Device for adjusting the phase angle of a camshaft of an internal combustion engine

A device for adjusting the phase angle of a camshaft of an internal combustion engine with a drive gear for driving a camshaft accommodated in a coaxial arrangement relative to the camshaft and with an electric motor communicating with the camshaft via a Harmonic Drive having a roller bearing with an elliptical inner ring, an externally toothed, flexible gear arranged on said roller bearing and a rigid, internally toothed gear engaging the externally toothed gear. A simple and compact structure is achieved by providing the electric motor with a housing which is rigidly connected to the drive gear and which is at least partially arranged therein and by also providing the electric motor with a rotor configured as a disk rotor accommodated in a slot between two housing halves of the electric motor.

BACKGROUND OF THE INVENTION
 The invention relates to a device for adjusting the phase angle of a
 camshaft of an internal combustion engine with a drive gear for driving a
 camshaft accommodated in a coaxial arrangement relative to the camshaft,
 and with an electric motor communicating with the camshaft via a Harmonic
 Drive having a roller bearing with an elliptical inner ring, an externally
 toothed, flexible gear arranged on said roller bearing and a rigid,
 internally toothed gear engaging the externally toothed gear.
 To obtain optimum values for fuel consumption and exhaust emissions in
 different areas of the internal combustion engine's operating
 characteristics, the valve timing must be varied in dependence of
 different operating parameters. An elegant manner of varying the valve
 timing is realized by rotating the camshaft relative to its driving gear.
 The camshaft of an internal combustion engine is usually driven by a
 sprocket wheel, which is connected to the crankshaft via a drive chain, or
 by a drive gear configured as a pulley, which is connected to the
 crankshaft via a toothed belt.
 DESCRIPTION OF THE PRIOR ART
 In GB 2 221 513 A a camshaft adjusting mechanism is described wherein an
 electric motor operates a set of link arms turning the camshaft relative
 to its driving gear. To this purpose an actuating element carrying the
 pivoted arms is shifted in axial direction. This solution however involves
 considerable expense and play on account of the large number of bearings.
 In DE 41 10 088 C1 and DE 39 29 619 A1 adjusting mechanisms are described
 wherein an adjusting element is provided between a member connected to the
 camshaft and a member connected to the drive gear, which element has two
 helical threads meshing with corresponding threads of the camshaft or the
 drive gear. By axially displacing this adjusting element, the camshaft can
 be rotated relative to its drive gear. Axial displacement of the adjusting
 element may be obtained by actuating a hydraulic plunger which is operated
 in dependence of the desired adjustment. The disadvantage of this solution
 is that the forces required can only be attained with a large hydraulic
 plunger necessitating considerable constructional expense. Moreover, a
 comparatively large quantity of oil is required for operating the plunger,
 which will necessitate a suitably sized pump and thus add to the engine
 load. As a further drawback of this known type of mechanism, adjustment of
 the camshaft is possible only between two extreme positions.
 An electric adjusting device also is presented in DE 41 01 676 A1 wherein
 an electric motor is provided for displacing the adjusting element by
 means of a threaded spindle. As the adjusting element rotates essentially
 at camshaft speed, an axial thrust bearing must be provided between the
 electric motor and the adjusting element, which takes up the relative
 movement between the non rotating and the rotating member. In the above
 solution, the thrust bearing is more or less permanently subject to load
 throughout the entire operating period, since the torsional moments acting
 between drive gear and camshaft will produce a force acting on the
 adjusting element in axial direction. For this reason the thrust bearing
 is a critical component which will limit the useful life of the engine. A
 similar solution is disclosed in DE 33 20 835 A1, wherein the same
 disadvantages are encountered.
 In DE 36 07 256 A a mechanism is described, wherein a stepping motor is
 provided for adjusting the camshaft, said stepping motor being connected
 to both camshaft and drive gear. As the stepping motor must take up the
 entire driving torque for the camshaft, such a solution cannot be achieved
 within reasonable limits of expense.
 EP 0 596 860 A discloses a device for adjusting the valve opening times in
 which the camshaft has a hollow configuration and comprises an inner
 shaft. The cams are bipartite, wherein each single cam section can be
 turned relative to the other by a determined angle. The rotation of the
 two cam sections is executed by a revolving electric motor, which is
 supplied via slip rings. Similar solutions are disclosed in U.S. Pat. Nos.
 5,417,186 and 4,770,060.
