothed annulus gear, which does not mesh with the radial-outer planet gears  36 , but rather with the inner planet gears  38  of the planet carrier  34 . The function and the connection to the superimposing gear train  26  correspond to  FIG. 2 . 
       FIG. 4  shows the power-dividing gear train assemblies U and U z  having superimposing gear train  26 , which correspond to  FIGS. 2 and 3 , as Wolf symbolic representations. The summation shaft (shown as a double line) of the reduced negative-ratio gear train U z  is identical to the planet carrier  34  of the differential U. As is further apparent, the output shaft  14  and the output shaft  16  of the differential U are identical to the delineated output shafts  14 ,  16  of the reduced epicyclic gear train U z . 
       FIGS. 5 and 6  show further power-dividing gear train assemblies, which are described only as far as they substantially differ from the above-presented power-dividing gear train assemblies. Parts having the same functions are provided with the same reference numerals. 
     According to  FIG. 5 , the external-toothed central gear  42 ″ meshes with planet gears  38   a  that are coaxially arranged on the inner planet gears  38  of the differential U or rather the planet carrier  34 , wherein the pitch circle diameter of the planet gears  38   a  is different from the adjacent planet gears  38 . This enables a further torque reduction via the epicyclic gear train U z  by using structurally simple means. 
     According to  FIG. 6 , the central gear  42 ′″ is in engagement with further planet gears  68 , which also mesh with the outer planet gears  36  of the planet carrier  34 . As is apparent, the planet gears  68  are also rotatably supported on the planet carrier  34 ′. The interleaved connection of the further planet gears  68  enables a still further design of the torque reduction of the additional epicyclic gear train U z . 
     As is apparent from the Wolf symbolic representations according to  FIG. 7 , the summation shaft (double line) of the reduced positive-ratio gear train assembly U z  shown herein is identical with the sun gear  40 . 
       FIG. 8  shows a further alternative power-dividing gear train assembly, which is described only as far as it substantially differs from the preceding embodiments. Same parts are provided with the same reference numerals. 
     According to  FIG. 8 , two additional epicyclic gear trains U z  and U z2  are provided, which have two central gears  42 ,  42 ′″ interacting with the differential U. The central gears  42 ,  42 ′″ are drivingly connected with the superimposing gear train  26  (not illustrated). Here, the one external-toothed central gear  42 ′″ meshes with planet gears  68  that are separately supported on the single planet carrier  34 ″ of the differential U in order to form a positive-ratio gear train, which planet gears  68  also mesh with the radial-outer planet gears  36  of the differential U. The other external-toothed central gear  42  acts directly on the outer planet gears  36  as a negative-ratio gear train. 
     Due to the integration of the two epicyclic gear train U z  and U z2  into the differential U, the transmission ratio at the electric motor  48  is further increased or rather the control torque is further reduced. If opposing torques are applied to the two central gears  42 ,  42 ′″, the inner planet gears  38  will be driven “doubly” in one direction and, therefore, the torque distribution of the differential U is accordingly affected. 
     Instead of the depicted embodiment according to  FIG. 8 , the drive coupling with the two central gears  42  can also be designed with the radial-inner planet gears  38  in connection with a set of reverse planet gears  68 . 
       FIG. 9  shows the Wolf symbolic representation for the power-dividing gear train assembly according to  FIG. 8  in the described version, or in versions derived therefrom. 
       FIGS. 10 and 11  show further alternative power-dividing gear train assemblies, which are again described only as far as they substantially differ from the preceding embodiments. Parts having the same functions are provided with the same reference numerals. 
     According to  FIG. 10 , the central gear specified with the preceding reference numeral  42  is designed as an internal- and external-toothed ring gear  42 ″″, which meshes in an overhung manner with the outer planet gears  36  of the differential U and, functioning as an internal-toothed outer gear, with the planet gears  70  of a further, integrated epicyclic gear train U z2 . The planet gears  70  are rotatably supported on carriers  72  that are attached to the sun gear  40 ′. The sun gear  74  of the second epicyclic gear train U z2  is also the planet carrier  58 ′ of the superimposing gear train  26 . Further, the ring gear  42 ″″ of the epicyclic gear train U z  forms the additional planet carrier  52 ′ of the superimposing gear train  26  which, apart from that, is designed according to  FIG. 2  and is not described again. 
     Due to the use of the second complete epicyclic gear train U z2 , an even higher transmission ratio for the electric motor  48  results. Now there is a second running carrier  72 , but this is identical with the sun gear  40 ′ of the differential U. As a result of this, a very simple construction is possible. The efficiency is very good with regard to the realizable transmission ratio. The concept has a very narrow axial structure and consists of relatively few components. 
