WO2005083300A1

WO2005083300A1 - Toroidal gearbox for a motor vehicle

Info

Publication number: WO2005083300A1
Application number: PCT/EP2005/001620
Authority: WO
Inventors: Wolfgang Elser; Steffen Henzler; Recep Tevetoglu
Original assignee: Daimlerchrysler Ag
Priority date: 2004-02-18
Filing date: 2005-02-17
Publication date: 2005-09-09
Also published as: DE102004007789A1

Abstract

The invention relates to a toroidal gearbox with a stepless variator and a multi-row planet set. Usually an axial support is necessary for the support of a planet gear (62) against adjacent components (68, 72) of a carrier, for example by means of running discs (74, 75). According to the invention, the planet gear (62) has planets (60, 61), provided with an inclined toothing. The angles of inclination ßl and ß2 are chosen such that the axial components of the toothing forces are exerted straight upwards on the planet gear (62). Toroidal gearbox for motor vehicles.

Description

Toroidal gear for a motor vehicle

The invention relates to a toroidal transmission for a motor vehicle according to the preamble of claim 1.

A toroidal transmission for a motor vehicle is known from the German patent specification DE 101 21 042 Cl of the applicant, which has a stepless variator. The power flow between the variator and the transmission output shaft runs over a multi-row planetary set, which has several stepped planets. In the case of the stepped planet, two planets are connected to each other in a drive-proof manner.

The components of the toroidal gear are exposed to mechanical stress during operation. The cause of the stresses are in particular drive torques to be transmitted, which result in radial forces, axial forces, bending or torsional moments. Another cause of mechanical stresses are normal forces which have to be applied to ensure that a scooter presses against the toroidal washers of the variator to ensure reliable transmission of a drive torque in the variator.

The above-mentioned stresses have the consequence that components such as shafts, gears, support elements and bearings have to be designed to be correspondingly strong. Furthermore, the stresses can lead to deformation of components

BESTATIGUNGSKOPIE lead of the toroidal gear, which impair the function and / or the service life of the toroidal gear.

The present invention is based on the object of proposing a toroidal transmission which is improved with regard to the mechanical loads and the deformations caused by these.

The object on which the invention is based is achieved by the features of patent claim 1.

According to the invention, the stepped planet has a first planet and a second planet (other planets can of course also be provided). The planets each have helical teeth. The helical gears have different helix angles ßx and ß ₂ . According to the invention, the advantages of helical teeth such as increased smoothness and / or load-bearing capacity are thus used. The different helix angles ß _x and ß ₂ dictate an axial position of the stepped planet, even if the planets are free of drive torque when the motor vehicle is at a standstill. During operation of the motor vehicle, axial forces act on the planets due to the helical teeth. According to the invention, the helix angles ß _x and ß ₂ are adapted such that the forces acting on the first planet and on the second planet in the axial direction approximately cancel each other out. In this way, components for specifying and determining the axial position of the stepped planet, for example axial bearings, can be designed for lower loads or can be omitted entirely.

The solution according to the invention leaves the usual paths of the person skilled in the art, who is entrusted with the task at hand here and usually with toroidal gears, ten US 6,302,819 Bl, US 5,372,555, US 6,306,059 Bl, US 6,616,564 B2 and US 6,517,461 B2 can provide clues for dimensioning the forces acting in the toroidal gear. With the advantage of a simplified manufacture achieved according to the invention and u. U. a shortened axial design, the invention also leaves the path taken in US Pat. No. 5,807,203 to avoid axial forces occurring in the toothing by using an arrow toothing. The present invention transfers individual features known from the publications WO 99/67 168, DE-PS 1 157 049 and DE-OS 28 21 320 to the technological field of toroidal gears for motor vehicles.

According to a preferred embodiment of the invention, the quotient from the diameters of the first planet and the second planet is first determined for the design of the helix angle. The quotient from the aforementioned diameters corresponds approximately to the quotient of the tangent of the assigned helix angles ßi or ß ₂ . This provides a simple approximation formula for the design.

