Gear train with split torque

A gear train is provided that includes at least one side gear comprising a helical face gear and a plurality of helical pinions in meshing engagement with the helical face gear. The gear train may further include an absorber configured to provide an axial force on at least one of the plurality of helical pinions. A differential may also be provided including a differential case and a gear train disposed in the differential case. The gear train includes at least one side gear comprising a helical face gear and a plurality of helical pinions in meshing engagement with the helical face gear. The gear train in the differential may further include a means for providing an axial force on at least one of the plurality of helical pinions.

TECHNICAL FIELD

The present invention relates to a gear train which features torque sharing among multiple pinions, including an improved gear train with split torque and an improved interaction of the pinions with other elements of the gear train.

BACKGROUND

Gear trains that require torque sharing among multiple pinions may be found in the automotive industry (e.g., automobile differentials, automobile transmissions, etc.), the aerospace industry (e.g. helicopter transmissions, etc.), the epicyclical transmissions of electric wind-power stations, as well as in many other fields of engineering.

Gear trains may encounter errors caused by the manufacturing of gear train components, the assembly of the gear train, and/or the elastic deformation of gear train components under an operating load, all of which may be unavoidable and may cause unequal torque sharing among the pinions of a gear train.

Increased accuracy for manufacturing of pinions and side gears that make up a gear train may be a straightforward way to reduce instances of unequal torque sharing among multiple pinions caused by misalignment and/or deviation from proper meshing between the pinions and side gears. However, the increase in manufacturing accuracy can be extremely costly and may not be commercially viable in the high volume production of gears for gear trains with split torque.

It may be desirable to design a gear train that is capable of substantially equal torque sharing among all pinions of the gear train without requiring costly changes in manufacturing methods. Moreover, substantially equal torque sharing in a gear train with split torque may make it possible to at least double torque density through the gear train.

SUMMARY

A gear train is provided that includes at least one side gear comprising a helical face gear and a plurality of helical pinions in meshing engagement with the helical face gear. The gear train may further include an absorber configured to provide an axial force on at least one of the plurality of helical pinions.

A differential may also be provided including a differential case and a gear train disposed in the differential case. The gear train includes at least one side gear comprising a helical face gear and a plurality of helical pinions in meshing engagement with the helical face gear. The gear train in the differential may further include a means for providing an axial force on at least one of the plurality of helical pinions.

DETAILED DESCRIPTION

Reference will now be made in detail to embodiments of the present invention, examples of which are described herein and illustrated in the accompanying drawings. While the invention will be described in conjunction with embodiments, it will be understood that they are not intended to limit the invention to these embodiments. On the contrary, the invention is intended to cover alternatives, modifications and equivalents, which may be included within the spirit and scope of the invention as embodied by the appended claims.

Referring toFIG. 1, a gear train10in accordance with the present invention may comprise one or more side gears12, a plurality of pinions14, and an absorber16configured to provide an axial force on at least one of the plurality of pinions14. The absorber16may also be configured to absorb manufacturing errors that may cause misalignment of the pinions14and side gear12and/or deviation from proper meshing between the pinions14and side gear12.

Although a single side gear12is illustrated inFIG. 1, gear train10may include a plurality of side gears12in accordance with an embodiment of the invention. The number of side gears12may be two. Accordingly, the plurality of pinions14may engage two side gears12. Although two side gears are mentioned in detail, there may be fewer or more side gears in other embodiments. Side gears12may transmit torque from the pinions14to an output (e.g., axle shafts). Each side gear12may have an axis of rotation13. Each side gear12may also have an inner axially aligned opening15through which the axle shaft (not shown) may connect to the side gear12via a splined interconnection. Referring now toFIG. 2A, in an embodiment, each side gear12may have a face with helical teeth17. In an embodiment, a substantially flat surface may oppose the face with helical teeth17on the side gear12.