 A former suggestion of the applicant, published in EP-A 0 903 471, presents
 an adjusting mechanism for the phase angle of a camshaft with a planetary
 gear set, in which the adjustment is executed by an electric motor that is
 supplied with current by means of slip contacts.
 In most of the solutions described above, the electromotive adjustment is
 made via a set of gears with a big gear reduction ratio in order to
 maintain the torque on the engine in an acceptable range. To transmit the
 force via an axially slidable element with screw-shaped teeth has the
 advantage that very big gear reduction ratios may obtained quite easily.
 The disadvantage of these solutions however is that the friction is quite
 high and that accordingly high forces originate in axial direction which
 have to be absorbed by the bearings. On the other side, the transmission
 of force via a planetary gear set is quite complicated and it is difficult
 to achieve big gear reduction ratios.
 By using a so-called Harmonic Drive, a big gear reduction ratio may be
 obtained easily. If for example the number of teeth of the externally
 toothed gear is smaller by two than the number of teeth of the internally
 toothed gear which amount to for example 50, a step-up ratio of about 1:25
 is achieved.
 In the concrete embodiment of a device of the type mentioned above however
 and on top of the already described questions, the following groups of
 problems have to be overcome in construction: the bearing of the electric
 motor has to be as simple as possible, so that the Harmonic Drive is not
 submitted to inadmissible big loads. Furthermore, the bearing of the
 electric motor is not allowed to hinder the necessary sealing between the
 oil guiding sections (camshaft and set of gears) and those sections that
 do not guide oil. It also is necessary that the bearing of the drive gear
 be realized in the simplest possible way. Particular attention should
 hereby be paid to the restricted building space available on the front
 side of the cylinder head of internal combustion engines.
 Another constructional aim to be achieved is to guarantee an emergency
 operation of the motor in case of failure of the adjusting device and
 generally, to restrict the adjusting range to allowable angles.
 SUMMARY OF THE INVENTION
 It is an object of the present invention to develop a device as described
 above in such a manner that the above mentioned requirements may be met
 with as little expenditure as possible.
 According to the invention, the electric motor is provided with a housing
 rigidly connected to the drive gear and at least partially arranged
 therein and the electric motor also is provided with a rotor configured as
 a disk rotor accommodated in a slot between two housing halves of the
 electric motor.
 Disk rotors are electric motors with an axial air gap, their rotors
 consisting of a disk-shaped base body on which the windings are applied in
 the shape of a thin layer. In many cases, the windings are applied in the
 form of punching bodies applied onto a disk of artificial resin or metal
 like aluminum or steel. In the case of direct current motors, the windings
 are energized by carbon brushes pressed in axial direction against the
 disk. Such disk rotors have been described in DE 31 07 834 A, DE 32 34 274
 A or in DE 32 42 394 A for example. The disk rotor can be manufactured by
 making a winding of a copper wire, putting the winding into a mold and
 pressing it to a disk-shaped form and injecting plastic material into the
 mold to embed the winding. The sliding planes for the carbon brushes may
 be produced by grinding the disk in the respective area.
 The configuration according to the invention makes it possible to give the
 adjusting device the smallest possible axial measurements. Since the rotor
 of the electric motor is configured as a disk rotor, a high torque may be
 achieved by a relatively large diameter. A particularly advantageous
 solution is achieved when the disk rotor is arranged within the drive
 gear. The building space inside the drive gear is hereby used for the
 adjusting device.
 Furthermore, the drive gear may be configured as a pulley or toothed belt
 gear supported on the camshaft by roller bearings. The advantage thereof
 is that the electric motor does not need any bearing since it is
 configured integral with the drive gear. By having the drive gear borne on
 the camshaft, the relative movement in the area of the Harmonic Drive is
 minimized so that excessive load caused by oscillations or the like may be
 excluded.
 For motors having two camshafts per cylinder row it is particularly
 advantageous to have the pulley provided with two engagement surfaces with
 different outer diameters, one of said engagement surfaces being designed
 to receive a toothed belt which is driven by a crankshaft of the internal
 combustion engine and the other engagement surface being designed to
 receive a toothed belt which is provided for driving another camshaft. In
 this way, only one camshaft has to be driven directly by the crankshaft,
 which simplifies the guiding of the corresponding toothed belt. Thus, the
 spacing between the two camshafts may be reduced.
 A particularly space-saving solution is achieved when slip rings designed
 for the supply of the electric motor are configured on a front surface of
 the drive gear.
