WO2015010841A1

WO2015010841A1 - Planetary gearbox

Info

Publication number: WO2015010841A1
Application number: PCT/EP2014/063142
Authority: WO
Inventors: Tobias DIECKHOFF; Thomas Dögel
Original assignee: Zf Friedrichshafen Ag
Priority date: 2013-07-23
Filing date: 2014-06-23
Publication date: 2015-01-29
Also published as: DE102013214351A1

Abstract

The invention relates to embodiments of a planetary gearbox (30) comprising at least one planetary gear (34) having a planetary gear toothing for meshing with a ring gear (68); and a ring gear (68) having a ring gear toothing which is in meshed engagement with the planetary gear toothing of the planetary gear (34). The planetary gearbox (30) is characterised in that in the planetary gear (34) a planetary gear contact surface (93-2) is constructed axially offset with respect to the planetary gear toothing, and that in the ring gear (68) ring gear contact surface (93-1) corresponding to the planetary gear contact surface (93-2) is constructed axially offset with respect to the ring gear toothing, wherein the planetary gear contact surface (93-2) and the ring gear contact surface (93-1) roll on one another during a relative rotation of the planetary gear (34) with respect to the ring gear (68) and form a radial bearing between the ring gear (68) and the planetary gear (34).

Description

Planetenqetriebe

Embodiments of the present invention relate to a planetary gear with at least one planetary gear with a Planetenradverzahnung for meshing with a ring gear, and with a ring gear with a ring gear, which meshes with the Planetenradverzahnung of the planetary gear or at least can be brought.

From the German patent application DE 10 201 1 007 1 18 A1 is an assembly in

Form of a torsional vibration damping arrangement, which divides a in an input area, for example by a crankshaft of a drive unit, introduced torque into a guided over a first torque transmission torque component and a guided over a second Drehmomentübertragungsweg torque component. In this torque distribution not only a static torque is divided. Also included in the torque to be transmitted oscillations or rotational irregularities, for example, generated by periodically occurring ignitions in a drive unit, are proportionately divided between the two torque transmission paths. In a coupling or overlay arrangement, which may be embodied as a planetary gearbox transmission, the torque components transmitted via the two torque transmission paths are recombined and then introduced as a total torque into an output range, for example a friction clutch, transmission or the like.

In at least one of the torque transmission paths, a phase shifter arrangement is provided with an input element and an output element, which is constructed in the manner of a vibration damper, ie with a primary side and by a compressibility of a spring arrangement with respect to this rotatable secondary side, especially if this vibration system in a supercritical state goes over, that is excited with vibrations that are above a resonant frequency of the vibration system, a phase shift of up to 180 ° can occur. This means that at maximum phase shift from the Vibration components emitted phase shifted by 180 ° with respect to the vibration components recorded by the vibration system. Since the vibration components conducted via the other torque transmission path experience no or possibly a different phase shift, the vibration components contained in the torque components combined by means of the coupling arrangement can be destructively superimposed on one another, so that in an ideal case the total torque introduced into the output region essentially follows is no vibration component static torque contained.

FIG. 1 schematically shows a torsional vibration damping arrangement 10, which operates on the principle of power branching or torque branching. The torsional vibration damping arrangement 10 may be in a drive train of a vehicle between a drive unit 12 and a subsequent part of the drive train, so for example a starting element 14, such. As a friction clutch, a hydrodynamic torque converter or the like can be arranged. The torsional vibration damping assembly 10 includes an input portion, generally designated 16. In the input region 1 6, a torque absorbed by the drive unit 12 branches into a first torque transmission path 18-1 and a second torque transmission path 18-2. In the area of an overlay unit generally designated by the reference numeral 20, which is also referred to as a coupling arrangement in the following, the torque components guided via the two torque transmission paths 18-1 and 18-2 are connected by means of a first coupling arrangement input part 22 which, for example, comprises a planet or ring gear carrier 24 can and a second coupling arrangement input part 26, which may have a drive sun gear 28, introduced into the coupling assembly 20 and merged there again. Differently constructed coupling arrangement input parts and coupling arrangement output parts are also possible. In this case, the coupling arrangement 20 may be embodied, for example, as a planetary gear 30. On the planet carrier 22, for example, a first Pianetenrad 32 and a second planetary gear 34 can be mounted radially successively and axially overlapping or overlapping rotatably. The first planetary gear 32 can mesh on the one hand with the drive sun gear 28 and on the other hand with the second planet gear 34. In the according to FIG 1 only lent exemplified arrangement, the second planetary gear 34 is used to reverse the direction of rotation. From the second planet gear 34, the converged torque via an output member 36 which may include, for example, a Abtriebshohlrad 38, which also meshes with the second planet gear 34 and is rotatably connected to an output portion 40, to the starting element 14, such as a clutch or a Geared.

In the first torque transmission path 18-1, a vibration system indicated generally by the reference numeral 42 is integrated. The vibration system 42 is operative as a phase shifting arrangement and includes a primary mass 44 to be connected to the prime mover 12, an input member 46 non-rotatably connected to the primary mass 44, and a spring assembly 48 connected to the input member 46. An output member 50 of the spring assembly 48 is further provided with a Intermediate element 52 is connected, which forms the planetary wheel carrier 24 by way of example here and on which the first planetary gear 32 and the second planetary gear 34 are rotatably mounted. Thus, according to the torsional vibration damping arrangement 10 of FIG. 1, the pianoforte support 24 is positioned, for example, in the first torque transmission path 18-1, which has a phase shift from the rotational nonuniformities directed via the first torque transmission path 18-1 with respect to the rotational nonuniformities conducted via the second torque transmission path 18-2. Characterized in that the output member 50 of the spring assembly 48 is rotatably coupled to the Planentenradträger 24, form the phase shifter assembly 42 and the coupling assembly 20 a, in axial extent, compact unit. Another positive for a decoupling is that inertia of the Planetenradt- carrier 24 and the first and second planetary gear 32, 34 enter into the inertia of the intermediate element 52.

A torque curve in the first torque transmission path 18-1 can extend from the drive unit 12 via the primary mass 44 and the input element 46 into the spring arrangement 48. From the spring assembly 48, the first torque is transmitted via the output member 50 of the spring assembly 48 and the intermediate member 52 to the Pianetenradträger 24, which the first planetary gear 32 and the second te planetary gear 34 primarily receives, led. In this case, the output element 50, the intermediate element 52 and the planet carrier 24 are rotatably coupled together.

In the second torque transmission path 18-2, the second torque is transmitted from the drive unit 12 to a drive sun gear 28 rotatably connected thereto. The drive sun gear 28 meshes with the first planetary gear 32 and thereby guides the second torque to the first planetary gear 32 of the coupling assembly 20.

Consequently, via the two torque transmission paths 18-1 and 18-2, the first and second torques reach the first planetary gear 32 and are brought together there again. The second planetary gear 34, which meshes with the first planetary gear 32, serves to reverse the direction of rotation, before the converged torque is guided by the second planetary gear 34 via the output ring gear 38 to the output region 40, to which the starting element 14, for example a friction clutch, a transmission or a torque converter is attached, which are not shown here.

In the event that the inertia of the intermediate element 52 is insufficient to achieve a decoupling quality, an additional mass element 54 can be fastened in a rotationally fixed manner to the intermediate element 52. An additional improvement of the decoupling can be achieved by the positioning of a known mass pendulum 56 on the intermediate element 52.

Such torsional vibration damping assemblies 10 may, in addition to hydrodynamic torque converters between a converter lock-up clutch and a

Output unit, such as. B. a transmission input shaft can be switched. For this purpose, converter lock-up clutch, torsional vibration damping arrangement and hydrodynamic torque converter in a common housing, for. B. within a bell housing, are. While engine torques to be transmitted on the one hand steadily increase, on the other hand an available space in the bell housing on the other hand decreases noticeably. FIG. 2 a shows a torsional vibration damping arrangement 10 'according to a similar principle as described in FIG. 1, as an application in conjunction with a hydrodynamic torque converter 90 as a starting element. The resulting starting element comprises predominantly the torque converter 90 with a converter lock-up clutch 62 and the torsional vibration damping arrangement 10 ', which is arranged between the converter lock-up coupling 62 and a power take-off unit, such as a power take-off. B. is arranged a transmission input shaft. The torsional vibration damping arrangement 10 'comprises, as already described with reference to FIG. 1, a first and a second torque transmission path 18-1 and 18-2, a phase shift arrangement 42 and a coupling arrangement 20 in the form of a planetary gear. For better clarification of the operating principle of the starting element shown in Figure 2a, Figure 2b shows a torque curve with closed converter lerüberrückkupplung 62, while the figure 2c shows a torque curve when the converter lock-up clutch 62 is open. Figures 2b and 2c can be seen with reference to the descriptions of Figure 2a.