According to a preferred embodiment of the toroidal gear, the quotient of the helix angle β ₂ / β (or the diameter D ₂ / D _x ) is in the range between 1.05 and 1.30. Investigations by the applicant have shown that the aforementioned range of values forms an optimum with regard to the transmission ratios in the toroidal gear, the smooth running and the load-bearing capacity of the toothings. For example, at least one of the helix angles β _x and β _{2 is} between 10 ° and 24 °.

According to a development of the invention, the second planet of the stepped planet meshes only radially on the inside with a sun gear, while the first planet meshes radially on the inside with another sun gear and radially on the outside with another gear element. Depending on the switching state of the toroidal transmission, the power flow from the second planet can optionally take place to the other sun gear, to the further gear element and / or simultaneously to the other sun gear and the further gear element. According to the invention, the design for an automatic compensation of the axial forces is thus not only adapted to a gear state or a driving range, but equally for different switching states of the toroidal gear.

The stepped planet and the further planet are preferably mounted in a common carrier, which is formed in one or more pieces. This results in a simple structure with a compact and mechanically optimized design.

According to a special embodiment of the toroidal gear, the carrier has a front, a middle and a rear support plane. The support planes can be designed in any way, for example by means of a basically disk-shaped or ring-shaped contour, if necessary with suitable recesses or with suitable struts. According to this configuration, the further planet is mounted opposite the middle and / or an outer support plane (front or rear support plane). The stepped planet is mounted opposite the front and / or the rear support plane. This results in a support of the bearing forces axially closely adjacent to the forces that occur, so that there are low support moments or low radial forces. It is also conceivable that the stepped planet is alternatively or additionally mounted in the middle support plane, so that an additional support plane results or an outer support plane is omitted can. It is also conceivable that the further planet is supported in the front and rear support planes instead of in the middle support plane.

According to a further embodiment, the middle support plane and the outer support plane adjoining the other planets are connected to one another via connecting webs in the toroidal transmission, which are used for a rigid construction of the carrier, in particular for absorbing axial forces, radial forces and moments. An increase in the rigidity and / or the possibility of a smaller dimensioning of the connecting webs results if the connecting webs are arranged radially on the outside or evenly distributed in the direction of the force flow. According to the invention, an existing space between several other planets is used, since connecting webs are arranged in the circumferential direction between several other planets.

According to a further embodiment of the invention, the middle support plane and the outer support plane which are not adjacent to the other planets are connected to one another via connecting areas which have an approximately cylindrical or conical surface area. By means of a conical outer surface, different diameter ranges of the middle support plane and the outer support plane can be connected to one another in a simple manner, resulting in an optimal rigidity. The cylindrical or conical outer surface also represents an optimum in terms of the space required. The outer surface has a radial recess in the area of the stepped planet. The radial recess enables the stepped planet to be mounted in at least a partial radial direction. According to this configuration, a particularly rigid design of the carrier is made possible with a simple installation option. According to an alternative embodiment of the invention, an axial assembly of the stepped planet is made possible. For this purpose, the outer supporting plane adjacent to the other planets is formed separately from other components of the carrier. If the outer supporting plane is removed from the carrier, an axial insertion of the stepped planet from the side of the removed outer supporting plane is possible, the stepping planet being insertable into the middle supporting plane up to the opposite outer supporting plane.

Advantageous further developments of the toroidal transmission result from the description and the drawings.

A preferred embodiment of the toroidal transmission according to the invention is explained in more detail below with reference to the drawing. The drawing shows:

1 is a wheel plan of an embodiment of a toroidal transmission according to the invention in longitudinal section,

2 shows a detailed representation of the wheel plan shown in FIG. 1 with a step planet and a carrier in longitudinal section,

3 shows a detailed representation of a further embodiment according to the invention with a stepped planet, another planet and a carrier in a spatial representation and

Figures 4 to 7 different views of an alternative embodiment of a carrier for use in a toroidal gear. The toroidal gear according to the invention is used in motor vehicles, in particular with a standard drive.