The side gears12may be located on opposing sides of the pinions14. The face of each side gear12that has helical teeth17may be configured for engagement with the pinions14. The helical teeth17of the side gear12may have tooth flanks19that feature complex geometry. Helical teeth of a conventional design on a side gear may be vulnerable to tooth pointing and/or to tooth undercutting. Referring now to the schematic illustration ofFIG. 2B, tooth pointing may occur at the outer diameter of a conventional side gear and tooth undercutting may occur at the inner diameter of a conventional side gear. In particular, tooth pointing may result in the pointing of the top profile of the tooth, such that the angle φoof the side profile of the pointed tooth is greater than angle φ of the side profile of the normal tooth. Tooth undercutting may result in the increased flattening of the top profile of the tooth, such that the angle φfof the side profile of the undercut tooth is less than the angle φ of the side profile of the normal tooth. Both tooth pointing and tooth undercutting are generally undesirable. In particular, side gear tooth pointing may reduce the torque capacity of a gear train with split torque and should be eliminated. Furthermore, side gear tooth undercutting may prohibit the forging of side gears, which is essential for high volume gear manufacturing.

The use of a helical face gear for the side gear12in an embodiment, as well as computation of appropriate design parameters for the pinions14and side gear12, may help address issues such as tooth pointing and tooth undercutting in accordance with an embodiment of the invention. Referring now toFIG. 2C, a schematic illustration of the lines of contact (e.g., contact lines L) on a tooth flank19of helical teeth17of side gear12of a gear train10is shown. Although contact line L references a single line of contact as illustrated, there may be different contact lines L at various instants of time. As the pinion14and side gear12rotate, the line of contact L travels across the tooth flank19and occupies different positions within the tooth flank19of the side gear12, as generally illustrated inFIG. 2C. Because of the geometry of the tooth flank19, the line of contact L may not only travel (e.g., migrate) across the tooth flank19and occupy different positions within the tooth flank19of the side gear12, the line of contact L may change shape.FIG. 2Cillustrates both the travel (e.g., migration) and the changing shape of the lines of contact L. The family of lines of contact L within the tooth flank19of the side gear12correspond to various different instants of time. The longer the lines of contact L, the higher the contact ratio. In an embodiment, the desired contact ratio is in the range of 1<u<1.2. A contact ratio u less than 1 (i.e., u<1) is generally undesirable as the gear mesh between the side gear12and pinions14may be interrupted under such a scenario. A contact ratio u greater than 1.2 (i.e., u>1.2) may entail a corresponding increase in helix angle on the pinion14, as well as of a spiral angle on the side gear12, and generally does not provide advantages.

The number of helical teeth17(i.e., the gear tooth number) of each side gear12may be equal to the number of pinions14times an integer number. With this tooth number, the pinions14may be distributed evenly in the circumferential direction around the side gear12. The even distribution of pinions14around the circumference of the side gear12(e.g., wherein the pinions are equi-angularly spaced around the circumference of the side gear12) may be preferred. However, in other embodiments, the pinions14may be distributed in various other configurations, including those having an uneven spacing of pinions14around the circumference of the side gear12.

The plurality of pinions14may be provided to transmit torque to the one or more side gears12and/or from one side gear12to another side gear. The number and size of the plurality of pinions14may vary. However, there is at least two pinions14in gear train10since the gear train10is designed to improve torque sharing among multiple pinions. For example, and without limitation, the number of pinions in gear train10may be six to eight pinions. Although this particular range for the number of pinions is mentioned in detail, there may be fewer or more pinions in other embodiments of the invention. In an embodiment, each of the plurality of pinions14may be generally cylindrical in shape. Further, each of the plurality of pinions14may have a first end18, a second opposing end20, and a longitudinal axis21. The pinions14may be configured to provide flexibility with respect to the number of helical teeth and gear tooth geometry. The number of teeth on the pinions14may be considered a low tooth count relative to the size of the pinions14.

Referring now toFIG. 3, a schematic illustrating an apex angle θpof a pinion14in accordance with an embodiment of the invention is shown. The apex angle θpmay be very low (e.g., less than about 20°) as compared to other gear designs. The apex angle θpmay be subject to the following equation:

θp≤sin-1⁡(do.p-dl.pdo.sg-di⁢⁢n.sg),
where do.pis the outer diameter of the pinion18, dl.pis the limit diameter of the pinion, do.sgis the outer diameter of the side gear, and din.sgis the inner diameter of the side gear. Each of the plurality of pinions14may comprise a helical pinion (i.e., have helical teeth23). The teeth23of the pinion14may appear as shown inFIG. 4, which is an illustration of the tooth23on a pinion14of a gear train10in accordance with an embodiment of the invention. The plurality of pinions14may be assembled in a housing (not shown). For example and without limitation, the housing may comprise a differential case. A ring gear (not shown) may be connected to an input source and/or drive source (not shown) in a conventional manner known in the art and may also be connected to the housing.

