CN114630977A

CN114630977A - Power split type stepless transmission device

Info

Publication number: CN114630977A
Application number: CN202180006009.1A
Authority: CN
Inventors: 菲利普·雷兴巴赫; 拉斐尔·希默尔斯巴赫; 米夏埃尔·韦克斯; 约翰尼斯·卡尔滕巴赫; 斯特凡·贝克; 马克斯·巴赫曼; 伊丽莎白·恩德尔; 塞缪尔·威廉斯
Original assignee: ZF Friedrichshafen AG
Current assignee: ZF Friedrichshafen AG
Priority date: 2020-03-17
Filing date: 2021-03-10
Publication date: 2022-06-14
Also published as: DE102020203391A1; WO2021185643A1

Abstract

The invention relates to a power split continuously variable transmission (100; 200; 300; 400; 500; 600; 700; 800; 900; 1000) having a variator (16). The first planetary set has a first sun gear (30), a first carrier (32), respective first planet gears (36), and a first ring gear (34). The second planetary gear set has a second sun gear (40), a second planet carrier (42), second and third planet gears (46, 48) and a second ring gear (44). The first and second planet carriers (32, 42) are connected in a rotationally fixed manner to one another. The second planet wheels (46) are in mesh with the second sun wheel (40) and the second ring gear (44). The second and third planet gears (46, 48) are each in meshing pair with one another. The first and third planet gears (36, 48) are each coupled to one another in a rotationally fixed manner. The variator (16) is mechanically coupled to the drive (12) and to the first sun gear (30). The drive end (12) can be mechanically operatively connected to the first toothed ring (34).

Description

Power split type stepless transmission device

Technical Field

The invention relates to a power split type stepless transmission device. The invention also relates to a drive unit having such a power split continuously variable transmission.

Background

Power split continuously variable transmissions, such as hydrostatic mechanical power split transmissions, are often used in the field of work machines. An example of a work machine is a tractor. Such hydrostatic mechanical transmissions allow for stepless adjustment of the transmission ratio at relatively high efficiency, but have a relatively low transmission ratio. The transmission ratio value of a transmission is understood to be the ratio between the maximum and minimum transmission ratios. In order to increase the transmission ratio, it is known to provide a plurality of driving ranges in the continuously variable transmission, which have different transmission ratio ranges. Such a transmission is known, for example, from DE 19741510 a 1.

Disclosure of Invention

A first aspect of the invention relates to a power split continuously variable transmission. The transmission may be configured for use in a work machine. The transmission ratio can be adjusted steplessly in a continuously variable transmission. The power split may be a hydrostatic mechanical power split, for example. Alternatively or additionally, for example, electromechanical power splitting is also possible. The transmission is capable of achieving high transmission ratio ratios with high efficiency. The transmission comprises a drive end, at which a variable to be transmitted is fed into the transmission. The transmission may have a drive shaft, which may form the drive end at one end. The drive shaft may extend through the transmission and enable extraction of power output power at an end remote from the drive end, for example for driving a work implement of a work machine. The transmission also includes a driven end at which the variable transmitted by the transmission is output. At the output end, for example, a torque for the driving drive, in particular for the rotary drive of the respective wheel of the work machine, can be provided. The driving end and the driven end may be arranged axially parallel to each other. However, other designs are also conceivable, such as a shaft arrangement.

The transmission has a variator. The variator can be designed for the stepless variation of the transmission ratio, for example, within the driving range. The driving range of the transmission can be understood as the state in which a fixed mechanical transmission ratio exists between the input and output of the transmission, wherein the transmission ratio of the transmission can be varied steplessly by a variator within the driving range. The variator can have two energy converters, which are designed, for example, as hydraulic actuators. The energy converter of the variator can however also be designed as an electric motor. The energy converter can be formed coaxially with the drive end of the gear. It is also conceivable for them to be formed parallel to the drive end axis. The first energy converter may for example be a hydraulic machine with a fixed quantity of feed per revolution and the second energy converter a hydraulic machine with an adjustable quantity of feed per revolution.

The transmission has a planetary assembly. The planetary assembly has a first and a second planetary gear set, each of which has a sun gear, a planet carrier with planet gears rotatably mounted thereon, and a ring gear. The first and second planetary gear sets are designed, for example, as negative planetary gear sets. The negative planetary gear set has a negative dead axle gear ratio.

If two elements are mechanically coupled, they are directly or indirectly coupled to each other such that movement of one causes the other to react. The mechanical connection can be provided, for example, by a form-locking or friction-locking connection. The mechanical connection can correspond, for example, to the meshing of corresponding teeth of the two elements. In the case of two elements which mesh with one another, their toothing is, for example, in engagement. The two elements may roll over each other. Between these elements, further elements, for example one or more spur gear stages, can be arranged. In contrast, a permanent rotationally fixed connection of two elements is understood to mean a connection in which the two elements are rigidly coupled to one another in all normal states of the transmission. In this case, these elements can be present as single parts connected to one another in a rotationally fixed manner or can also be present in one piece.

If a clutch is provided between two elements of the transmission, the rotary elements are not permanently connected to one another in a rotationally fixed manner, but can be connected to one another in a rotationally fixed manner via the clutch. The clutch may be a switching element. The anti-rotation connection is only brought about by actuating the clutch located in the middle. In this case, actuating the clutch means that the clutch is transferred into the closed state, so that the components directly coupled to the clutch are balanced with respect to their rotational movement. The respective switching element can be switched, for example, at least between an open and a closed state. If the shift element concerned is designed as a form-locking shift element, the structural elements connected in a rotationally fixed manner directly to one another via the shift element concerned operate at the same rotational speed. In the case of friction-locking shift elements, a rotational speed difference may occur between the components even after actuation thereof. Nevertheless, this desired or undesired state is referred to within the scope of the invention as a rotationally fixed connection of the respective structural elements. In a friction-locking connection, for example, a certain rotational speed difference between two elements connected to one another may occur due to slip. In this case, despite the slip, the rotational speeds of the two components are generally approximately the same. Accordingly, such a connection is considered herein to be anti-rotation despite the slip.

The respective planetary gear sets described herein may be devoid of additional elements other than those described herein, such as additional ring gears, planet carriers, planet gears, and sun gears. The planetary assemblies and the gear mechanisms have, for example, no further planetary gear sets of the planetary gear sets described here. The respective ring gear can be designed as a housing element of the transmission or of the planetary assembly. In the transmission, it can be provided that the planetary components or all planetary wheel sets thereof are designed as planetary rollers. The planetary rollers may be compact and robust. The design as a planetary roller can mean that the planetary assembly has only coaxial elements. In addition, the design as a planetary roller can mean that the planetary assembly is free of cylindrical gears. The planetary assembly can be arranged coaxially with the driven end, for example.

The first planetary gear set is composed of a first sun gear, a first carrier rotatably supporting a set of first planetary gears, and a first ring gear. The second planetary gear set is composed of a second sun gear, a second planet carrier and a second ring gear, which rotatably support a set of second planet gears and a set of third planet gears. The second planet gears (which mesh with the second sun gear, the second ring gear, and the third planet gears) have a larger pitch diameter and a shorter axial length than the third planet gears. The pitch circle diameter of a gear is the tooth diameter defined in the literature. This diameter describes the average diameter of the teeth. The planet wheels of a set of planet wheels can have, for example, respective axes of rotation arranged on the same circumference, alternatively or additionally have respective functional connections of the same type, alternatively or additionally be designed as a common part. The respective sets of planet gears are arranged, for example, at regular intervals from one another on the circumference of the transmission.

The first and second planet carriers are permanently connected to each other in a rotationally fixed manner. The two planet carriers can be formed, for example, in one piece or as a common component. The first planet gear meshes with the first sun gear and the first ring gear. The second planet gear is meshed with the second sun gear and the second ring gear. The second planet gears and the third planet gears are respectively meshed with each other in pairs. The first planet wheel and the third planet wheel are each permanently rotationally fixed to one another in pairs. A corresponding rotation of the first planet wheel can be achieved if the respective third planet wheel rolls over the respective second planet wheel.

The variator is mechanically operatively coupled to the drive end and the first solar turbine. The drive end can be mechanically operatively connected to the first toothed ring. A particularly high transmission ratio is achieved by these connections of the planetary assemblies. The transmission ratio is greater than, for example, an embodiment in which the drive input can be mechanically operatively connected to an element of a second planetary gear set having two planetary gear sets. The transmission is configured to transmit torque from the second planetary gear set to the driven end. The transmission can take place mechanically, for example, via the respective hollow shaft, the spur gear stage, the shift element and the further planetary gear set.

All clutches of the transmission according to the invention can be designed as friction-locking clutches, for example as diaphragm clutches. Likewise, some or all of the clutches can also be designed as form-locking clutches, for example as claw clutches or synchronizers. A synchronizer is understood as a positive clutch between two shafts, wherein the shafts are put into operation simultaneously, i.e. synchronized, before the positive connection is established.

A power split transmission allows multiple ranges of travel to be provided with fewer shift elements and thus a larger transmission ratio. In this case, the transitions between the driving ranges can be synchronized. This can mean that, when the lower or upper limit of the rotational speed of the respective driving range is reached, the next driving range can be engaged without an adjustment being carried out on the variator. The variator does not have to change its rotational speed or position, for example, in order to carry out a driving range change. Furthermore, the variators may also be in respective maximum positions at one or both limits of the respective driving range. In the respective driving range, a small number of clutches can be in their respective disengaged position, as a result of which drag losses are reduced and transmission efficiency can be increased.