With a closed torque converter lock-up clutch 62, as shown in FIG. 2b, a total torque Mg, which may come from a drive unit 12, for example an internal combustion engine, passes via a crankshaft 19 to a converter housing 95. Further, the total torque Mg is guided from the converter housing 95 via a converter clutch drive 63 into the converter lock-up clutch 62. As a result of a torque converter lock-up clutch 62 closed in accordance with FIG. 2b, the total torque Mg is also conducted via a converter clutch output 64 into the torsional vibration damping arrangement 10 ', in this case to a guide plate 59 of a radially inner spring set or inner spring set 58 which is non-rotatably connected to the Wandierkupplungsabtrieb 64. Accordingly, the guide plate 59 can also be regarded as the input region 16 of the torsional vibration damping arrangement 10 '. From the guide plate 59, the total torque Mg is divided into a first torque Mg1 and a second torque Mg2. The first torque Mg1 passes from the guide plate 59 to a inner spring set 58. From the inner spring set 58, the first torque Mg1 is led via a hub disc 61 to an outer spring set 57 which is arranged radially further outward relative to the inner spring set 58 within the converter housing 95 is. From the outer spring set 57, the first torque Mg1 passes through a stop element 65 and an intermediate element 52, which is exemplified here as a drive hollow gear carrier of the planetary gear or the coupling assembly 20 and rotatably connected to the stop element 65, to a Antriebshohlrad 68, which in turn rotatably is connected to the Antriebshohlradträger 52 and is rotatable about an axis A. Here, the drive ring gear 68 meshes with a first gear segment 81 -1 of a planetary gear 34 and thus performs the first torque Mg1 to the planetary gear 34th

The second torque Mg2 passes via the guide plate 59 to a drive sun gear carrier 17 connected in a rotationally fixed manner to the guide plate 59. A drive sun gear 28 is attached in a rotationally fixed manner to the drive sun gear carrier 17. The drive sun gear carrier 17 and the drive sun gear 28 can also be manufactured as one component. As a result, the second torque Mg2 is supplied to the drive sun gear 28. The drive sun gear 28 meshes with a second gear segment 81 -2 of the planetary gear 34 and thus guides the second torque Mg2 to the planetary gear 34. Thus, the first torque Mg1 and the second torque Mg2 are brought together again at the planetary gear 34. In this case, a vibration component in the first torque Mg1, which is passed through the first torque transmission path 18-1 through the phase shifter assembly 42, by means of the phase shift in the ideal case by 180 ° to the vibration component in the second torque Mg2, which is not passed through the phase shifter 42, phase-shifted , Consequently, ideally the planetary gear 34 would destructively overlap the first torque Mg1 with a vibration component phase-shifted by 180 ° and the second torque Mg2, so that the total torque Mg without torsional vibration components will be applied to a planetary carrier 24 on the output side. The planet carrier 24 can also be regarded here as the output region 40 of the torsional vibration damping arrangement 10 '. The Planentenradträger 24 is rotatably connected according to Figure 2a, b, c with a Abtriebsflansch 86 to which in turn a transmission input shaft, not shown here, can be rotatably coupled and the total torque Mg, ideally without vibration components, to a transmission, not shown here, can forward. In order to increase a moment of inertia of the intermediate element or of the drive hollow wheel carrier 52, which can have a positive effect on the phase shift, a turbine wheel 75 is non-rotatably connected via a carrier riveted to the intermediate element 52, which is thus non-rotatably connected to the intermediate element 52 the intermediate member or the Antriebshohlradträger 52 coupled. In addition, additional masses 76 coupled to the carrier 71 can be provided which increase the mass moment of inertia of the intermediate element 52 or the secondary side of the phase shifter 42 and thus have a positive effect on the phase shift. The turbine wheel 75 of the torque converter 90 also forms a connection to a thrust bearing 72. In the illustration according to FIG. 2a, b, c, an additional thrust bearing 72 is inserted between a thrust washer 77 and the output flange 86, so that one additionally rotates with the turbine wheel 75 connected bearing disk 78 between rolling elements of the bearing 72 is guided axially. Thus, not only an axial bearing of a stator 66 which is non-rotatably connected to the pressure plate 77, guaranteed, but also in addition an axial bearing of the turbine 75 and the components attached thereto, both with respect to the output flange 86, as compared to a freewheel 91 of the Leitrads 66 and the converter housing 95 reaches. A plain bearing or a differently designed rolling bearing would also be possible as thrust bearing 72. The axial bearing point 72 should, however, essentially absorb the axial forces of the turbine wheel 75 during converter operation and define the axial position of the intermediate element or the drive hollow wheel carrier 52. A radial bearing of the coupling arrangement 20, 30 takes place here via the toothed segments 81 -1, 81 -2 of the planetary gear 34 as so-called flying bearing.

One way of being able to realize a required for the function of the torsional vibration damping assembly 10 'stationary translation between the Antriebssonnenrad 28 and the Antriebshohlrad 68 with a smaller radial space requirement than in Figure 1, is a use of the planetary gear 34 with two different gear segments 81 -1 and 81 -2, as shown in Fig. 2a. In this case, a central axis B forms a rotary and central axis both for the toothed segment 81-1 and for the toothed segment 81-2. Furthermore, the two toothed segments 81 -1 and 81 -2 may overlap partially axially (ie in the direction of the axis of rotation A or B), so that the toothed segments 81 -1 and 81 -2 each have 180 angular axes. can be executed. The use of the planetary gear 34 with two different, partially axially overlapping toothed segments 81 -1 and 81 -2 is possible because a rotation angle about the rotational axis B of the planetary gear 34 is sufficiently low. Characterized in that the toothed segment 81 -2, which meshes with the Antriebssonnenrad 28, is greater than the toothed segment 81 -1, which meshes with the Antriebshohlrad 68, the amount of stationary translation increases compared to a transmission with known planetary gears at the same off - and dimensions. For a better utilization of the axial space, the two toothed segments 81 -1 and 81 -2 of the planetary gear 34 may also, as shown, partially offset axially relative to each other.

In an open converter clutch 62 with the torque curve, shown in Figure 2c, a total torque Mo on the converter housing 95 and a connecting plate 67 is further passed to a pump 74 of the torque converter 90. In this case, the impeller 74 rotationally fixed, for example by means of a welded connection, connected to the connecting plate 67. The connecting plate 67 is in turn rotationally fixed, for example by means of a welded joint, with the transducer housing 95, respectively. The torque converter 90 thus applies the total torque Mo to the impeller 74. Depending on a design of the hydrodynamic torque converter 90, as well as the applied total torque Mo and an applied speed at the impeller 74, a torque Mt is applied to the turbine wheel 75. Since the turbine wheel 75 is non-rotatably coupled to the drive hollow wheel carrier or the intermediate element 52, the torque Mt is forwarded from the turbine wheel 75 to the intermediate element 52. From the intermediate element 52, the torque Mt is divided into two torque components Mt1 and Mt2. The one torque component Mt2 is applied to the drive ring gear 68, which is non-rotatably coupled to the intermediate element 52. The other torque component Mt1 is guided via the intermediate element 52 and the stop element 65 to the outer spring set 57. From the outer spring set 57, this torque component Mt1 passes via the hub disk 61 to the inner spring set 58 and further from the inner spring set 58 via the guide plates 59 to the drive sun gear carrier 17 and consequently to the drive sun gear 28. Since both the drive sun gear 28 and the drive ring gear 68 follow meshing with the planetary gear 34, the two torque components Mt1 and Mt2 are re-connected to the planetary gear 34. brought together. Via the output-side planet carrier 24, on which the planet gear 34 is rotatably mounted, the combined torque Mt is forwarded to the output flange 86, which is non-rotatably connected, for example by means of a welded connection, to the planet carrier 24. It is also possible to design the output-side output flange 86 and the planet carrier 24 as an output-side component. From the output-side output flange 86, the combined torque Mt can be forwarded to a transmission, not shown here or a similar component.

For storage of the components on the output side of the phase shifter 42 is a thrust bearing 72 in the area or using the usually existing in converter designs freewheel storage known. A radial bearing can - as described - be designed as a floating bearing of the ring gear 68 on a planetary gear 34. In the event that the flying bearing does not provide sufficient radial guidance, an additional radial bearing or a use of combined radial / thrust bearings in the area of the thrust bearing point 72 may be considered. However, such measures, especially when occurring very high radial loads, may not be sufficient.

Based on this, it is an object of the present invention to develop a planetary gear so that it is improved over known planetary gears, in particular taking into account a compact (axial) space.

This object is achieved by a planetary gear with the features of the independent claim.

Some advantageous embodiment and developments are the subject of the dependent claims.

Embodiments relate to a special form of radial bearing, which can be used in planetary gears in general and thus especially in the torque nonuniformity reduction (DU reduction) by power split. Likewise, embodiments provide specifically for a An additional possibility is available to intercept additional high axial loads, as can occur, for example, in the case of a strong deceleration of a vehicle due to a mass of parts to be stored.

According to a first aspect, embodiments provide a planetary gear, which can be used, for example, as a coupling gear for a power or torque split, i. can merge a first and a second torque transmission path and forward to an output. The planetary gear in this case comprises at least one planetary gear with a planetary gear toothing for combing with a ring gear and optionally with a sun gear. Further, the planetary gear has a ring gear with a ring gear, which is in mesh with the planetary gear teeth of the planetary gear. According to embodiments, the planetary gear is characterized in that in the planetary gear axially offset from the Planetenradverzahnung a Planetenradkontaktierungsflä- surface is formed. In the ring gear, a ring gear contacting surface corresponding to the planetary gear contacting surface is formed axially offset from the ring gear contacting surface so that the planetary gear contacting surface and the ring gear contacting surface can form a radial bearing between the ring gear and the planetary gear during a relative rotation of the planetary gear to the ring gear. In this case, this radial bearing can be designed for example as a known plain bearing and or as a rolling bearing, in which also known additional sliding or additional rolling elements can be applied. Additionally or alternatively, a sun gear contacting surface corresponding to or to a further Planetenradkontaktierungs- surface may be formed in the sun gear axially offset from the sun gear so that the Planetenradkontaktierungsfläche and Sonnenradkontaktierungsfläche roll on a relative rotation of the planet gear to the sun and a radial bearing between the planet and Sun wheel can form. This radial bearing can also be designed in the already mentioned embodiments.