In the power flow between a central input shaft 10 and a coaxial output shaft 11, a stepless toroidal gear 12 is arranged, which has a planetary gear summation 13 and a planetary gear reversal gear 22. Coaxial and fixed to the input shaft 10, a central intermediate shaft 14 is provided, which is connected to the one central drive pulley 15 of the variator 99, which is designed according to the two-chamber principle, and to a two-walled planet carrier 16, which forms a first transmission element of the sum transmission 13. The planet carrier 16 is additionally connected to the other central drive pulley 15a of the variator 99 so as to make the coaxial power passage possible. A concentric intermediate shaft 17 is arranged coaxially to the input shaft 10 and concentrically to the central intermediate shaft 14, which connects the two central driven disks 18, 18a of the variator 99 to an inner central wheel 19 forming a second gear member of the summation gear 13 in a rotationally fixed manner.

The sum transmission 13 has a third transmission element in the form of an outer central wheel 20, an indirect or direct drive connection 27 between the third transmission element and the output shaft 11 being able to be established by a first switching element in the form of a clutch (Kl) for a lower driving range at lower driving speeds ,

The sum transmission 13 has a fourth transmission link in the form of an inner central wheel 21, with an indirect drive connection 39 between the fourth transmission link and the output shaft 11 can be produced by a second shift element in the form of a clutch (K2) in an upper driving range with higher driving speeds.

The input shaft 10 is engageable by-passing the variator 99 by activation of a third ^'control element in the form of a clutch (K3) for a gear ratio i = l (direct gear) to the output shaft 11 in driving connection.

The clutch (K3) for the direct gear is connected on the one hand with the drive to the central intermediate shaft 14. The drive connection 27 is coupled to the output shaft via the reversing gear 22. The clutch (Kl) is rotatably connected to one (here central wheel 26a) of two outer central wheels 26 and 26a of the reversing gear 22. The other central wheel (here central wheel 26) is connected to the output shaft 11 in a rotationally fixed manner. The central wheels 26, 26a lie axially on both sides of a radial support web 23a of the planet carrier 23, by means of which the latter is fixed non-rotatably relative to a non-rotating housing part 31 of the gear housing. On the planet carrier 23, planets 30 are rotatably mounted, the two toothed rings of which mesh with one of the outer central wheels 26, 26a, which have the same number of teeth and therefore the

Ensure ratio 1: 1 between input shaft 10 and output shaft 11. The ring gears are connected to one another in a rotationally fixed manner, so that the planets 30 are designed as a stepped planet.

The planet carrier 16 has double planets 44 and a radial drive web 49 which is connected in a rotationally fixed manner to the central intermediate shaft 14. The double planet 44 consist of a main and secondary planet 45 and 46, which together Comb and are referred to below as the first planet 60 and another planet. The main planets 45 have a first ring gear 47 lying on the side of the drive web 49 facing away from the toroid gear 12 and a second ring gear 48 lying on the side of the drive web 49 facing the toroid gear 12. The gear rings 47, 48 are connected to one another in a rotationally fixed manner, so that the main planet 45 is designed as a stepped planet. The secondary planet 46 mesh with the outer central wheel 20. With the main planet 45, the first ring gear 47 meshes with the inner central wheel 21 and the second ring gear 48 with the inner central wheel 19.

The sprockets 47 and 48 of the main planet 45 have unequal numbers of teeth, the sprocket 47 having the larger number of teeth. For different operating ranges of the toroidal gear shown, a force flow between input shaft 10 and output shaft 11 takes place as follows:

A geared-neutral function means that when the second clutch (K2) and the third clutch (K3) are disengaged, the respective speed of the output shaft 11 and that directly connected to the clutch (Kl) when the first clutch (Kl) is engaged Gear elements initially equal to zero and the partial transmission in variator 99 set to a predetermined value.

The first clutch (Kl) remains engaged in the subsequent lower driving range with lower speeds of the output shaft 11. When driving forward, the power flows via the direct path of the central intermediate shaft 14 to the summation gear 13 and is branched, part flowing via the first clutch (Kl) to the output shaft 11 and the other part via the variator 99 to the intermediate shaft 14 or flows back to the planet carrier of the summation gear. Circulation power thus occurs in the gear arrangement, the power in at least one of the paths being higher than the gear input power.