Because every pinion14is engaged in mesh with both side gears12simultaneously, the axial thrust of a first pinion14in mesh with a first side gear12is substantially equal to the axial thrust of a second pinion14in mesh with a second, opposing side gear12. In this way, the axial load that is acting on the housing (i.e., a differential case) from the pinions14may be of approximately zero value. By having the axial load that is acting on the housing (i.e., a differential case) be of approximately zero value, the housing may be designed with thinner walls as compared to conventional housings for gear trains (e.g., differential cases).

Torque may be transmitted in the gear train10(i.e., from the ring gear to the side gears12) either with or without the use of a torque ring25(see, e.g.,FIG. 5). When torque is transmitted in gear train10with a torque ring25, the torque ring25may be mounted within and enclosed by the housing (e.g., the differential case). The torque ring25may be generally ring shaped and may be provided for locating one or more pinions14between the side gears12. The torque ring25may include a plurality of radially inwardly extending holes27extending into the torque ring25from an outer radial surface of the torque ring25. The pinions14may be disposed within the holes27. In this way, the pinions14may be circumferentially spaced around the torque ring25. The pinions14may freely rotate in the holes27of the torque ring25. The pinions14may be axially trapped between the inner surface of the housing and a radially inwardly portion of the torque ring25. The housing and the torque ring may thus substantially or fully restrain the pinions14from axial movement. The torque ring may further include channels in its side surfaces. The channels may be configured to allow and/or confine the pinions14to be in meshing engagement with the side gears12. This is because the helical face of each side gear12may extend into one of the channels in the side surfaces of the torque ring25, and the helical teeth23of each pinion14may also extend into the opposed channels in the side surfaces of the torque ring25for engagement with the side gears12. In an embodiment, the torque ring25may exert pressure on the pinions14to move them radially about an axial center line13of the side gears12. Due to the meshing engagement between the pinions14and the side gears12, the side gears12may be forced to turn about their axial centerline13. Because the output (e.g., axle shafts) are coupled to the side gears12, the motor vehicle may move.

When torque is transmitted in gear train10without a torque ring25, the pinions14may be configured to rotate on axles (not shown), for example, as is conventional in the art. Whether the torque is transmitted with or without a torque ring25in gear train10, each of the plurality of pinions14may be in meshing engagement with two side gears12simultaneously. The side gear12and the pinions14may thus share torque via gear meshing.

With a possibility for manufacturing errors with conventional gear trains, misalignment of the axis of the pinions14and the axis13of the side gear12, as well as other deviations from proper meshing between the pinions14and the side gear12, may occur. Such misalignment and deviation may result in unequal sharing among the pinions14of the torque that is transmitted through the gear train10. Accordingly, one of the plurality of pinions14may be loaded heavier, while the rest of the plurality of pinions14may be loaded to a lesser degree. In some circumstances, some pinions14may be idle (i.e., not loaded at all). When torque is not shared equally among the plurality of pinions14, the most heavily loaded pinion14may bear the critical load and may be the first and most likely pinion to break under the load. Equal torque sharing among the plurality of pinions14may be especially advantageous for a gear train that features high torque density.

Rather than solely looking to increase accuracy in connection with the manufacturing of pinions14and side gears12that make up a gear train10(i.e., to avoid manufacturing errors, which may include significant costs), the inventive gear train10may include an absorber16configured to provide an axial force on at least one of the plurality of pinions14. The absorber16may also be configured for, among other things, absorbing manufacturing errors that may cause misalignment of the pinions14and side gear12

Referring again toFIG. 1, absorber16may comprise one or more elastic absorbers or other means for providing an axial force on at least one of the plurality of pinions14. A first absorber161may be placed at the first end18of each pinion14. A second absorber162may be placed at the second end20of each pinion14. Each absorber161,162may have a predetermined dimension (e.g., thickness) in order to absorb the manufacturing errors that may cause misalignment. The dimension of each absorber161,162is therefore, a function of, or is related to, the manufacturing error that is intended to be absorbed. In an embodiment, each absorber161,162may also exhibit a predetermined stiffness and/or rigidity, which can be expressed in terms of the design parameters of the gear train10(e.g., the manufacturing error that is intended to be absorbed) and of the amount of torque being transmitted. The amount of torque being transmitted may depend on the predetermined value of the differentiation between side gears12. Each absorber161,162may also allow slight movement for variances in the meshing engagement between the pinions14and the side gear12.