 It is particularly preferable when the electric motor is provided with a
 shaft that is directly connected to the elliptical inner ring of the
 roller bearing. It provides benefits to have the flexible gear of the
 Harmonic Drive directly connected to the camshaft. In this case, the
 internally toothed gear of the Harmonic Drive is connected to the drive
 gear. In this way, a particularly simple structure of the solution
 according to the invention may be brought forth.
 The disadvantage of conventional Harmonic Drives is that they only can be
 loaded to a limited extent because of their relatively fine teeth. Such
 gears are particularly sensitive to impulsive loads. In order to avoid
 this disadvantage, the gear may be made of the following component parts:
 a first plane of action arranged on the inner periphery of a first
 engaging part, a second plane of action arranged on the outer periphery of
 a flexible engaging part and engaging the first plane of action and a
 driving member arranged coaxially to the first engaging part and to the
 flexible engaging part, a roller bearing provided with a non circular
 inner ring being accommodated on said driving member and having a flexible
 outer ring connected to the flexible engaging part and pushing it at
 preferably two points against the first engaging part, wherein the first
 plane of action of the first engaging part frictionally engages the second
 plane of action of the flexible engaging part.
 In such a gear, the evident allocation of the different component parts
 with regard to the phases and the exact transmission ratio of a toothed
 gearing is no longer given, but resistance to overload may thus be
 achieved, which is not possible with a toothed gearing. Furthermore, the
 gear according to the invention is unaffected by dirt and requires little
 lubrication. A further advantage of the invention is that the eccentricity
 of the wave generator may be considerably smaller than in a toothed
 gearing of the art. In those conventional toothed gearings, it is
 necessary to make the eccentricity so big that the teeth of the first and
 of the second plane of action do not touch each other outside the engaging
 areas. In the solution according to the invention, the eccentricity is
 only defined by the small path needed to establish a frictional
 engagement. That is why the deformation of the flexible engaging part
 during operation is considerably smaller, which reduces losses and
 increases service life.
 In principle it is possible to have the first and the second plane of
 action meshing at one, two, three or more points. It proved particularly
 advantageous however to provide two diametrically opposite meshing points.
 In such a solution, the inner ring of the roller bearing is essentially
 elliptical in cross section.
 Secure transmission of force may particularly be achieved by giving the
 first plane of action a conical shape with a small aperture angle.
 Eventual wear can thus be compensated, too. In this connection, it is
 particularly advantageous to have the aperture angle amounting to between
 1.degree. and 1.degree., preferably between 2.degree. and 6.degree.. In
 this case, the wave generator preferably is embodied in such a manner that
 the outer ring of the roller bearing has got a conical shape with a small
 aperture angle. The aperture angle of the outer ring should thereby
 essentially match the aperture angle of the first plane of action.
 It is particularly advantageous to provide a pressure means that presses
 the first plane of action in axial direction against the second plane of
 action. An independent adjustment of the pressure force and with it of the
 transmissible torque may thus be achieved. The pressure means preferably
 is provided with a spring that prestresses the first engaging part and the
 flexible engaging part against each other in axial direction.
 A particularly favorable constructional solution is achieved by having the
 retaining elements configured as bolts that are guided through oblong
 holes in the form of circular arcs provided in the drive gear. A
 particularly advantageous lateral guiding of the drive gear is achieved
 when the drive gear is by one side adjacent to a shoulder of the camshaft
 and by the other side to shoulders of sleeves being pressed against the
 camshaft by the bolts.
 A particularly good exploitation of the building space at the biggest
 possible torque is achieved when the diameter of the disk rotor amounts to
 between 60 and 90% of the outer diameter of the drive gear.
 In a particularly preferred embodiment of the invention a means for
 limiting rotation is provided between the camshaft and the drive gear,
 said means consisting of a stop element received in a groove so as to be
 capable of moving in a restricted way. The rotation of the camshaft is
 thus limited to allowable values. This measure also makes certain of
 ensuring a defined position of the camshaft in case of failure of the
 adjusting device or in case of cold start. Actually, the Harmonic Drive is
 self-locking due to the big gear reduction ratio, but, due to the
 ineluctable rotational oscillations in the system, a position of the
 camshaft corresponding to the latest possible valve opening time is almost
 immediately reached when the electric motor is idle. It is therefore
 possible to tune the electronics of the motor in such a way that, under
 these conditions, acceptable emergency operation is possible.