In one embodiment of the transmission, it can be provided that the respective first planet gears and the respective third planet gears are each designed as long planet gears in pairs. The long planetary gear can be mechanically engaged with two axially spaced elements, for example. The long planetary gear can have elements for the operative connection with other elements in two axial regions on the outer circumference, for example. For example, the long planetary gear can have teeth at two opposite ends for mechanical engagement. However, the long planetary gear may also have successive teeth which are in mechanical engagement with the first and second further elements in different axial regions. The long planetary gear can be produced for example inexpensively in one piece. To reduce weight, the middle section may be milled down into the shaft connection bracket. The long planetary gear can also be formed, for example, by two toothed wheels, which are connected in a rotationally fixed manner to the hollow shaft of the long planetary gear. The two gears are rotatably mounted on the associated planet carrier by means of the hollow shaft. The long planetary gear can have the same working diameter or even the same toothing in the two regions for the mechanical functional connection, for example. A large degree of nesting can be provided by the long planet gears, whereby the transmission can be particularly compact. Furthermore, the long planetary gear can be manufactured inexpensively. Alternatively, the respective first planet gears and the respective third planet gears can each be designed as stepped planets in pairs.

In one embodiment of the transmission, it can be provided that the transmission has only two first planet gears. Alternatively or additionally, the transmission can have only two second planet gears. Alternatively or additionally, the transmission may have only two third planet gears. By providing only two planet wheels, the transmission can be made more compact. The second and third set of planet wheels can thus simply be arranged in the radial direction between the second sun wheel and the second ring gear. The second planet gears and the third planet gears can therefore be arranged in a good radial overlap, i.e. for example at least partially on the same circumference. The transmission can, for example, have only two long planetary gears, which form the two first planetary gears and the two third planetary gears.

In one embodiment of the transmission, it can be provided that the second ring gear can be mechanically operatively connected to the output side by means of the first clutch. The second ring gear can thus form an output shaft of the planetary assembly, which can be used in a switchable manner to provide the driven variable at the output end. The first clutch can be arranged coaxially with the drive input or the drive output, for example. The first clutch can be arranged axially on the output side relative to the planetary assembly, for example. The first clutch can be mechanically operatively connected to the second ring gear, for example, by means of a first spur gear stage. The first spur gear stage can be formed, for example, in a single stage. The first clutch can be mechanically operatively connected to the output, for example, via the countershaft, for example, exclusively by means of the third spur gear stage or alternatively by means of the third or fourth spur gear stage. The selection of the active connection can be effected by means of a further clutch. The third spur gear stage can be designed, for example, in a single stage or in two stages. The fourth spur gear stage can be designed, for example, as a single stage or as a double stage. The mechanical coupling via the third spur gear stage may cause the driven end to rotate in the opposite direction compared to the coupling via the fourth spur gear stage.

In one embodiment of the transmission, it can be provided that the second sun gear can be mechanically connected to the output side by means of a second clutch. The second sun gear can thus form an output shaft of the planetary assembly, which output shaft can be used in a switchable manner to provide the driven variable at the output end. The second clutch can be arranged coaxially with the drive input or the output, for example. The second clutch can be arranged axially on the output side, for example, relative to the planetary assembly and alternatively or additionally relative to the first clutch. The first clutch can be mechanically connected to the second sun gear, for example, by means of a second spur gear stage. The second spur gear stage can be formed, for example, in a single stage. The second clutch can be mechanically connected to the output side, for example, exclusively by means of the third spur gear stage or alternatively by means of the third or fourth spur gear stage, for example, via an intermediate shaft. In this case, a third and a fourth spur gear stage and an intermediate shaft may also be provided, which mechanically connect the third clutch to the output. However, additional spur gear stages and additional countershafts are also possible. The selection of the active connection can be effected by means of a further clutch. The third spur gear stage can be of single-stage or double-stage design, for example. The fourth spur gear stage can be designed, for example, as a single stage or as a double stage. The mechanical coupling via the third spur gear stage may cause the driven end to rotate in the opposite direction compared to the coupling via the fourth spur gear stage. The first and second spur gear stages may have the same gear ratio. The second spur gear stage can be arranged axially on the output side relative to the first spur gear stage and the second clutch. The third spur gear stage may be arranged on the driven side along the axis relative to the second spur gear stage. The fourth spur gear stage may be arranged axially on the driven side relative to the third spur gear stage.

The first clutch and the second clutch can be designed as a common double shift element. The dual switching element may be compact and simple in construction. The double switching elements can be actuated simply by means of a common single actuator, for example. The double shifting element can, for example, as a shift state, alternatively provide an active connection of the shifting element formed by one of the two double shifting elements, i.e., for example, an active connection of the first or second clutch that has been closed. In one embodiment, the double switching element may have a neutral position. The neutral position may correspond to, for example, the first and second clutches being open. The first clutch and the second clutch may be nested axially or radially. The first and second clutches can be mechanically operatively connected to the output side together via one or alternatively via optionally selected spur gear stages and an alternative or additional intermediate shaft.

In one embodiment of the transmission, it can be provided that the second planet carrier can be mechanically operatively connected to the output side by means of a third clutch. The second planet carrier can thus form an output shaft of the planetary assembly, which can be used to provide the driven variable at the output side in a switchable manner. The third clutch may be arranged coaxially with the drive or driven end, for example. The third clutch can be arranged axially on the output side, for example, relative to the planetary assembly and possibly relative to the first and second clutches. The third clutch may be connected to the second planet carrier, for example, in a rotationally fixed manner. The third clutch may be connected to the driven end, for example, in a rotationally fixed manner. Accordingly, the second planet carrier can be connected in a rotationally fixed manner to the output side by means of a third clutch.

In one embodiment of the transmission, it can be provided that the second sun gear can be mechanically connected to the output side by means of a fourth clutch. The second sun gear can thus form the output shaft of the planetary assembly, which can be used to provide the driven variable at the output side in a switchable manner. For example, the mechanical operative connection of the second sun gear can alternatively be provided by means of a second and a fourth clutch. The mechanical operative connection via the second and fourth clutches may, for example, have different gear ratios, so that an additional driving range may be provided. One or more spur gear stages can be provided, for example, in the active connection by means of the second clutch. The second sun gear can be connected, for example, in a rotationally fixed manner to the output side by means of the operative connection of the fourth clutch. The fourth clutch may be arranged coaxially with the drive or driven end, for example. The fourth clutch can be arranged axially on the output side, for example, relative to the planetary assembly and possibly relative to the first, second and third clutches. The fourth clutch may be connected to the second planet carrier, for example, in a rotationally fixed manner. The fourth clutch may be connected to the driven end, for example, in a rotationally fixed manner.

The third clutch and the fourth clutch may be designed as a common double shift element. The third clutch and the fourth clutch may be nested axially or radially. The third and fourth clutches can be mechanically operatively connected to the output side together via one or alternatively via optionally selected spur gear stages and an alternative or additional countershaft.

In one embodiment of the transmission, it can be provided that the transmission has a driving direction changing arrangement which is designed to change the direction of rotation of the driven end relative to the direction of rotation of the driven end in at least one transmission driving range. This allows, for example, the direction of travel to be adjusted. For this purpose, the drive direction changing assembly can have, for example, at least one shift element and a spur gear stage or a planetary gear set. The driving direction changing assembly can be arranged on the drive side, for example. In the case of a drive-side arrangement, the drive input can be mechanically operatively connected to the first ring gear, for example, by means of a drive direction changing assembly. In the case of a drive-side arrangement, the torques acting on the drive direction changing assembly and in particular on its switching elements are low. In the case of an arrangement on the driven end side, the planetary assembly can be mechanically operatively connected to the driven end, for example, by means of a drive direction changing assembly. In the case of an arrangement on the output side, the driving direction changing assembly can be integrated more easily into the transmission if necessary. In the case of a driving direction changing assembly arranged on the output side or on the drive side, the driving direction changing assembly can be designed to provide all forward driving ranges as well as reverse driving ranges. In the case of an arrangement on the driven end side, the first, second, third and fourth clutches can be mechanically operatively connected to the driven end via the drive direction changing assembly. In contrast, in the case of a nested arrangement, the load advantage of the driving direction changing assembly can be achieved, wherein, however, it may not be possible to provide all forward driving ranges and all reverse driving ranges.

In one embodiment of the transmission, it can be provided that the drive direction changing arrangement has a single-stage first spur gear stage, a two-stage second spur gear stage, a first clutch and a second clutch. The drive input can be mechanically operatively connected to the output in at least one driving range by means of a first clutch of the driving direction changing arrangement via a single-stage first spur gear stage and in at least one other driving range by means of a second clutch of the driving direction changing arrangement via a double-stage second spur gear stage. The respective clutch and spur gear stage of the drive direction changing assembly can be a different clutch from the first to fourth clutches and spur gear stages already described (which serve to connect the respective output shaft of the planetary assembly to the output side).

