U.S. patent application number 11/491994 was filed with the patent office on 2006-11-23 for synchronous drive apparatus and methods.
This patent application is currently assigned to LITENS AUTOMOTIVE. Invention is credited to Witold Gajewski.
Application Number | 20060264285 11/491994 |
Document ID | / |
Family ID | 26988561 |
Filed Date | 2006-11-23 |
United States Patent
Application |
20060264285 |
Kind Code |
A1 |
Gajewski; Witold |
November 23, 2006 |
Synchronous drive apparatus and methods
Abstract
A synchronous drive apparatus and method, wherein the apparatus
comprises a plurality of rotors comprising at least a first and a
second rotor. The first rotor has a plurality of teeth for engaging
the engaging sections of an elongate drive structure, and the
second rotor has a plurality of teeth for engaging the engaging
section of the elongate drive structure. A rotary load assembly is
coupled to the second rotor. The elongate drive structure engages
about the first and second rotors. The first rotor is arranged to
drive the elongate drive structure and the second rotor is arranged
to be driven by the elongate drive structure. One of the rotors has
a non-circular profile having at least two protruding portions
alternating with receding portions. The rotary load assembly is
such as to present a periodic fluctuating load torque when driven
in rotation, in which the angular positions of the protruding and
receding portions of the non-circular profile relative to the
angular position of the second rotor, and the magnitude of the
eccentricity of the non-circular profile, are such that the
non-circular profile applies to the second rotor an opposing
fluctuating corrective torque which reduces or substantially
cancels the fluctuating load torque of the rotary load
assembly.
Inventors: |
Gajewski; Witold; (Richmond
Hill, CA) |
Correspondence
Address: |
PILLSBURY WINTHROP SHAW PITTMAN, LLP
P.O. BOX 10500
MCLEAN
VA
22102
US
|
Assignee: |
LITENS AUTOMOTIVE
Woodbridge
CA
|
Family ID: |
26988561 |
Appl. No.: |
11/491994 |
Filed: |
July 25, 2006 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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11101597 |
Apr 8, 2005 |
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11491994 |
Jul 25, 2006 |
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10294933 |
Nov 15, 2002 |
7044875 |
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11101597 |
Apr 8, 2005 |
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60333118 |
Nov 27, 2001 |
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60369558 |
Apr 4, 2002 |
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Current U.S.
Class: |
474/141 ;
474/148 |
Current CPC
Class: |
F01L 2820/01 20130101;
F01L 1/022 20130101; F16H 9/24 20130101; F01L 1/02 20130101; F01L
2810/03 20130101; F16H 7/023 20130101; F16F 15/264 20130101; F16H
55/084 20130101; Y10T 74/19884 20150115; F02B 75/06 20130101; F16H
2035/003 20130101; F01L 1/46 20130101; F01L 1/356 20130101; F01L
1/024 20130101; F02B 67/06 20130101; F16H 35/02 20130101; F16H
57/0006 20130101; Y10T 74/1987 20150115; F16H 55/171 20130101 |
Class at
Publication: |
474/141 ;
474/148 |
International
Class: |
F16H 55/30 20060101
F16H055/30; F16H 7/00 20060101 F16H007/00 |
Claims
1. An improved rotor for use in a synchronous drive apparatus, said
drive apparatus comprising a continuous-loop elongate drive
structure having a plurality of engaging sections; a plurality of
rotors comprising at least a first and a second rotor, the first
rotor having a plurality of teeth for engaging the engaging
sections of the elongate drive structure, and the second rotor
having a plurality of teeth for engaging the engaging section of
the elongate drive structure; a rotary load assembly coupled to the
second rotor; the rotary load assembly being such as to present a
periodic fluctuating load torque when driven in rotation; the
elongate drive structure being engaged about the first and second
rotors, the first rotor being arranged to drive the elongate drive
structure and the second rotor being arranged to be driven by the
elongate drive structure, and one of the rotors having a
non-circular profile having at least two protruding portions
alternating with receding portions; wherein the improved
non-circular profile rotor has at least two reference radii, each
reference radius passing from the centre of the non-circular
profile rotor and through the centre of a protruding portion of the
non-circular profile, the angular position of the non-circular
profile being related to a reference direction, the reference
direction being the direction of a vector that bisects an angle
about which the elongate drive structure is wrapped about the rotor
having the non-circular profile, the angular position of a
reference radius is within a range of 90.degree. to 180.degree.
from the reference direction taken in the direction of rotation of
the rotor on which the non-circular profile is formed.
2. Apparatus according to claim 1, in which the angular position of
the reference radius being within a range of 130.degree. to
140.degree. from the reference direction taken in the direction of
rotation of the rotor on which the non-circular profile is
formed.
3. Apparatus according to claim 1, in which the angular position of
the reference radius is substantially at 135.degree. from the
reference direction taken in the direction of rotation of the rotor
on which the non-circular profile is formed.
Description
[0001] This application is a continuation application of Ser. No.
11/101,597, filed Apr. 8, 2005, which claims benefit of and is a
divisional of application Ser. No. 10/294,933, filed Nov. 15, 2002,
which claims priority to U.S. Provisional Application Nos.
60/333,118, filed Nov. 27, 2001 and 60/369,558, filed Apr. 4, 2002,
the entirety of each of the these applications is hereby
incorporated into the present application by reference thereto,
respectively.
FIELD OF INVENTION
[0002] The present invention relates to a synchronous drive
apparatus, a method of operating a synchronous drive apparatus and
a method of constructing a synchronous drive apparatus. The
invention relates to the elimination or reduction of mechanical
vibrations, in particular but not exclusively in internal
combustion engines.
BACKGROUND OF INVENTION
[0003] Synchronous drive systems, such as timing belt-based
systems, are widely used in motor vehicles, as well as in
industrial applications. In motor vehicles, for example, timing
belts or chains are used to drive the camshafts that open and close
the engine intake and exhaust valves. Also other devices such as
water pumps, fuel pumps etc. can be driven by the same belt or
chain.
[0004] Internal combustion engines produce many types of mechanical
vibrations during their operation, and these vibrations are usually
transmitted through the timing belt or chain in the synchronous
drive system. A particularly intense source of mechanical
vibrations is given by the intake and exhaust valves and the
camshafts that open and close those intake and exhaust valves.