Embodiments therefore relate to a special embodiment of a roller bearing, in which a ring gear and one or more planets can be provided with additional (rolling) surfaces, which are arranged coaxially to the respective axes of rotation can, roll on each other and can transmit radial forces between the planetary gear and the ring gear via an emerging line contact.

According to some embodiments, the mutually corresponding annular or rolling surfaces may be formed only on one side of the meshing teeth to each other, for example, to be able to save axial space. In order to achieve a more symmetrical load distribution, which may be advantageous in particular at high radial loads, but also according to some embodiments with occurring axial loads, some embodiments provide axially to form on both sides of the intermeshing toothings corresponding planetary and HohlVSonnenradkontaktierungsflächen. These can optionally be designed such that, in addition to the radial bearing, the two-sided arrangement of the mutually corresponding contacting surfaces also forms an axial bearing between the hollow sun gear and the planetary gear.

Depending on the requirements, a toothing and an axially offset along the respective axis of rotation contacting surface of the planetary gear and / or Hohlbzw. Sunwheel be formed in one piece or several pieces. The former can primarily a higher strength, the latter mean a better or cost-effective manufacturability.

According to some embodiments, a Planetenradkontaktierungsfläche be formed by means of an axially offset from the planetary gear toothing and radially outwardly facing annular shoulder of the planetary gear. In this case, the Planetenradringschulter (or the rolling surface thus formed) project beyond the planetary gear or a pitch circle of the planetary gear radially outward or else be set back radially inward. Accordingly, the corresponding hollow or Sonnenradkontaktierungsfläche can be formed by means of an axially offset from the hollow / sun gear and arranged in the direction Planetenradringschulter facing annular shoulder of HohlVSonnenrads. Again, the HohlVSonnenradringschulter (or the rolling surface formed thereby) the hollow / sun gear or the pitch circle thereof protrude radially outward or in contrast to be set back radially inward. Through this Orders arise, in addition to the radial bearing formed by the rolling surfaces rolling against each other, in particular by means of the axially axially supporting and encompassing annular shoulders, in addition, a thrust bearing of the meshing drive and driven components. The mutually corresponding contacting surfaces on the one hand convex (eg on the planet gear) and on the other hand concave (eg on the ring gear) may be formed. Two mutually corresponding convex contacting surfaces are also conceivable, for example if the sun gear and the planetary gear are to bear against one another radially.

Additionally or alternatively, embodiments may provide that the planet gear is supported by a planet carrier and that the hollow or sun gear is supported by a hollow or sun gear carrier axially offset from the planet carrier, the planet carrier and / or the hollow / sun carrier may be formed such that defined between the planet carrier and the hollow / sun carrier when exceeding a Axialkraftschwellenwerts defined axial bearing points for mutual axial Abstützung. Such defined Axiallagerstellen can be provided for example by axial projections or lugs on the planet and / or the hollow / sun carrier, which provide defined contact points or surfaces between the planet and hollow / sun carrier. In some embodiments, the ring gear carrier may be coupled to a torque converter. Embodiments may thus provide one or more additional bearing points with respect to the axial bearing, which may intervene in the event that e.g. in a strong braking maneuver an existing axial bearing is not sufficient.

In the event that it comes through the predefined contact points or Axiallagerstellen to touch between the rotating planet carrier (output) and hollow / sun carrier (drive), it may be advantageous if the defined thrust bearing a material with a lower coefficient of friction and / or have a lower wear than adjacent sites. For example, a low coefficient of friction can reduce fuel consumption, while low wear can extend maintenance intervals. Additionally or alternatively, a first axis of rotation of the hollow or sun gear and a second axis of rotation of the planetary gear in an plane spanned by the two axes of rotation oblique, ie, not parallel to each other. According to some

Embodiments is therefore proposed to tilt the second axis of rotation of the planetary gear relative to the first axis of rotation. In particular, the second axis of rotation relative to the first axis of rotation can be tilted such that a space radially within a previously described inner spring set a torsional vibration damping arrangement and the associated cover or guide plates can be better used. By means of appropriate inclination or tilting the radially inner sun gear can be axially mounted on its sun gear closer to the inner spring set or its guide plates, whereby the torsional vibration damping arrangement and in particular the torsional vibration damping arrangement comprehensive starting elements can be built axially narrower. Thus, such embodiments allow the trend to follow ever decreasing space in the bell housing.

According to embodiments, the first and the second rotation axis are tilted relative to each other such that the first and the second rotation axis extend obliquely in a plane spanned by the two axes of rotation. Starting from an axial direction, which is defined by the first axis of rotation, the second axis of rotation comprises, in addition to an axial component parallel to the first axis of rotation, an additional directional component which is oriented perpendicular to the axial direction defined by the first axis of rotation. This can be for example a radial component. Depending on the specific structural requirements, an angle between the two axes of rotation may be in a range of 0 ° to 45 °, in particular of 5 ° to 20 °. According to exemplary embodiments, an inclination or tilting of the two axes of rotation relative to one another is selected such that a radially inner part of the planetary gear or a sun gear meshing therewith axially move closer together with an input area or an (inner) spring set of a torsional vibration damping arrangement comprising the planetary gear can.

According to some embodiments, the planetary gear may have a first planetary gear part having a first gear diameter and a second planetary gear part having a second of the first different gearing diameter. While the first and second Planetenradteil according to some embodiments by different arranged coaxially along the second axis of rotation planetary gears can be realized with different gearing diameters, embodiments may also be preferred in which the first Planetenradteil by a first circular segment of the planetary gear with the first gearing diameter and the second Planetary gear is formed by a second circular segment of the planetary gear with the second gear diameter. In particular, the last-mentioned embodiments enable a significant axial space gain in an efficient manner. The different gear diameters of the first and second Planetenradteils gear ratios between a first torque transmission path and a second torque transmission can be made variable, which can be advantageous to the design of an entire torsional vibration damping arrangement and thereby can provide a space advantage.

Accordingly, in some embodiments, the ring gear may be coupled to a first torque transfer path of a torsional vibration damping assembly, the sun gear may be coupled to a second torque transfer path of the torsional vibration damping assembly, and the planetary gear may form a coupling assembly for superimposing torques over the torque transfer paths.

According to some embodiments, the ring gear located in the first torque transmission path may mesh with the first planetary gear part and the sun gear located in the second torque transmission path may mesh with the second planetary gear part. In order to arrange the ring gear on the one hand and the sun gear on the other hand space reasons in different axial planes, the two Planetenradteile in the direction of the first and / or the second axis of rotation axially (ie, in the respective axial direction) offset from each other. Of course, embodiments are conceivable in which the two Planetenradteile are arranged in the axial direction, ie in the direction along the first and / or the second axis of rotation in the same axial plane. Such embodiments allow in particular a simple and cost-effective production of the planetary gear.

According to some embodiments, it may be provided that the first torque transmission path comprises a phase shifter arrangement for generating a phase shift of rotational irregularities conducted over the first torque transmission path with respect to rotational irregularities conducted via the second torque transmission path. In at least one of the torque transmission paths can thus be provided with an input element and an output element, a phase shifter assembly, which may be constructed in the manner of a vibration damper, ie with a primary side and a compressibility of a spring arrangement with respect to this rotatable secondary side. In particular, when this vibration system is in a supercritical state, that is excited with vibrations that are above the resonant frequency of the vibration system, a phase shift between the two torque transmission paths of up to 180 ° may occur. This means that at maximum phase shift, the vibration components emitted by the vibration system are phase-shifted by 180 ° with respect to the vibration components picked up by the vibration system. Since the vibration components conducted via the other torque transmission path experience no or possibly a different phase shift, the vibration components contained in the merged torque components and then phase-shifted with respect to each other can be destructively superimposed on one another, so that in an ideal case the total torque introduced into the output region has essentially no vibration components containing static torque. The spring arrangement of the phase shifter arrangement may comprise at least one spring set, which advantageously comprises a helical spring. When using at least two spring sets, these can be arranged both in parallel and in serial mode of action.

In order to bring about further improvements with regard to a required axial installation space, a secondary side of the phase shifter arrangement, which is coupled with its primary side via the spring arrangement, can essentially be provided by an integral mass main body to provide a desired mass moment of inertia. be formed. Compared to conventionally multi-part or multi-piece masses and / or additional masses to provide the desired mass moment of inertia offers a one-piece mass body in particular space advantages. In order to be able to save even more axial and / or radial installation space, some embodiments propose to mold a ring gear for meshing with the planetary gear in the secondary-side one-piece grounding body. In embodiments of this type, the one-piece grounding body can thus simultaneously serve as a drive ring gear for introducing a torque guided via the first torque transmission path into the planetary gear, in which the two torque transmission paths are brought together before being transmitted via an output-side planetary or ring gear carrier to a torque output of the torsional vibration damping arrangement become.

