Patent Application
20140309075
This patent claims priority from German Patent Application No. 10 2013 206 749.6, filed Apr. 16, 2013, which application is incorporated herein by reference its entirety.
The invention relates to a differential gear having a gear housing, an epicyclic gear housing which is arranged in the gear housing in a manner allowing rotation about a gear axis, and a planetary carrier which sits in the epicyclic gear housing, wherein the drive power applied to the epicyclic gear housing is split by means of this differential gear, and the planetary carrier and the epicyclic gear housing can be selectively coupled to each other by a friction fit via a coupling device.
Differential gears are generally constructed as planetary wheels, and most commonly serve the purpose of splitting or distributing an input power, supplied by a power input, to two drive shafts. Differential gears are most frequently used in the building of automobiles as so-called axle differentials. In this case, drive power supplied by a drive motor is distributed via the differential gear to wheel drive shafts of driven wheels. The two wheel drive shafts leading to the wheels in this case are each driven at the same torque, meaning they are balanced. When the vehicle drives straight forward, both wheels rotate at the same speed. When the vehicle travels a curve, the rotation speeds of each wheel are different. The axle differential makes this rotation speed difference possible. The rotation speeds are able to adjust themselves freely; only the average of the two speeds is unchanged.
In certain applications, particularly in all-wheel drive vehicles, differential gears are used which enable a switchable decoupling when the all-wheel drive function is not necessary, and additionally enable a separation in the drivetrain in order to drive the vehicle via only one axle, thereby reducing friction loss in the drive system, the same being not necessary at the moment, but otherwise driven anyway. Such a differential gear is known from DE 10 2008 037 885 A1, by way of example.
The problem addressed by the invention is that of creating a differential gear which enables a switchable release of the drive connection between the power input and the two power outputs, and which is characterized by a robust construction which can be realized in a cost-effective manner, wherein the differential gear generates the least possible drag torque when idling—meaning when the connection between the power input and the two power outputs is released.
The problem named above is addressed according to the invention by a differential gear, having a gear housing, an epicyclic gear housing which is arranged in the gear housing in a manner allowing rotation about a gear axis, a planetary carrier which is arranged in the epicyclic gear housing coaxial to the gear axle, a first output sun gear, a second output sun gear, a planetary arrangement accommodated in the planetary carrier for the purpose of coupling the two output sun gears in a manner allowing rotation in opposite directions, a brake device having a brake disk pack which generates a coupling torque which couples the planetary carrier to the epicyclic gear housing, and an actuating mechanism for the purpose of generating an axial force which engages the brake device, wherein the planetary arrangement comprises multiple revolving planets which as such are able to rotate about planetary axes which are oriented parallel to the gear axle, and the brake device and the planetary carrier are matched to each other in such a manner that the brake disk pack is positioned at the radial distance of the planetary axes.
In this way, it is advantageously possible to create a differential gear wherein it is possible for the axial forces which are necessary to bring about the coupled state of the brake device to be at least partially directed via the planetary wheel pins, the same providing support axially, in a manner which is advantageous for the structural mechanics of the differential gear.
The differential gear according to the invention is advantageously constructed in such a manner that the brake disk pack is composed of multiple annular brake disks. These brake disks can be designed as flat steel sheet metal hollow disks, which are optionally coated with a friction material layer. The brake disk pack in this case can be constructed in such a manner that it comprises brake disks which are kinematically coupled to the planetary carrier via an inner peripheral contour, in a manner allowing axial displacement, but which is nevertheless non-rotatable. In addition, the brake disk pack then also comprises brake disks which are kinematically coupled to the epicyclic gear housing via an outer peripheral contour, in a manner allowing axial displacement, but which is nevertheless non-rotatable.
According to one particularly preferred embodiment of the invention, the axial support of the brake disk pack is realized in interaction with a pressure ring element, wherein this pressure ring element is supported on the end faces of the planet pins.
The epicyclic gear housing included for the purpose of receiving the planetary carrier is preferably designed as a two-part bowl housing which is composed of a first bowl element and a second bowl element, wherein the first bowl element has a base section which extends inward radially. This base section of the first bowl element can be configured with cylindrical passages which pass through the base section sequentially at points evenly distributed around the periphery. Plunger elements of a first set of plunger elements can be received in these passages, wherein the plunger elements are guided in the passages in a manner allowing axial sliding toward the planet axes. These plunger elements function as pressure transmission organs between the interior of the bowl housing and the exterior of the same.
