U.S. patent application number 15/334362 was filed with the patent office on 2017-06-29 for oil ring.
The applicant listed for this patent is HONDA MOTOR CO., LTD.. Invention is credited to Naokazu Kawase, Hajime Nakagawa.
Application Number | 20170184198 15/334362 |
Document ID | / |
Family ID | 59087718 |
Filed Date | 2017-06-29 |
United States Patent
Application |
20170184198 |
Kind Code |
A1 |
Kawase; Naokazu ; et
al. |
June 29, 2017 |
OIL RING
Abstract
Provided is an oil ring (13) that can minimize the frictional
resistance without impairing the oil control function such as the
scraping of lubricating oil from the cylinder wall surface by the
oil ring. The oil ring consists of an upper side rail (16), a lower
side rail (17) and an annular expander (18) interposed between the
two side rails, and is received in an oil ring groove (8) with a
prescribed vertical clearance (20) that permits tilting of the side
rails. An outer peripheral surface (16E, 17E) of each side rail is
provided with a radially inward slant. Therefore, during the upward
stroke of the piston, the upper side rail tilts in the radially
outward direction so that the effective slant angle of the outer
peripheral surface of the upper side ring is adjusted to an optimum
value.
Inventors: |
Kawase; Naokazu; (Wako-shi,
JP) ; Nakagawa; Hajime; (Wako-shi, JP) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
HONDA MOTOR CO., LTD. |
Tokyo |
|
JP |
|
|
Family ID: |
59087718 |
Appl. No.: |
15/334362 |
Filed: |
October 26, 2016 |
Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F16J 9/206 20130101;
F16J 9/066 20130101; F16J 9/203 20130101 |
International
Class: |
F16J 9/20 20060101
F16J009/20 |
Foreign Application Data
Date |
Code |
Application Number |
Dec 25, 2015 |
JP |
2015-253302 |
Claims
1. An oil ring configured to be received in an oil ring groove
formed in an outer circumferential surface of a piston, comprising:
an upper side rail and a lower side rail each consisting of an
annular plate member; and an annular expander interposed between
the upper side rail and the lower side rail such that the upper
side rail, the lower side rail and the expander are disposed in a
coaxial relationship; wherein an outer peripheral surface of the
upper side rail includes a slanted surface with a radially inward
slant slanting by a first slant angle with respect to an axial line
of the piston; wherein an outer peripheral surface of the lower
side rail includes a slanted surface with a radially inward slant
slanting by a second slant angle smaller than the first slant angle
with respect to the axial line; wherein a clearance is defined
between an upper wall surface of the oil ring groove and an upper
end surface of the upper side rail such that the upper side rail
can tilt in a radially outward direction by a prescribed tilt angle
during an upward stroke of the piston; wherein the first slant
angle is an angle of the slanted surface of the outer peripheral
surface of the upper side rail with respect to the axial line of
the piston when the upper side rail is not tilting; and wherein the
first slant angle is greater than the tilt angle.
2. (canceled)
3. The oil ring according to claim 1, wherein the first slant angle
is greater than the tilt angle by 0.5 degrees to 4.5 degrees.
4. The oil ring according to claim 3, wherein the first slant angle
is greater than the tilt angle by 1.0 degree to 3.0 degrees.
5. The oil ring according to claim 3, wherein the first slant angle
is between 2.5 degrees and 10.5 degrees.
6. The oil ring according to claim 5, wherein the first slant angle
is between 4.0 degrees and 7.5 degrees.
7. The oil ring according to claim 3, wherein the second slant
angle is between 0.5 degrees and 4.5 degrees.
8. The oil ring according to claim 7, wherein the second slant
angle is between 1.0 degree and 3.0 degrees.
9. The oil ring according to claim 3, wherein the tilt angle is
between 2.0 degrees and 6.0 degrees.
10. The oil ring according to claim 9, wherein the tilt angle is
between 2.0 degrees and 4.5 degrees.
11. The oil ring according to claim 1, wherein the outer peripheral
surface of the upper side rail is connected to a lower end surface
thereof via a smooth curve, and the outer peripheral surface of the
lower side rail is connected to a lower end surface thereof via a
smooth curve.
12. The oil ring according to claim 1, wherein the outer peripheral
surface of the upper side rail is provided with a barrel shape
having a vertically intermediate part thereof bulging radially
outward, and the outer peripheral surface of the lower side rail is
provided with a barrel shape having a vertically intermediate part
thereof bulging radially outward.
Description
TECHNICAL FIELD
[0001] The present invention relates to an oil ring for a piston of
a reciprocating machine such as an internal combustion engine, and
in particular to a composite oil ring including a pair of side
rails and an expander interposed between the two side rails.