To further space optimization, the one-piece ground mass body can continue to be used as a radial support for the (Au Ben) spring set the phase shifter assembly, whereby conventional components such. B. guide plates and stop members for the spring assembly can be saved. Additionally or alternatively, the mass base body can also have webs projecting into a spring channel, which can serve as a spring in the spring arrangement as stops in the circumferential direction (that is, tangentially to the first axis of rotation). Thus, more components and thus ultimately also further space can be saved.

Further space-saving potential, in particular in the axial direction, can be achieved in that a hub disk of a primary side of an (outer) torsional vibration damper or spring set engages from radially inward to radially outward into the spring assembly. Compared to conventional designs, this measure allows the torsional vibration damping arrangement to be made axially narrower.

Additionally or alternatively, according to some embodiments, a turbine wheel of a torque converter arranged axially next to the torsional vibration damping arrangement can be non-rotatably connected about an axis of rotation, eg a transmission rotational axis, with a first output range (with respect to the first torque range). Mentübertragungswegs) of the torsional vibration damping arrangement and the phase shifter be coupled. A spline is a possible shape in a shaft-hub connection. It is a multiple driver connection, with a torque is transmitted from the tooth flanks. The shaft is external and the hub is internally toothed. The two parts can be basically move axially against each other, which is exploited in many applications. By means of the splines, the turbine wheel or the turbine with the first output region can indeed be connected to it in a rotationally fixed manner, but not axially, so that deformations of a turbine blade due to excessive axial forces can be avoided. Likewise, a shift or deformation of the transmission by the axial thrust of the turbine should be avoided. According to some exemplary embodiments, the first output region of the torsional vibration damping arrangement can comprise a carrier for the ring gear, wherein the ring gear carrier is then coupled to the turbine wheel in a rotationally fixed manner via the spline but in an axially indeterminate manner.

According to a further aspect, further exemplary embodiments also provide a motor vehicle having an exemplary torsional vibration damping arrangement.

In the following, some exemplary embodiments will be explained in more detail with reference to the attached figures. Show it:

Figure 1 is a schematic diagram of a torsional vibration damping arrangement with two planetary gears, which are mounted at the output of a phase shifter assembly;

Figure 2a shows a torsional vibration damping arrangement in application in connection with a hydrodynamic torque converter;

FIG. 2b shows a torque curve of the arrangement according to FIG. 2a with the converter clutch closed; FIG. 2c shows a torque curve of the arrangement according to FIG. 2a with the converter clutch open;

3 shows a section through a torsional vibration damping arrangement according to an embodiment;

Figure 4a, b is a sectional view of a segmented planetary gear with two different gear diameters according to an embodiment;

FIG. 5 shows a starting element with a torsional vibration damping arrangement according to a further exemplary embodiment, which is arranged between a torque converter

Bypass clutch and a torque converter is arranged;

6 shows a starting element with a torsional vibration damping arrangement according to a further embodiment;

FIG. 7 shows a starting element with a torsional vibration damping arrangement according to an exemplary embodiment with an integral mass base body for providing a mass moment of inertia;

8 shows a starting element with a torsional vibration damping arrangement according to an embodiment with a hub disc, which engages from radially inward to radially outward in a spring arrangement of a phase shifter assembly;

Figure 9, 10 further embodiments of starting elements with a torsional vibration damping arrangement for obtaining additional axial space;

Figure 1 1 shows an embodiment with a radial bearing in a coupling or planetary gear according to an embodiment;

Figure 12 shows some design variants of the radial bearing in the coupling or planetary gear; FIG. 13 shows an exemplary embodiment of an axial bearing of a turbine wheel, a torque converter and an input ring gear of a power branch adjacent thereto;

FIG. 14 shows an exemplary embodiment with a spline coupling between a turbine wheel of a torque converter arranged axially next to a torsional vibration damping arrangement and an output region of the torsional vibration damping arrangement.

Various embodiments will now be described in more detail with reference to the accompanying figures, in which some embodiments are shown. In the figures, the thickness dimensions of lines, layers and / or regions may be exaggerated for the sake of clarity.

In the following description of the attached figures, which show only some exemplary embodiments, like reference characters may designate the same or similar components. Further, summary reference numerals may be used for components and objects that occur multiple times in one embodiment or in a drawing but are described together in terms of one or more features. Components or objects which are described by the same or by the same reference numerals may be the same, but possibly also different, in terms of individual, several or all features, for example their dimensions, unless otherwise explicitly or implicitly stated in the description.

Although embodiments may be modified and changed in various ways, exemplary embodiments are illustrated in the figures as examples and will be described in detail herein. It should be understood, however, that it is not intended to limit embodiments to the particular forms disclosed, but that embodiments are intended to cover all functional and / or structural modifications, equivalents and alternatives that are within the scope of the invention. Note that an element referred to as being "connected" or "coupled" to another element may be directly connected or coupled to the other element, or intervening elements may be present. Conversely, when an element is referred to as being "directly connected" or "directly coupled" to another element, there are no intervening elements. Other terms used to describe the relationship between elements should be interpreted in a similar fashion (eg, "between" versus "directly in between,""adjacent" versus "directly adjacent," etc.).

The terminology used herein is for the purpose of describing particular embodiments only and is not intended to limit the embodiments. As used herein, the singular forms "a," "a," "an," and "the" are also meant to include the plural forms unless the context clearly indicates otherwise. Further, it should be understood that the terms such as "including," "including," "having," and / or "having," as used herein, indicate the presence of said features, integers, steps, operations, elements, and / or components. but does not preclude the presence or addition of one or more features, integers, steps, operations, elements, components, and / or groups thereof.

Unless otherwise defined, all terms (including technical and scientific terms) used herein have the same meaning as commonly assigned to one of ordinary skill in the art to which the embodiments pertain. Further, it should be understood that terms, e.g. those that are defined in commonly used dictionaries are to be interpreted as having the meaning consistent with their meaning in the context of the relevant art, and not to be interpreted in an idealized or overly formal sense, unless this is so is explicitly defined.

FIG. 3 shows an exemplary embodiment of a torsional vibration damping arrangement 100 which, by way of example, is integrated into a converter housing 95 together with a converter lockup clutch 62 and a hydrodynamic torque converter 90. is grated and forms a starting element. The torsional vibration damping arrangement 100 thus forms an assembly of a drive train arranged axially next to or adjacent to the converter 90. An output 64 of the converter clutch 62 forms a to be driven or driven for rotation about a first axis A input portion 1 6 of the torsional vibration damping arrangement 1 00. A planet carrier 24, which may be coupled, for example by means of a welding ßverbindung with a Abtriebsflansch 86 to the transmission input shaft, forms an output range As already described above with reference to FIGS. 1 and 2, the torsional vibration damping arrangement 100 according to FIG. 3 also comprises a first torque transmission path 1 8-1 extending from the input area 16 to the output area 40, and also one of the input area 1 6 to the output area 40 extending second torque transmission path 1 8-2 and thus provides a power split. Connected to the output region 40 is a coupling arrangement 20 for superimposing torques conducted via the two torque transmission paths 18-1, 18-2. According to embodiments, the coupling arrangement 20 comprises a planetary gear 30 with a planetary gear 34, which is rotatable about a second axis of rotation B, which is arranged radially outwardly of the first axis of rotation A, which may be formed for example by a transmission input shaft. The torsional vibration damping arrangement 100 according to FIG. 3 differs from the torsional vibration damping arrangement 10 'according to FIGS. 2a, b, c, in particular in that the first rotation axis A and the second rotation axis B run obliquely relative to one another. Otherwise, the functions are similar, which is why a repeated detailed explanation of the operation is omitted. The reader is referred to the description of Figures 2a - c. The term "oblique" can be understood to mean that the first axis of rotation A and the second axis of rotation B extend obliquely or tilted relative to one another in a plane spanned by the two axes of rotation A, B. The two axes of rotation A, B can thus be arranged according to embodiments such that they span a common plane. This plane may have an axial component (in the direction of the first axis of rotation A) and a radial component (radially away from the first axis of rotation A toward the second axis of rotation B). In this common plane, the two axes of rotation A, B can include an angle other than 0 °. In particular, can the amount enclosed by the two axes of rotation A, B is in the range from 0 ° to 45 °, in particular from 5 ° to 20 °,

As can further be seen on the basis of the exemplary embodiment of FIG. 3, a rotation axis of a drive ring gear 68 of the coupling arrangement 20 located in the first torque transmission path 18-1 which meshes with the inclined planet gear 34 can run parallel to the first rotation axis A. Similarly, a rotation axis of a arranged in the second torque transmission 18-2 sun gear 28 of the coupling assembly 20 which meshes with the inclined planetary gear 34, also parallel to the first axis of rotation A. In some embodiments, the axes of rotation of the drive sprocket 68 and / or the sun gear 28 may be coupled to the axis of rotation A, e.g. can be formed by a transmission input shaft, coincide. Due to the inclination of the planetary gear 34, sun gear 28 and drive ring gear 68 may be in different axially disposed planes, i. H. in different axially along the axis of rotation A offset planes. Compared with the arrangement explained with reference to FIG. 2a-c, according to the arrangement of FIG. 3, the sun gear 28 is substantially closer to the radially extending inner guide plates 59 of the inner torsional vibration damper 58 in the axial direction (ie in the direction of axis of rotation A). In particular, the sun gear 28 may now be in the immediate axial vicinity of a connecting pin 69 between the inner damper guide plates 59 and the bearing flange 17. Especially in radial proximity to the first axis of rotation A thus considerable axial space can be saved.