For example, a first clutch of the drive direction changing assembly can be used to establish an operative connection via a first spur gear stage for providing the respective forward driving range. For example, a second clutch of the drive direction changing assembly can be used to establish an operative connection via a second spur gear stage for providing the respective reverse driving range. The respective clutch of the drive direction changing assembly can be arranged coaxially with the drive input or the output. In the case of a coaxial arrangement with the drive end, the pressure oil feed can be simpler. For example, a hollow shaft, which may be necessary, for example, if the planetary assembly is arranged coaxially with the planetary assembly, can be dispensed with, and oil can be fed from the housing via the first sun gear to the first ring gear.

The first and second clutches of the drive direction changing assembly can be designed as a common double shift element. The first and second clutches of the direction change assembly may be nested axially or radially. The first and second clutches of the drive direction changing assembly can be mechanically operatively connected to the output side together via one or alternatively via an optionally selected spur gear stage and an alternative or additional intermediate shaft.

The respective clutch of the drive direction changing assembly can be arranged coaxially with the planetary assembly. For example, a common part of the first and second clutches, which connects the output shaft of the planetary assembly to the output side, can be used in the clutch. The first and second clutches can be designed as common parts, for example, with the first and second clutches of the drive direction changing assembly.

The first and second clutches of the drive direction changing assembly can be arranged axially between the respectively associated spur gear stages. The first clutch of the driving direction changing assembly can be arranged axially on the drive side or on the output side relative to the second clutch of the driving direction changing assembly, for example. If their associated clutches are arranged, the respectively associated spur gear stage of the two clutches of the direction-of-travel changing assembly can be arranged axially opposite the other clutch of the direction-of-travel changing assembly. The first spur gear stage of the driving direction changing assembly can be arranged, for example, on the drive side along the axis relative to the first clutch of the driving direction changing assembly, the first clutch of the driving direction changing assembly being arranged on the drive side axially relative to the second clutch of the driving direction changing assembly, and the second spur gear stage of the driving direction changing assembly being arranged on the driven side axially relative to the second clutch of the driving direction changing assembly. This axial arrangement can also be reversed, for example.

In one embodiment of the transmission, it can be provided that the second ring gear can be mechanically operatively connected to the output by means of the drive direction changing arrangement via the first clutch. The second sun gear can be mechanically operatively connected to the output by means of the driving direction changing assembly via the second clutch. The second planet carrier can be connected with the driven end in a rotation-proof manner by means of a third clutch. The second sun wheel can be connected with the driven end in a rotation-proof manner by means of a fourth clutch. In such a nested arrangement, for example, the driving range which can be provided only by the closed first and second clutches, respectively, can be provided not only as a reverse driving range but also as a forward driving range. For example, only two reverse driving ranges can be set. The driving direction changing assembly can therefore be arranged axially on the output side, but nevertheless only a lower torque is applied than, for example, an arrangement on the other output side.

In one embodiment of the transmission, it can be provided that the second ring gear can be connected in a rotationally fixed manner to the countershaft by means of the first clutch. The second sun wheel can be connected in a rotationally fixed manner to the intermediate shaft by means of a second clutch. The intermediate shaft can be mechanically connected to the output shaft by means of a fourth clutch. This design allows the first to fourth clutches to be arranged coaxially with the planetary structural assembly. The first to fourth clutches may be formed as planetary rollers together with the planetary assembly.

In one embodiment of the transmission, it can be provided that the transmission has a third sun gear, a third planet carrier and a third ring gear, which form a third planetary gearset. The third planet carrier can be rotatably mounted with the respective planet wheels associated therewith. The transmission may have a brake. The intermediate shaft can be connected in a rotationally fixed manner to the third planet carrier by means of a fourth clutch. The second planet carrier can be connected to the third planet carrier in a rotationally fixed manner by means of a third clutch. The intermediate shaft can be permanently connected to the third sun gear in a rotationally fixed manner. The third ring gear can be fixed to the stationary component by means of a brake. The third planet carrier can be permanently connected to the output side in a rotationally fixed manner. Thus, the third planetary gear set and the brake may provide additional ranges of travel. Furthermore, the third planetary set can also be designed as part of a planetary roller. The third planetary gearset may be configured as a negative planetary gearset. The brake can be designed as a form-locking or friction-locking shift element. The component can be fixed by means of a brake so that it can no longer rotate. The stationary component may be, for example, a component that does not move relative to the rotating elements of the respective planetary gear sets. The stationary component can be configured, for example, as a housing.

Furthermore, a synergistic effect is achieved when the electric machine is used as part of a variator, which is mechanically operatively connected to the drive input. The motor can for example replace a hydraulic machine with variable feed per revolution. The driven end can be purely electrically driven. For this purpose, for example, if the driving direction changing assembly is arranged on the drive side, the first and second clutches of the driving direction changing assembly are closed, for example. They thus block the first ring gear, so that the first sun gear forms the input shaft of the planetary gear set. The driving range can additionally be provided by engaging at least one further shift element, for example a first or second clutch and an additional brake, in order to be able to transmit torque to the output. In addition, the motor may require an external energy supply, such as a battery. An electrically driven pump may be provided in order to ensure lubrication and cooling oil supply. Which can take over the supply of lubricating and cooling oil in the event of a stoppage of the motor connected to the drive end.

In one embodiment of the transmission, it can be provided that the planetary assembly has a central shaft for the supply of lubricating oil. The central shaft can be configured, for example, as a solid shaft with respective lubricant ducts. The central shaft can be connected to the first sun gear, for example, permanently against rotation, or can be formed by the first sun gear. Such a central shaft can be supported in particular simply. The central shaft can be connected to the second planet carrier, for example, permanently fixed in rotation, or can be formed by the second planet carrier. The second planet carrier can rotate all the time in all driving ranges when the motor is running, thereby facilitating the supply of lubricating oil.

In one embodiment of the transmission, it can be provided that the fixed-axis transmission ratio of the first planetary gear set is from-2.7 to-3. Furthermore, the dead axle ratio of the second planetary gear set may be-3.5 to-3.8. The layout of the planetary structural assembly with the first and second planetary gear sets requires complex coordination. The fixed-axis gear ratio can realize excellent transmission gear ratio under the condition of higher efficiency. The dead axle ratio of the first planetary gear set-2.83 and the dead axle ratio of the second planetary gear set-3.6 are very suitable. An advantageous coordination is achieved in the design of the planetary assembly with only two first planet wheels, two second planet wheels and three third planet wheels. Suitable tooth numbers, alternative or additional modules for the respective gears of the planetary gearset are already known from the front sectional views of the first planetary gear set and the second planetary gear set.

In order to engage the driving range and to control the various clutches and brakes, the transmission can have a control device designed for this purpose. The control device can be designed and set up to control the respective states of the clutch and the brake. The control device may have a respective actuator for this purpose and alternatively or additionally a position sensor and alternatively or additionally a rotational speed sensor.

A second aspect relates to a drive unit having a power split continuously variable transmission according to the first aspect. The drive unit has a motor which is mechanically operatively connected to the drive end for driving the drive end. The motor may be, for example, an internal combustion engine. The drive unit may be adapted to drive the motor at a substantially constant rotational speed. The respective driving speed and the selectable driving direction can be controlled, for example, by means of the selected driving range and the variator. Another aspect relates to a work machine having a power split continuously variable transmission according to the first aspect and a drive unit according to the second aspect.

Drawings

Fig. 1 shows a schematic view of a transmission according to a first embodiment;

FIG. 2 shows a switching matrix of the transmission according to FIG. 1;

fig. 3 shows a schematic perspective view of a long planet;

FIG. 4 shows a front sectional view through a second planetary gear set of the transmission;

FIG. 5 shows a front cross-sectional view of a first planetary gear set through the transmission;

fig. 6 shows a schematic view of a transmission according to a second embodiment;

fig. 7 shows a schematic view of a transmission according to a third embodiment;

fig. 8 shows a schematic view of a transmission according to a fourth embodiment;

fig. 9 shows a schematic view of a transmission according to a fifth embodiment;

fig. 10 shows a schematic view of a transmission according to a sixth embodiment;

fig. 11 shows a schematic view of a transmission according to a seventh embodiment;

fig. 12 shows a schematic view of a transmission according to an eighth embodiment;

FIG. 13 shows a switching matrix of the transmission according to FIG. 2;

fig. 14 shows a schematic view of a transmission according to a ninth embodiment;

fig. 15 shows a schematic view of a transmission according to a tenth embodiment;

fig. 16 shows a schematic view of a transmission according to an eleventh embodiment;

fig. 17 shows a schematic view of a transmission according to a twelfth embodiment;

fig. 18 shows a schematic representation of a transmission according to a thirteenth embodiment.

Detailed Description

Fig. 1 shows a schematic representation of a first embodiment of a power split continuously variable transmission 100 having a drive input 12, a drive output 14, a variator 16 and a planetary assembly 18. The planetary assembly 18 has two planetary gear sets and is used to provide different driving ranges. The drive end 12 is operatively connected to a motor. The transmission 100 has a drive shaft that extends through the transmission 100. At the driven end side end portion, the drive shaft constitutes a power take-off shaft 20. The output shaft 14 is arranged axially parallel to the drive shaft 12 and can be operatively connected to a differential, for example.

The planetary assembly 18 is currently designed as a planetary roller.