BACKGROUND ART
[0002] The oil ring used in the pistons of conventional
reciprocating internal combustions typically consists of a three
piece oil ring including a pair of annular side rails positioned
one above the other and a space expander interposed between the two
side rails. See JP2003-194222A, for instance.
[0003] In recent years, there has been an increasing desire to
reduce the frictional resistance between the oil ring and the
associated cylinder wall surface in view of improving fuel economy.
The frictional resistance may be reduced by decreasing the pressure
exerted on the cylinder wall surface by the oil ring, but it may
impair the oil control function of the oil ring such as the
scraping of the lubricating oil off the cylinder wall surface.
Therefore, the pressure of the oil ring on the cylinder wall
surface cannot be reduced beyond a certain limit.
BRIEF SUMMARY OF THE INVENTION
[0004] In view of such problems of the prior art, a primary object
of the present invention is to provide an oil ring that can
minimize the frictional resistance without impairing the oil
control function such as the scraping of lubricating oil from the
cylinder wall surface by the oil ring.
[0005] To accomplish such objects, the present invention provides
an oil ring (13) configured to be received in an oil ring groove
(8) formed in an outer circumferential surface of a piston (4),
comprising an upper side rail (16) and a lower side rail (17) each
consisting of an annular plate member; and an annular expander (18)
interposed between the upper side rail and the lower side rail such
that the upper side rail, the lower side rail and the expander are
disposed in a coaxial relationship; wherein an outer peripheral
surface (16E) of the upper side rail includes a slanted surface
with a radially inward slant (such that an upper part thereof
recedes more radially inward than a lower part thereof) slanting by
a first slant angle (.theta.1) with respect to an axial line of the
piston; and wherein an outer peripheral surface of the lower side
rail includes a slanted surface with a radially inward slant (such
that an upper part thereof recedes more radially inward than a
lower part thereof), slanting by a second slant angle (.theta.2)
smaller than the first slant angle with respect to the axial
line.
[0006] In this arrangement, because the outer peripheral surface of
the upper side rail is slanted by the first slant angle .theta.1
and the outer peripheral surface of the lower side rail is slanted
by the second slant angle .theta.2, during the upward stroke of the
piston, each side rail is subjected to a lift that moves the side
rail away from the cylinder wall surface owing to a relatively
thick buildup of lubricating oil between the outer peripheral
surface of the side rail and the cylinder wall surface. Thereby,
the lubricating oil is not excessively scraped off the cylinder
wall by the oil ring, and hence the shear resistance of the
lubricating oil can be reduced. Meanwhile, during the downward
stroke of the piston, because very little wedge effect is produced
between the outer peripheral surfaces of the two side rails and the
cylinder wall surface, the lubricating oil is scraped off the
cylinder wall surface in an effective manner and the thickness of
the oil film is appropriately controlled. During the upward stroke
of the piston, owing to the presence of a clearance between the oil
ring and the oil ring groove and the compressive deformation of the
expander, the upper side rail tilts radially outward (the inner
peripheral part thereof is raised in comparison with the outer
peripheral part thereof). Therefore, during the upward stroke of
the piston, the effective slant angle of the outer peripheral
surface is reduced from the first slant angle by this tilting
action. However, according to the present invention, because the
first slant angle is originally greater than the second slant
angle, this tilting action causes the actual slant angle of the
outer peripheral surface of the upper side rail to be reduced from
the first slant angle to a value closer to the second slant angle.
Therefore, both the upper and lower side rails are allowed to
engage the cylinder wall surface with an optimum slant angle which
may be approximately equal to the second slant angle.
[0007] According to a preferred embodiment of the present
invention, a clearance (20) is defined between an upper wall
surface of the oil ring groove and the upper end surface of the
upper side rail such that the upper side rail can tilt in a
radially outward direction by a prescribed tilt angle during an
upward stroke of the piston.
[0008] Thus, the tilt angle that is caused during the upward stroke
of the piston can be accurately controlled by determining the size
of the clearance and the resiliency of the expander.
[0009] The first slant angle is greater than the tilt angle
preferably by 0.5 degrees to 4.5 degrees, and more preferably by
1.0 degree to 3.0 degrees. The first slant angle is between 2.5
degrees and 10.5 degrees, and more preferably between 4.0 degrees
and 7.5 degrees. The second slant angle is preferably between 0.5
degrees and 4.5 degrees, and more preferably between 1.0 degree and
3.0 degrees. The tilt angle is preferably between 2.0 degrees and
6.0 degrees.