Opening and closing the intake and exhaust valves leads to a type
of vibration known as torsional vibration. When the frequency of
these vibrations is close to natural frequency of the drive, system
resonance occurs. In resonance the torsional vibrations and the
span tension fluctuations are at their maximum.
[0005] As flexible mechanical structures, timing belts and chains
are particularly susceptible to the deleterious effects of
mechanical vibrations. Mechanical vibrations transmitted through
the timing belt or chain cause fluctuations in belt or chain
tension, which can lead to increased wear and reduced belt or chain
life. Vibrations may also cause timing errors, and result in
undesirable amounts of noise.
[0006] Conventional techniques to attenuate the vibrations include
increasing the tension on the belt or chain and installing camshaft
dampers. Camshaft dampers connect a source of inertia to a camshaft
sprocket by a vibration-absorbing rubber or silicone. However,
increasing the belt or chain tension increases the noise level and
reduces the useful life of the belt or chain. Installing camshaft
dampers is also an undesirable solution, because of their cost
and/or because of lack of space.
[0007] In DE-A-195 20 508 (Audi AG), there is disclosed a wrapped
belt drive for an internal combustion engine, the timing belt being
wrapped around two driven pulleys coupled to the camshaft of the
engine, and one drive pulley coupled to the crankshaft of the
engine. The objective of the invention is to counter the torsional
vibrations which are found in such belt drives. It is proposed to
provide an additional torsional vibration through which the
critical resonance can be moved to a range where it can either be
tolerated, or does not arise. It is proposed in the citation to
produce torsional vibrations by an "out of round" pulley, which is
shown as consisting of one of the camshaft pulleys. The out of
round pulley which is shown has four protruding portions and four
receding portions arranged regularly around the pulley. It is said
that the variations in the pulley profile introduce torsionals to
the timing belt at the incoming or outgoing spans of the driven
pulleys, which are superimposed on the dynamics of the combustion
engine, and thus shift or eliminate the critical resonance range. A
figure is shown which is said to show a graph of torsional
vibrations of the timing drive in degrees camshaft over the RPM of
the crankshaft. The total amplitude is shown, and also the dominant
vibration of the second order and the less relevant vibrations of
the fourth order are shown. A single example of a magnitude of
eccentricity of an out of round pulley is given, but no teaching is
given as to how to select the magnitude of the eccentricity, and
the angular alignment of the out of round rotor relative to the
other rotors, for any given conditions of type of engine, type of
drive belt, and type of load. As has been mentioned, the objective
of the invention in the citation is to counter the torsional
vibrations in the belt drive, and not to deal with the source of
the vibrations.
[0008] In Japanese Utility Model JP 62-192077 (Patent Bulletin No.
HEI 1-95538) of 1987 (Hatano et al/Mitsubishi), there is disclosed
a tension equalising driving device which transmits the rotation of
a drive pulley to a driven pulley by a belt drive such as a timing
belt in an internal combustion engine. There is shown a timing belt
arrangement in which a toothed pulley of the drive shaft of a
camshaft is driven by an oval timing belt driving sprocket
connected to the drive shaft of an internal combustion engine. The
teaching of the citation is that the drive pulley is made oval in
shape so as to give the drive belt a tension fluctuation with a
phase opposite to that of the tension fluctuation in the belt
produced by the rotation of the internal combustion engine. It is
said that the drive pulley is installed in such a way that it gives
the drive belt a tension fluctuation with a phase opposite to that
of the tension fluctuation of the belt already present. The oval
drive sprocket is said to be a tension equalising device, and is
provided to equalise the tension in the drive belt. A figure is
shown of a graph illustrating the tension caused by the valve train
torque and the tension caused by the tension equalising device (the
oval drive sprocket), the two tensions being shown of the same
magnitude and opposite phase. There is no specific teaching given
as to how to determine the magnitude of the eccentricity of the
oval drive pulley, nor how to relate the angular position of the
drive pulley to the camshaft pulley which is driven by the belt. In
addition, as discussed in Japanese Application No. HEI 9-73581
(Patent Bulletin No. HEI 10-266868) of 1997 (Kubo/Mitsubishi), it
was subsequently determined by the Applicant in JP 62-192077 (HEI
1-95538) that the use of an oval sprocket as a crank sprocket has a
number of difficulties and problems and is thus not desirable.
SUMMARY OF INVENTION
[0009] In accordance with the present invention in a first aspect,
there is provided a synchronous drive apparatus, comprising a
continuous-loop elongate drive structure having a plurality of
engaging sections. A plurality of rotors comprising at least a
first and a second rotor, wherein the first rotor has a plurality
of teeth for engaging the engaging sections of the elongate drive
structure, and the second rotor has a plurality of teeth for
engaging the engaging section of the elongate drive structure. A
rotary load assembly is coupled to the second rotor. The elongate
drive structure engages about the first and second rotors. The
first rotor is arranged to drive the elongate drive structure and
the second rotor is arranged to be driven by the elongate drive
structure. One of the rotors has a non-circular profile having at
least two protruding portions alternating with receding portions.
The rotary load assembly is such as to present a periodic
fluctuating load torque when driven in rotation, in which the
angular positions of the protruding and receding portions of the
non-circular profile relative to the angular position of the second
rotor, and the magnitude of the eccentricity of the non-circular
profile, are such that the non-circular profile applies to the
second rotor an opposing fluctuating corrective torque which
reduces or substantially cancels the fluctuating load torque of the
rotary load assembly.
[0010] In preferred forms of the apparatus, the non-circular
profile is such as to produce the opposing fluctuating corrective
torque by periodic elongation and contraction of the spans of the
elongate drive structure adjoining the rotor on which the
non-circular profile is formed. The elongate drive structure has a
drive span on the tight side of the rotor on which the non-circular
profile is formed, the angular position of the non-circular profile
being within +/-15 degrees (preferably within +/-5 degrees) of an
angular position for which a maximum elongation of the drive span
coincides with a peak value of the fluctuating load torque of the
rotary load assembly. Most preferably the angular position of the
non-circular profile is that for which a maximum elongation of the
drive span substantially coincides with a peak value of the
fluctuating load torque of the rotary load assembly.
[0011] Also in preferred forms of the apparatus, the magnitude of
the eccentricity of the non-circular profile is such that the
fluctuating corrective torque has an amplitude in the range of 70%
to 110% (preferably in the range 90% to 100%) of the amplitude of
the fluctuating load torque at a predetermined selected set of
operating conditions of the synchronous drive apparatus. Most
preferably, the amplitude of the fluctuating corrective torque is
substantially equal to the amplitude of the fluctuating load
torque.
