The present invention relates to pistons and piston rings for internal
Piston rings are employed on a piston both to prevent combustion
gases escaping from the combustion space, and which is known as blowby, and to
limit the amount of lubrication oil entering the combustion space which would
otherwise cause excessive oil consumption. In order to achieve these aims, three
or more rings are generally provided. Modern internal combustion engines often
employ a three-ring piston ring pack comprising two compression rings in the upper
and centre positions and an oil control ring in the lower position. It has generally
been considered that for practical purposes, two compression rings in contact
with the cylinder wall are the minimum necessary for efficient control of combustion
gas blowby allied to acceptable oil consumption.
The ability to produce an efficient piston having only one compression
ring in contact with the cylinder wall would bring many advantages. Firstly, there
are the dimensional consequences such as the piston compression height and weight
being reduced and secondly, benefits in reduced friction. A reduction of the compression
height, that is, the distance between the piston crown and gudgeon pin axis, allows
an overall reduction of engine height which consequently means lower bonnet lines
in cars allowing improved aerodynamics, etc. However, piston weight also leads
to reductions in stresses imposed on gudgeon pins, connecting rods, bearings,
crankshaft, etc., which may all be made smaller and, therefore, lighter.
Various methods of improving the gas sealing of compression rings
have been proposed. These include rings within a single groove but having multiple
elements bearing against the cylinder wall. Such an arrangement is described by
Schunichi in European Patent Application No. 0 069 175. Other methods comprise
the provision of an auxiliary ring which lies in two grooves, one in the inner
axial face of the piston ring and the other in the axial wall of the piston ring
groove, the auxiliary ring thus forms a seal between the main ring and the piston.
Examples of this type of ring assembly are described by Graham in GB 2,117,868A
and Williams in US 2,228,495. The embodiments described by Graham show a relatively
deep groove in the piston ring groove for the auxiliary ring comparable with the
groove depth in the inner axial face of the piston ring which bears against the
cylinder wall. The auxiliary ring in some embodiments is shown as a free floating
ring in the grooves, neither expanding outwardly against the outer piston ring,
nor contacting the inner wall of its own groove in the piston. In one instance
the auxiliary ring is shown as a two-piece assembly substantially as two semi-circular
elements. The auxiliary ring is also shown in two embodiments as being a rigid
flange-like feature extending outwardly from the inner axial face of the piston
Williams describes an arrangement in which an inner ring fits into
an annular groove cut into the axial wall of the piston ring groove and again
co-operates with a corresponding groove in the inner axial face of the piston ring
which bears against the cylinder wall.
It is necessary in Williams, however, that the inner auxiliary or
sealing ring fulfils an expander function to increase the force with which the
main piston ring bears against the cylinder walls. Additionally, it is necessary
that the inner sealing ring is a "snug" fit in its grooves in the piston wall
and in the main ring in order to prevent tilting and limit movement of the main
ring in the piston axial direction. For the sealing ring to be a "snug" fit in
both grooves it is necessary that the axial depth of each groove is equal. The
sealing ring groove in the piston wall should also be of greater radial depth
than the corresponding groove in the piston ring. It is clear, therefore, that
the sealing action of the inner ring is mostly due to it being a "snug" fit in
its grooves and also due to a labyrinth sealing effect.
Both Graham and Williams show the inner sealing ring to have straight,
parallel and concentric inner and outer axial faces and indeed Williams makes a
point of this in the text.
It has now been found that unexpectedly good results are given when
the inner auxiliary ring and the outer main ring are so constructed that the inner
ring does not limit the movement of the outer ring in its groove and furthermore
that the inner ring is able to tilt and form seals with the outer main ring.
According to the present invention, a piston for an internal combustion
an annular piston ring groove having two radially extending walls and an axially
a compression ring received in the piston ring groove for sealing engagement with
an associated cylinder wall or liner;
a recess in the compression ring extending in an outwardly direction;
an auxiliary sealing ring in permanent but deflectable contact with the axially
extending wall of the annular piston ring groove and penetrating the recess in
the compression ring;
the arrangement being such that the auxiliary sealing ring does not physically
restrict the movement of the compression ring in the annular piston ring groove.
The recess may be in the form of a groove in the inner axial face
of the compression ring or in the form of a rebate in the top face of the compression
The upper and lower radial walls of such a groove in the inner face
of the compression ring may not necessarily be parallel, but may lie at an angle
of between 0° and 4° to a radial plane passing through the groove. The upper groove
face may extend upwardly and to the outside of the groove and the lower groove
face may extend downwardly and to the outside of the groove. In section, therefore,
the groove may be frusto-conical.