The first planetary gear set is formed by a first sun gear 30, a first carrier 32 and a first ring gear 34. A set of first planet gears 36 is rotatably mounted on the first carrier 32. The second planetary gear set of the planetary assembly 18 is formed by a second sun gear 40, a second planet carrier 42 and a second ring gear 44. A second planetary gear set 46 and a third planetary gear set 48 are rotatably mounted on the second planetary gear carrier 42. The second planet gears 46 (which mesh with the second sun gear 40, the second ring gear 44, and the third planet gears 48) have a larger pitch circle diameter and a shorter axial length than the third planet gears 48. The pitch circle diameter of a gear is the tooth diameter defined in the literature. This diameter describes the average diameter of the teeth.

First carrier 32 and second carrier 42 are permanently rotationally fixed to one another. The first planet gears 36 mesh with the first sun gear 30 and the first ring gear 34. This is also shown in the front sectional view of the first planetary gear set according to fig. 5. Fig. 5 also shows that the group of first planet wheels 36 has only two first planet wheels 36.

The second planet gears 46 mesh with the second sun gear 40 and the second ring gear 44. The second planetary gear 46 and the third planetary gear 48 mesh with each other in pairs. This is shown in a front sectional view of the second planetary gear set according to fig. 4. It can be seen here that the set of second planet wheels 46 and the set of third planet wheels 48 each have only two gears. The second planet 46 has a larger effective diameter than the third planet 48. The first planet 36 and the third planet 48 have the same effective diameter. The second and third

planetary gears

46, 48 are each arranged side by side in the circumferential direction. The rotational axis of the second planet 46 is arranged on the same circumference as the rotational axis of the third planet 48. However, they may also be arranged on different circumferences and thus radially spaced apart from the central axis.

The first planet 36 and the third planet 48 are each permanently connected to one another in pairs in a rotationally fixed manner. In the transmission 100, for this purpose, the first and third planet gears 48 are each formed in pairs as a common long planet gear, as shown in fig. 3. In the embodiment shown, the long planetary gear is formed in one piece, wherein the respective teeth at its two end regions are of identical design. The teeth at the respective end regions can be made, for example, together on one gripper. The intermediate region without teeth can be made before the teeth are made. The teeth of the long planetary gear at one end region mesh with the first sun gear 30 and the first ring gear 34. The long planet is in toothed engagement with one of the second planet wheels 46 in the opposite end region. Based on the long planet gears, the transmission has a total of only four planet gear members, namely two long planet gears and two second planet gears 46. As shown in fig. 3, the long planetary gear can be embodied as a spur planetary gear, but there is also the possibility of the long planetary gear being embodied as a helical planetary gear.

At least for the transmission 100, the front sectional views according to fig. 4 and 5 are at least to scale with respect to the respective tooth parameters. In particular, in fig. 4 and 5, the modules and the numbers of teeth of the elements of the two planetary gear sets are proportional and shown correctly with respect to their number.

The variator 16 has a first energy converter 80 and a second energy converter 82, which are coupled to each other. The two

energy converters

80, 82 are designed as hydraulic machines, wherein the first energy converter 80 has a fixed feed per revolution and the second energy converter 82 has a variable feed per revolution. The first energy converter 80 is mechanically operatively connected to the first sun gear 30 in order to transmit torque between the first sun gear 30 and the energy converter 80 and thus the variator 16. In the exemplary embodiment shown, the connection is indirect, i.e. by means of a single spur gear stage 84. One of the spur gears is connected to the shaft of the first energy converter 80 for this purpose in a rotationally fixed manner. For this purpose, the other spur gear is connected in a rotationally fixed manner to the first sun gear 30. In this case, the spur gear stage 84 is arranged axially on the drive side relative to the planetary assembly 18. The second energy converter 82 is mechanically operatively connected to the drive end 12 to transmit torque between the drive end 12 and the second energy converter 82 and thus the variator 16. In the exemplary embodiment shown, the connection is indirect, i.e. by means of a single spur gear stage 86. The spur gear stage 86 is arranged axially on the output side, in particular at the level of the power take-off shaft 20 and thus with respect to the planetary assembly 18. The spur gear stage 86 is formed by two meshing spur gears, one of which is designed as a fixed gear and is connected in a rotationally fixed manner to the shaft of the second energy converter 82. The other spur gear is designed as a counter gear and is permanently connected to the drive end 12 in a rotationally fixed manner.

The drive end 12 is mechanically connected to the first ring gear 34 via a drive direction changing assembly 50. The travel direction changing assembly 50 includes a first clutch KV and a second clutch KR. The two clutches KV, KR are arranged coaxially with the drive input 12 and are connected directly thereto. The drive end 12 is mechanically operatively connected to the first ring gear 34 by means of the clutch KV via the single-stage spur gear stage 52. One gear of the spur gear stage 52 is permanently connected in a rotationally fixed manner to the first ring gear 34, while the other gear of the spur gear stage 52 can be connected in a rotationally fixed manner to the drive end 12. Spur gear stage 50 is arranged axially between spur gear stage 84 and the first set of planet gears. The drive input 12 is mechanically operatively connected to the first ring gear 34 via a double-stage spur gear stage 54 by means of a clutch KR. The fixed gear of the spur gear stage 54 is permanently connected in a rotationally fixed manner to the first ring gear 34, while the other gear can be mechanically connected to the drive input 12 by means of a clutch KR. Radially between these two gearwheels, a further gearwheel of the spur gear stage 54 is arranged, which meshes with these two gearwheels.

The spur gear stage 54 of the drive direction changing assembly 50 is arranged axially on the output side relative to the spur gear stage 84 that mechanically couples the variator 16 to the first sun gear. The clutch KR is arranged axially on the driven side relative to the spur gear stage 54. The clutch KV is arranged on the driven side in the axial direction with respect to the clutch KR. The clutch KV and the clutch KR are nested axially and form a double shift element. The spur gear stage 52 is arranged axially on the output side relative to the clutch KV, which is arranged axially on the drive side relative to the first planetary gear set.

The transmission 100 is configured for transmitting torque from the second planetary gear set to the driven end 14. For this purpose, the second sun gear 40 forms the output shaft of the planetary assembly 18 and is connected to the fourth friction-locking clutch K4. The second sun gear 40 is connected in a rotationally fixed manner to the output 14 by means of a fourth clutch K4. The second planet carrier 42 forms a further output shaft of the planetary gear set 18 and is connected to a third friction-locking clutch K3. The second planet carrier 42 is connected in a rotationally fixed manner to the output side 14 by means of a third clutch K3. The third and fourth clutches K3, K4 are axially nested and coaxially disposed with the drive end 14. The third and fourth clutches K3, K4 form double shift elements. The third clutch K3 is disposed axially on the drive side relative to the spur gear stage 86 that connects the variator 16 to the drive end 12. The fourth clutch K4 is disposed on the driven side in the axial direction with respect to the third clutch K3.

The second ring gear 44 forms a further output shaft of the planetary assembly 18. The second ring gear 44 is mechanically operatively connected to the first friction-locking clutch K1 via the single-stage first spur gear stage 60. For this purpose, the gearwheels of the first spur gear stage 60 are permanently connected in a rotationally fixed manner to the second ring gear 44 and the further gearwheels of the first spur gear stage 60 are connected to the first clutch K1. The first spur gear stage 60 can be operatively connected to the countershaft 70 by means of the first clutch K1. The countershaft 70 is mechanically connected to the driven end 14 by the third spur gear stage 64. The intermediate shaft 70 is disposed coaxially with the drive end 12. One gear of third spur gear stage 64 is permanently connected in a rotationally fixed manner to countershaft 70 and the other gear of third spur gear stage 64 is permanently connected to driven end 14. Accordingly, second ring gear 44 is mechanically connectable to driven end 14 via first spur gear stage 60 and third spur gear stage 64 by means of first clutch K1.

The second sun gear 40 as the output shaft of the planetary assembly 18 can also be mechanically connected to the output side 14 by means of a second friction-locking clutch K2. One gear of the second spur gear stage 62 of the single stage is permanently connected in a rotationally fixed manner to the second sun gear 40, while the other gear of the second spur gear stage 62 is permanently connected in a rotationally fixed manner to the second clutch K2. The second spur gear stage 62 can be connected to the countershaft 70 by means of a second clutch K2. Correspondingly, second sun gear 40 can be operatively connected to output 14 via second spur gear stage 62, and countershaft 70 and third spur gear stage 70 can be operatively connected by means of a second clutch K2.

The first spur gear stage 60 is arranged axially on the output side relative to the second planetary gear set and axially on the drive side relative to the first clutch K1. The first clutch K1 is arranged on the drive side in the axial direction relative to the second clutch K2. The first and second clutches K1, K2 are coaxially disposed with the drive end 12 and are axially nested. The first and second clutches K1, K2 form a double shift element. The second spur gear stage 62 is arranged axially on the output side relative to the second clutch K2 and axially on the drive side relative to the fourth clutch K4. The third spur gear stage 64 is arranged axially on the output side relative to the third clutch K3 and axially on the drive side relative to the spur gear stage 86 that mechanically couples the variator 16 to the drive input 12.

The friction clutches K1, K2, K3, K4, KV and KR are designed as disk clutches. The respective double shifting element allows an operative connection to be optionally provided by means of a respective one of the two clutches forming them. In one embodiment of the double switching element, it also allows a neutral position to be provided.