A roller guide ring is preferably seated on the plunger elements, on a side of the same which is opposite the brake disk pack, and said roller guide ring can be loaded axially via an annular piston element for the purpose of axially compressing the brake disk pack. This annular piston element is preferably received in a circular chamber which is concentric to the gear axle, and able to move axially according to the magnitude of a fluid pressure applied to the circular chamber. This annular piston presses the rollers of the roller guide ring, wherein said rollers run on the same, into the planetary carrier—meaning in the direction of the brake disk pack—when the fluid pressure is adequate.
According to one particularly advantageous embodiment, the planetary carrier is likewise supported on the side thereof which is opposite the brake disk pack by plunger elements, which in turn are guided through a base surface of the epicyclic gear housing, and supported on the end face thereof on a second roller ring. These plunger elements can have the same construction as the first plunger elements named above.
A ring gear is preferably seated on the epicyclic gear housing. The power input into the epicyclic gear housing is realized via this ring gear. A right-angled drive can be implemented with an interface to this ring gear. This construction is particularly suitable for use as a rear differential which can be reversibly disengaged. It is also possible to arrange a spur gear on the epicyclic gear housing in the place of the ring gear, for the purpose of inputting the drive power.
The inner differential included in this case, the same accommodated in the interior of the epicyclic gear housing, is designed as a spur gear differential having two output sun gears which are able to rotate in opposite directions, via a planetary arrangement. This spur gear differential is, according to a particularly preferred embodiment of the invention, designed with a Wildhaber/Novikov toothing. Details on the geometries which are preferably implemented on the respective toothed wheels in this case, as well as the addendum and foot circle diameters of the output sun gears and the revolving planet gears which engage with the same, are explained in the description of the figures.
Further details and features of the invention are found in the following description, with reference to the drawing, wherein:
The illustration according to
The differential gear further comprises a first output sun gear 1, a second output sun gear 2, and a planetary arrangement P accommodated in the planetary carrier 3, for the purpose of coupling the two output sun gears 1, 2 in a manner allowing rotation in opposite directions. A brake device is positioned in the differential gear, which in this case is designed as a brake disk pack BLP, for the purpose of generating a coupling torque which selectively couples the planetary carrier 3 to the epicyclic gear housing U, according to the magnitude of an axial force F engaging the brake disk pack BLP.
In addition, the differential gear according to the invention has an actuating mechanism 5 for the purpose of generating the axial force F applied to the brake disk pack 4a, 4b. The brake disk pack BLP is integrated into the differential gear in such a manner that it couples the planetary carrier 3 to the epicyclic gear housing U with a friction fit when there is a corresponding axial load. As a result of this approach, it is possible to release the drive connection between the planetary carrier 3 and the epicyclic gear housing U by unloading the brake disk pack BLP and/or to couple the planetary carrier 3 to the epicyclic gear housing U with a friction fit by means of loading the brake disk pack BLP axially.
The differential gear which in this case includes the planetary carrier 3, the planetary arrangement P and the output sun gears 1, 2 is designed as a spur gear differential with two output sun gears 1, 2. The planetary arrangement P has multiple revolving planets P1, P2 which are mounted as such on planet pins 6. The differential gear according to the invention is characterized in that the brake disk pack BLP and the planetary carrier 3 are matched to each other in such a manner that the brake disk pack BLP is positioned at the radial or track distance of the planet axes XP which are parallel to the gear axle X. As a result of this special construction, the axial forces F acting on the brake disk pack BLP can be divided axially through the planetary carrier 3 while incorporating the planet pins 6, the same oriented parallel to the gear axle X.
The brake disk pack BLP has a set of first, annular brake disks 4a which engage with the planetary carrier 3 via an inner edge contour thereof, in a non-rotatable manner, but nevertheless allowing axial sliding. The brake disk pack BLP has a set of second brake disks 4b which engage with the epicyclic gear housing U via an outer edge contour, in a non-rotatable manner, but nevertheless allowing axial sliding. These brake disks 4a, 4b are designed as flat steel sheet metal hollow disks, and preferably are coated with a friction material layer.
The axial support of the brake disk pack BLP on the planet pins 6 is realized with the integration of a pressure ring element 4d which is supported on the end faces of the planet pins 6. The planetary carrier 3 and the brake disk pack BLP are matched to each other in such a manner that the radial distance of each of the planet pin axes XP from the gear axle X is greater than the inner diameter of a brake disk 4a, 4b, and also is smaller than the outer diameter of the brake disk 4a, 4b.