A synchronous point can be set for changing the driving range, at which the differential speed at the second clutch (K2) is approximately zero, so that a smooth change of drive from the first clutch (Kl) to the second clutch (K2) is made possible. In this upper driving range, the transmission input power is generally divided into two parallel paths, so that the power component in both paths (variator 99 on the one hand and central intermediate shaft 14 on the other) is smaller than the transmission input power. Current performance does not occur.

Another synchronous point can also be set when the change gear arrangement is in the position with constant overall gear ratio, i.e. the third clutch (K3) is engaged. This means that the drive change from the clutch (K3) to the clutch (K2) in the upper driving range can also be carried out smoothly.

With regard to further configurations of the power flow, the switching elements and the wheel map, reference is made in full to the applicant's publication DE 101 21 042 C1.

Sprocket 47 is formed by a first planet 60 and sprocket 48 by a second planet 61. The planets 60, 61 are rotatably connected to one another to form a step planet 62.

In a first plane 63, the first planet 60 meshes radially on the inside with the inner central wheel 21 and radially on the outside. with the secondary planet 46. In a second plane 64, the second planet 61 meshes radially on the inside with the inner central wheel 19. The second plane 64 is arranged axially between the variator 99 and the first plane 63.

Between the levels 63, 64 there is a central support level 65 which contains (at least partially) the planet carrier 16 and the drive web 49. A front support level 66 is arranged between the variator 99 and the second level 64, while a rear support level 67 is arranged on the side of the level 63 facing away from the variator 99.

The support plane 66 is formed with a circular ring body 68 (FIG. 2). The variator 99 is supported in the axial direction on the annular body 68. The annular body has an extension 69 pointing radially on the outside in the direction of the variator 99. The toroidal disc 15a is supported axially displaceably relative to the extension 69. The end face of the circular ring body 68 facing the variator 99 delimits a working space, via which the toroidal disc 15a can be pressed against the roller. With regard to the connection and the functional interaction of the annular body 68 and the torus disk 15a, reference is made in full to the documents DE 102 06 204 AI, DE 101 32674 AI, which is hereby made the subject of the application.

The drive web 49 is formed with a disk body 70 which has radially inwardly directed recesses 71 or has an axial bore into which the stepped planet 62 can be inserted from the outside or through which the stepped planet 62 is inserted in the axial direction for assembly can be. The support plane 67 is formed with an annular body 72. The inner diameter of the annular bodies 68, 72 is selected such that they can be displaced in the axial direction via the intermediate shaft and possibly the central wheels 19, 21 for assembly purposes. The support planes 66, 67, the annular bodies 68, 72 and the disk body 70 are oriented approximately parallel to one another and transversely to the longitudinal axis of the intermediate shaft 14.

The planets 60, 61 are held by a planet bolt 73 shown in dashed lines in FIG. 2. The planet pin 73 is supported on the axial bores of the annular bodies 68, 72. The step planet 62 is rotatably supported relative to the annular bodies 68, 72. In this case, a plain bearing or a roller bearing can be provided between the planet 60, 61 and the planet pin 73 or between the planet pin 73 and the annular bodies 68, 72. If the stepped planet 62 is axially displaceable with respect to the planet pin 73 or the entire stepped planet 62 with planet pin 73 with respect to the annular bodies 68, 72, axial bearings or thrust washers are located between facing end faces of the first planet 60 and the annular body 72 and the second planet 61 and the annular body 68 74, 75 interposed. The thrust washers 74, 75 are preferably made of at least one of the materials steel, brass or bronze. These can be flat or with any contour, for example with pockets for receiving a lubricant. The thrust washers 74, 75 are formed separately from or in one piece with the surrounding components, for example as a projection of a planet or the annular body 68, 72. The drive web 49 is formed in one or more pieces with the intermediate shaft 14. The drive web 49 is preferably welded to the intermediate shaft 14 (deviating from FIG. 2). The central wheels 19, 21 and the planets 60, 61 are helically toothed. The helix angle of the central wheel 19 and the second planet 61 is β ₂ , while the helix angle of the central wheel 21 and the first planet 60 is 60 β. The acute angle ß SSI and ₂ are for the planetary 60, 61 in the drawing Fig. 2 is preferably opened in the direction of the variator 99th According to the illustrated embodiment, the diameter Di is the first planetary 60 larger than the diameter D ₂ of the second planetary 61. The helix angle is smaller than the helix angle ßi ß. ₂