The predetermined stiffness and/or rigidity for each absorber161,162may be expressed mathematically. In an embodiment, to derive the mathematical expression for the stiffness and/or rigidity for each absorber16, the number of pinions14in gear train10may be designated as n, wherein n is ≧2. The total error to be absorbed by absorber16(e.g., absorbers161,162) may be designated as δT. Due to the various manufacturing errors, one of the pinions14is the first to be loaded, and another pinion14is the last to be loaded with a corresponding portion of the operating load. With multiple pinions14, it may not be known at what instant of time the remainder of the pinions14of gear train10are loaded and/or how the rest of the pinions are loaded. The actual load sharing among the pinions14in gear train10may be specific for each individual gear train. If the number of pinions14is sufficiently large, then a normal (i.e., Gaussian) distribution of the load among the pinions may be utilized. If the number of pinions14is not sufficiently large (e.g., there are only a few pinions14as in the design of a gear train10with split torque in accordance with an embodiment of the invention), then a reasonable assumption regarding the load sharing among pinions may be made. The pinion14that is the first to transmit torque in gear train10may be denoted as nmax, and may be more heavily loaded as compared to the remainder of the pinions14. The pinion14that is the last to transmit torque in gear train10may be denoted as nmin, and may be the least loaded pinion14as compared to the remainder of the pinions14.

The maximum torque being transmitted through the gear train may be designated TΣ. Accordingly, the maximal torque being transmitted by one pinion Tiis equal to the following equation:

Ti=T∑n.
The portion Tiof the entire torque TΣ(i.e., the maximal torque being transmitted by one pinion) is transmitted by the most heavily loaded pinion nmax. The rest of the pinions may be loaded with lower torque. The pinion that is the last to transmit torque in the gear train10(i.e., nmin) may transmit a portion of the torque that is denoted as ti. A permissible difference between Tiand timay be specified by a torque factor kT, which may be equal to the following equation:

kT=Ti-tiTi.
For example and without limitation, the torque factor kTmay be equal to kT=0.1. This torque factor kTcorresponds to that variation of the pinion loading which does not exceed 10%. This also means that torque actually transmitting through the gear train10may be 10% less than the desired torque TΣ. When the pinion14that is the first to transmit torque in gear train10(i.e., nmax) is transmitting the maximal portion Tiof the torque, then the pinion14that is the last to transmit torque in the gear train10(i.e., nmin) is transmitting the portion tiof the torque. The portion tican be expressed in terms of the gear train parameters as set forth in the following equation: ti=Ti(1−kT). For any particular application, the rest of the pinions14in the gear train10(e.g., n−2 pinions) can be equally loaded, each carrying a portion of torque τi. The portion of torque loaded onto the remainder of the pinions14can be set forth in the following equation:

τi=Ti+ti2.
Substituting for tias set forth above, the expression for the computation of the portion of torque loaded onto the remainder of the pinions14is set forth in the following equation:

τi=Ti+ti2=Ti+Ti⁡(1-kT)2=Ti⁢1-kT2.
Under this scenario, the total loadTΣ=Ti+ti+τi(n−2)<TΣ. As of the instant of time when the pinion14that is the last to transmit torque in the gear train10(i.e., nmin) is starting to transmit torque, the pinion14that is the first to transmit torque in gear train10(i.e., nmax) moves relative to the rest of the pinions14at a distance that is equal to the resultant error δT. While absorbing the error δT, torque through the pinion nmaxincreases kTtimes (i.e., the torque increases by kT·Ti). Therefore, the required stiffness qTof the absorbers161,162is equal to the following equation:

qT=kT·TiδT.
The torque capacity of the gear train10may be improved by proper load sharing among the pinions14.