The epicyclic gear housing U included for the purpose of receiving the planetary carrier 3 is designed as a two-part bowl housing, and composed of a first bowl element U1 and a second bowl element U2, wherein the first bowl element U1 has a base section U1a which extends inward radially. This base section U1a of the first bowl element U1 in this case is configured with circular passages D1 which pass through said base section U1a in sequential positions at equal distances around the periphery. The plunger elements Q1 of a first set of plunger elements sit in these passages D1. These plunger elements Q1 are guided in the passages D1 in a manner allowing axial sliding in the direction of the planet axes XP. These plunger elements Q1 function as pressure transmission organs between the inner region of the bowl housing U and the outer region of the same. A roller guide ring R1 is seated on a side of the plunger elements Q1, on these plunger elements Q1, said side being opposite the brake disk pack BLP, wherein said roller guide ring R1 can be loaded axially via the annular piston RK, for the purpose of axially pressing the brake disk pack BLP together.
The annular piston element RK is received in a circular chamber RC which is concentric with the gear axle X, and can be moved axially according to the magnitude of a fluid pressure applied to the circular chamber RC via a fluid channel C1. This annular piston element RK impels the rollers R1a of the roller guide ring R1, said rollers running toward the same, toward the planetary carrier 3, meaning in the direction of the brake disk pack BLP. The circular chamber RK above is directly molded into the gear housing G in this case.
The planetary carrier 3 is likewise supported by plunger elements Q2 on the side thereof which is opposite the brake disk pack BLP, said plunger elements in turn being guided through a base surface U2a of the epicyclic gear housing U and supported on the end face thereof by a second roller ring R2. These plunger elements Q2 are designed with the same construction as the first plunger elements Q1 named above.
The roller ring R2 carries a roller arrangement R2a which runs directly to an end face of an outer bearing ring L4a of a bearing L4 which supports the epicyclic gear housing U. The roller ring R2 can be designed in such a manner that it is centered by the inner bearing ring L4i of this bearing L4.
The plunger elements Q1, Q2 can be seated sectionally in suitable receiving pockets of the roller rings R1, R2, and optionally secured, to prevent them from falling out, in the same by means of a press-fit, by way of example. A locking device can be implemented by means of the plunger elements Q1, Q2 in relation to the roller rings R1, R2, such that the roller rings R1, R2 are able to travel together axially with the plunger elements functionally assigned to the same, but are not able to rotate with respect to the epicyclic gear housing U.
The first and second revolving planets P1, P2 engage directly with each other, and are therefore coupled in a driving relationship to each other, in such a manner that they rotate in opposite directions, as will be explained below in greater detail. In this embodiment, there are a total of three revolving planets P1 which engage with the first output sun gear 1. These revolving planets P1 which engage with the first output sun gear 1 form a first set of revolving planets. In addition, in this embodiment, there are a total of three revolving planets P2 which engage with the second output sun gear 2. These revolving planets P2 which engage with the second output sun gear 2 form a second set of revolving planets. Each revolving planet P1 of the first set engages with one revolving planet P2 of the second set. The engagement of the revolving planets P1 of the first set with the revolving planets P2 of the second set is realized at the same tooth plane as the engagement of the revolving planets P1 of the first set with the output sun gear 1.
The first output sun gear 1 and the second output sun gear 2 are matched to each other, in regards to the tooth geometry thereof, in such a manner that the addendum circle of the spur gear toothing 1 a of the first output sun gear 1 is smaller than the root circle of the output sun gear toothing 2a of the second output sun gear 2. The revolving planets P1 of the first set engage with the revolving planets P2 of the second set in the region of the tooth plane of the first output sun gear 1. The two output sun gears 1, 2 are directly adjacent to each other.
The two output sun gears 1, 2 are designed in such a manner that that the output sun gear toothing la of the first output sun gear 1, and the output sun gear toothing 2a of the second output sun gear 2 have the same number of teeth. The revolving planets P1 of the first set and the revolving planets P2 of the second set also have the same number of teeth. The input of the drive power into the differential gear is realized via the ring gear 7 and the epicyclic gear housing U. A symmetrical division of torque and a division of power to the output sun gears 1, 2 is realized via the revolving planets P1, P2. Flange sections 1b, 2b are constructed on the output sun gears 1, 2. These flange sections 1b, 2b are produced by extrusion in a molding process, and are configured with an inner toothing 1c, 2c. End segments of wheel drive shafts or other power transfer components of the respective wheel drivetrain can be inserted into this inner toothing 1c, 2c, said end segments accordingly having complementary toothing. In place of the inner toothing shown here, other connection geometries can also be possible for the transmission of rotational torque, and for centering and receiving corresponding components.