[0012] In this specification, unless otherwise indicated, the term
amplitude of a periodically varying item means peak-to-peak
amplitude.
[0013] Thus, the magnitude of the eccentricity of the non-circular
profile is determined with reference to the amplitude of the
fluctuating load torque of the rotary load assembly. In some
arrangements the amplitude of the fluctuating load torque may be
substantially constant, and in other arrangements the amplitude of
the fluctuating load torque may vary. Where the amplitude of the
fluctuating load torque is constant, the magnitude of the
eccentricity is determined with reference to that substantially
constant amplitude of fluctuating load torque. Where the amplitude
of the fluctuating load torque varies, the value thereof which is
used to determine the magnitude of the eccentricity will be
selected according to the operating conditions in which it is
desired to eliminate or reduce the unwanted vibrations. For example
where the fluctuating load torque of the rotary load assembly
varies, the eccentricity may be determined with reference to the
amplitude of the fluctuating load torque when determined at
conditions such that it is a maximum, or for example when
determined at the natural resonance frequency of the apparatus. For
example in a diesel internal combustion engine, the most
troublesome region for vibration may be at the maximum fuel
delivery by the fuel pump. In these conditions, the eccentricity is
determined with reference to the amplitude of the fluctuating load
torque when determined at these conditions. Similarly in a petrol
or gasoline internal combustion engine, the most troublesome region
may be at the region of natural resonance of the timing drive, and
in such a case the eccentricity is determined with reference to
such conditions.
[0014] It is to be appreciated that the invention finds application
in many forms of synchronous drive apparatus other than in internal
combustion engines. Also, the non-circular profile may be provided
in many different locations within the drive apparatus. For example
a non-circular profile may be provided on the first rotor (which
drives the elongate drive structure), and/or on the second rotor
(which is driven by the elongate drive structure), and/or may be
provided on a third rotor, for example an idler rotor urged into
contact with the continuous loop elongate drive structure.
[0015] However, the invention finds particular use when installed
in an internal combustion engine and the first rotor comprises a
crankshaft sprocket. In some arrangements the internal combustion
engine is a diesel engine, and the rotary load assembly comprises a
rotary fuel pump. As has been mentioned in such arrangements, it
may be arranged that the magnitude of the eccentricity of the
non-circular profile is such that the fluctuating corrective torque
has an amplitude substantially equal to the amplitude of the
fluctuating load torque when determined at conditions of maximum
delivery of the fuel pump. In other arrangements, the internal
combustion engine may be a petrol or gasoline engine and the rotary
load assembly may be a camshaft assembly.
[0016] In determining the angular position of the non-circular
profile, consideration may be given to various reference parameters
of the profile and the rotor on which it is formed. In some
arrangements the non-circular profile has at least two reference
radii, each reference radius passing from the centre of the rotor
on which the non-circular profile is formed and through the centre
of a protruding portion of the non-circular profile, and the
angular position of the non-circular profile is related to a
reference direction of the rotor on which the non-circular profile
is formed, the reference direction being the direction of the hub
load force produced by engagement of the elongate drive structure
with that rotor. The angular position of the non-circular profile
is such that, when the fluctuating load torque of the rotary load
assembly is at a maximum, the annular position of a reference
radius is preferably within a range of 90.degree. to 180.degree.
from the reference direction taken in the direction of rotation of
the rotor on which the non-circular profile is formed. Preferably,
the range comprises a range of 130.degree. to 140.degree.. Most
preferably, the angular position of the reference radius is
substantially at 135.degree. from the reference direction taken in
the direction of rotation of the rotor on which the non-circular
profile is formed.
[0017] It will be appreciated that many different forms of
non-circular profile may be provided, for example a generally oval
profile, or a profile having three or four protruding portions
arranged regularly around the rotor. The choice of profile will
depend upon other components of the synchronous drive apparatus.
Examples which may be provided include the following, namely: the
internal combustion engine is a 4-cylinder inline combustion engine
and the crankshaft sprocket has an oval contoured profile; the
internal combustion engine is a 4-cylinder inline combustion engine
and the camshaft sprocket has a generally rectangular contoured
profile; the internal combustion engine is a 4-cylinder inline
combustion engine, and the camshaft sprocket has a generally
rectangular contoured profile and the crankshaft sprocket has an
oval contoured profile; the internal combustion engine is a
3-cylinder inline combustion engine and the camshaft sprocket has a
generally triangular contoured profile; the internal combustion
engine is a 6-cylinder inline combustion engine and the crankshaft
sprocket has a generally triangular contoured profile; the internal
combustion engine is a 6-cylinder V6 combustion engine and the
camshaft sprocket has a generally triangular contoured profile; the
internal combustion engine is an 8-cylinder V8 combustion engine
and the camshaft sprocket has a generally rectangular contoured
profile; or the internal combustion engine is a 2-cylinder
combustion engine and the camshaft sprocket has an oval contoured
profile.
[0018] In most embodiments of the invention as set out above, the
protruding portions and receding portions will be generally of the
same magnitude, giving a regular non-circular profile. However
depending upon the circumstances of the torsional vibrations to be
removed, a non-regular profile may be provided. Furthermore, the
protruding portions referred to above may constitute major
protruding portions and the receding portions constitute major
receding portions, and the non-circular profile may include
additional minor protruding portions of lesser extent than the
major protruding portions. These minor protruding portions may be
adapted to produce additional, minor, fluctuating corrective torque
patterns in the torque applied to the second rotor, for the purpose
of reducing or substantially cancelling subsidiary order
fluctuating load torque presented by the rotary load assembly, in
particular for example in order to reduce or substantially cancel
fourth order fluctuating load torques presented by the rotary load
assembly.
[0019] It is to be appreciated that where features of the invention
are set out herein with regard to apparatus according to the
invention, such features may also be provided with regard to a
method according to the invention (namely a method of operating a
synchronous drive apparatus, or a method of constructing a
synchronous drive apparatus), and vice versa.
[0020] In particular, there is provided in accordance with another
aspect of the invention a method of operating a synchronous drive
apparatus which comprises a continuous-loop elongate drive
structure having a plurality of engaging sections. A plurality of
rotors comprises at least a first and a second rotor. The first
rotor has a plurality of teeth engaging the engaging sections of
the elongate drive structure, and the second rotor has a plurality
of teeth engaging the engaging section of the elongate drive
structure. A rotary load assembly is coupled to the second rotor.