Fig. 2 shows a switching matrix of the transmission 100 according to fig. 1. The corresponding columns show the switching states of the elements that should be switched. Transmission 100 has four forward ranges of travel, which are indicated by FB1, FB2, FB3, and FB 4. The driving ranges are numbered in order of their highest speed. That is, the travel range FB2 is capable of achieving higher speeds than the travel range FB1, while the travel range FB3 has higher speeds than FB1 and FB 2. FB4 is the fastest forward driving range. The transmission 100 likewise has four reverse travel ranges FB1R, FB2R, FB3R and FB 4R. The travel range FB4R is the fastest reverse travel range, and is similar in rank to the forward travel range according to speed. The speed range is shown next to the column identifying the driving range. The driving ranges FB1 and FB1R start at 0 km/h. In the case of the limit speed a, switching can be made between the driving range FB1 and the driving range FB 2. The same applies to the subsequent driving range and its maximum or minimum speed. The driving speed in the backward driving range is the same as the driving speed in the forward driving range, wherein, however, the driven end 14 rotates in the opposite direction with the rotational direction of the drive end 12 remaining unchanged.

The forward driving range is provided in principle by the closed clutch KV of the driving direction changing assembly 50. The reverse driving range is provided in principle by the closed clutch KR of the driving direction changing assembly 50. The driving range FB1 is provided by the additionally closed first clutch K1. The driving range FB2 is provided by the additionally closed second clutch K1. The driving range FB3 is provided by the additionally closed third clutch K3. The driving range FB4 is provided by the additionally closed fourth clutch K4. The reverse driving range is provided by means of the respective closed clutch partners, which are identical to the forward driving range with the same numbering, wherein for this purpose the clutch KR is closed instead of the clutch KV of the driving direction changing assembly 50. The transmission 100 can provide four forward driving ranges and four reverse driving ranges with less mechanical effort. The subsequent driving range can be switched by engaging and disengaging the clutch.

In the transmission 100, the fixed-axis gear ratio of the first planetary gear set is-2.9 and the fixed-axis gear ratio of the second planetary gear set is-3.7.

Further figures show further embodiments of the power split continuously variable transmission. Only the differences relating to each of the other embodiments are discussed here. Thus, components having the same function and also having the same design have the same reference numerals and will not be described further. The switching matrix is also applicable to the other embodiments unless otherwise specified. The respective differences can also be combined with one another in other embodiments. The additional planetary gear set and the additional brake of the transmission according to fig. 12 can also be combined, for example, with a system of drive direction changing assemblies of the transmission according to fig. 6 and 7.

Fig. 6 shows a second embodiment of the transmission 200. This embodiment differs from the transmission 100 in that the axial positioning of the clutches KV and KR and the spur gear stages 52 and 54 of the drive direction changing assembly 50 is interchanged. The clutch KV is now arranged axially on the drive side relative to the clutch KR. The double-stage spur gear stage 54 is now arranged axially on the output side relative to the clutch KR. The single-stage spur gear stage 52 is now arranged axially on the drive side relative to the clutch KV.

Fig. 7 shows a third embodiment of the transmission 300. The difference between this embodiment and the transmission 100 lies primarily in the positioning of the clutches KV, KR of the drive direction changing assembly 50. In the transmission 300, the two clutches KV, KR are arranged coaxially with the driven end 14 and thus with the planetary assembly 18. Accordingly, the clutch KR is connected to the drive input 12 by means of a double-stage spur gear stage 54. Thus, the fixed gear of the spur gear stage 54 is permanently connected to the drive end 12 in a rotationally fixed manner. Accordingly, the clutch KV is connected to the drive input 12 by means of a single-stage spur gear stage 52. Thus, the fixed gear of the spur gear stage 52 is permanently connected to the drive end 12 in a rotationally fixed manner. The axial system of the components of the direction of travel changing assembly 50 corresponds to the axial system of the transmission 100.

The arrangement of the drive direction changing assembly 50 in the transmission 300 is such that its respective clutch KV, KR has a plurality of common parts in relation to the double shifting element (which is formed by the first and second clutches K1, K2). Furthermore, the respective clutches KV, KR of the drive direction changing assembly 50 are loaded with a lower torque and are implemented smaller than the corresponding clutches KV, KR in the transmission 100. The transmission 300 is provided with additional pressure oil feed portions to the clutches KV and KR of the traveling direction changing structural assembly 50, as compared with the transmission 100. The additional pressurized oil supply extends in the transmission 300 through the housing, the central shaft of the first sun gear 30 and the hollow shaft of the first ring gear 34.

Fig. 8 shows a fourth embodiment of a transmission 400. This embodiment differs from transmission 300 in the positioning of clutches KV, KR and spur gear stages 52 and 54 of drive direction changing assembly 50. They are positioned the same as in the transmission 200 and the description made there applies equally to the transmission 400.

Fig. 9 shows a fifth embodiment of the transmission 500. This embodiment shows that the direction change assembly is docked on the driven side instead of the drive side.

In the transmission 500, the drive input 12 is mechanically connected to the first ring gear 34 by means of a spur gear stage 502. One gear of the spur gear stage 502 is permanently connected to the drive input 12 in a rotationally fixed manner, and the other gear of the spur gear stage 502 is connected to the first ring gear 34. The spur gear stage 502 is arranged axially on the driven side relative to the spur gear stage 84 that mechanically connects the variator 16 to the first sun gear 30. The spur gear stage 502 is arranged axially on the drive side relative to the first planetary wheel set.

The clutches KV, KR of the drive direction changing assembly are arranged coaxially with the driven end 14 and the planetary assembly 18. The second planet carrier 42 can be connected in the transmission 500 in a rotationally fixed manner to the further countershaft 504 by means of a third clutch K3. The second sun gear 40 can be connected in the transmission 500 in a rotationally fixed manner to the further countershaft 504 by means of a fourth clutch K4. The intermediate shaft 504 can be connected to the driven end 14 in a rotationally fixed manner by means of a clutch KV of the drive direction changing assembly. The second planet carrier 42 can therefore be connected in a rotationally fixed manner to the output side 14 via the third clutch K3 and the clutch KV.

In the transmission 500, the single-stage spur gear stage 52 forms a mechanically active connection between the intermediate shaft 70 and a further intermediate shaft 504. One gear of the single-stage spur gear stage 52 is permanently connected in a rotationally fixed manner to the countershaft 70, and the other gear is connected to the other countershaft 504. The second ring gear 44 can thus be mechanically operatively connected to the output via the first spur gear stage 60, the first clutch K1, the intermediate shaft 70, the single-stage spur gear stage 52, the further intermediate shaft 504 and the clutch KV. The second sun gear 40 can thus be mechanically operatively connected to the output via the second spur gear stage 60, the second clutch K2, the countershaft 70, the spur gear stage 52, the further countershaft 504 and the clutch KV. One gear of the two-stage spur gear stage 54 is permanently connected in a rotationally fixed manner to the countershaft 70. The other gear of the two-stage spur gear stage 54 is connected to the clutch K3 of the drive direction changing assembly.

In the transmission 500, the third spur gear stage 64 is replaced by a single-stage spur gear stage 52 and a two-stage spur gear stage 54, the drive input 12 additionally being connected via a spur gear stage 502. The operative connection can be optionally adjusted in order to provide a corresponding forward driving range and reverse driving range.

The single-stage spur gear stage 52 is arranged on the driven side with respect to the third clutch K3. The clutch KV is arranged on the output side relative to the single-stage spur gear stage 52. Between the single-stage spur gear stage 52 and the clutch KV, a further countershaft 504 is arranged coaxially with the driven end 14. The clutch KR is arranged on the driven side with respect to the clutch KV. The two-stage spur gear stage 54 is arranged on the driven side with respect to the clutch KR.

The so-called power reversal of the work machine can be improved by the design of the transmission 500. The work machine can be driven forward and, when changing from forward to reverse, can utilize the clutch KV or KR in order to reverse the drive power of the motor connected to the drive end 12. This may cause a rotational direction on driven end 14 that is opposite to the current direction of travel. The clutches KV and KR of the drive direction changing assembly 50 are subjected to a higher torque in the transmission 500 than in the transmission 100 and are therefore designed to be larger.

Fig. 10 shows a sixth embodiment of a transmission 600. In this embodiment, only two reverse travel ranges are provided. The driving direction changing assembly 500 is embodied in an interposed arrangement.

The clutch KV of the driving direction changing assembly 50 is connected to the intermediate shaft 70. The clutch KV is connected to the output side 14 via an associated single-stage spur gear stage 52. Accordingly, one gear of the single-stage spur gear stage 52 is permanently connected in a rotationally fixed manner to the output side 14, while the other gear of the single-stage spur gear stage 52 can be connected to the countershaft 70 by means of the clutch KV.

The clutch KR of the drive direction changing assembly 50 is connected to the intermediate shaft 70. The clutch KR is connected to the output side 14 via an associated double-stage spur gear stage 54. Accordingly, one gear of the double-stage spur gear stage 54 is permanently connected in a rotationally fixed manner to the output shaft 14, while the other gear of the double-stage spur gear stage 54 can be connected to the countershaft 70 by means of the clutch KR.

Thus, the first and second clutches K1, K2 can be mechanically connected to the output side 14 by means of the clutch KV of the drive direction changing assembly 50 in order to provide the variable to be transmitted as the forward drive range. Thus, the first and second clutches K1, K2 can be mechanically operatively connected to the output 14 by means of the clutch KR of the drive direction changing assembly 50 in order to provide the variable to be transmitted as the reverse drive range. Accordingly, the transmission 600 may provide the reverse travel ranges FB1R and FB 2R.