The ring gear 7 seated on the epicyclic gear housing U in a non-rotatable manner is driven via a primary drive sprocket 8. The ring gear 7 and the primary drive sprocket 8 form a right-angled drive. The embodiment shown here is therefore particularly suitable as an axle differential for a rear axle which can be selectively decoupled from the primary drivetrain. In place of the transmission of rotary torque via a right-angled drive indicated here, it is also possible for a spur gear to be configured on the epicyclic gear housing U, which is driven via a further spur gear, by way of example. Such a variant is then particularly suitable for direct installation in a vehicle transmission.
The planetary carrier 3 sits between the ring element R and the brake disk pack BLP. The axial force transmission between the ring element R and the brake disk pack BLP is realized in this case primarily via the planet pins 6 and the planetary carrier 3 itself, braced by the same.
The planetary carrier 3 is composed of two carrier jackets 3a, 3b and a carrier pin 3c. The carrier jackets 3a, 3b are each produced as molded sheet metal part. These two carrier jackets 3a, 3b and the carrier pins 3c are welded to each other. For this purpose, rods are formed on the first carrier jacket 3a, which as such bridge the tooth region. The first carrier jacket 3a forms an inner bore hole in which an extension of the first output sun gear 1 is accommodated in a manner allowing rotation. The brake disk pack BLP sits on the carrier pin 3c. When the brake disk pack BLP is fully braked, the planetary carrier 3 can therefore be coupled by a friction fit to the epicyclic gear housing U. The brake disk pack BLP and the actuating mechanism 5 which is configured to load the same axially are designed in such a manner that it is possible for the drive torque applied to the ring gear 7 to be transmitted to the planetary carrier 3 via the brake disk pack BLP when the same is loaded axially.
The mounting of the planetary carrier 3 in the epicyclic gear housing U is realized via a first needle roller bearing L1 and a second needle roller bearing L2. The mounting of the epicyclic gear housing U in the gear housing G is realized via angular ball bearings L3, L4. These angular ball bearings L3, L4 direct the gear reaction force components, which engage the ring gear 7 and are oriented both radially and axially, into the gear housing G. Neither the bearing L1 nor the bearing L2 need to convey axial forces. The primary purpose of these bearings L1, L2 is to center and mount the planetary carrier 3 in the epicyclic gear housing U. A shaft seal ring is inserted into the seat S2 in order to accordingly establish a seal between the housing and the insert shaft (not illustrated).
The functionality of the differential gear according to the invention is as follows: The ring gear 7 is driven by the primary drive sprocket 8. The ring gear 7 is fixed to the epicyclic gear housing U in a manner preventing rotation. Accordingly, the epicyclic gear housing U is made to rotate via the ring gear 7. This epicyclic gear housing U is arranged concentrically with a gear axle X, and mounted in the gear housing G via the first and the second bearings L3, L4 in a manner allowing rotation.
The brake disk rings 4b of the brake disk pack BLP which are also coupled to the epicyclic gear housing U in a non-rotatable manner rotate together with the same. The brake disk pack BLP is loaded axially by the pressure ring 4d and the annular plate 4c according to the magnitude of the axial force generated by the actuating mechanism 5, and are thereby optionally brought into a coupled state in which the epicyclic gear housing U and the planetary carrier 3 are coupled by a friction fit. A division of power is realized inside the planetary carrier 3 via the planets P1, P2 and the output sun gears 1, 2.
The planetary gear train accommodated in this case in the epicyclic gear housing U forms a spur gear differential, as already described. In the embodiment shown here, the output sun gear 1, 2 and the planet gears P1, P2 of the planetary arrangement P are configured with a Wildhaber/Novikov toothing. The first output sun gear 1 in this case has a toothing with a small addendum circle and concave tooth flank surfaces. The second output sun gear 2 has a toothing with a large addendum circle and convex tooth flank surfaces. The addendum circle diameter of the first output sun gear 1, and theoretic root circle of the second output sun gear 2 approximately correspond to the same, identical semicircle diameter. Both gears 1, 2 have the same number of teeth. The first output sun gear 1 engages with the revolving planets P1. The second output sun gear 2 engages with the revolving planets P2. The revolving planets P1 have a large addendum circle diameter and form convex tooth flanks The revolving planets P2 have a small addendum circle diameter and form concave tooth flanks The revolving planets P1, P2 engage with each other in pairs.
The engagement occurs in the engagement plane of the first revolving planets P1 with the first output sun gear 1. The first revolving planets P1 have an axial length which corresponds substantially to the axial length of the toothing 1a of the first output sun gear 1. The second revolving planets P2 have an axial length which corresponds substantially to the sum of the axial lengths of the toothings 1a, 2a of the two output sun gears 1, 2.
| Number | Date | Country | Kind |
|---|---|---|---|
| 102013206749.6 | Apr 2013 | DE | national |
