Endless belt power transmission systems or components – Positive drive pulley or guide roll – Having nonmetallic component
Reexamination Certificate
2000-12-01
2002-04-16
Bucci, David A. (Department: 3682)
Endless belt power transmission systems or components
Positive drive pulley or guide roll
Having nonmetallic component
C474S160000, C474S156000
Reexamination Certificate
active
06371875
ABSTRACT:
BACKGROUND OF THE INVENTION
The present invention relates to the automotive timing chain art. It finds particular application in conjunction with a roller chain sprocket for use in automotive camshaft drive applications and will be described with particular reference thereto. However, the present invention may also find application in conjunction with other types of chain drive systems and applications where reducing the noise levels associated with chain drives is desired.
Roller chain sprockets for use in camshaft drives of automotive engines are typically manufactured according to one or more international standards such as DIN, JIS, ISO, etc. The ISO-606:1994(E) (International Organization for Standardization) standard specifies requirements for short-pitch precision roller chains and associated chain wheels or sprockets.
FIG. 1
illustrates a symmetrical tooth space form for an ISO-606 compliant sprocket. The tooth space has a continuous fillet or root radius R
i
extending from one tooth flank (i.e., side) to the adjacent tooth flank as defined by the roller seating angle &agr;. The flank radius R
f
is tangent to the roller seating radius R
i
at the tangency point TP. A chain with a link pitch P has rollers of diameter D
1
in contact with the tooth spaces. The ISO sprocket has a chordal pitch also of length P, a root diameter D
2
, and Z number of teeth. The pitch circle diameter PD, tip or outside diameter OD, and tooth angle A (equal to 360°/Z) further define the ISO-606 compliant sprocket. The maximum and minimum roller seating angle &agr; is defined as:
&agr;
max
=(140°−90°)/Z and &agr;
min
=(120°−90°)/Z
With reference to
FIG. 2
, an exemplary ISO-606 compliant roller chain drive system
10
rotates in a clockwise direction as shown by arrow
11
. The chain drive system
10
includes a drive sprocket
12
, a driven sprocket
14
and a roller chain
16
having a number of rollers
18
. The sprockets
12
,
14
, and chain
16
each generally comply with the ISO-606 standard.
The roller chain
16
engages and wraps about sprockets
12
and
14
and has two spans extending between the sprockets, slack strand
20
and taut strand
22
. The roller chain
16
is under tension as shown by arrows
24
. A central portion of the taut strand
22
may be guided between the driven sprocket
14
and the drive sprocket
12
with a conventional chain guide (not shown). A first roller
28
is shown at the onset of meshing at a 12 o'clock position on the drive sprocket
12
. A second roller
30
is adjacent to the first roller
28
and is the next roller to mesh with the drive sprocket
12
.
Chain drive systems have several components of undesirable noise. A major source of roller chain drive noise is the sound generated as a roller leaves the span and collides with the sprocket during meshing. The resultant impact noise is repeated with a frequency generally equal to that of the frequency of the chain meshing with the sprocket. The loudness of the impact noise is a function of the impact energy (E
A
) occurring during the meshing process. The impact energy (E
A
) is related to engine speed, chain mass, and the impact velocity between the chain and the sprocket at the onset of meshing. The impact velocity is affected by the chain-sprocket engagement geometry, of which an engaging flank pressure angle &ggr; (
FIG. 3
) is a factor, where:
E
A
=
wP
2000
V
A
2
;
V
A
=
π
nP
30000
sin
(
360
Z
+
γ
)
;
γ
=
180
-
A
-
α
2
;
and
E
A
=
Impact
Energy
[
N
·
m
]
V
A
=
Roller
Impact
Velocity
[
m
/
s
]
γ
=
Engaging
Flank
Pressure
Angle
n
=
Engine
Speed
[
RPM
]
w
=
Chain
Mass
[
Kg
/
m
]
Z
=
Number
of
Sprocket
Teeth
A
=
Tooth
Angle
(
360
°
/
Z
)
α
=
Roller
Seating
Angle
P
=
Chain
Pitch
(
Chordal
Pitch
)
The impact energy (E
A
) equation presumes the chain drive kinematics will conform generally to a quasi-static analytical model and that the roller-sprocket driving contact will occur at a tangent point TP (
FIG. 3
) of the flank and root radii as the sprocket collects a roller from the span.
As shown in
FIG. 3
, the pressure angle &ggr; is defined as the angle between a line A extending from the center of the engaging roller
28
, when it is contacting the engaging tooth flank at the tangency point TP, through the center of the flank radius R
f
, and a line B connecting the centers of the fully seated roller
28
, when it is seated on root diameter D
2
, and the center of the next meshing roller
30
, as if it were also seated on root diameter D
2
in its engaging tooth space. The roller seating angles &agr; and pressure angles &ggr; listed in
FIG. 4
are calculated from the equations defined above. It should be appreciated that &ggr; is a minimum when a is a maximum. Thus, the exemplary 23-tooth, ISO-606 compliant, drive sprocket
12
shown in
FIGS. 2 and 3
will have a pressure angle &ggr; in the range of 14.13 to 24.13° as listed in the table of FIG.
4
.
FIG. 3
also shows the engagement path (phantom rollers) and the driving contact position of roller
28
(solid) as the drive sprocket
12
rotates in the direction of arrow
11
.
FIG. 3
depicts the theoretical case with chain roller
28
seated on root diameter D
2
of a maximum material sprocket with both chain pitch and sprocket chordal pitch equal to theoretical pitch P. For this theoretical case, the noise occurring at the onset of roller engagement has a radial component F
R
as a result of roller
28
colliding with the root surface R
i
and a tangential component F
T
generated as the same roller
28
collides with the engaging tooth flank at point TP as the roller moves into driving contact. It is believed that the radial impact occurs first, with the tangential impact following nearly simultaneously. Roller impact velocity V
A
is shown to act through, and is substantially normal to, engaging flank tangent point TP with roller
28
in driving contact at point TP.
The impact energy (E
A
) equation accounts only for a tangential roller impact during meshing. The actual roller engagement, presumed to have a tangential and radial impact (occurring in any order), would therefore seem to be at variance with the impact energy (E
A
) equation. The application of this quasi-static model, which is beneficially used as a directional tool, permits an analysis of those features that may be modified to reduce the impact energy occurring during the tangential roller-sprocket collision at the onset of meshing. The radial collision during meshing, and its effect on noise levels, can be evaluated apart from the impact energy (E
A
) equation.
Under actual conditions as a result of feature dimensional tolerances, there will normally be a pitch mismatch between the chain and sprocket, with increased mismatch as the components wear in use. This pitch mismatch serves to move the point of meshing impact, with the radial collision still occurring at the root surface R
i
but not necessarily at D
2
. The tangential collision will normally be in the proximity of point TP, but this contact could take place high up on the engaging side of root radius R
i
or even radially outward from point TP on the engaging flank radius R
f
as a function of the actual chain-sprocket pitch mismatch.
Reducing the engaging flank pressure angle &ggr; reduces the meshing noise levels associated with roller chain drives, as predicted by the impact energy (E
A
) equation set forth above. It is feasible but not recommended to reduce the pressure angle &ggr; while maintaining a symmetrical tooth profile, which could be accomplished by simply increasing the roller seating angl
Bucci David A.
Charles Marcus
Cloyes Gear & Products, Inc.
Fay Sharpe Fagan Minnich & McKee LLP
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