 The present invention will be described more explicitly in the following
 with the help of the embodiment illustrated in the figures.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
 FIG. 1 shows a device according to the invention for adjusting a camshaft 1
 via an electric motor 2. The camshaft 1 is driven by a drive gear 3
 configured as a pulley for a toothed belt. One end of the camshaft 1 is
 rigidly connected to one drive part 4 which is sealed against the housing
 5 by means of a shaft seal 6.
 Two roller bearings 7a, 7b configured as ball bearings support the drive
 gear 3 on the drive part 4. The electric motor 2 consists of an
 essentially cylindrical housing 2a that is rigidly connected to the drive
 gear 3. At its front sides, the housing 2a is terminated by housing halves
 2b and 2c. The windings of the electric motor 2 are arranged in the
 housing halves 2b and 2c. A shaft 14 also is arranged in the housing
 halves 2b, 2c via roller bearings, said shaft being rigidly connected to
 the elliptical inner ring 16 of a roller bearing 17. In the literature,
 this part of a Harmonic Drive is mostly called a wave generator. The outer
 ring 18 of the roller bearing 17 is connected to a flexible, externally
 toothed gear 19 that is configured as a whole like a pot and that
 communicates with the camshaft. The external thread of the gear 19 meshes
 with a rigid, internally toothed gear 22 at two points, namely at the
 apexes of the ellipse formed by the wave generator. The externally toothed
 gear is rigidly connected to the drive gear 3 via screws 23.
 A disk-shaped support 13 is rigidly connected to a front side of the drive
 gear 3 via screws 13a. Slip rings 10 communicating with stationary slip
 contacts 12 configured as carbon brushes are arranged on the support 13.
 The drive gear 3 has two engagement surfaces 3a and 3b with different
 diameters. The engagement surface 3a has the larger diameter and is
 designed to receive a toothed belt (not shown) for driving the drive gear
 3a via the crankshaft of the internal combustion engine which is not
 illustrated in the drawing herein. The engagement surface 3b has a smaller
 diameter and is designed to receive a toothed belt (not shown either)
 driving another camshaft. The slip rings 12 are located in the free space
 within this toothed belt that is not illustrated in the drawing herein.
 A disk rotor 25 is accommodated between the housing halves 2b and 2c of the
 electric motor 2. The outer diameter d of the disk rotor 25 amounts to
 approximately 80% of the outer diameter D of the drive gear 3. It is
 hereby completely accommodated within the drive gear 3 and altogether more
 than half the electric motor 2 is located inside the drive gear 3.
 Operation of the device of the present invention will be described in the
 following. When the electric motor 2 is not energized, the camshaft 1 is
 not adjusted relative to the drive gear 3. But when the shaft 14 of the
 electric motor completes one revolution, the rotation of the wave
 generator causes all the teeth of the flexible gear 18 to engage
 successively the internally toothed gear 22. If for example the externally
 toothed gear 18 has forty-eight teeth, whereas the internally toothed gear
 22 has fifty teeth, a relative motion of these two gears takes place to
 the extent of two teeth, i.e. of one twenty-fifth revolution. This also
 corresponds to the angle of rotation by which the camshaft 1 rotates
 relative to the drive gear 3. Thanks to the big gear reduction ratio, the
 torque that has to be produced by the electric motor 2 is quite small. In
 the structure according to the invention, the bearing of the drive gear 3
 is very simple and, since the connection is made directly via the roller
 bearings 7a, 7b, the clearance occurring in the Harmonic Drive is small.
 The screws 23 are arranged in mating recesses 26 having the shape of a
 ring segment and being located in the supporting part 4 and act as means
 to limit rotation. This simultaneously delimits the allowable adjusting
 angle of the camshaft in order to prevent excessive adjusting movement.
 FIG. 2 shows details of a Harmonic Drive in an axonometric exploded view.
 The inner ring 11 of the roller bearing 10 is elliptic with a slight
 eccentricity. The outer ring 9 is directly supported by the inner side of
 a flexible gear 6. This gear 6 meshes at two diametrically opposite points
 with a rigid, internally toothed gear 8, which has got a circular shape.