One of the rotors has a non-circular profile having at least two
protruding portions alternating with receding portions. The rotary
load assembly presents a periodic fluctuating load torque when
driven in rotation.
[0021] The method comprises the steps of engaging the elongate
drive structure about the first and second rotors, driving the
elongate drive structure by the first rotor, and driving the second
rotor by the elongate drive structure, and applying to the second
rotor by means of the non-circular profile an opposing fluctuating
corrective torque which reduces or substantially cancels the
fluctuating load torque of the rotary load assembly.
[0022] In accordance with yet another aspect of the invention,
there may be provided a method of constructing a synchronous drive
apparatus, comprising: [0023] (i) assembling components comprising
a continuous-loop elongate drive structure having a plurality of
engaging sections, a plurality of rotors comprising at least a
first and a second rotor, the first rotor having a plurality of
teeth for engaging the engaging sections of the elongate drive
structure, and the second rotor having a plurality of teeth for
engaging the engaging section of the elongate drive structure, and
a rotary load assembly coupled to the second rotor; and [0024] (ii)
engaging the elongate drive structure about the first and second
rotors, the first rotor being arranged to drive the elongate drive
structure and the second rotor being arranged to be driven by the
elongate drive structure, and one of the rotors having a
non-circular profile having at least two protruding portions
alternating with receding portions, the rotary load assembly being
such as to present a periodic fluctuating load torque when driven
in rotation; and [0025] (iii) determining the angular positions of
the protruding and receding portions of the non-circular profile
relative to the angular position of the second rotor, and the
magnitude of the eccentricity of the non-circular profile, to be
such that the non-circular profile applies to the second rotor an
opposing fluctuating corrective torque which reduces or
substantially cancels the fluctuating load torque of the rotary
load assembly.
[0026] In a preferred form of the method of constructing the
synchronous drive apparatus, the method includes: [0027] (i)
arranging the non-circular profile to produce the opposing
fluctuating corrective torque by periodic elongation and
contraction of the spans of the elongate drive structure adjoining
the rotor on which the non-circular profile is formed, the elongate
drive structure having a drive span between the rotor on which the
non-circular profile is formed and the second rotor, the drive span
being positioned on the tight side of the rotor on which the
non-circular profile is formed; and [0028] (ii) determining the
angular positions of the protruding and receding portions of the
non-circular profile by arranging the angular position of the
non-circular profile to be within +/-15 degrees of an angular
position for which a maximum elongation of the drive span coincides
with a peak value of the fluctuating load torque of the rotary load
assembly.
[0029] Also in a preferred form of the invention the method of
constructing a synchronous drive apparatus includes determining the
magnitude of the eccentricity of the non-circular profile is
determined by the following steps: [0030] (i) measuring the
amplitude of the fluctuating load torque of the rotary load
assembly at a predetermined selected set of operating conditions of
the synchronous drive apparatus; [0031] (ii) calculating the
required amplitude of periodic elongation and contraction of the
drive span by the following formula: L = T rk ##EQU1## [0032] L=the
amplitude of the periodic elongation and contraction of the said
drive span; [0033] T=the amplitude of the fluctuating load torque
of the rotary load assembly at a predetermined selected set of
operating conditions of the synchronous drive apparatus; [0034]
r=the radius of the second rotor: [0035] k=the stiffness
coefficient of the elongate drive structure defined as k = .times.
dF .times. d .times. .times. L ##EQU2## [0036] where dF is the
force required to produce an increase of length dL in the length of
the structure. [0037] (iii) producing and recording data to relate
empirically a series of values of (a) the divergence from circular
of the protruding and receding portions of the non-circular profile
and (b) the resulting amplitude of the periodic elongation and
contraction of the drive span; and [0038] (iv) selecting from the
data the corresponding eccentricity to give the required amplitude
of the periodic elongation and contraction of the drive span.
[0039] The present invention arises from an understanding that the
best way to eliminate or reduce torsional vibrations in a
synchronous drive system is to arrange a non-circular profile on
one of the rotors which is such as to cancel or reduce the
fluctuating load torque in the load assembly, rather than trying to
cancel or reduce the varying tension in the continuous loop drive
structure, as was attempted in the prior art. Indeed it is found
essential to provide a varying tension in the elongate drive
structure, in order to cancel or reduce the fluctuating load torque
in the load assembly. The present invention allows the
cancellation, or reduction, of the source of the torsional
excitation, rather than endeavouring to deal with the effects of
torsionals by cancelling variations in tension in the elongate
drive structure.
[0040] Thus although it has been known to provide a non-circular
profile on one of the rotors in a synchronous drive assembly, the
methods chosen to determine the magnitude of the eccentricity, and
the timing of the protruding and receding portions of the
non-circular profile, have not been such as to produce the required
result. By way of example, in a typical internal combustion engine,
if the eccentricity is chosen such as to try to equalise the
tension in a drive belt, the eccentricity will typically be
considerably too great to cancel the torsional vibrations in the
load assembly. In a typical international combustion engine, there
will be a resonant frequency at, say, 2000 to 2500 rpm. If the
eccentricity of the non-circular profile is chosen to attempt to
cancel any tension variation in the drive belt in the region of
resonance, then typically the eccentricity will be set at much more
tension than is required to cancel the vibrations. The result will
be excessive wear in the drive belt and the various sprockets, and
also the system will not be successful in reducing vibration.
[0041] Considering another manner in which the prior art
arrangements were deficient, it is important to arrange the timing
(translated into angular position) of the non-circular profile, to
be correctly related to the timing (translated into angular
positioning) of the fluctuations in load torque in the load
assembly. Conveniently the relative timing of the non-circular
profile and the fluctuating load torque of the rotary load assembly
is determined in relation to a periodic elongation and contraction
of a drive span of the elongate drive structure between the first
and second rotors on the tight side of the first rotor. The most
preferable arrangement in accordance with the invention is that the
angular position of the non-circular profile is that for which a
maximum elongation of the drive span of the elongate drive
structure substantially coincides with a peak value of the
fluctuating load torque of the rotary load assembly. However, the
invention can provide substantial reduction in vibration if the
timing is set within a range of plus/minus 15.degree. of the
preferred angular position. A particularly preferred range is
plus/minus 5.degree. of the preferred angular position.