[0010] Preferably, the outer peripheral surface of the upper side
rail is connected to a lower end surface thereof via a smooth curve
(16G), and the outer peripheral surface of the lower side rail is
connected to a lower end surface thereof via a smooth curve
(17G).
[0011] In a particularly preferred embodiment of the present
invention, the outer peripheral surface of the upper side rail is
provided with a barrel shape having a vertically intermediate part
thereof bulging radially outward, and the outer peripheral surface
of the lower side rail is provided with a barrel shape having a
vertically intermediate part thereof bulging radially outward.
[0012] An oil ring configured as discussed above can minimize
frictional resistance without impairing the function to control the
oil film thickness.
BRIEF DESCRIPTION OF THE DRAWINGS
[0013] FIG. 1 is a sectional view of a piston fitted with an oil
ring embodying the present invention;
[0014] FIG. 2 is a cross sectional view of the oil ring;
[0015] FIG. 3a is an enlarged cross sectional view of an upper side
rail of the oil ring with different lateral and vertical
magnification factors (the lateral dimension being five times more
enlarged than the vertical dimension);
[0016] FIG. 3b is an enlarged cross section view of a lower side
rail of the oil ring with different lateral and vertical
magnification factors (the lateral dimension being five times more
enlarged than the vertical dimension);
[0017] FIG. 4 is a view similar to FIG. 2 showing the state of the
oil ring when the piston is moving upward;
[0018] FIG. 5 is a view similar to FIG. 2 showing the state of the
oil ring when the piston is moving downward;
[0019] FIG. 6a is a diagram illustrating a planar pad moving over a
planar surface at a speed U;
[0020] FIG. 6b is a graph showing the relationship between the load
bearing capacity and a parameter m representing the angle of the
planar pad relative to the planar surface;
[0021] FIG. 7 is a graph showing the relationship between the
friction force and the angle between the outer peripheral surface
of the lower rail and the cylinder wall surface for two different
oil film thicknesses; and
[0022] FIG. 8 is a graph comparing the lubricating consumptions of
the present invention and the prior art.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT(S)
[0023] An oil ring embodying the present invention is described in
the following with reference to FIG. 1.
[0024] As shown in FIG. 1, a cylinder block 2 of an internal
combustion engine 1 is provided with a cylinder 3 having a circular
cross section and extending along a prescribed axial line (which is
assumed to be extending vertically for the convenience of
description), and a piston 4 is slidably received in the cylinder
3. A combustion chamber is defined by an upper part of the cylinder
3, the top surface of the piston 4 and a cylinder head not shown in
the drawing. The outer peripheral part of the piston 4 is formed
with a first ring groove 6, a second ring groove 7 and a third ring
groove 8, in that order from the top. Each of these ring grooves is
annular in shape. The first ring groove 6 and the second ring
groove 7 receive a first pressure ring 11 and a second pressure
ring 12, respectively, and the third ring groove 8 receives an oil
ring 13.
[0025] The third ring groove 8 includes a bottom surface 8A defined
by a circumferential surface centered around the axial line of the
piston 4 and having a prescribed vertical width (in the axial
direction of the piston 4), an annular upper wall surface 8B
extending radially outward from the upper edge of the bottom
surface 8A and an annular lower wall surface 8C extending radially
outward from the lower edge of the bottom surface 8A. The upper
wall surface 8B and the lower wall surface 8C are both defined by
planes that are perpendicular to the axial line of the piston 4.
Thus, the third ring groove 8 has a rectangular cross section. An
oil ejection passage 14 extends from the corner part defined
between the lower wall surface 8C and the bottom surface 8A of the
third ring groove 8 to the back or inner side of the piston 4.
[0026] In the illustrated embodiment, the internal combustion
engine 1 consists of an automotive engine. The diameter of the
cylinder 3 may be in the range of 68 mm to 92 mm, and the stroke of
the piston 4 may be in the range of 60 mm to 100 mm.
[0027] As shown in FIGS. 1 and 2, the oil ring 13 consists of a
three-piece oil ring including an upper side rail 16, a lower side
rail 17 and an expander (space expander) 18 interposed between the
upper side rail and the lower side rail 17.