Due to the inclination of the planet gear 34 and its axis of rotation B, which can be defined by a pin 79, an internal toothing of the drive ring gear 68 and / or an external toothing of the sun gear 28 can also be formed obliquely. In other words, this means that a plane formed by a pitch circle of the internal teeth of the drive ring gear 68 and / or a pitch circle of the outer teeth of the sun gear 28 is perpendicular to the second rotation axis B (and thus obliquely to the rotation axis A), as well as or the pitch circles of the external teeth of the planet gear 34th As can be seen from FIG. 3, some exemplary embodiments provide a planetary gear 34 which has a first planetary gear part (above the second rotation axis B) with a first gear diameter and a second planetary gear part (below the second rotation axis B) with a second one from the first one includes different gear diameter. While, deviating from the embodiment shown here, different sized and along the second axis of rotation B axially offset from each other arranged planet gears are conceivable, of which, for example, the larger with the sun gear 28 and the smaller can be in mesh with the Antriebshohlrad 68, preferred embodiments in that the first planetary gear part is formed by a first circular segment of the planetary gear 34 with the first gear diameter and the second planetary gear part by a second circular segment of the planetary gear 34 with the second gear diameter.

FIG. 4 a shows a possible embodiment of the planetary gear 34 with two different toothed segments 81 -1 and 81 -2 in plan view. In this case, the central or rotational axis B of the toothed segments 81 -1 and 81 -2 may be the same. In the embodiment shown here, the respective toothing (circle) segment 81 -1 and 81 -2 executed with 180 degrees. Not shown here, but also the toothed segments 81 -1 and 81 -2 can be performed with different degrees, such as the toothed segment 81 -1 with 150 degrees and the toothed segment 81 -2 with 210 degrees. The sum of the angular degrees of the toothed segments 81 -1 and 81 -2 can also be less than 360 degrees, but a maximum of 360 degrees.

FIG. 4b shows a possible planetary gear 34 with two different toothed segments 81-1 and 81-2 in section and in plan view. Both gear segments 81 -1 and 81 -2 have the same central axis or axis of rotation B. In this case, the gear segment 81 -1 may be formed with approximately 90 degrees and the gear segment 81 -2 with approximately 100 degrees. Both toothed segments 81 -1 and 81 -2 may partly overlap in the axial direction (along the axis of rotation B) (see FIG. 4b, left). It is good to see that comparatively much Mass and / or material can be saved when using toothed segments.

Figures 3 and 4 show how can be saved by embodiments axial space in torsional vibration damping arrangements and thus coupled start-up elements. Here, the axis of rotation B of the planet 34 of the coupling assembly 20 is tilted slightly relative to the axis of rotation A of the transmission. In this way, the space can be used radially within the inner spring set 58 partially for the sun gear 28 and the corresponding toothed segment 81 -2 of the planet 34, which allows a greater width of the teeth.

In the previously presented constructive versions of the power split in torque converter, the functional elements on the secondary side of the phase shifter 42, ie cover plate 52 of the spring set 57, ring gear 68 and additional mass 76 were considered as separate components connected by a joining process, eg rivets, directly or via connecting plates were. When using bent sheet metal parts, the bending radii and other restrictions in the design lead to free spaces between the parts. This is disadvantageous when mass is to be used as effectively as possible for moment of inertia, as this means to fill the space radially outward as close as possible with material and leave there no open spaces. The additional mass 76 is arranged outside the power flow and thus acts only in its function to increase the mass moment of inertia of the secondary side of the phase shifter 42. The material does not contribute to the strength or rigidity of the construction, but also causes an additional burden on the surrounding parts. The connection of the input ring gear 68 with a separate, located radially within the outer damper 57 connection with the cover plate 52 limits the diameter of the tooth pitch circle and thus also leads to the mass of the ring gear 68 is disposed on a radially smaller radius and thus not so generates much moment of inertia as on a larger radius. As a solution approaches for optimized design of the corresponding components are presented below, in which in particular the mass arrangement is optimized in terms of moment of inertia and power line. The mass moment of inertia on the output side of the radially outer spring accumulator 57 is a critically important factor in terms of power-split reduction in both the quality of the phase shift and the decoupling of the oscillatory components of the torque branch 18-1 routed via the phase shifter 42 significantly influenced. In general, better decoupling results can be achieved with high mass moment of inertia and matched spring sets and gear ratios than with low ones. On the other hand, however, there are the demands for the lowest possible weight of the entire converter and low total mass moment of inertia for reasons of driving dynamics. It is therefore necessary to provide a maximum permissible mass moment of inertia with as little mass as possible at the output of the phase shifter 42. In the sense of a modular principle, the option can also be provided to vary the mass moment of inertia by adding or omitting elements. In a construction according to FIGS. 2a-c, these requirements are already taken into account by connecting existing masses or moments of inertia, such as turbine 75, to the output side of phase shifter 42, and by providing additional additional mass 76, for example in the form of variable sizes a sheet and / or a Massering was provided. However, the conventional manufacturing method mainly riveted together sheet metal parts allows optimal use of space in which as much mass sits on a large radius, only conditionally. In addition, due to the additional mass 76 connected as a separate component, no force flow takes place, thus the very massively executed additional mass 76 does not contribute to the strength or stiffening of the assembly.

A further exemplary embodiment is shown in FIG. 5, which differs in particular from a more compact design of the secondary side of the phase shifter arrangement 42 or of the outer spring set 57 from previously explained embodiments. Further, the input gear 68 is connected radially outside an inner diameter of the outer damper 57 to the secondary side of the damper 57, resulting in a higher moment of inertia.

As already explained above with reference to FIGS. 2 a - c, the first torque transmission path 18 - 1 may comprise a phase shifter arrangement 42 for achieving generation of a phase shift of rotational irregularities conducted over the first torque transmission path 18-1 with respect to the rotational nonuniformities conducted by the second torque transmission path 18-2. The operation of the phase shifter assembly 42 has already been explained in detail at the beginning, which is why a new explanation is omitted here. The phase shifter assembly 42 has a radially outer (outer) spring set 57. This outer spring set 57 couples a primary side formed by the hub disc 61 with a secondary side formed by the intermediate element 52. The intermediate element 52 coupled to a stop element 65 is connected to a secondary-side ground body 82, for example by means of a welded connection. In order to increase a moment of inertia of the integrally formed mass base body 82, which can have a positive effect on the phase shift, it is connected via a coupled to the grounding body 82 and from radially outward to radially inwardly extending carrier 71 which is rotatably connected to the grounding body 82, with a turbine wheel 75 of an axially adjacent torque converter arranged rotationally fixed. In addition, additional masses 76 can also be provided here, which increase the mass moment of inertia of the main body 82 and thus have a positive effect on the phase shift.

Much of the mass of the spring set 57 downstream assembly is formed by the one-piece grounding body 82, which may be prepared for example by massive forming or casting. The grounding body 82 constitutes a connecting link between the guide plate or intermediate element 52 of the outer spring set 57, which can be constructed simpler here than in the original construction according to FIGS. 2a-c. which allows on the modular principle to adapt the moment of inertia of the assembly to different applications. In the section shown in FIG. 5, a common axially extending rivet 83 connects the components 65, 68, 82 and 76 radially outwardly of an inner diameter of the outer damper 57. At other locations along the circumference of the outer spring set 57 where no contact element 65 is positioned, There may be other such joints where then only the other parts are connected together accordingly. A torsion stop 70, which limits a twist angle of the outer spring accumulator 57 and thus protects the spring set 57 against block load can Here, preferably between the input-side hub disc 61 and the output-side guide plate or Zwischeneiement 52 are provided and may be located on the engine side (or torque flow upstream) of the spring set 57. This z. B. on both components 61 and 52 on the motor side corresponding tabs are formed, which overlap to the same extent and thus abut each other after a defined angle of rotation.

FIG. 6 shows a further optional modification of the construction, in which an integration of a ring gear toothing 68a into the one-piece grounding base 82 has taken place. An inner diameter of the ground body 82 and thus also the ring gear 68a may be greater than an inner diameter of the outer spring set 57, which leads to a higher moment of inertia. The guide plate 52 of the outer spring set 57 may be designed such that it is drawn from the outer spring set 57 next to the latter and between the latter and the Massegrundkorper 82 radially in the direction of axis of rotation A and with a radially inwardly facing portion on an axial plane surface of the axially adjacent grounding body 82nd is applied. At the points of its circumference, in which the stop elements 65 are positioned, corresponding recesses may be provided, so that nesting with the stop elements 65 in the circumferential direction is possible. The support plate 71, which is arranged axially between Massegrundkorper 82 and an additional mass 76 may be drawn in the axial direction as far in the direction of turbine wheel 75 that a connection of the support plate 71 with the Massegrundkorper 82, the additional mass 76 and - depending on the position on the circumference - either the guide plate 52 or the stop element 65 - for example, by riveting - on a pitch circle with all of the above components axially sweeping rivets is possible.