[0042] In contrast, in the prior art it has been attempted to set
the eccentricity of the non-circular profile with reference to the
tension in the elongate drive structure. However in a typical
internal combustion engine the peak tension in the drive belt
varies in its timing according to the region of the rpm range which
is examined. Typically the peak tension in the drive belt occurs at
one timing stage for the resonant frequency of the engine, and
occurs at an earlier timing in the cycle for the rev range below
resonance, and occurs at a later part of the timing cycle for the
region of the rev range above the resonant condition. Thus,
depending upon which conditions are selected in the prior art in
order to attempt to equalise the tension in the drive belt, the
timing of the eccentricity of the non-circular profile may be ahead
of, or may lag behind, the preferred position for cancelling the
fluctuating load torque in the load assembly.
[0043] Thus to summarise, the present invention provides for the
correct selection of the eccentricity and the timing of the
non-circular profile, to be that which most advantageously cancels
or reduces the fluctuating load torque in the load assembly.
DESCRIPTION OF THE DRAWINGS
[0044] Embodiments of the invention will now be described by way of
example with reference to the accompanying drawings in which:
[0045] FIG. 1 is a schematic illustration of a synchronous drive
apparatus for a motor vehicle internal combustion engine, embodying
the invention;
[0046] FIG. 2 is an enlarged view of the crankshaft sprocket shown
in FIG. 1;
[0047] FIG. 3 is a schematic illustration of the synchronous drive
apparatus of an internal combustion engine in DOHC engine
configuration;
[0048] FIG. 4a shows a graph of a fluctuating load torque at the
camshaft of an SOHC internal combustion engine and a fluctuating
corrective torque generated by an oval crankshaft sprocket
illustrated in FIGS. 1 and 2, all graphs being taken over one
crankshaft revolution;
[0049] FIG. 4b shows a graph of a fluctuating load torque which
arises from the intake cam of an DOHC internal combustion engine, a
fluctuating load torque which arises from the exhaust cam, and a
fluctuating corrective torque generated by an oval crankshaft
sprocket in the engine illustrated in FIG. 3, all graphs being
taken over one crankshaft revolution;
[0050] FIGS. 5a to 5d show different combinations of crankshaft and
camshaft sprockets embodying the invention in 4-cylinder and
3-cylinder engines;
[0051] FIGS. 6a to 6d show different combinations of crankshaft and
camshaft sprockets embodying the invention in 6-cylinder,
8-cylinder and 2-cylinder engines;
[0052] FIG. 7a is a graph illustrating the magnitude of torsional
vibrations in an internal combustion engine at different engine
speeds, the vertical axis indicating the amplitude of torsional
vibrations in degrees of movement of the camshaft, and the
horizontal axis indicating engine speed in rpm, the graph
indicating the situation in a known engine, having a round
crankshaft sprocket;
[0053] FIG. 7b is a graph illustrating the magnitude of torsional
vibrations in an internal combustion engine at different engine
speeds, the vertical axis indicating the amplitude of torsional
vibrations in degrees of movement of the camshaft, and the
horizontal axis indicating engine speed in rpm, the graph
indicating the situation for a synchronous drive apparatus
embodying the invention, utilising an oval crankshaft sprocket;
[0054] FIG. 8a is a graph illustrating the magnitude of tensions in
an internal combustion engine at different engine speeds, the
vertical axis indicating the amplitude of the belt tension, and the
horizontal axis indicating engine speed in rpm, the graph
indicating the situation in a known engine, having a round
crankshaft sprocket;
[0055] FIG. 8b is a graph illustrating the magnitude of tensions in
an internal combustion engine at different engine speeds, the
vertical axis indicating the amplitude of the belt tension, and the
horizontal axis indicating engine speed in rpm, the graph
indicating the situation for a synchronous drive apparatus
embodying the invention, utilising an oval crankshaft sprocket;
[0056] FIGS. 9a and 9b show respectively the fluctuations in
tension in the drive belt over one revolution of the crankshaft at
1500 RPM, for an engine according to the prior art, having a round
crankshaft sprocket, FIGS. 9a and 9b showing respectively the belt
tension variations on the tight side and the slack slide of the
crankshaft sprocket respectively;
[0057] FIGS. 10a and 10b show respectively the fluctuations in
tension in the drive belt over one revolution of the crankshaft at
2500 RPM, for an engine according to the prior art, having a round
crankshaft sprocket, FIGS. 10a and 10b showing respectively the
belt tension variations on the tight side and the slack slide of
the crankshaft sprocket respectively;
[0058] FIG. 11 show respectively the fluctuations in tension in the
drive belt over one revolution of the crankshaft at 3500 RPM, for
an engine according to the prior art, having a round crankshaft
sprocket, FIGS. 11a and 11b showing respectively the belt tension
variations on the tight side and the slack slide of the crankshaft
sprocket respectively;
[0059] FIG. 12 is a three-dimensional graph showing the
distribution of camshaft torsional vibrations in a known internal
combustion engine having a round crankshaft sprocket, in which the
X-axis indicates various harmonic orders of vibration, the Y-axis
indicates engine speed in RPM, and the Z-axis indicates the
amplitude of the camshaft torsional vibrations;
[0060] FIG. 13 is a three-dimensional graph showing the
distribution of camshaft torsional vibrations in an engine
embodying the invention and having an oval crankshaft sprocket, in
which the X-axis indicates various harmonic orders of vibration,
the Y-axis indicates engine speed in RPM, and the Z-axis indicates
the amplitude of the camshaft torsional vibrations;
[0061] FIG. 14a shows a graph of fluctuating load torque on a
rotary load assembly such as a camshaft;
[0062] FIG. 14b shows how a non-circular profile 19 may be derived
to cancel the torque fluctuations of FIG. 14a, in an embodiment of
the invention; and
[0063] FIGS. 15, 16 and 17 show a computer generated virtual
representation of an oval crankshaft profile embodying the
invention, the profile being stepped on by an angular advance of
one tooth in FIG. 16 relative to FIG. 15, and in FIG. 17 relative
to FIG. 16.