The helix angles ßi, ß ₂ are preferably in the range between 10 ° and 24 °. The following relationship applies for a rough determination of the helix angle: tan (ß ₂ ) / tan (ß _x ) = D ₂ / Dι,

The preferred range of values for the helix angle is: tan (ß ₂ ) / tan (ß ₁ ) = 1.05 ... 1.30.

According to FIG. 3, the circular ring body 68 and the disk body 70 are connected radially on the outside via a cylindrical outer surface 76. In the area of the recesses 71 in the area of the recesses 71 from the disk body 70 there are recesses 77 from the jacket surface 76 which extend completely between the disk body 70 and the circular ring body 68 in the axial direction or at a distance from the circular ring body 68 (cf. FIG. 3). end up. 3, a recess 77 is selected with a semicircular contour. The first planet 60 meshes with the secondary planet 46 between the disk body 70 and the annular body 72, the first planet 60 and the secondary planet 46 being arranged approximately adjacent to one another in the circumferential direction in such a way that the secondary planet 46 is spaced radially (slightly) further from the intermediate shaft 14 as the first planet 60. The secondary planet 46 is supported by a planet pin 78, which in turn is supported in relation to the annular body 72, the disk body 70 and / or the annular body 68, with a degree of freedom of rotation of the secondary planet 46 about the longitudinal axis of the planet pin 78, for example by means of a roller bearing or a plain bearing. In the space between such pairs of planet 60 and secondary planet 46, connecting webs 79 are arranged radially on the outside, which firmly connect the disk body 70 to the annular body 72. With the circular ring body 68, the extension 69, the disk body 70, the circular ring body 72, the lateral surface 76 and the connecting webs 79, a support 80 designed as a rigid and as rigid as possible component is created in one or more pieces.

The fact that - in contrast to the document DE 101 21 042 Cl - the wide plane 64 is arranged between the first plane 63 and the variator 99, the outer central wheel 20 can be made axially shorter. This reduces the inertia of the outer central wheel 20. The lever arms of the toothing forces acting on the outer central wheel 20 become smaller.

The risk of the outer central wheel tipping is also reduced.

The circular ring body 68 is exposed to high mechanical stresses, in particular by those acting in the axial direction Contact forces between torus disc 15a and roller. The inventive design of the carrier 80 keeps the deformations of the carrier 80 caused thereby low. This has advantages for the stresses on the components, in particular the contact area between the toroidal disc 15a and the annular body 68, the wear and fatigue of which are kept low, as a result of which the service life of the transmission is increased. Such a rigid design is preferably achieved by distributing the material as uniformly as possible in the area of the lateral surface 76, the size of the recesses 77 being kept as small as possible.

For the exemplary embodiment shown in FIG. 3, the stepped planet 62 can be inserted into the recess 77 from the radial direction or obliquely from above. The recess 77 preferably has a large radius in order to keep notch stresses low. If the stepped planet 62 and possibly the secondary planet 46 are mounted in the axial direction from the side facing away from the variator 99, the recesses 77 can be completely omitted. For such an assembly, it is advantageous if the circular ring body 62 is connected to further components of the carrier 80 via a detachable connection.

A particularly advantageous embodiment of the invention results if the lateral surface 76 is partially conical, the cone opening in the direction of the variator 99. In the event that the secondary planet 46 (in contrast to FIG. 3) is supported with respect to the toroidal body 68 and the toroidal body 72, the central supporting plane 65 with the disc body 70 can be omitted, in which case at least one of the toroidal bodies 68, 72 is rotationally fixed must support against the intermediate shaft 14. The configuration according to the invention results in an increased service life. There is a reduced tendency to pitting, a reduced tendency to seizure or other signs of failure. The torque capacity of the variator 99 and thus of the transmission can be increased. As a result of the lower deformations, there may be a lower oil requirement in the hydraulic pressing device when the pressure is increased. Furthermore, a higher rigidity and strength of the carrier 80 is ensured with a lower weight and a lower inertia. Compared to single-row planetary gears, it is possible to reduce the stationary gear ratio at comparable speeds on the planet and with a comparable installation space. The aforementioned increase in stationary gear ratio causes an increase in the output torque.