[0028] As shown in FIG. 2, the expander 18 includes an annular
expander main body 18A which is made of sheet metal, and is
provided with a wavy shape as it extends along the circumferential
direction. The main body is provided with a circumferential end gap
therein. The wavy shape of the expander main body 18A is thus
defined by upper projecting parts 18B and lower projecting parts
18C created by bending the sheet metal material of the expander
main body 18A in an alternating manner along the circumferential
direction. Each upper projecting part 18B is provided with an upper
ear portion 18D projecting upward from a radially inner part of the
upper surface thereof, and each lower projecting part 18C is
provided with a lower ear portion 18E projecting downward from a
radially inner part of the lower surface thereof. Thus, the upper
ear portions 18D jointly define a radially outwardly facing
shoulder surface forming an obtuse angle with respect to the upper
surface of the upper projecting parts 18B, and the lower ear
portions 18E jointly define a radially outwardly facing shoulder
surface forming an obtuse angle with respect to the lower surface
of the lower projecting parts 18C.
[0029] The side rails 16 and 17 are generally planar annular
members each defining an upper end surface 16A, 17A and a lower end
surface 16B, 17B that are in parallel to each other, and an outer
peripheral surface 16C, 17C and an inner peripheral surface 16D,
17D which are concentric to each other.
[0030] FIG. 3a is an enlarged cross sectional view of an outer
peripheral part of the upper side rail 16 of the oil ring 13, and
FIG. 3b is an enlarged cross sectional view of an outer peripheral
part of the lower side rail 17 of the oil ring 13. In each of these
drawings, the lateral and vertical dimensions are enlarged with
different magnification factors; i.e., the lateral dimension is
five times more enlarged than the vertical dimension. In each of
these side rails 16 and 17, a major outer peripheral surface 16E,
17E which is located in a vertically intermediate part of the outer
peripheral surface of the corresponding side rail 16, 17 is slanted
with respect to the axial line in a radially inward direction or in
such manner that the upper part of the major outer peripheral
surface 16E, 17E recedes away from the opposing cylinder wall
surface or toward the central axial line of the cylinder 3 in
comparison with the lower part of the major outer peripheral
surface 16E, 17E. In other words, each major outer peripheral
surface 16E, 17E defines the outer peripheral surface of a cone
with an upward taper. The slant angle of the major outer peripheral
surface 16E of the upper side rail 16 is defined as a first slant
angle .theta.1, and the slant angle of the major outer peripheral
surface 17E of the lower side rail 17 is defined as a second slant
angle .theta.2. The outer peripheral surface 16C, 17C of each of
the side rails 16 and 17 may be defined as a true conical surface
or a conical surface with a barrel shaped bulge (where a vertically
middle part protrudes radially outward as compared to an outer
peripheral surface of a true cone as is the case with the
illustrated embodiment). The first slant angle .theta.1 and the
second slant angle .theta.2 may also be considered as the average
values of the slant angles of various parts of the outer peripheral
surfaces of the respective side rails 16 and 17.
[0031] In each side rail 16, 17, an upper outer peripheral surface
16F, 17F is defined as a curve (in the cross sectional view) that
smoothly connects the corresponding major outer peripheral surface
16E, 17E with the corresponding upper end surface 16A, 17A.
Similarly, in each side rail 16, 17, a lower outer peripheral
surface 16G, 17G is defined as a curve (in the cross sectional
view) that smoothly connects the corresponding major outer
peripheral surface 16E, 17E with the corresponding lower end
surface 16B, 17B. In each side rail 16, 17, the major outer
peripheral surface 16E, 17E accounts for a much larger part of the
outer peripheral surface 16C, 17C than the corresponding upper
outer peripheral surface 16F, 17F and/or the corresponding lower
outer peripheral surface 16G, 17G. Each upper outer peripheral
surface 16F, 17F has a greater radius of curvature (in the cross
sectional view of FIGS. 3a and 3b) than the corresponding lower
outer peripheral surface 16G, 17G. If each major outer peripheral
surface 16E, 17E is barrel shaped, and is hence provided with a
radius of curvature, this radius of curvature is greater than that
of the corresponding upper outer peripheral surface 16F, 17F.
[0032] The upper side rail 16 and the lower side rail 17 are shaped
identically except for the different configurations of the outer
peripheral surfaces 16C and 17C thereof.
[0033] As shown in FIG. 2, the upper side rail 16, the lower side
rail 17 and the expander 18 are combined in a mutually coaxial
relationship. The lower end surface 16B of the upper side rail 16
abuts the upper ends of the upper projecting parts 18B, and the
inner peripheral surface 16D of the upper side rail 16 abuts the
radially outward facing sides of the upper ear portions 18D. The
upper end surface 17A of the lower side rail 17 abuts the lower
ends of the lower projecting parts 18C, and the inner peripheral
surface 17D of the lower side rail 17 abuts the radially outward
facing sides of the lower ear portions 18E. In the initial
condition of the oil ring 13, as shown in FIG. 2, the upper end
surfaces 16A and 17A of the upper side rail 16 and the lower side
rail 17 are substantially parallel to each other, and the vertical
dimension (the distance between the upper end surface 16A of the
upper side rail 16 and the lower end surface 17B of the lower side
rail 17) of the oil ring 13 is at a minimum value. Alternatively,
in the initial condition of the oil ring 13, the upper end surfaces
16A and 17A of the upper side rail 16 and the lower side rail 17
are provided with a radially outward slant and a radially inward
slant, respectively. A vertical clearance 20 which is present
between the oil ring 13 and the third ring groove 8 as will be
discussed hereinafter is thus taken up equally by the tilting of
the upper side rail 16 and the lower side rail 17 in the opposite
directions in the initial condition of the oil ring 13.