DESCRIPTION OF THE INVENTION
[0064] FIG. 1 is a diagrammatic representation of a synchronous
drive apparatus for a motor vehicle internal combustion engine,
embodying the invention. The apparatus comprises a continuous loop
elongate drive structure 10, first and second rotors 11 and 12, and
further rotors 13, 14 and 17. The continuous loop elongate drive
structure 10 is provided by a conventional timing belt having teeth
15 together with intervening valleys which constitute a plurality
of engaging sections of the continuous loop elongate drive
structure. Each rotor 11 and 12 is provided by a sprocket having a
plurality of teeth 16 for engaging the valleys between the teeth 15
of the timing belt 10. The sprocket 11 is coupled to the crankshaft
(not shown) of an internal combustion engine, and the sprocket 12
is coupled to a rotary load assembly (not shown) which is
constituted by a camshaft 26 of the internal combustion engine. The
timing belt 10 is engaged about the first and second rotors 11 and
12, the first rotor 11 being arranged to drive the belt 10 and the
second rotor 12 being arranged to be driven by the belt 10. The
rotor 14 also has teeth 16 and consists of a sprocket for driving
other elements of the internal combustion engine, such as a water
pump, and the rotor 13 is preferably for a belt tensioner bearing
on a non-toothed side of the timing belt 10, to tension the belt in
known manner. Rotor 17 is preferably for a fixed idler pulley
bearing on the non-toothed side of timing belt 10.
[0065] In a known form of a synchronous drive apparatus, the
crankshaft sprocket would have a circular profile. In such a case,
the synchronous drive apparatus is prone to vibrations, known as
torsional vibrations, which arise from the opening and closing of
the intake and exhaust valves of the internal combustion engine by
the overhead camshaft. The source of the excitations is illustrated
in FIGS. 4a and b. FIG. 4a illustrates the fluctuating load torque
103 applied to the camshaft in a SOHC engine and FIG. 4b
illustrates the same for a DOHC engine. FIG. 4b shows the variation
of camshaft torque over a single cycle of the engine, indicating
how the intake torque shown by the curve 101 varies with degrees of
rotation of the engine, and how the exhaust torque profile 102
varies in the same way.
[0066] In accordance with the embodiment of the present invention
shown in FIG. 1 for a SOHC engine, the crankshaft sprocket 11 has a
non-circular profile (as shown in exaggerated form in FIG. 2)
indicated generally by reference numeral 19. The non-circular
profile 19 is, in the particular embodiment described, an oval
having a major axis 20 and a minor axis 21. The profile 19 has two
protruding portions 22 and 23 and has two receding portions 24 and
25.
[0067] The provision of the oval profile 19 on the sprocket 11 as
shown in FIG. 2, generates a fluctuating corrective torque, which
is applied by the belt 10 to the second rotor 12. This fluctuating
corrective torque is shown at 104 in FIG. 4a. In the preferred
situation, the total fluctuating load torque 103 is opposed by the
overall corrective torque 104. Preferably the corrective torque 104
is 180.degree. out of phase with the overall load torque 103, and
the peak to peak amplitude of the fluctuating corrective torque 104
is made equal to the peak to peak amplitude of the overall
fluctuating load torque 103.
[0068] In accordance with the embodiment of the invention using the
oval profile 19 shown in FIG. 2, the angular positions of the
protruding and receding portions 22 to 24 of the non-circular
profile 19 relative to the angular position of the second rotor 12,
and the magnitude of the eccentricity of the non-circular profile
19, are such that the non-circular profile 19 applies to the second
rotor 12 an opposing fluctuating corrective torque 104 which
substantially cancels the fluctuating load torque 103 of the rotary
load assembly 26.
[0069] The determination of the timing and magnitude of the
eccentricity of the non-circular profile 19 will now be described
in more detail. In FIG. 1 the spans between the various rotors are
indicated as 10A between rotor 12 and rotor 14, 10B between rotor
14 and rotor 11, 10C between rotor 12 and rotor 13, and 10D between
rotor 13 and rotor 17 and 10E between rotor 17 and rotor 11. The
span between the first rotor 11 and the second rotor 12, indicated
as 10A, 10B, is referred to as the drive span between the two
rotors, it being positioned on the tight side of the first rotor 11
on which the non-circular profile 19 is formed. The span between
the first rotor 11 and second rotor 12 which is indicated as 10C,
10D, 10E is referred as the slack side, although of course the belt
is under tension on both sides. The torsional vibrations to be
eliminated are formed by the fluctuating load torque on the rotary
load assembly (the camshaft 26) and in accordance with the present
invention this is reduced or substantially cancelled by the
application of an opposing fluctuating corrective torque to the
camshaft 26 by means of the timing belt 10. The opposing
fluctuating corrective torque is produced by the non-circular
profile 19 by periodic elongation and contraction of the spans 10A
10B and 10C 10D 10E, adjoining the rotor 11 on which the
non-circular profile is formed. In preferred forms of the
invention, the angular position of the non-circular profile 19 is
set as closely as possible to be that for which a maximum
elongation of the drive span 10A 10B substantially coincides with a
peak value of the fluctuating load torque of the camshaft 26. It
may not always be possible to arrange this exactly, and advantage
is obtained in accordance with the invention if the angular
position of the non-circular profile is within +/-15 degrees of the
preferred angular position, more preferably within +/-5
degrees.
[0070] With regard to the particular case illustrated, and
referring to FIGS. 1 and 2, the oval profile 19 has two reference
radii 20a and 20b, which together form the major axis 20 of the
oval. Each reference radius 20a, 20b passes from the centre of the
rotor 11 and through the centre of the respective protruding
portion 22, 23. The angular position of the non-circular profile 19
is related to a reference direction of the rotor 11, the reference
direction being the direction of a vector or imaginary line 27 that
bisects the angle or sector of wrap of the continuous loop drive
structure 10 around the rotor 11. This vector that bisects the
angle of wrap is in the same direction as the hub load force
produced by engagement of the belt 10 with the rotor 11 when the
belt drive system is static. It should be appreciated, however,
that the hub load force direction changes dynamically during
operation of the belt drive system. The timing of the non-circular
profile 19 is set to be such that, at the time when the fluctuating
load torque on the second rotor 12 is at a maximum, the angular
position of the reference radius 20a is within a range of
90.degree. to 180.degree. from the reference direction of the angle
of wrap bisection 27, taken in the direction of rotation of the
rotor 11, preferably within a range of 130.degree. to 140.degree..
Assuming that the assembly of FIG. 1 is shown at the instant when
the fluctuating load torque on the second rotor 12 is at a maximum,
then the preferred timing of the non-circular profile 19 is as
shown in FIG. 1, namely that the angle between the reference radius
20a and the bisection direction 27 is 135.degree., as indicated by
the angle .theta..