 FIGS. 3A, B, C and D show the mode of operation of this Harmonic Drive. In
 the position shown in FIG. 3B, the inner ring 11 is rotated 90.degree.
 clockwise relative to the position shown in FIG. 3A. The FIG. 3C shows a
 further rotation by 90.degree. and FIG. 3D one complete revolution by
 360.degree.. For the sake of clarity, an arrow 11a was introduced into the
 FIG. 3A, B, C and D. The number of teeth of the flexible gear 6 is smaller
 by two than the number of teeth of the internally toothed gear 8. A small
 difference in angular velocity between gear 6 and gear 8 arises out of it.
 As may be seen in the Figures, the sign 6a that alludes to the gear 6 is
 moved slowly counterclockwise while the inner ring 11 is turning. As a
 whole, the rotating angle corresponds to the central angle of two teeth of
 the gear 6.
 FIGS. 4 through 6 and 6A show alternative variants of a Harmonic Drive that
 may be used in an embodiment of the invention according to FIG. 1.
 The gear of FIG. 4 consists of a motor shaft 100 provided on its end with a
 bearing surface 101 on which a roller bearing 102 configured as a ball
 bearing is wedged up. The roller bearing 102 has an inner ring 103 whose
 outer periphery is elliptical. The flexible outer ring 105 gets its
 elliptical shape by the rollers 104. Since the outer periphery of the
 outer ring 105 is slightly beveled, the ring as a whole adopts the shape
 of an elliptical cone. A flexible engaging part 106 sits close to the
 outer ring 105. The second plane of action 107 is arranged on the outer
 periphery of the engaging part 106, said second engaging part engaging a
 first engaging part 108 arranged on the inner periphery of a first
 engaging part 109. The first plane of action 109 has got the shape of a
 circular cone. An output shaft 110 is communicating with the engaging part
 106. A thin-walled inner ring of uniform thickness that is fixed on an
 elliptical bearing surface may be used instead of an elliptical inner ring
 103, so that an elliptical circumferential surface is obtained by
 resilient deformation.
 The way of operation of the device according to the invention is explained
 more thoroughly in the following. In the position illustrated in FIG. 10,
 the planes of action 107 and 108 are in contact in the areas 111 and 112,
 which are opposite one another. When the motor shaft 100 rotates, these
 areas 111 and 112 creep along the first plane of action 108 until they
 reach their original position. The second plane of action 107 thereby
 rolls off on the first plane of action 108. Since the circumference
 U.sub.1 of the first plane of action 108 is slightly bigger than the
 circumference U.sub.2 of the second plane of action 107, the engaging part
 106 turns slightly in the opposite direction of rotation from the motor
 shaft 100. The transmission ratio i, which is defined by the speed of the
 motor shaft over the speed of the output shaft, corresponds to the
 reciprocal value of the eccentricity .epsilon., when the latter is defined
 according to the following equation:
EQU .epsilon.=(U.sub.1 -U.sub.2)/U.sub.1
 Depending on the material chosen, values of i=1/.epsilon. may be achieved
 in a range of 100 through 300 and more. Generally speaking, the harder the
 substances used for the planes of action 107 and 108, the bigger the
 transmission ratio may become.
 FIG. 6 shows a set of gears according to the invention that is part of an
 adjusting device for the camshaft of an internal combustion engine, which
 has not been illustrated in detail. A frictional wheel 206, configured as
 a thin-walled cylinder and constituting the flexible engaging part, is
 fixed to the camshaft 201 by means of a screw 204. A sleeve 202 having a
 sprocket wheel 203 integrated to it is rotatable relative to the camshaft
 201. A roller bearing 210 is given an elliptical shape so that the
 frictional wheel 206 engages an engaging area of the sleeve 202 at only
 two diametrically opposite points. The circumference of the frictional
 wheel 206 is slightly smaller than the circumference of the plane of
 action 208, which is arranged on a ring 220 that may be displaced in axial
 direction relative to the sleeve 202 and that is configured as a hollow
 gear. The inner ring of the roller bearing 210 is rigidly connected in a
 drive element 212 communicating with an adjusting motor (not shown). To
 make sure that the required pressure of the frictional wheel 206 acting
 against the ring 220 is permanent, a spring 221 is provided that
 prestresses the ring 220 relative to the sleeve 202.
 In FIG. 6A, an exploded view shows the structure of the above mentioned set
 of gears. Although the ring 220 is axially slidable in guides 219 on the
 sprocket wheel 203, it is non rotatably borne. The drive element 212 is
 connected to a support 214 for the inner ring of the roller bearing 210
 via a ring 213. The connection is secured by a disk 215 and by a Seeger
 circlip ring.