[0034] As shown in FIG. 1, the oil ring 13 is fitted into the third
ring groove 8 such that the upper end surface 16A of the upper side
rail 16 opposes the upper wall surface 8B, and the lower end
surface 17B of the lower side rail 17 opposes the lower wall
surface 8C. The expander 18 is pre-stressed in a direction to
expand in the radially outward direction so that the upper ear
portions 18D press the inner peripheral surface 16D of the upper
side rail 16 in the radially outward direction, and the lower ear
portions 18E press the inner peripheral surface 17D of the lower
side rail 17 in the radially outward direction. In other words, the
expander 18 urges both the upper side rail 16 and the lower side
rail 17 in the radially outward direction. Urged by the expander
18, the outer peripheral surfaces 16C and 17C of the upper side
rail 16 and the lower side rail 17 abut the wall surface 3A of the
cylinder 3. Under this condition, the upper and lower side rails 16
and 17 and the expander 18 are positioned in a coaxial relationship
to the axial line of the cylinder 3 and the axial line of the
piston 4.
[0035] The vertical dimension (the sum of the thicknesses of the
upper side rail 16, the expander 18 and the lower side rail 17) of
the oil ring 13 in the initial condition is smaller than the
vertical dimension of the third ring groove 8 so that a prescribed
vertical clearance (vertical gap) 20 is created between the oil
ring 13 and the third ring groove 8 when the oil ring 13 is fitted
in the third ring groove 8. Owing to this clearance 20, each side
rail 16, 17 is enabled to tilt in such a manner that the outer
peripheral part thereof is higher or lower than the inner
peripheral part thereof. Even if the upper side rail 16 and the
lower side rail 17 are tilted in the radially outward direction in
the initial condition, and the clearance 20 may not be apparent,
the following description applies substantially equally.
[0036] As shown in FIG. 4, when the piston 4 moves upward in the
cylinder 3 toward the top dead center, the side rails 16 and 17 are
pushed downward by the lubricating oil interposed between the outer
peripheral surfaces 16C and 17C of the side rails 16 and 17 and the
opposing wall surface 3A of the cylinder 3. As a result, the lower
end surface 17B of the lower side rail 17 makes a surface contact
with the lower wall surface 8C, and the upper end surface 17A and
the lower end surface 17B of the lower side rail 17 are placed in a
horizontal state or extend perpendicularly to the axial line of the
cylinder 3 (piston 4). Meanwhile, the expander 18 is pushed
downward by the upper side rail 16 so that the lower projecting
parts 18C are brought into contact with the upper end surface 17A
of the lower side rail 17. The upper side rail 16 is caused to tilt
radially outward such that the outer peripheral part thereof is
lower than the inner peripheral part thereof owing to the presence
of the clearance 20 between the upper wall surface 8B and the upper
end surface 16A of the upper side rail 16. The angle defined
between the upper end surface 16A of the upper side rail 16 and a
plane perpendicular to the axial line of the cylinder 3 (piston 4)
at this time is defined as a first tilt angle .theta.3.
[0037] As shown in FIG. 5, when the piston 4 moves downward in the
cylinder 3 toward the bottom dead center, the side rails 16 and 17
are pushed upward by the lubricating oil interposed between the
outer peripheral surfaces 16C and 17C of the side rails 16 and 17
and the opposing wall surface 3A of the cylinder 3. As a result,
the upper end surface 16A of the upper side rail 16 makes a surface
contact with the upper wall surface 8B, and the upper end surface
16A and the lower end surface 16B of the upper side rail 16 are
placed in a horizontal state or extend perpendicularly to the axial
line of the cylinder 3 (piston 4). Meanwhile, the expander 18 is
pushed upward by the lower side rail 17 so that the upper
projecting parts 18B are brought into contact with the lower end
surface 16B of the upper side rail 16. The lower side rail 17 is
caused to tilt such that the outer peripheral part thereof is
higher than the inner peripheral part thereof owing to the presence
of the clearance 20 between the lower wall surface 8C and the lower
end surface 17B of the lower side rail 17. The angle defined
between the lower end surface 17B of the lower side rail 17 and a
plane perpendicular to the axial line of the cylinder 3 (piston 4)
at this time is defined as a second tilt angle .theta.4.