     Unlike in  FIG. 10 , the central gear in  FIG. 11  is designed as an internal-toothed annulus gear  42 ′″″, which meshes with the radial-inner, broadened planet gears  38  of the differential U and with the planet gears  70  of the integrated epicyclic gear train assembly U z2 . 
       FIG. 12 , in turn, shows the Wolf symbols according to  FIGS. 10 and 11  for a better overview of the gear train couplings. 
     Power-dividing gear train assemblies are described with  FIGS. 13 to 15 , in which the epicyclic gear train U z  is produced from the combination of two elementary gear trains U 1  and U 2 , two of its three shafts being connected with two of the three shafts of the differential U. 
     According to  FIG. 13 , the two elementary gear trains are designed as a reduced gear set with two negative-ratio gear trains U 1  and U 2 . 
     Accordingly, the sun gear  76  of the negative-ratio gear train U 1  is identical with that of the negative-ratio gear train U 2 . The planet gears  78  mesh with the sun gear  76  and the outer gear  80 . This is connected to the planet carrier  34  of the differential U. The planet carrier  82  of the planet gears  78  rotates freely. 
     The second elementary gear train U 2  has an outer gear  84 , which is also drivingly connected with the sun gear  76  via the planet gears  78 . The planet carrier  82  is shared by the elementary drives U 1  and U 2 . 
     The two elementary drives U 1  and U 2 , in turn, are drivingly coupled to the superimposing gear train  26  (corresponding to  FIG. 2 ), wherein the outer gear  84  also represents the planet carrier  58  and the sun gear  76  also represents the planet carrier  52 . 
     The complete epicyclic gear trains U 1  and U 2  are arranged coaxially in a structurally and spatially advantageous manner. The outer gear  84  of the negative-ratio gear train U 2  is connected with the planet carrier  58  of the epicyclic gear train  44  (corresponding to  FIG. 2 ) to form one unit. Moreover, the sun gear  76  of the negative-ratio gear train U 1  is integrally designed with the planet carrier  52  of the epicyclic gear train  46 . 
     The described arrangement and coupling of the elementary gear trains U 1  and U 2  allows for the provision of a particularly effective torque-reducing gear train combination U z  with a high transmission ratio, which ensures a very precise, smooth redistribution of the output torque to the wheels  22 ,  24  of the motor vehicle, if a fast rotating electric motor  48  is used with correspondingly low driving torque. 
       FIG. 14 , in turn, shows the Wolf symbolic representation corresponding to the power-dividing gear train assembly according to  FIG. 13 . 
     Alternatively to  FIG. 13 ,  FIG. 15  shows an embodiment of the power-dividing gear train assembly with two elementary gear trains, both being designed as negative-ratio gear trains U 1  and U 2 . The power-dividing gear train assembly is described only as far as it substantially differs from the embodiment according to  FIG. 13 . Parts having the same function are provided with the same reference numerals. 
     Unlike in  FIG. 13 , the outer gear  84 ′ of the elementary gear train U 2  in  FIG. 15  is connected to the sun gear  40 ′ of the differential U via a shaft  86 , on which sun gear  40 ′ the planet gears  56  of the epicyclic gear train  44  of the superimposing gear train  26  are also rotatably supported via a carrier  88  (according to  FIG. 2 ). 
     The planet gears  78  of the elementary gear train U 1  are supported on the common planet carrier  82 ′ with the planet gears  90  of the epicyclic gear train U 2 . The planet gears  90  mesh with the sun gear  76 ′ a ; the planet gears  78  mesh with the sun gear  76 ′ b . The shaft  76  shared by the sun gears  76 ′ a  and  76 ′ b  interacts with the superimposing gear train  26 . 
     The Wolf symbolic representation of the power-dividing gear train assembly according to  FIG. 15  corresponds again to  FIG. 14 . 
     Based on the exemplary embodiments described above, other couplings of the elementary gear trains U 1  and U 2  are also designable in accordance with the required efficiencies and transmission ratios. 
     When the output shafts  14 ,  16  are rotating synchronously, the epicyclic gear trains U and U z  or U 1 , U 2  rotate as a block, i.e. there is no loss of efficiency due to rolling-off. As this driving state occurs predominately in motor vehicles, the teeth of the gear elements may be implemented in a less expensive manner, e.g., as simple, robust straight teeth. 
     The teeth of the superimposing gear train  26 , on the contrary, are designed to be quieter (e.g. helical teeth) and with good rolling-off characteristics. Due to the torque-reducing connection of the superimposing gear train  26 , the structure may be designed lighter; the same applies to the drive motor, which may be implemented as a hydraulic motor or preferably an electric motor. If necessary, the gears as well as the other gear train elements of the superimposing gear train  26 , including the not-illustrated housings, may be made, at least partially, of high strength plastic. As gears made of plastic run quieter, straight teeth could also be used for the superimposing gear train  26 , if necessary.