In contrast, the third and fourth clutches K3, K4 are directly connected to the driven end 14. Furthermore, the drive input 12 is mechanically connected to the first ring gear 34 by means of the spur gear stage 602. The direction of rotation of the respective driving ranges provided by the third and fourth clutches K3, K4 cannot be reversed in the transmission 600 at the output side 14. Accordingly, while the forward travel ranges FB3 and FB4 are provided in the transmission 600, the corresponding reverse travel ranges FB3R and FB4R are not provided.

The drive direction changing assembly 50 is located on the output side, but is connected only to the first and second clutches K1, K2. As a result, the drive direction changing assembly 50 is subjected to a lower torque than, for example, the transmission 500, so that the clutches KV and KR are implemented smaller. For this purpose, however, only two reverse driving ranges are present.

The spur gear stage 602 is arranged axially between the spur gear stage 84, which mechanically connects the variator 16 to the first sun gear 30, and the first planetary gear set. One gear of the cylindrical gear stage 602 is permanently connected in a rotationally fixed manner to the first ring gear 34, and the other gear of the cylindrical gear stage 602 is permanently connected in a rotationally fixed manner to the drive end 12. The clutches KV and KR of the drive direction changing assembly 50 are arranged coaxially with the drive input 12. The single-stage spur gear stage 52 is arranged axially on the output side relative to the third clutch K3 and axially on the drive side relative to the clutch KV. The clutch KR is arranged on the driven side in the axial direction with respect to the clutch KV. The two-stage spur gear stage 54 is arranged axially on the driven side relative to the clutch KR and axially on the drive side relative to the spur gear stage 86 that connects the variator 16 to the drive end 12. In the transmission 600, the spur gear stages 52, 54 of the drive direction changing arrangement 50 are instead transmitted by means of the action according to the third spur gear stage 64 in the transmission 100.

Fig. 11 shows a seventh embodiment of a transmission 700. This embodiment is based on a transmission 600 in which, however, the KV/KR spur gear chain is replaced. The transmission 700 differs from the transmission 600 in the same way as the transmission 200 differs from the transmission 100, whereby the axial positioning of the clutches KV and KR and the spur gear stages 52, 54 of the drive direction changing assembly 50 is reversed.

In the transmission 700, the double-stage spur gear stage 54 is arranged axially on the output side relative to the third clutch K3 of the drive direction shifting assembly 50 and axially on the drive side relative to the clutch KR. The clutch KR is arranged axially on the drive side relative to the clutch KV of the driving direction changing assembly 50. The single-stage spur gear stage 52 is arranged axially on the output side relative to the clutch KV and axially on the drive side relative to the spur gear stage 86 that mechanically connects the variator 16 to the drive input 12.

Fig. 12 shows an eighth embodiment of the transmission 800. This embodiment differs from the first embodiment in the design and arrangement of the respective clutches K1, K2, K3, K4, so that these clutches are explained in their entirety despite their respective commonality. Furthermore, the transmission 800 has an additional third planetary gear set, which forms planetary rollers together with the planetary assembly 18, and a brake B. The connection of the variator 16 and the direction of travel changing assembly 50 is the same as in the transmission 100 and will not be described further herein. In the transmission 800, the dead axle ratio of the first planetary gear set is-2.78, the dead axle ratio of the second planetary gear set is-3.60, and the dead axle ratio of the third planetary gear set is-3.60. The fixed-axis transmission ratio of the second planetary gear set is the same as that of the third planetary gear set

The second ring gear 44 can be connected in a rotationally fixed manner to the intermediate shaft 70 by means of a first clutch K1, wherein the intermediate shaft 70 is arranged coaxially with the output end 14 in the transmission 800. The second sun gear 40 can be connected in a rotationally fixed manner to the intermediate shaft 70 by means of a second clutch K2. The first clutch K1 and the second clutch K2 are again designed as double shift elements, but are radially nested in the transmission 800 and are formed coaxially with the output side 14 and the planetary assembly 18. The first clutch K1 is arranged radially outside with respect to the second clutch K2. The intermediate shaft 70 can be mechanically connected to the output side 14 by means of a fourth clutch K4.

The fourth clutch K4 is arranged on the driven side in the axial direction with respect to the first and second clutches K1, K2. The first and second clutches K1, K2 are arranged axially on the output side relative to the planetary assembly 18. The intermediate shaft 70 extends axially from the first and second clutches K1, K2 to the fourth clutch K4.

The transmission 800 has a central shaft for the lubricant oil supply, which central shaft is formed by the first sun gear 30 in the transmission 800. The other portions of the planetary rollers are supported on the central shaft. By using the first sun gear 30 as a central shaft, the central shaft is simply supported.

The transmission 800 has a third sun gear 850, a third planet carrier 852 and a third ring gear 854, which form a third set of planet gears. Rotatably mounted on the third planet carrier 852 are respective planet gears 856. The third ring gear 854 can be fixed to a stationary member by means of the brake B. The countershaft 70 can be connected in a rotationally fixed manner to the third carrier 852 via a fourth clutch K4. The second planet carrier 42 can be connected in a rotationally fixed manner to the third planet carrier 852 by means of a third clutch K3 via a hollow shaft 858. The hollow shaft 858 is disposed radially outwardly relative to the intermediate shaft 70. The intermediate shaft 70 is permanently connected to the third sun gear 850 for rotation therewith. The third carrier 852 is permanently connected to the driven end against relative rotation.

The third planetary gear set is configured as a negative planetary gear set. The respective planet gears 856 of the third set mesh with a third sun gear 850 and a third ring gear 854.

The third and fourth clutches K3, K4 are again designed as double shift elements, wherein they are nested radially in the transmission 800. The third and fourth clutches K3, K4 are arranged on the drive side in the axial direction with respect to the third planetary gear set. The third clutch K3 is arranged radially outside with respect to the fourth clutch K3. The brake B is arranged axially on the output side relative to the third ring gear 854 and axially on the drive side relative to the spur gear stage 86 that mechanically couples the variator 16 to the drive input 12.

The transmission 800 is extended from the transmission 100 by the third planetary gear set and the brake B. In the transmission 800, a higher final speed is achieved within the respective forward driving range. In the transmission 800, the respective reverse travel ranges do not correspond exactly mirror-imaged with respect to their speeds. In the illustrated embodiment, the reverse driving range has a lower final speed than the corresponding forward driving range. This makes it possible to provide a higher torque in the rear driving range.

In the transmission 800, the direction of travel changing assembly 50 is arranged on the drive side as in the transmission 100. Thus, clutches KV and KR are loaded with less torque and they can be very small. The high torque only loads the third planetary set, which is optimal for component optimization. There, larger shafts and gears may be provided which are required for higher torques.

Fig. 13 shows a switching matrix of the transmission 800 according to fig. 12. The basic structure of the switching matrix is the same as the switching matrix of fig. 2 and the categorization and labeling of the various ranges of motion. The respective speeds identified with double lines are smaller than the corresponding speeds identified with single lines. The maximum speed A 'of the reverse travel range FB1R of the transmission 800 is slower than the maximum speed A' of the forward travel range FB 1. The speed identified with a single reticle is higher compared to the corresponding speed without reticle identification within the switching matrix according to fig. 2. The speed identified with double markings is lower compared to the corresponding speed without marking in the switching matrix according to fig. 2. The speeds A of FB1 and FB1R in transmission 100 are slower than the speed A 'of FB1 of transmission 800 and faster than the speed A' of FB1R of transmission 800.

In the transmission 800, each driving range is provided by three closed shift elements. The driving range FB1 is provided by the closed first clutch K1, the closed brake B and the closed clutch KV. The driving range FB2 is provided by the closed second clutch K2, the closed brake B and the closed clutch KV. The driving range FB3 is provided by the closed second clutch K2, the closed third clutch K3 and the closed clutch KV. The driving range FB4 is provided by the closed second clutch K2, the fourth clutch and the closed clutch KV. The reverse driving ranges FB1R to FB4R are provided with the closed clutch KR instead of the closed clutch KV, and are provided with the same closed shift element. In the transmission 800, only one shift element is switched on and off each time a continuous driving range is changed.

Fig. 14 shows a ninth embodiment of a transmission 900. This embodiment is based on the eighth embodiment according to fig. 12. The switching matrix according to fig. 13 applies in the same way to the transmission 900.

Transmission 900 differs from transmission 800 in that it is used as a central shaft for the supply of lubricating oil. In the transmission 800, the central shaft is permanently connected in a rotationally fixed manner to the second planet carrier 42 and thus also to the first planet carrier 32, in that the central shaft is formed by the second planet carrier 42. The second carrier 42 is always rotated, and the intermediate rotation speed of the first sun gear 30 in each driving range is 0. For the lubrication oil supply, a rotating shaft is preferred compared to a stationary shaft.

Fig. 15 shows a tenth embodiment of the transmission 1000. This embodiment is based on the eighth embodiment according to fig. 12. The transmission 1000 differs from the transmission 800 in the energy converter of its variator 16. The switching matrix according to fig. 13 applies in the same way to the transmission 1000.

In the transmission 1000, the variator 16 has a first energy converter 1080 instead of the energy converter 80. Furthermore, the variator 16 has a second energy converter 1082, which replaces the energy converter 82. The two

energy converters

1080, 1082 are embodied as electric motors in the transmission 1000. The transmission 1000 is an electric power split transmission instead of a hydraulic power split transmission. The efficiency of the transmission 1000 is better than that of a hydraulic power split transmission. Furthermore, the electrical power of the intermediate circuit can be supplied to the additional consumers. For this reason, the costs and the installation space requirements are more favorable in a hydraulic power split transmission.