[0038] The first tilt angle .theta.3 and the second tilt angle
.theta.4 can be selected freely, and may be between 2.0 degrees and
6.0 degrees, for instance. More preferably, the first tilt angle
.theta.3 and the second tilt angle .theta.4 may be between 2.0
degrees and 4.0 degrees. The first tilt angle .theta.3 and the
second tilt angle .theta.4 can be adjusted by selecting the
vertical dimensions of the upper side rail 16, the lower side rail
17 and/or the expander 18 for the given vertical dimension of the
third ring groove 8, and/or the flexibility of the expander 18. The
first tilt angle .theta.3 and the second tilt angle .theta.4 may be
equal to each other, or may differ from each other. In the
illustrated embodiment, the first tilt angle .theta.3 and the
second tilt angle .theta.4 are both 2.5 degrees.
[0039] The first slant angle .theta.1 may be greater than the
second slant angle .theta.2 (Condition 1). The first slant angle
.theta.1 may be greater than the first tilt angle .theta.3
(Condition 2). When Conditions 1 and 2 are met, during the upward
stroke of the piston 4, even though the upper side rail 16 tilts
radially outward by the first tilt angle .theta.3, the angle
defined between the major outer peripheral surface 16E of the upper
side rail 16 and the wall surface 3A of the cylinder 3 is still
greater than zero.
[0040] The first slant angle .theta.1 may be preferably greater
than the first tilt angle .theta.3 by an angle between 0.5 degrees
and 4.5 degrees (Condition 3). More preferably, the first slant
angle .theta.1 may be greater than the first tilt angle .theta.3 by
an angle between 1.0 degree and 3.0 degrees. When all of Conditions
1 to 3 are met, it is particularly preferable if the first slant
angle .theta.1 is between 2.5 degrees and 10.5 degrees (Condition
4), and the second slant angle .theta.2 is between 0.5 degrees and
4.5 degrees (Condition 5). More preferably, the first slant angle
.theta.1 may be between 4.0 degrees and 7.5 degrees, and the second
slant angle .theta.2 may be between 1.0 degree and 3.0 degrees.
[0041] In a thrust bearing, the load bearing capacity coefficient
Kw for an infinite plane pad having a width D and moving at a
relative velocity of U can be expressed by the following
equation.
Kw = 6 ( m - 1 ) 2 [ ln m - 2 ( m - 1 ) m + 1 ] ( 1 )
##EQU00001##
where m (=hi/ho) is a parameter representing the slanting of the
pad. hi is the thickness of the oil film at the inlet of the pad,
and ho is the thickness (minimum oil film thickness) of the oil
film at the outlet of the pad as shown in FIG. 6a. The load bearing
capacity coefficient Kw represents the wedge effect (lift) for a
unit surface area, and is known to take a maximum value when m is
about 2.2 and decrease in value with an increase in the slant angle
of the pad as shown in FIG. 6b. When the vertical width (thickness)
of the side rails 16 and 17 is 500 .mu.m, and ho is 3 .mu.m under a
normal rpm condition of the internal combustion engine 1, m=2.2 can
be achieved by setting the slant angle to about 0.4 degrees.
Therefore, the angle between the main part of the outer peripheral
surface 16C, 17C of each of the upper and lower side rails 16 and
17 and the wall surface 3A of the cylinder 3 during the upward
stroke of the piston 4 is preferably 0.5 degrees or more. By noting
the fact that the load bearing capacity coefficient Kw decreases
sharply with the decrease in the value of m when m is less than
2.2, a margin of 0.1 degrees may be allowed for the optimum angle
of 0.4 degrees. Also, in view of variations in the configurations
of the upper and lower side rails 16 and 17, and variations in the
operating condition of the engine, a margin of 0.5 degrees may be
allowed for the optimum angle of 0.4 degrees so that the angle
between the main part of the outer peripheral surface 16C, 17C of
each of the upper and lower side rails 16 and 17 and the wall
surface 3A of the cylinder 3 during the upward stroke of the piston
4 may be preferably 1.0 degree or greater.