Figure 6 also shows an alternative design of the stopper 65 for protecting the outer spring set 57, wherein radially inwardly of the outer spring set 57 bent and cooperating tabs 84 and 85 of the components 61 and 65 define a Verdrehwinkei. In this case, a bent-out tab 84 of the primary-side hub disc 61 points essentially in the direction of the planetary gear 34. A tab 85 of the stop element 65 corresponding thereto is connected by a radially inward-pointing end disc. cut of the axially extending stop member 65 formed above. Other specific designs are of course possible.

FIG. 7 shows a further optional modification of the construction, in which, in turn, a further functional or component integration has been implemented in order to simplify the use of space, assembly and manufacturability. Although, according to FIG. 7, the second axis of rotation B is only slightly tilted or not tilted relative to the first axis of rotation A, the construction shown in FIG. 7 can readily be combined with exemplary embodiments with first and second axes of rotation A, B running at an angle to each other. According to FIG. 7, the one-piece ground body 82 is again preferably produced as a massive forming part. In comparison to other embodiments, the grounding body 82 virtually forms the secondary side of the outer spring set 57 and assumes functions of the intermediate element 52 and stop elements 65. The grounding body 82 is here shaped such that it can ensure radial support of the outer spring set 57 as well as in FIG may have the spring channel projecting webs, which provide the springs a stop in the circumferential direction. According to embodiments, the grounding body 82 may thus have in a spring channel of the spring assembly 57 projecting webs, which serve as a spring of the spring assembly 57 as stops in the circumferential direction. The spring itself can, as shown in the lower section, in FIG. 7, run in a slide track plate 87, which can be arranged radially within an axial lip of the ground body 82 pointing in the direction of the motor. As a result, the main body 82 can be made simpler, since no spherical contour is necessary. In order to hold the slide sheet 87, so that it can not slip out axially on the motor side, at several points on the circumference - off the webs for the spring stop - the axial lip of the body 82 can be inflected radially inwardly in these areas, as in the lower section of Figure 7 can be seen.

The ring gear 68 can again be designed as a separate component and be pressed with the grounding body 82, wherein an additional positive engagement, for example by means of a spline, determine the position and secure against rotation. An inner diameter of the mass body 82 and thus also the ring gear 68a may be greater than an inner diameter of the External spring set 57. The spline (also called splined) is a possible shape in a shaft-hub connection. It is a multi-drive connection, whereby the torque is transmitted by the tooth flanks. The shaft is outside and the hub is internally toothed. Other joining or connecting method between grounding body 82 and ring gear 68 or integration as a single component are of course also conceivable.

According to the embodiment of FIG. 7, a torsion stop for block protection of the outer spring set 57 can be provided such that fingers of the primary side hub disc 61, which reach between the individual springs of the outer spring set 57 in order to drive them, with their tips into an axial groove 88 immerse in the main body 82, which can limit the rotation due to interruptions in the circumferential direction accordingly.

FIG. 8 shows a further embodiment, which differs from the exemplary embodiments described above in that the primary-side hub disk 61, which is coupled on the input side to the output 64 of the converter lockup clutch 62 and moves radially outward in the direction of the inner spring set 58 Outer spring set 57 extends, radially engages from the inside out into the outer spring set 57 of the outer torsional vibration damper. In this case, the guide plate or intermediate element 52 of the outer spring set 57 may be shaped such that it guides the springs axially and radially on the motor side. According to the embodiment shown in Figure 8, the guide plate 52 has a substantially Ω-shaped cross-section. To provide the springs in the circumferential direction a stop, also at several points in the circumferential direction (for example, between two springs or spring sets connected in series) - segments of the guide plate 52 may be bent radially inward into the spring channel. A separate stop element is therefore not necessary. A connection with the grounding body 82 may, for example, as shown, carried out by pressing and / or welding. A torsion stop can here, analogous to the embodiment of Figure 6, by mutually corresponding formations on the hub disc 1 1 and the guide plate 52 done. The connecting member of the secondary-side components with each other in turn forms the grounding main body 82. In addition to the already described connection to Guide plate 52, the ring gear 68 and optionally an additional mass 76 by a joining method, such as pressing and / or pinning, be connected to the grounding body 82. The inner diameter of the ground body 82 and thus also the ring gear can be significantly larger than an inner diameter of the outer spring set 57. The support member 71, which forms the connection to the turbine 75 and the thrust bearing 72 may also be attached to the base 82, for example by pressing and / or welding, as shown in FIG.

FIGS. 9 and 10 show further embodiments of assemblies with torsional vibration damping arrangements, which are coupled to a torque converter 90. Although in the illustration of Figures 9 and 10, the two axes of rotation A, B are not or only slightly tilted to each other, the constructions of Figure 9 and 10 can be combined with exemplary embodiments of the unproblematic in which the two axes A, B obliquely to each other .

With the aim of releasing axial space for the toothing of the linkage 20, 30, various measures can be taken to move the inner spring set 58 further towards the engine or crankshaft 19 and in particular the free space radially inside or below the converter clutch 62nd better to use. The two figures 9 and 10 represent such variants, which can be combined without problems with other embodiments.

9 shows a modification of the control of the converter clutch 62. A channel for a fluid (for example oil) which presses an actuating piston 89 against the clutch 62 in order to actuate the clutch 62 is usually formed by beading in the piston carrier 99. Here, beads refer to manually or mechanically produced channel-shaped recesses. However, according to one embodiment, the fluid channel may be displaced into the housing 95 such that the piston carrier 89 may be made axially shallower and thereby narrower by the height of the channel. Accordingly, the inner spring set 58 can be displaced in the direction of the engine and the result for the transmission 20 is the space gain identified in FIG. In addition, for a simple assembly of the rivet 69 to a radius outside the sun gear head circle be shifted and for the radially inwardly facing guide plates 59 of the inner spring set 58 and possibly the hub disc 61 are adjusted.

10 shows a further optional modification of the converter clutch 62, in which, in addition to the above-proposed displacement of the oil passage into the housing 95, the clutch 62 itself is radially outwardly offset so that there is no radial overlap between the converter clutch 62 and the inner spring set 58 or whose guide plates 59 are. Due to the additional space thus gained, a point of contact between the actuating piston 88 and the piston carrier 89 can also be displaced radially outwards (approximately to the radial height of the inner spring set 58) and axially in the direction of the engine or crankshaft 19. Accordingly, the inner spring set 58 can likewise be displaced axially in the direction of the engine and release the additional installation space for the coupling gear 20, 30 marked in FIG. As a further advantage, in the arrangement of Figure 10, the connection of the converter clutch 62 to the spring accumulator, d. H. the clutch output 64, done by appropriate design of the spring accumulator cover plate 59 itself. According to FIG. 10, the converter output 64 and the cover plate 59-1 are designed as a single component which can be coupled in a rotationally fixed manner to the hub disc 61 by means of an axial pin 98.

In the foregoing, essentially two possibilities have been mentioned in order to radially guide the secondary-side components to the phase shifter 42 and to absorb the radial forces acting on them: on the one hand a flying bearing on the toothing between planet gears 32, 34 and ring gear 68, and on the other hand optionally to the bearing point 72 an additional radial bearing or bearing with radial and axial effect in common design.

As a further possibility, a special embodiment of a roller bearing is presented in the following, in which the ring gear 68 and the planet or 34 can be provided with additional annular surfaces which are arranged coaxially to the respective axes of rotation, roll on each other and thereby on the resulting line contact radial forces can transfer. This special roller bearing can be combined with all the embodiments described above. the. With respect to an axial bearing an additional bearing is also presented, which can intervene in the event that, for example, in a strong braking maneuver an existing axial bearing is not sufficient.

Figure 1 1 shows an embodiment of the linkage 20, which is designed as a planetary gear 30 and axially offset from the teeth of the ring gear 68 and the planetary gear 34 and a planetary segment, which meshes with the ring gear 68, cylindrical surfaces 93-1, 93rd -2 are provided, which can roll on each other in operation. The cylindrical surfaces 93-1, 93-2 each have surface normal in the radial direction. It is in the embodiment shown in Figure 1 1 quasi a combination of planetary gear and bearings, wherein one or more planetary gears 34 with one or more cylindrical or annular surfaces 93-2 quasi rolling elements, which correspond to the corresponding cylindrical or annular surfaces 93-1 of the ring gear 68 can roll. For this purpose, a cylindrical surface 93-2 of the planetary gear 34 (Planetenradkontaktie- rungsfläche) convex and each coaxially with the planet gear 34 and its axis of rotation may be formed. A cylindrical surface 93-1 of the ring gear (Hohlradkontaktierungsfläche), however, may be concave and coaxial with the ring gear 68 and its axis of rotation may be formed. Accordingly, the embodiment of FIG. 1 1, a planetary gear 30 with at least one planet gear 34 with a Planetenradverzahnung for combing with a ring gear 68. The ring gear 68 includes a ring gear which meshes with the pinion gear of the planetary gear 34. In the planetary gear 34, a Planetenradkontaktierungsfläche 93-2 is formed axially offset from the Planetenradverzahnung. In the ring gear 68, a ring gear contacting surface 93-1 corresponding to the Planetenradkontaktierungs- surface 93-2 is axially offset from the ring gear contacting surface 93-2 formed, the Planetenradkontaktierungsfläche 93-2 and the Hohlradkontaktierungsfläche 93-1 on a relative rotation of the planet gear 34 to the ring gear 68 to each other and can form a radial bearing between ring gear 68 and planet gear 34.