[0071] It is to be appreciated that in this specification, where
the term "reference radius" is used for a non-circular profile 19,
the reference parameter measured is the radius of a notional circle
passing through the associated protruding portion, and is not a
radius of the entire profile, since this entire profile is
essentially non-circular. The term reference radius is used merely
to indicate the distance between the centre of the axis of the
rotor on which the profile is formed, to the maximum extent of the
profile at the relevant protruding portion.
[0072] Consideration will now be given to the determination of the
magnitude of the eccentricity of the non-circular profile 19 in the
specific embodiment shown. In summary, the magnitude of the
eccentricity of the profile 19 is preferably set to be such that
the fluctuating corrective torque 104 shown in 4a has an amplitude
substantially equal to, and phase substantially opposite to, the
amplitude of the fluctuating load torque 103 shown in FIG. 4a.
However advantage is still found in embodiments where the amplitude
of the fluctuating corrective torque 104 is in the range of 75% to
110% of the amplitude of the fluctuating load torque 103, more
preferably in the range 90% to 100%. Where the fluctuating load
torque 103 has a substantially constant amplitude over the rev
range of the engine, the amplitude of the corrective torque 104 is
merely made substantially equal to the constant amplitude of the
fluctuating load torque.
[0073] The practical steps of determining the magnitude of the
eccentricity may be as follows. First the amplitude of the
fluctuating load torque 103 of the camshaft 26 is measured at the
selected set of operating conditions, in this case at the maximum
amplitude of the fluctuating load torque. Next there is calculated
the required amplitude of period elongation and contraction of the
drive span 10a, 10b by the following formula: L = T rk ##EQU3##
where: [0074] L=the amplitude of the periodic elongation and
contraction of the drive span which is required; [0075] T=the
amplitude of the fluctuating load torque of the camshaft 26, which
has been measured at maximum amplitude; [0076] r=the radius of the
second rotor 12: and [0077] k=the stiffness coefficient of the belt
10.
[0078] The stiffness coefficient k is obtained from the formula k =
dF ##EQU4## d .times. .times. L ##EQU4.2## where dF is the force
required to produce an increase of length dL in the of the
structure.
[0079] By way of example of the calculations above, the amplitude
of the fluctuating load torque T may be 10 Nm (zero to peak), and
the radius of the rotor 12 may be 50 mm. This gives a maximum force
F required to provide the required fluctuating corrective torque of
F=200N. In the example discussed, the required change in span
length is obtained by dividing the tension of 200N by the stiffness
coefficient k, which for example for a typical belt may be 400
N/mm. This gives required amplitude of elongation and contraction
of the timing belt of 0.5 mm (zero to peak).
[0080] The next step is to calculate the eccentricity required to
provide this length of elongation and contraction at a timing stage
when the major axis 20 of the ellipse is set at .theta.=135.degree.
as shown in FIG. 1. A theoretical calculation of this value is
difficult to achieve, so that the calculation of the eccentricity
is arrived at by the equivalent of a "look-up" table. This is done
by producing and recording data to relate empirically a series of
values of (i) the divergence from circular of the protruding and
receding portions of the non-circular profile and (ii) the
resulting amplitude of the periodic elongation and contraction of
the drive span. The required eccentricity is then selected from the
data to give the required amplitude of the periodic elongation and
contraction of the drive span.
[0081] The data bank which is produced, to provide the "look-up"
table consists of a table of values of the amplitude of elongation
and contraction of the drive span 10A and 10B, for various values
of the eccentricity of the oval profile 19 along the major axis.
Examples of such data are given in the following table, Table 1.
The reference circle used for comparison is a circle having a
diameter equal to the average of the major axis length 20 and the
minor axis length 21. The eccentricity of the oval profile 19 can
be determined, in the example shown, by considering the divergence
of the outline from the reference circle at the major axis 20.
TABLE-US-00001 Difference between Amplitude of periodic selected
oval reference elongation and contraction outline and reference
circle of drive span 10A, 10B 0.5 mm 0.25 mm 1.0 mm 0.49 mm 1.5 mm
0.74 mm
[0082] This table may be derived for example by producing a
computer simulation of the oval profile 19, and stepping this
through a series of angular advancements of, say one tooth at a
time, for example as shown in FIGS. 15, 16 and 17. For each of
these steps, the computer simulation is arranged to provide an
indication of the elongation or contraction of the equivalent drive
span 10A, 10B, for a particular length of major axis giving the
radius 20A. On the computer simulation, the reference radius 20A is
then varied, and a further series of data are produced for the new
radius 20A. The purpose of stepping the profile through the
positions shown at FIGS. 15, 16 and 17, is to determine empirically
the position at which the maximum extension of the corresponding
drive span 10A, 10B takes place. Having determined that, the
appropriate data is extracted, for the maximum length of the span
10A, 10B, which is set against the corresponding eccentricity of
the reference radius 20A. FIGS. 15, 16 and 17 show how the
amplitude of elongation may be determined by using virtual
prototyping.
[0083] FIGS. 5a to 5d show different combinations of crankshaft and
camshaft sprockets for 4-cylinder and 3-cylinder engines. FIGS. 6a
to 6d show different combinations of crankshaft and camshaft
sprockets for 6-cylinder, 8-cylinder and 2-cylinder engines.
[0084] FIG. 7a shows the amplitude of camshaft torsional vibrations
in degrees of rotary vibration versus the engine speed in rpm for a
round crankshaft sprocket. FIG. 7b shows the amplitude of camshaft
torsional vibrations in degrees of rotary vibration versus the
engine speed in rpm for an oval crankshaft sprocket. FIG. 7b shows
that the torsionals are significantly reduced. Only torsionals
coming from the crankshaft remain. The resonance has been
cancelled.
[0085] FIG. 8a shows the tight side tension fluctuation versus the
engine speed in rpm for a round crankshaft sprocket. FIG. 8b shows
the tight side tension fluctuation versus the engine speed in rpm
for an oval crankshaft sprocket. FIG. 8b also shows that resonance
has been cancelled. Tension fluctuations are still present in the
whole rpm range, but they need to be there to provide cancelling
torque.