[0042] It can be appreciated from the graph of FIG. 6b that m is
desired to be less than 15 to increase the lift owing to the wedge
effect. When the vertical dimension (thickness) of each side rail
16, 17 is 500 .mu.m, and ho in the normal rpm range of the internal
combustion engine 1 is 3 .mu.m, m is 15 when the slant angle is
about 4.8 degrees. Therefore, the angle between the main part of
the outer peripheral surface 16C, 17C of each of the upper and
lower side rails 16 and 17 and the wall surface 3A of the cylinder
3 during the upward stroke of the piston 4 is preferably 4.5
degrees or less. To further increase the lift owing to the wedge
effect, the angle between the main part of the outer peripheral
surface 16C, 17C of each of the upper and lower side rails 16 and
17 and the wall surface 3A of the cylinder 3 during the upward
stroke of the piston 4 is preferably 3.0 degrees or less.
[0043] In the oil ring 13 of the illustrated embodiment, because
the major outer peripheral surface 16E, 17E of each side rail 16,
17 is given with the first slant angle .theta.1 or the second slant
angle .theta.2, during the upward stroke of the piston 4, the major
outer peripheral surface 16E, 17E is slanted with respect to the
wall surface 3A of the cylinder 3, and the resulting wedge effect
causes a lift that pushes the side rail 16, 17 away from the wall
surface 3A of the cylinder 3. Therefore, the scraping of the
lubricating oil from the cylinder wall surface 3A by the side rail
16, 17 is made less active. Also, because the film thickness of the
lubricating oil between the side rail 16, 17 and the wall surface
3A of the cylinder 3 is increased, the shear resistance of the
lubricating oil is reduced, and the fuel economy is hence
improved.
[0044] The upper side rail 16 tilts by the first tilt angle
.theta.3 during the upward stroke of the piston 4. However, because
the first slant angle .theta.1 is greater than the first tilt angle
.theta.3, the major outer peripheral surface 16E of the upper side
rail 16 is slanted with respect to the wall surface 3A of the
cylinder 3 such that a lift owing to the wedge effect is applied to
the upper side rail 16. When the first slant angle .theta.1 is
between 2.5 degrees and 6.5 degrees, and the tilt angle is 2.0
degrees, the angle between the major outer peripheral surface 16E
of the upper side rail 16 and the wall surface 3A of the cylinder 3
during the upward stroke of the piston 4 will be between 0.5
degrees and 4.5 degrees so that the upper side rail 16 receives a
relatively large lift directed in the radially inward direction
owing to the wedge effect. When the first slant angle .theta.1 is
between 6.5 degrees and 10.5 degrees, and the tilt angle is 6.0
degrees, the angle between the major outer peripheral surface 16E
of the upper side rail 16 and the wall surface 3A of the cylinder 3
during the upward stroke of the piston 4 will be between 0.5
degrees and 4.5 degrees. When the second slant angle .theta.2 is
between 0.5 degree and 4.5 degrees, the angle between the major
outer peripheral surface 17E of the lower side rail 17 and the wall
surface 3A of the cylinder 3 during the upward stroke of the piston
4 will be between 0.5 degrees and 4.5 degrees so that the lower
side rail 17 receives a relatively large lift directed in the
radially inward direction owing to the wedge effect. When the
second slant angle .theta.2 is between 1.0 degree and 3.0 degrees,
the angle between the major outer peripheral surface 17E of the
lower side rail 17 and the wall surface 3A of the cylinder 3 during
the upward stroke of the piston 4 will be between 1.0 degree and
3.0 degrees.
[0045] The wedge effect produced in each side rail 16, 17 during
the downward stroke of the piston 4 is small, and the distance
between the side rail 16, 17 and the wall surface 3A of the
cylinder 3 is kept small so that the lubricating oil on the wall
surface 3A of the cylinder 3 is effectively scraped off, and a
prescribed thickness of lubricating oil is maintained on the wall
surface 3A of the cylinder 3.
[0046] FIG. 7 is a graph showing the frictional force associated
with the lower side rail 17 and the minimum oil film thickness ho
(the oil film thickness at the lower end (outlet) of the major
outer peripheral surface 17E of the lower side rail 17) during the
upward stroke of the piston 4 in relation with changes in the
second slant angle .theta.2 of the lower side rail 17. This graph
was obtained by experimental tests conducted under the condition
where the diameter of the cylinder 3 is 73 mm, the stroke of the
piston 4 is 78.7 mm, the tension of the oil ring 13 is 14.5 N, and
the vertical dimension of the lower side rail 17 is 500 The
internal combustion engine 1 was operated at the rotational speeds
of 1,500 rpm and 6,000 rpm. As can be appreciated from FIG. 7, when
the engine rotational speed is 1,500 rpm, the minimum oil film
thickness ho increases with an increase in the second slant angle
.theta.2 when the second slant angle .theta.2 is smaller than about
0.7 degrees, and decreases with an increase in the second slant
angle .theta.2 when the second slant angle .theta.2 is greater than
about 0.7 degrees. Thus, when the engine rotational speed is 1,500
rpm, the friction is minimized when the second slant angle .theta.2
is about 0.7 degree where the minimum oil film thickness ho attains
the maximum value, and increases thereafter with an increase in the
second slant angle .theta.2. Similarly, when the engine rotational
speed is 6,000 rpm, the minimum oil film thickness ho increases
with an increase in the second slant angle .theta.2 when the second
slant angle .theta.2 is smaller than about 1.2 degrees, and
decreases with an increase in the second slant angle .theta.2 when
the second slant angle .theta.2 is greater than about 1.2 degrees.