Referring now toFIG. 5, a schematic illustrating the axial forces acting on each pinion14of a gear train10in accordance with an embodiment of the invention is shown. Referring now toFIGS. 1 and 5, the absorbers16may create two axial forces Paxin, Paxoutacting on every pinion14. The axial forces Paxin, Paxoutmay cancel each other out (i.e., Paxin=Paxout) when no manufacturing errors are observed. Under this circumstance, the axial forces Paxin, Paxoutare of the same magnitude and may be oppositely pointed or directed. The axial force Paxinis pointed or directed radially inwardly, toward the axis13of the side gear12. The axial force Paxoutis pointed or directed radially outwardly, away from the axis13of the side gear12. Under this scenario, equal torque sharing among all pinions14may occur. However, this scenario (i.e., in which no manufacturing errors are observed are ideal only). When a manufacturing error is present, then the equality Paxin=Paxoutis no longer valid. The inequality Paxin≠Paxoutis valid instead. The absorbers16may be configured to push the corresponding pinion14in an axial direction through a distance at which the equality Paxin=Paxoutmay be restored. In this way, equal torque sharing among the pinions14may be ensured. Absorbers16may thus axially restrain each pinion14under some compressive load so that the load is equalized on each pinion14.

Referring now toFIG. 6, a top view of several examples of a design of the absorbers161,162that may be used in a gear train10in accordance with various embodiments of the invention are illustrated. For example,FIG. 6(a) illustrates an absorber161,162comprising an elastic shim bended in a diametrical direction. Generally, to function as absorber161,162, an elastic shim must be modified to have a predetermined dimension (e.g., thickness) and predetermined stiffness.FIG. 6(b) illustrates an absorber161,162shaped in the form of a “dish” (e.g., a truncated cone)FIG. 6(c) illustrates an absorber161,162made of elastic material, such as rubber, plastic, or other similar material.FIG. 6(d) illustrates an example of a wavy absorber161,162made of steel, for example.FIG. 6(e) illustrates an absorber161,162comprising a spring.FIG. 6(f) illustrates an absorber161,162comprising a two-chamber cylinder with a piston that acts on compressed air. The thickness and rigidity and/or stiffness of each absorber may vary. Although these examples are shown and illustrated in detail for the absorbers161,162, other absorbers may used in accordance with other embodiments of the invention.

Referring now toFIGS. 7 and 8, an exploded view and a cross-sectional view of a differential22including a gear train10are respectively shown. As illustrated, differential22may comprise a locking differential. Although a locking differential is mentioned and shown in detail, the differential22may comprise other types of differential in other embodiments of the invention. For example, the differential22may comprise an open differential, a limited slip differential, and/or a locking differential in various embodiments. Differential22, in any of its various embodiments, may include a gear train10in accordance with an embodiment of the invention. Differential22may also include a differential case24, cover26, and one or more fasteners28for connecting the different case24and cover26. Differential22may be provided to allow a motor vehicle to negotiate turns, while maintaining power to both the left and right wheels of a drive axle. An open differential may allow two axle wheels to rotate at different speeds, but may provide the greatest risk of a motor vehicle remaining stuck because of the inability to force torque transfer to the wheel that has the most traction. A limited slip differential may include a means to limit the slippage associated with an open differential (e.g., a clutch pack configured to limit the amount of slippage by transferring a portion of the power from one wheel to another wheel). A locking differential may be configured to fully lock (e.g., automatically or selectably) when excessive wheel slippage occurs in order to provide full power to both wheels.

The foregoing descriptions of specific embodiments of the present invention have been presented for purposes of illustration and description. They are not intended to be exhaustive or to limit the invention to the precise forms disclosed, and various modifications and variations are possible in light of the above teaching. The embodiments were chosen and described in order to explain the principles of the invention and its practical application, to thereby enable others skilled in the art to utilize the invention and various embodiments with various modifications as are suited to the particular use contemplated. The invention has been described in great detail in the foregoing specification, and it is believed that various alterations and modifications of the invention will become apparent to those skilled in the art from a reading and understanding of the specification. It is intended that all such alterations and modifications are included in the invention, insofar as they come within the scope of the appended claims. It is intended that the scope of the invention be defined by the claims appended hereto and their equivalents.