[0086] FIGS. 9a and b show the tight side and slack side tension
fluctuations over one revolution of the round crankshaft sprocket
at 1500 rpm. FIGS. 10a and b show the tight side and slack side
tension fluctuations over one revolution of the round crankshaft
sprocket at the system resonance (2500 rpm). FIGS. 11a and b show
the tight side and slack side tension fluctuations over one
revolution of the round crankshaft sprocket at 3500 rpm.
[0087] FIG. 12 shows the camshaft torsional vibrations for a round
crankshaft sprocket presented as a spectral analysis where:
x-axis=harmonics orders; y-axis=engine rpm; and z-axis=amplitude of
the camshaft torsional vibrations.
[0088] FIG. 13 shows the camshaft torsional vibrations for an oval
crankshaft sprocket presented as a spectral analysis where:
x-axis=harmonics orders; y-axis=engine rpm; and z-axis=amplitude of
the camshaft torsional vibrations. Only second order torsionals are
eliminated by the oval profile. Using a more complex profile, as
shown in FIG. 14 will cancel simultaneously second and fourth order
torsionals.
[0089] FIGS. 14a and 14b show, in greatly exaggerated form, how a
non-circular profile 19 of one of the rotors in a synchronous drive
apparatus embodying the invention can be shaped to accommodate two
different orders of torsional fluctuations in the torque of a
rotary load assembly. FIG. 14 consists of two FIGS. 14a and 14b.
FIG. 14a shows in curve 110 a second order fluctuating load torque,
equivalent to the second order peak shown in FIG. 12. The curve 111
shows a fourth order fluctuating load torque equivalent to the
fourth order peak shown in FIG. 12. Curve 112 shows the combined
fluctuating load torque on the rotary load assembly.
[0090] In FIG. 14b there is shown at 19A in greatly exaggerated
form a generally oval profile suitable for use on a crankshaft
rotor 11 in FIG. 1, having protruding portions 22 and 23. These
protruding portions produce a corrective fluctuating load torque
which can be applied to cancel the second order fluctuating load
torque 110 in FIG. 14a. A second profile indicated at 19B is shaped
to have four minor protruding portions which, if it were to be used
as a profile of crankshaft sprocket 11, would produce a corrective
torque equivalent to the fourth order fluctuating load torque 111
in FIG. 14a. In FIG. 14b, a non-circular profile embodying the
invention is indicated at 19C, which is a combination of the two
profiles 19A and 19B. The combined profile 19C has two major
protruding portions, and two minor protruding portions. The
combined profile 19C produces a fluctuating corrective torque which
can be made to cancel the combined fluctuating torque 112 shown in
FIG. 14a.
[0091] Thus in FIG. 14, there is shown a modification of the oval
rotor in which additional minor protruding portions of the profile
are provided. The reason for this is to take account of fourth
order harmonic torsional vibrations which are illustrated in FIGS.
12 and 13. In FIG. 12, there is shown the torsional vibrations
which arise from the second, fourth and sixth order harmonics, with
a synchronous drive apparatus having a circular crankshaft
sprocket. FIG. 13 shows the torsional vibrations remaining after
use of an oval crankshaft drive sprocket in accordance with the
invention. It will be seen that the fourth order harmonic torsional
vibrations remain. These vibrations can be reduced or eliminated by
providing on the non-circular profile of the crankshaft sprocket
additional protruding portions. The minor protruding portions are
of lesser extent than the major protruding portions, and are
arranged to produce lesser fluctuating corrective torque patterns
in the torque applied to the second rotor, to reduce or
substantially cancel the fourth order fluctuating load torque
presented by the rotary load assembly.
[0092] Returning now to a general consideration of the operation of
embodiments of the invention, it is known to provide in a
synchronous drive system for an internal combustion engine a
crankshaft sprocket of oval profile. The present invention provides
for the correct selection of the eccentricity and the timing of the
non-circular profile, to be that which advantageously cancels or
reduces the fluctuating load torque in the load assembly, rather
than endeavouring to equalise the tension in the drive belt, has as
been done in the prior art arrangements.
[0093] The invention can be understood by considering Newton's
second law, that the presence of an unbalanced force will
accelerate an object. For linear examples this provides:--
Acceleration=Force/Mass In rotary motion:
Acceleration=Torque/Inertia In an ordinary internal combustion
engine the torque from the valve train or diesel fuel pump
fluctuates, causing the speed to fluctuate, causing angular
displacement to fluctuate (also known as torsional vibration). By
using an ellipsoidal crankshaft sprocket that is pulling the belt
(at appropriate instant) additional torque can be created that has
such amplitude and phase that the combined torque acting on the
camshaft is zero. Absence of torque means absence of acceleration
by first Newton's law. Absence of acceleration means absence of
speed fluctuations, which means that no torsionals are present.
[0094] The opening and closing of the intake and exhaust valves is
a source of torque fluctuations. These torque fluctuations cause
the camshaft to be inflicted with speed fluctuations, which in
turn, causes angular position fluctuations otherwise know as
torsional vibrations. The best cure for that behaviour is to attack
the cause right at the source by introducing another torque acting
on the camshaft i.e. removing torque fluctuations at the camshaft.
One way of doing it is to use the oval sprocket at the crankshaft.
The oval sprocket, while rotating, will introduce fluctuations of
span length i.e. will pull and relieve two times per one crankshaft
revolution. When the tight side is being pulled, the slack side is
relieved and vice versa. Pulling and relieving the belt means that
a new, additional torque is generated at the camshaft. If this new
torque is of appropriate amplitude and phase it can balance the
first torque from the valve train. Absence of torque fluctuations
means absence of speed fluctuations and therefore absence of
torsionals.
[0095] In embodiments of the invention, when the torsional
vibrations in the camshaft are eliminated the belt tension still
varies. Indeed it is the variation in tension in the belt, which
causes the torsional vibrations in the camshaft to cease. In the
prior art, the objective is said to be the removal of tension
variation in the belt, which is not what is needed to remove
torsional vibration in the camshaft. The object is to remove the
variation in speed of the driven sprocket, which is caused by
variation in torque load in the driven sprocket. This is done by
varying the tension in the belt during the cycle of the driven
sprocket. At a time of increase of torque load on the driven
sprocket, there must be an increase in tension in the belt. At
moment when increase in tension is required the effective length of
the span must be increased. This is achieved by having the oval
positioned so that the long axis is moving from a position
perpendicular to the hub load, to position along the hub load
direction. At the moment when decrease in tension is required the
effective length of the span must be decreased. This is done while
the major axis moves from vertical to horizontal.
* * * * *