Thus, it can be concluded that setting the second slant angle
.theta.2 to a relatively small value such as less than 4.0 degrees
is beneficial in reducing the frictional resistance during the
upward stroke of the piston 4.
[0047] FIG. 8 is a graph showing the influences of the shapes of
the outer peripheral surfaces 16C and 17C of the side rails 16 and
17 on the consumption of lubricating oil. In particular, this graph
compares the lubricating oil consumptions (LOC) when the oil ring
13 of the illustrated embodiment is used, and the lubricating oil
consumptions (LOC) when a conventional oil ring is used. In regards
to the oil ring 13 of the illustrated embodiment, the first slant
angle .theta.1 of the upper side rail 16 was 4.5 degrees, the
second slant angle .theta.2 of the lower side rail 17 was 2.5
degrees, and the tilt angle was 2.5 degrees. The vertical dimension
of each side rail was 500 .mu.m. The conventional oil ring is
similar to the oil ring 13 of the illustrated embodiment except for
in the configurations of the outer peripheral surfaces 16C and 17C
of the upper and lower side rails 16 and 17. The upper and lower
side rails of the conventional oil ring were identical to each
other including the slant angles of the major outer peripheral
parts thereof. As indicated by the broken lines 100 in FIG. 3b, the
major outer peripheral surfaces 16E and 17E of the conventional oil
ring were each given with a zero slant angle, and were each barrel
shaped or have a vertical middle part that bulge out radially
outward.
[0048] The internal combustion engine 1 was operated at the
rotational speed of 6,800 rpm in Test (1), and at a low rpm that
changes cyclically so as to simulate an engine brake condition in
Test (2). In Test (1), the speed of the movement of the oil ring 13
was so fast that a relatively thick oil film was formed, and each
side rail 16, 17 was significantly lifted away from the wall
surface 3A of the cylinder 3. On the other hand in Test (2), the
speed of the movement of the oil ring 13 was so slow that a
relatively thin oil film was formed, and each side rail 16, 17 was
substantially in contact with the wall surface 3A of the cylinder
3. Therefore, the lubricating oil consumption was significantly
higher in Test (1) than in Test (2). In Test (2), the consumption
of lubricating oil consumption was attributed largely to the fact
that the throttle valve was substantially closed, and the negative
pressure thereby created in the combustion chamber caused the
lubricating oil to be drawn into the combustion chamber primarily
via the end gap of the oil ring 13.
[0049] By comparing the results of Tests (1) and (2), it was
confirmed that the oil ring 13 of the illustrated embodiment allows
the lubricating oil consumption to be reduced as compared with the
conventional oil ring. It is surmised that the oil ring 13 of the
illustrated embodiment was effective in reducing the lubricating
oil consumption in Test (1) because the oil ring 13 of the
illustrated embodiment scrapes upward the lubricating oil on the
wall surface 3A of the cylinder during an upward stroke of the
piston 4 to a less extent than the convention oil ring. Also, it is
surmised that the oil ring 13 of the illustrated embodiment was
effective in reducing the lubricating oil consumption in Test (2)
because the oil ring 13 of the illustrated embodiment prevents or
minimizes the deposition of lubricating oil on top of the piston.
Therefore, even when the negative pressure in the combustion
chamber is significant as was the case in the condition of Test
(2), the amount of lubricating oil drawn into the combustion
chamber by the negative pressure was reduced, and the consumption
of lubricating oil was minimized.
[0050] Although the present invention has been described in terms
of a preferred embodiment thereof, it is obvious to a person
skilled in the art that various alterations and modifications are
possible without departing from the scope of the present invention
which is set forth in the appended claims. The contents of the
original Japanese patent application on which the Paris Convention
priority claim is made for the present application as well as the
contents of the prior art references mentioned in this application
are incorporated in this application by reference.
* * * * *