According to FIG. 11, the planetary and ring gear contacting surfaces 93 which correspond to one another axially on both sides of the meshing toothings can be formed, for example in such a way that the two-sided arrangement tion of the mutually corresponding Planetenrad- and Hohlradkontaktierungsflä- surfaces 93 in addition to the radial bearing provided by them also an axial bearing between ring gear 68 and planet gear 34 is formed. This can be achieved by the contacting surfaces 93-1 and / or 93-2 are formed by peripheral surfaces or lateral surfaces of Radialborden axially adjacent the meshing teeth. In other words, for example, the Planetenradkontaktierungsfläche 93-2 by means of an axially offset from the planetary gear arranged and radially outwardly facing or projecting or projecting annular shoulder (radial board) of the planetary gear 34 are formed. The Planetenradkontaktie- surface 93-2 may then correspond, for example, an outer circumferential surface of this annular shoulder. Analogously, in addition or as an alternative, the corresponding ring gear contacting surface 93-1 can be formed by means of a radially inwardly pointing or projecting annular shoulder (radial flange) of the ring gear 68 arranged axially offset from the ring gear teeth. The ring gear contacting surface 93-1 may then correspond, for example, to an inner circumferential surface of this annular shoulder on which the Planetenradkontaktierungsfläche 93-2 can roll in operation. By lateral rims in addition to the toothing can be achieved in addition to the radial bearing and an axial definition of the meshing teeth or the related components.

Although a rolling and a force transmission in the contact region of the planetary gear 34 and the ring gear 68 already takes place between the respective gears, so that a solution to increase the carrying capacity would be to make the gears correspondingly wider. However, the presented solution with the cylindrical rolling surfaces 93-1, 93-2 offers the following advantages:

While the toothing between planetary gear 34 and ring gear 68 is designed primarily with an engagement angle of approximately 20 ° to the transmission of tangential forces, the transmission between the described bearing surfaces 93-1, 93-2 takes place in the radial direction. Radial loads thus lead to lower loads.

• As a result of elastic deformations under load and speed as well as by manufacturing inaccuracies, it can lead to a misalignment between the gears of the planetary gear 34 and ring gear 68, which reduces an effective load capacity with wide gears, so that a further widening of the teeth only is not very effective. With the bearing surfaces 93-1, 93-2 directly next to the toothing is counteracted, since these tilting moments can be supported around the teeth. • Compared to a classic radial bearing at the usual bearing point 72, the advantage of a much shorter power flow, since the expected loads mainly by centrifugal forces on the heavy parts on the larger radii - ie in the vicinity of the ring gear teeth - arise. The tolerance chains are also correspondingly short.

FIG. 12 shows different variants A, B, C, how a combination of toothing and bearing raceway or contacting surfaces 93-1, 93-2 can be implemented constructively. While on the left a possible construction of the ring gear 68 is shown with Hohlrad- contacting surface 91 -1, the corresponding construction of the planetary gear 34 is shown with Planetenradkontaktierungsfläche 93-2 right to it. Although it is in principle possible, as shown in Figure 1 1, contacting surface 93-1 or 93-2 and toothing to produce each one component, production technology, however, the division into two components may be useful, which then with conventional joining methods such as presses, Gluing or welding can be joined together. Such two-part designs between contacting surface and toothing are each shown in FIG.

In all variants A, B, C, at least one Planetenradkontaktierungsfläche 93-2 is formed by means of an axially offset from the planetary gear toothing and radially outwardly facing annular shoulder 94 of the planetary gear 34. In variant A annular shoulder or board 94 and planet 34 are formed in two parts, while it is in the variants B and C respectively one-piece or one-piece constructions of planet or carrier and board 94. Likewise, in all variants A, B, C, at least one ring gear contacting surface 93-1 is formed by means of an annular shoulder 97 of the ring gear 68 arranged axially offset from the ring gear teeth and pointing radially inward. The radial extent of the annular shoulders 94, 97 may be different in each case pronounced and ranging from a complete radial overlap with the adjacent teeth up to no radial overlap. It should always be ensured the mountability. With the exception of the planetary gear 34 of the variant A, thus, the ring gear 68 and the planetary gear 34 each split into a flange-shaped carrier with the track 93 and the respective gear itself, wherein a board 94, 97 is formed by a flange. In the case of the planetary gear 34 according to variant A, a bearing ring can be placed on an offset axially on the toothing on the gear to form the shelf 94. Variants A and B are also only supported on one side, variant C on both sides. For reasons of mountability, the diameter of the respective raceway should only be smaller on one side of the ring gear 68 and on the planetary gear 34 on one side only greater than the root diameter of the respective toothing. In this way, planetary gear 34 and ring gear 68 can be pushed axially into one another during assembly, for example. The raceways 93-1, 93-2 themselves can be designed as cylindrical surfaces, or - be provided with the goal of a more even load distribution (edge pressing) with a logarithmic profile.

FIG. 13 shows a possible supplementary or alternative axial bearing 101 of the components, which are connected to the output side 52 of the phase shifter 42, opposite the output 24, 86 to the transmission. In particular, in vehicles with rear longitudinal driveline, a deceleration of the vehicle may cause these parts generate a force in the direction of the engine due to their inertia in their axial bearing and within the elasticity of the thrust bearing and the parts and their connection to the camp itself, move in that direction. In addition to a high bearing load it could also lead to collisions with other modules, such as the converter housing 95, if the elasticities allow too much movement. For this reason, it may be useful to introduce at least one additional support or bearing point 101 in this axial direction. Such additional support or bearing points 101 can be arbitrarily combined with all embodiments.

In one possible form, the additional thrust bearing 101 is formed by a corresponding design of the planet carrier 24 and the axially adjacent and radially overlapping support plate 71, which in turn is rotatably coupled to the guide plate or ring gear 52. Likewise, the leadership be further radially inwardly drawn so that it is arranged axially adjacent to the Planetenrad- carrier 24 or other output-side gear carrier (eg, a Ab- Abtriebshohlradträger) and radially overlapping. According to some embodiments, therefore, the planetary gear 34 can be supported by a Planetenrad- carrier 24. The ring gear 68 may be coupled to a carrier 52 or 71 arranged offset axially with respect to the planetary gear carrier 24, the planet carrier 24 and / or the carrier 52 or 71 being formed such that between the planet carrier 24 and the carrier 52 or 71 when an axial force threshold value (which is determined, for example, by a mass of the transducer 90) is exceeded, at least one defined axial bearing point results for the mutual axial support. In this case, the ring gear carrier 52 or the component 71 coupled thereto can form a secondary side of a phase shifter 42 arranged in a first torque transmission path 18-1, while the planet carrier 24 couples the coupling transmission 20, 30 with the output in the direction of the transmission. Furthermore, the ring gear carrier 52 or the component 71 coupled thereto can be coupled in a torque-proof manner to a turbine wheel 75 of a torque converter 90.

To form the additional axial bearing 101, the parts 24 and / or 71 can be formed on a radius, circumferentially or at least at several positions of its circumference, such that a defined point, line, and / or surface contact occurs, particularly when the vehicle is decelerating between the parts 24, 71 results. This can, for example, as shown, be solved so that the planet carrier 24 in the corresponding radial overlap region has a support 71 facing the planar surface and the carrier 71 has pronounced, for example, beads 96 in the direction of the mating surface.

As an alternative or in addition to the described embodiment of the bearing surfaces with the beads 96 on the carrier plate 71, the bearing 101 can also be optimized for friction and / or wear, for example by using additional components (bearings) made of special functional materials. This includes in particular the use of bearing elements made of plastic or metal with particularly favorable Reibungsund wear behavior that can be used between the two parts 24 and 71. Also would be a wear-reducing coating on the components 24th and / or 71 in the appropriate area. The defined Axiallagerstellen 101 may thus have a material with a lower coefficient of friction and less wear than adjacent thereto sites. However, in order to keep the friction of the system low in normal operation, these bearing surfaces should only touch if an overload of the thrust bearing or a collision actually threatens. Accordingly, a gap between the bearing surfaces of the two parts 24, 71 can be provided, which is bridged depending on the rigidity of the carrier 71 only when a defined axial force is exceeded.

In the foregoing, an additional axial bearing point 101 was used to present a solution to the problem that high axial forces can act in the direction of the motor due to an axial thrust of the turbine 75 during converter operation. However, a high axial force is not only important in terms of the bearing load, but may also lead to deformations of the turbine plate and the components attached thereto. This is critical if this results in an inadmissibly high tilting or displacement in the toothing of the power split transmission 20, 30.