Furthermore, the transmission 1000 also has the possibility of driving the working machine in a purely electrical manner when the motor at the drive input 12 is stopped, since the drive-direction changing assembly 50 with the clutches KV, KR is provided on the drive side. For this purpose, in one embodiment, the respective driving range is provided by the first clutch K1 and the brake B which are engaged, and by a further closed shift element. In a further embodiment, the respective driving ranges are provided for this purpose alternatively or additionally by the second clutch K2 and the brake B which are engaged, and by a further closed shift element.

Fig. 16 shows an eleventh embodiment of the transmission 1100. The drive input 12 is connected to the drive output shaft 20, wherein the drive output shaft 20 is connected via a spur gear stage 86 to the variator 16, which in turn is connected via a spur gear stage 84 to the first sun gear 30. The drive input 12 can be connected to the first ring gear 34 via the clutch KR and the spur gear stage 54 of the drive direction changing assembly 50 or via the clutch KV and the spur gear stage 52 of the drive direction changing assembly 50. Here, the travel direction changing structural assembly 50 is disposed on the power output shaft 20. The second ring gear 44 is connected via a spur gear stage 60 to the first clutch K1, wherein the spur gear stage 60 can be connected via the first clutch K1 to the countershaft 70 and/or the second clutch K2. The second clutch K2 can be connected to the second sun gear 40 and the fourth clutch K4 via the spur gear stage 62. The fourth clutch K4 is connected to the third clutch K3, the third and fourth clutches K3, K4 in turn being connected to the output 14. The planet carrier of the second planetary gear 46 is connected to the third clutch K3. The countershaft 70 is connected via a spur gear stage 64 to a third clutch K3 and the output 14.

In contrast to the embodiment shown in fig. 1, it is achieved by the embodiment of the transmission 1100 shown in fig. 16 that a high degree of application to common parts is achieved with regard to the first and second clutches K1, K2, or the clutches KV, KR of the drive direction changing assembly 50, since identical or very similar clutches can now be used. Furthermore, the arrangement shown here makes it possible to implement the pressurized oil supply for actuating the clutches K1, K2, KV, KR with little complexity, as a result of which at the same time a higher efficiency of the transmission is achieved. A new arrangement can also be used to achieve a smaller installation space requirement, since the distance between the driver 12 and the driver 16 can be reduced. The means for matching the spacing between the driven end 14 and the driving end 12 is also less extensive because fewer components need to be changed.

Fig. 17 shows a twelfth embodiment of a transmission 1200. The drive input 12 is connected to the drive output shaft 20, wherein the drive output shaft 20 is connected to the variator 16 via the spur gear stage 86. The variator is in turn connected to the first sun gear 30 via a spur gear stage 84. The drive end 12 is connected to the first ring gear 34 via the spur gear stage 52. The first clutch K1, the second clutch K2, the third clutch K3 and the two clutches KV, KR of the drive direction changing assembly 50 are arranged on the intermediate shaft 70. Via a third clutch K3 and the spur gear stage 64, the intermediate shaft 70 can be connected to the planet carrier of the third planetary gear 46. Furthermore, the countershaft 70 can be connected to the output side 14 and the fourth clutch K4 via the clutch KR and the spur gear stage 54 of the drive direction changing assembly 50. The countershaft 70 can be connected to the output side 14 and the fourth clutch K4 via the clutch KV and the spur gear stage 56 of the drive direction changing assembly 50. The second ring gear 44 is connected via the spur gear stage 60 to the first clutch K1, wherein the second ring gear 44 can in turn be connected via the first clutch K1 to the countershaft 70 and/or the second clutch K2. The second clutch K2 is connected via the spur gear stage 62 to the second sun gear 40, wherein the second clutch K2 can be connected via the spur gear stage 58 to the fourth clutch K4 and thus to the output 14. The reduced length of the gear mechanism in the embodiment shown here proves to be particularly advantageous, i.e. the installation space between the gear mechanism input and the gear mechanism output is reduced.

Fig. 18 shows a thirteenth embodiment of a transmission 1300. The drive input 12 is connected to the drive output shaft 20, wherein the drive output shaft 20 is connected via a spur gear stage 86 to the variator 16, which in turn is connected via a spur gear stage 84 to the first sun gear 30. The drive input 12 can be connected to the first ring gear 34 via the clutch KR and the spur gear stage 54 of the driving direction changing assembly 50 or via the clutch KV and the spur gear stage 52 of the driving direction changing assembly 50. Here, the driving direction changing structural assembly 50 is arranged on the power take-off shaft 20. The second sun gear 40 is connected via the spur gear stage 62 to the second clutch K2 and can be connected via this to the first clutch K1 and/or the driven end 14. The second sun gear 40 can be connected to the fourth clutch K4 via the spur gear stage 64 and can be connected to the output 14 via this. The first, second, third and fourth clutches K1, K2, K3, K4 are arranged on the driven end 14. The intermediate shaft 70 connects the planet carrier of the third planet gear 46 via the spur gear stage 60 to the third clutch K3, wherein the intermediate shaft 70 can thus be connected to the output 14. The second ring gear 44 is connected via the spur gear stage 56 to the first clutch K1, whereby the second ring gear 44 can be connected to the output 14.

With the embodiment of the transmission 1300 shown here, in addition to the further advantage, the advantageous effect is also obtained that the output side 14, on which the first to fourth clutches K1, K2, K3, K4 are arranged, is operated at a lower rotational speed than in the embodiment according to fig. 16 if the highest speed is achieved by the transmission 1300. This increases the life of the components used (e.g., bearings), while reducing drag losses based on the differential speed otherwise produced across the first and second clutches. In addition, a greater degree of flexibility is achieved in the use of the transmission in vehicles, since a high degree of variability is achieved by means of the spur gear stages, or a series of axial distances between the drive end 12 and the driven end 14 can be achieved in a simple manner. In particular, a small spacing between the driving end 12 and the driven end 14 can thus be achieved. In the embodiment shown here, the first to fourth clutches K1, K2, K3, K4 can also be advantageously equipped with a high degree of common use, like the clutches KR, KV of the driving range changing assembly 50. Advantageously, the spur gear stages 56, 62 and the spur gear stages 60, 64 can be equipped with common parts in a corresponding pair-wise manner.

The embodiments according to fig. 16 to 18 operate according to the switching matrix of fig. 2. The embodiment of the transmission 1200 according to fig. 17 is subject to the limitation, however, that the driving range FB4R cannot be realized due to the arrangement of the fourth clutch with the gear set, so that only four forward driving ranges FB1 to FB4 and three reverse driving ranges FB1R to FB3R can be shown.

List of reference numerals

100; 200 of a carrier; 300, respectively; 400, respectively; 500, a step of; 600, preparing a mixture; 700 of the base material; 800; 900; 1000 driving device

12 drive end

14 driven end

16 changing device

18 planetary structural assembly

20 power output shaft

30, 40; 850 sun gear

32. 42; 852 planet carrier

34, 44; 854 toothed ring

36, 46, 48; 856 planet wheel

50 driving direction changing structural assembly

70 intermediate shaft

504 another intermediate shaft

80. 82; 1080, 1082 energy converter

52. 54, 56, 58, 60, 62, 64, 84, 86; 502; 602 cylindrical gear stage

858 hollow shaft

K1 first clutch

K2 second clutch

K3 third clutch

K4 fourth clutch

KV clutch

KR clutch

B brake

The claims (modification according to treaty clause 19)

1. Power split continuously variable transmission (1100) having a drive input (12), a drive output (14), a variator (16) and a planetary assembly (18), wherein the transmission (1100) is configured to provide different driving ranges, wherein the planetary assembly (18) has a first sun wheel (30) forming a first planetary wheel set, a first planet carrier (32) rotatably supporting a first set of planet wheels (36), a first ring gear (34), and a second sun wheel (40) forming a second planetary wheel set, a second planet carrier (42) rotatably supporting a second set of planet wheels and a third set of planet wheels (46, 48), and a second ring gear (44), wherein the first and second planet carriers (32, 42) are permanently connected to one another in a rotationally fixed manner, wherein the first planet wheels (36) mesh with the first sun wheel (30) and the first ring gear (34), wherein the second planet wheels (46) mesh with the second sun wheel (40) and the second ring gear (44), wherein the second planet wheels (46) and the third planet wheels (48) mesh with one another in pairs, wherein the first planet wheels (36) and the third planet wheels (48) are permanently rotationally fixed to one another in pairs, wherein the variator (16) is mechanically operatively connected to the drive input (12) and the first sun wheel (30), wherein the drive input (12) is mechanically operatively connectable to the first ring gear (34), and wherein the transmission (1100) is designed to transmit a torque from the second planetary wheel set to the drive output (14),

-the drive end (12) is connected to a power take-off shaft (20), wherein the power take-off shaft (20) is connected to the variator (16) via a spur gear stage (86), which variator in turn is connected to the first sun gear (30) via a spur gear stage (84);

-the drive end (12) can be connected to the first toothed ring (34) via a clutch (KR) and a spur gear stage (54) of a driving direction changing assembly (50) or via a clutch (KV) and a spur gear stage (52) of the driving direction changing assembly (50), wherein the driving direction changing assembly (50) is arranged on the power take-off shaft (20);

-the second ring gear (44) is connected to a first clutch (K1) via a spur gear stage (60), wherein the spur gear stage (60) can be connected to an intermediate shaft (70) and/or a second clutch (K2) via the first clutch (K1), wherein the second clutch (K2) can be connected to the second sun gear (40) and a fourth clutch (K4) via a spur gear stage (62);

-the fourth clutch (K4) is connected to a third clutch (K3), wherein the third and fourth clutches (K3, K4) are in turn connected to the driven end (14);

-the planet carrier of the second planet wheels (46) is connected to the third clutch (K3);

-the intermediate shaft (70) is connected to the third clutch (k3) and the driven end (14) via a spur gear stage (64).