In the embodiment of the carrier 80 shown in FIGS. 4 to 7, the outer diameter of the disk body 70 forming the middle supporting plane is designed with a smaller diameter than the circular ring bodies 68, 72. In this case, the connecting webs 76, 79 are conical.

Claims

claims

1. Toroidal transmission for a motor vehicle with a continuously variable variator (99) and a multi-row planetary gear set (13), which has at least one stepped planet (62), in which two planets (60, 61) are connected to one another in a drive-proof manner, characterized in that a first Planet (60) has a helical toothing with a helix angle ß _x , a second planet (61) has a helical toothing with a helix angle ß ₂ and the helix angle ßi, ß ₂ are specified such that the on the first planet (60) and the approximately cancel the second planet (61) acting in the axial direction.

2. Toroidal gear according to claim 1, characterized in that the quotient from the tangent functions of the helix angle ß ₂ and ß corresponds approximately to the quotient from the diameters of the planets D ₂ and D _x .

3. Toroidal gear according to claim 2, characterized in that the quotient of the helix angle ß ₂ / ßi or the diameter D ₂ / Di is in the range between 1.05 and 1.30.

4. Toroidal transmission according to one of the preceding claims, characterized in that a helix angle ß _x and / or ß _{2 is} between 10 ° and 24 °.

5. Toroidal transmission according to one of the preceding claims, characterized in that the second planet (61) of the stepped planet (62) meshes exclusively radially on the inside with a sun gear (inner central wheel 19) and the first planet (60) on the radially inside with another sun gear ( inner central wheel 21) meshes and radially on the outside with another gear element.

6. Toroidal transmission according to claim 5, characterized in that the further transmission element is a further planet (secondary planet 46), which forms a double planet with the first planet (60).

7. Toroidal transmission according to claim 6, characterized in that the stepped planet (62) and the further planet (secondary planet 46) are mounted in a common carrier (80).

8. Toroidal gear according to claim 7, characterized in that the carrier (80) has a front, a middle and a rear support plane (66, 65 and 67), the further planet (secondary planet 46) relative to the middle support plane (65) and / or an outer support plane (66; 67) is mounted and the step planet (62) is mounted opposite the front support plane (66) and / or the rear support plane (67).

9. Toroidal gear according to claim 8, characterized in that the middle support plane (65) and the outer planet (adjoining planet 46) adjoining outer support plane (rear support plane 67) are connected to one another via connecting webs (79), which in the circumferential direction between several others Planets (minor planet 46) are arranged.

10. Toroidal gear according to claim 8 or 9, characterized in that the middle support plane (65) and the outer support plane (front support plane 66) which is not adjacent to the further planets (secondary planet 46) via connection areas with a cylindrical or conical outer surface (76). are connected to one another, which have radial recesses (77) in the area of the stepped planet (62), which enable the stepped planet (62) to be mounted in the radial direction.

11. Toroidal gear according to claim 8 or 9, characterized in that the middle support plane (65) and the outer support plane (66) not adjoining the further planets are connected to one another via a connecting region with a cylindrical or conical outer surface (76) and to the Another planet (secondary planet 46) adjacent outer support plane (67) is formed separately from other components of the carrier (80), so that an axial mounting of the stepped planet (62) in the carrier (80) is made possible.

12. Toroidal gear according to one of the preceding claims, characterized in that the outer support plane (67) and the connecting web (79) are formed in one piece.

13. Toroidal gear according to claim 12, characterized in that the outer supporting plane (67) and the connecting web (79) are made from a sheet metal.

14. Toriod gear according to claim 12 or 13, characterized in that the outer support plane (67) and the connecting web (79) by means of a material connection, in particular a weld, are connected to the central support plane (65; disk body 70).