In addition or as an alternative to the additional bearing point 101 for limiting the deformation, it is also possible to solve the problem by decoupling an axial freedom of movement or bearing of the turbine 75 and the input ring gear 68 from one another. In Figure 14, this is implemented such that the turbine 75 is still stored in the already known manner with respect to the output and the freewheel through the bearing 72, but via a spline 102 between support plate 71 and turbine 75 with the carrier 52 of the input ring gear 68 is rotatably connected to this, but no longer axially determined. Some exemplary embodiments therefore provide that a turbine wheel 75 of a torque converter 90 arranged axially next to the torsional vibration damping arrangement 100 is non-rotatably coupled about the rotational axis A to a first output region 52, 71 of the first torque transmission path 18-1 of the torsional vibration damping arrangement 100 via a spline 102. As a result, the rotational inertia of the turbine 75 remains usable as part of the intermediate mass for the phase shift, but no forces can be transmitted in the axial direction, so that the input ring gear 68 remains unaffected by an axial thrust of the turbine 75. However, since the axial position of the input ring gear 68 now also can no longer be determined via the axial bearing 72, an additional axial bearing is necessary, which can be provided by previously described annular shoulders 94, 97 and corresponding Wälzflächen 93-1, 93-2 in addition to the teeth between planetary and Eingangsshohlrad 34, 68. In this case, additional axial contact surfaces in the planetary gear 34 and the ring gear 68 are integrated by the annular shoulders 94, 97. In order to obtain a thrust bearing acting on both sides, it is also possible to add an axial bearing disk 103 which, for example, can be firmly connected to the ring gear 68 by a conventional joining method. The thrust bearing 103 can serve as a thrust washer to the axially immediately adjacent board 94 of the planetary gear 34.

The features disclosed in the foregoing description, the appended claims and the appended figures may be taken to be and effect both individually and in any combination for the realization of an embodiment in its various forms.

Although some aspects have been described in the context of a device, it should be understood that these aspects are also a description of the device

represent appropriate method, so that a block or a component of a device is to be understood as a corresponding method step or as a feature of a method step. Analogous to this, aspects that appear in the

Related to a or as a method step also described a description of a corresponding block or detail or feature of a corresponding device.

The embodiments described above are merely illustrative of the principles of the present invention. It will be understood that modifications and variations of the arrangements and details described herein will be apparent to others of ordinary skill in the art. Therefore, it is intended that the invention be limited only by the scope of the appended claims and not by the specific details presented in the description and explanation of the embodiments herein. Reference number, 10 'conventional torsional vibration damping drive unit

output unit

entrance area

Antriebssonnenradträger

torque transmission

crankshaft

coupling arrangement

first coupling arrangement input part

Planet'Hohlradträger

second coupling arrangement input part

input sun gear

planetary gear

first planetary gear

second planetary gear, planetary gear

output portion

output ring

output range

Vibration system, phase shift array primary mass

input element

spring assembly

output element

intermediate element

Additional mass element

mass pendulum

Outside spring set

Inside spring set

guide plate

radial bearings

hub disc Transducer lockup clutch Clutch drive drive Transducer clutch output Drive stop element

stator

connecting plate

drive internal gear

connecting bolts

torsion stop

support plate

thrust bearing

impeller

turbine

additional mass

thrust washer

bearing disk

bearing bolt

axial space

toothed segment

Mass base

rivet

flap

output flange

Gleitbahnblech

axial groove

actuating piston

torque converter

Freewheel. rolling elements

Outer ring radial bearing

contacting surface

Planet ring shoulder, board converter housing 96 bead

97 Hohlradringschulter, board

98 axial bolts

99 piston carrier

100 torsional vibration damping arrangement

101 axial bearing

102 spline

103 Thrust bearing washer

Claims

claims

1 . Planetary gear (30), with the following features:

at least one planet gear (34) having a planet gear for meshing with a ring gear (68); and

a ring gear (68) having a ring gear meshing with the pinion gear of the pinion gear (34),

characterized in that in the planetary gear (34) axially offset from the Planetenradverzahnung a Planetenradkontaktierungsfläche (93-2) is formed, and that in the ring gear (68) axially offset from the ring gear teeth to the Planetenradkontak- tierungsfläche (93-2) corresponding Hohlradkontaktierungsfläche (93-1) is formed, wherein the Planetenradkontaktierungsfläche (93-2) and the Hohlradkontaktierungsfläche (93-1) at a relative rotation of the planetary gear (34) to the ring gear (68) has a radial bearing between ring gear (68) and planet gear (34) form.

2. Planetary gear (30) according to claim 1, characterized in that axially on both sides of the meshing toothings mutually corresponding Planetenrad- and Hohlradkontaktierungsflächen (93) are formed, such that by the two-sided arrangement of the mutually corresponding Planetenrad- and Hohlradkontaktierungsflächen ( 93) in addition to the radial bearing an axial bearing between the ring gear and planetary gear is formed.

3. planetary gear (30) according to claim 1 or 2, characterized in that a toothing and an axially offset contacting surface (93) of the planetary gear (34) and / or the ring gear (68) are integrally formed or multi-piece.

4. Planetary gear (30) according to one of the preceding claims, characterized in that the Planetenradkontaktierungsfläche (93-2) arranged by means of an axially adjacent to the Planetenradverzahnung and radially outwardly facing annular shoulder (94) of the planetary gear (34) and / or the corresponding thereto Hohlradkontaktierungsfläche (93-1) by means of an axially offset from the ring gear arranged radially inwardly facing annular shoulder (97) of the ring gear (68) is formed.

5. Planetary gear (30) according to one of the preceding claims, characterized in that the planetary gear (34) by a Planetenradtrager (24) is supported and that the ring gear (68) from a relative to the Planetenradtrager (24) axially offset ring gear (52 71), wherein the Planetenradtrager (24) and / or the Hohlradträger (52; 71) is formed such that between the Planetenradträger (24) and the Hohlradträger (52; 71) when exceeding a Axialkraftschwellenwerts at least one defined Axiallagerstelle (101) for mutual axial support.

6. planetary gear (30) according to claim 5, characterized in that the defined Axiallagerstellen (101) have a material with a lower coefficient of friction and / or less wear than adjacent thereto points.

7. planetary gear (30) according to any one of the preceding claims, characterized in that a first axis of rotation (A) of the ring gear (68) and a second axis of rotation (B) of the planet gear (34) extend obliquely in a plane spanned by the two axes of rotation plane ,

8. Planetary gear (30) according to any one of the preceding claims, characterized in that the planetary gear (34) has a first Planetenradteil (81 -1) with a first tooth diameter and a second Planetenradteil (81 -2) with a second, different from the first Gear diameter includes.

9. planetary gear (30) according to claim 8, characterized in that the ring gear (68) with the first Planetenradteil (81 -1) and a sun gear (28) with the second Planetenradteil (81 -2) meshes.

10. Planetary gear (30) according to any one of claims 8 or 9, characterized in that the Planetenradteile (81 -1, 81 -2) in the direction of the second axis of rotation (B) in the axial direction offset from one another.

1 1. Planetary gear (30) according to any one of claims 8 to 10, characterized in that the first Planetenradteil (81 -1) by a first circular segment of the planetary gear (34) with the first gear diameter and the second Planetenradteil (81 -2) by a second circular segment of the planet gear (34) is formed with the second gear diameter.

12. Planetary gear (30) according to any one of the preceding claims, characterized in that the ring gear (68) with a first torque transmission path (18-1) of a torsional vibration damping arrangement (100), a sun gear (28) with a second torque transmission path (18-2) the torsional vibration damping arrangement (100) is coupled, and the planetary gear (30) forms a coupling arrangement (20) for superimposing on the torque transmission paths (18-1, 18-2) guided torques.

The planetary gear (30) according to claim 12, characterized in that the first torque transmission path (18-1) comprises a phase shifter assembly (42) for producing a phase shift of rotational irregularities directed over the first torque transmission path (18-1) with respect to the second torque transmission path (18-2) conducted rotational irregularities.

14, planetary gear (30) according to claim 13, characterized in that the phase shifter assembly (42) has a primary side and with this by a spring arrangement (57) coupled secondary side, wherein the secondary side formed by a one-piece grounding body (82) for providing a moment of inertia is, wherein in the ground body (82) the ring gear for combing with the planet gear (34) is formed.

15. planetary gear (30) according to claim 14, characterized in that the spring arrangement (57) radially on the ground body (82) is supported.

16. Planetary gear (30) according to claim 14 or 15, characterized in that the ground body (82) in a spring channel of the spring assembly (57) has projecting webs which serve as a spring of the spring arrangement as stops in the circumferential direction.

17, planetary gear (30) according to any one of claims 14 to 1 6, characterized in that a hub disc (61) engages the primary side from radially inward to radially outward into the spring arrangement (57) of the phase shifter assembly (42).

18. Planetary gear (30) according to claim 12, characterized in that a turbine wheel (75) of a torque converter (90) arranged axially next to the torsional vibration damping arrangement (I0O) can be rotationally fixed about an axis of rotation via a spline (102). A) is coupled to a first output region (52; 71) of the torsional vibration damping arrangement (100) with respect to the first torque transmission path (18-1).

19. Planetary gear (30) according to claim 18, characterized in that the first output region of the torsional vibration damping arrangement (100) comprises a carrier (52; 71) for the ring gear (68), wherein the carrier (52; 71) via the spline (102 ) is rotatably coupled to the turbine wheel.

20. Motor vehicle with a planetary gear (30) according to one of the preceding claims.