2. Drive unit with a power split continuously variable transmission (1100) according to claim 1 and a motor which is mechanically operatively connected with the drive (12) for driving the drive (12).

3. Power split continuously variable transmission (1200) having a drive input (12), a drive output (14), a variator (16) and a planetary assembly (18), wherein the transmission (1200) is designed to provide different driving ranges, wherein the planetary assembly (18) has a first sun wheel (30) forming a first planetary wheel set, a first planet carrier (32) rotatably carrying a first set of planet wheels (36), a first ring gear (34), and a second sun wheel (40) forming a second planetary wheel set, a second planet carrier (42) rotatably carrying a second set of planet wheels and a third set of planet wheels (46, 48), and a second ring gear (44), wherein the first and second planet carriers (32, 42) are permanently connected to one another in a rotationally fixed manner, wherein the first planet wheels (36) mesh with the first sun wheel (30) and the first ring gear (34), wherein the second planet wheels (46) mesh with the second sun wheel (40) and with the second ring gear (44), wherein the second planet wheels (46) and the third planet wheels (48) mesh with one another in pairs, wherein the first planet wheels (36) and the third planet wheels (48) are permanently rotationally fixed to one another in pairs, wherein the variator (16) is mechanically operatively connected to the drive input (12) and the first sun wheel (30), wherein the drive input (12) is mechanically operatively connectable to the first ring gear (34), and wherein the transmission (1200) is designed to transmit a torque from the second planetary wheel set to the driven input (14),

-the drive end (12) is connected to a power take-off shaft (20), wherein the power take-off shaft (20) is connected to the variator (16) via a spur gear stage (86), which variator in turn is connected to the first sun gear (30) via the spur gear stage (84);

-the drive end (12) is connected to the first ring gear (34) via a cylindrical gear stage (52);

-a first clutch (K1), a second clutch (K2), a third clutch (K3) and two clutches (KV, KR) of the drive direction changing assembly (50) are arranged on an intermediate shaft (70), wherein the intermediate shaft (70) can be connected to the planet carrier of the third planetary gear (46) via the third clutch (K3) and the spur gear stage (64), and the intermediate shaft (70) can be connected to the output (14) and a fourth clutch (K4) via the clutch (KR) and the spur gear stage (54) of the drive direction changing assembly (50);

-the intermediate shaft (70) can be connected to the driven end (14) and the fourth clutch (K4) via a clutch (KV) and a spur gear stage (56) of the drive direction change assembly (50);

-the second ring gear (44) is connected via a spur gear stage (60) to the first clutch (K1), wherein the second ring gear (44) can in turn be connected via the first clutch (K1) to the countershaft (70) and/or the second clutch (K2), wherein the second clutch (K2) is connected via a spur gear stage (62) to the second sun gear (40), wherein the second clutch (K2) can be connected via a spur gear stage (58) to the fourth clutch (K4) and thus to the output (14).

4. Drive unit with a power split continuously variable transmission (1200) according to claim 3 and a motor which is mechanically operatively connected with the drive (12) for driving the drive (12).

5. Continuously variable power transmission (1300) having a drive input (12), a drive output (14), a variator (16) and a planetary assembly (18), wherein the transmission (1300) is designed to provide different driving ranges, wherein the planetary assembly (18) has a first sun wheel (30) forming a first planetary wheel set, a first planet carrier (32) rotatably carrying a first set of planet wheels (36), a first ring gear (34), and a second sun wheel (40) forming a second planetary wheel set, a second planet carrier (42) rotatably carrying a second set of planet wheels and a third set of planet wheels (46, 48), and a second ring gear (44), wherein the first and second planet carriers (32, 42) are permanently connected to one another in a rotationally fixed manner, wherein the first planet wheels (36) mesh with the first sun wheel (30) and the first ring gear (34), wherein the second planet wheels (46) mesh with the second sun wheel (40) and with the second ring gear (44), wherein the second planet wheels (46) and the third planet wheels (48) mesh with one another in pairs, wherein the first planet wheels (36) and the third planet wheels (48) are permanently rotationally fixed to one another in pairs, wherein the variator (16) is mechanically operatively connected to the drive input (12) and the first sun wheel (30), wherein the drive input (12) is mechanically operatively connectable to the first ring gear (34), and wherein the transmission (1300) is designed to transmit a torque from the second planetary wheel set to the driven input (14), characterized in that,

-the second sun gear (40) is connected via a spur gear stage (62) to a second clutch (K2) and can therefore be connected to the first clutch (K1) and/or the output (14), wherein the second sun gear (40) is connected via a spur gear stage (64) to a fourth clutch (K4) and can therefore be connected to the output (14);

-the first, second, third and fourth clutches (K1, K2, K3, K4) are arranged on the driven end (14);

-an intermediate shaft (70) connecting the planet carrier of the third planet wheel (46) via a spur gear stage (60) with the third clutch (K3), wherein the intermediate shaft (70) is thus made connectable with the driven end (14);

-the second ring gear (44) is connected to the first clutch (K1) via a spur gear stage (56), whereby the second ring gear (44) can be connected to the driven end (14).

6. Drive unit with a power split continuously variable transmission (1300) according to claim 5 and a motor which is mechanically operatively connected with the drive (12) for driving the drive (12).

Claims

1. Power split continuously variable transmission (100; 200; 300; 400; 500; 600; 700; 800; 900; 1000, 1100, 1200, 1300) having a drive input (12), a drive output (14), a variator (16) and a planetary assembly (18), wherein the transmission (100; 200; 300; 400; 500; 600; 700; 800; 900; 1000, 1100, 1200, 1300) is designed to provide different driving ranges, wherein the planetary assembly (18) has a first sun wheel (30) forming a first planetary gear set, a first planet carrier (32) rotatably supporting a first set of planet wheels (36), a first ring gear (34), and a second sun wheel (40) forming a second planetary gear set, a second planet carrier (42) rotatably supporting a second set of planet wheels and a third set of planet wheels (46, 48), and a second ring gear (44), wherein the first and second planet carriers (32, 42) are permanently connected to one another in a rotationally fixed manner, wherein the first planet wheel (36) meshes with the first sun wheel (30) and the first ring gear (34), wherein the second planet wheel (46) meshes with the second sun wheel (40) and the second ring gear (44), wherein the second planet wheel (46) and the third planet wheel (48) each mesh in pairs, wherein the first planet wheel (36) and the third planet wheel (48) each mesh in pairs permanently in a rotationally fixed manner, wherein the variator (16) is mechanically connected to the drive end (12) and the first sun wheel (30), wherein the drive end (12) is mechanically connectable to the first ring gear (34), and wherein the transmission (100; 200; 300; 400; 500; 600; 700; 800; 900; 1000; 300; 400; 500; 600; 700; 800; 1000; 200; 300; 400; 500; 600; 1000; 700; 1000; 800; 1000; 1; 2; one of the drive end) is mechanically connectable to one of the drive end, respectively, in pairs, each other, and one of the other, respectively, and one of the other, and one of the other, respectively, are mechanically connectable to each other, and one of the other, respectively, and one of the other, respectively, and one of the other, and one of the same, and one of the other, respectively, and one of the other, and one of the same, and one of the other, respectively, and one of the other, respectively, are connected to each other, and one of the other, respectively, and one of the other, and one of the same, respectively, and one of the other, 1100. 1200, 1300) is configured for transmitting a torque from the second planetary gear set to the driven end (14).

2. The power-split continuously variable transmission (1100) according to claim 1,

3. The power split continuously variable transmission (1200) according to claim 1,

-the second ring gear (44) is connected via a spur gear stage (60) to the first clutch (K1), wherein the second ring gear (44) can in turn be connected via the first clutch (K1) to the countershaft (70) and/or the second clutch (K2), wherein the second clutch (K2) is connected via a spur gear stage (62) to the second sun gear (40), wherein the second clutch (K2) can be connected via a spur gear stage (58) to the fourth clutch (K4) and further to the driven end (14).

4. The power-split continuously variable transmission (1300) of claim 1,

-the drive end (12) can be connected to the first ring gear (34) via a clutch (KR) and a spur gear stage (54) of a driving direction changing assembly (50) or via a clutch (KV) and a spur gear stage (52) of the driving direction changing assembly (50), wherein the driving direction changing assembly (50) is arranged on the power take-off shaft (20);

5. Drive unit having a power split continuously variable transmission (100; 200; 300; 400; 500; 600; 700; 800; 900; 1000, 1100, 1200, 1300) according to any one of the preceding claims and a motor which is mechanically operatively connected to the drive end (12) for driving the drive end (12).