The depth of the groove in the compression ring is preferably not
less than 0.5mm and more preferably at not less than 1mm.
Preferably, there may be a shallow depression in the axially extending
wall of the annular piston ring groove, the depression being sufficient to locate
the inner auxiliary sealing ring in, for example, the centre of the axially extending
wall to facilitate fitting of the compression ring to the piston body and to limit
axial movement of the inner auxiliary ring in operation.
Preferably, the auxiliary sealing ring is in the form of a steel
rail having a rounded inner endface such that deflection by pivoting about the
inner end face is facilitated.
It is preferable that the inner auxiliary sealing ring extends beyond
the inner axial face of the compression ring to a distance of at least 0.5mm and
certainly not less than 0.25mm and more preferably to a depth of greater than
In order that the invention may be more fully understood, embodiments
will now be described by way of example only with reference to the accompanying
drawings of which:
Figure 1 shows an axial section through part of a stationary piston and piston
ring according to the invention within an engine;
Figure 2 shows the piston and piston ring of Figure 1 during a power stroke
in an internal combustion engine;
Figure 3 shows the piston and piston ring of Figure 1 sealing against a downward
flow of gas;
Figure 4 shows the piston and piston ring of Figure 1 sealing against an upward
flow of oil;
Figure 5 shows an axial section of a first alternative form of piston and piston
ring according to the invention;
Figure 6 shows an axial section of a second alternative form of piston and
piston ring according to the invention;
Figure 7 shows an axial section of a third alternative form of piston and piston
ring according to the invention;
Figure 8 shows an axial section of part of a fourth alternative form of piston
and piston ring according to the invention; and
Figure 9 which shows an axial section of part of a fifth alternative form of
piston and piston ring according to the invention.
Referring now to Figures 1 to 4 where the same features are denoted
by common reference numerals.
In Figure 1 a piston body is shown generally at 10. In the body 10
is a circumferential piston ring groove, also shown generally at 11. The groove
11 comprises two parallel radial faces 12 and 13 and an inner axially extending
face 14. In the centre of the face 14 is a shallow circumferential depression
15. Lying in the groove 11 is a main compression piston ring made from cast iron
and shown generally at 16. The ring 16 comprises an outer axial face 17 which
bears against a cylinder wall 18, two substantially parallel radial faces 19 and
20 and an inner axial face 21. In the inner axial face 21 is a groove, shown generally
at 22. The groove 22 comprises an axial face 23 and two radial faces 24 and 25
which lie at an angle of approximately 2° to horizontal planes 26 (shown as dotted
lines) passing through the groove 22. Seated in the depression 15 and also penetrating
the groove 22 is an auxiliary sealing ring shown generally at 27. The ring 27
is a circular split steel rail formed to a diameter less than that of the piston
10 such that the rail is permanently in contact with the piston 10 when fitted.
The ring 27 comprises two parallel radial faces 28 and 29 and rounded ends 30 and
In a typical piston of about 68mm diameter, the main compression
ring 16 may have an axial thickness of about 1.5mm and a radial thickness of about
3.25mm. The groove 22 has a radial depth of approximately 1.8mm and a minimum
axial depth, defined by the wall 23 of about 0.6mm. The ring 27 has a radial thickness
of about 2.5mm and an axial thickness of approximately 0.5mm. Penetration of the
ring 27 into the groove 22 varies between approximately 1.25 and 1.75mm under
normal running conditions. The depression 15 may be between 0.05 and 1.0mm in
radial depth but typically about 0.25mm.
Figure 2 shows the configuration assumed by the ring components during
a combustion stroke in an internal combustion engine. Gas pressure from the burning
fuel charge forces the main compression ring 16 downwards such that the face 13
of the piston ring groove 11 and face 20 of the ring 16 are in intimate contact.
The end 30 of the auxiliary sealing ring 27 is also forced in a downwardly directionby
pivoting about the end face 31 seated in the depression 15 to contact the face
25 of the groove 22. Gas pressure also forces the ring 16 in an outwardly direction
producing a radial force to seal the face 17 against the cylinder wall 18, thus
effecting an efficient gas seal. Blow-by around the back of the piston ring 16
is minimised by the seals formed between the end 31 and depression 15 and between
the end 30 and groove face 25.
It will be noted that the gas pressure forcing the ring 16 in an
outwardly direction to seal the face 17 against the cylinder wall 18 may substantially
only act radially on the upper portion of face 21 and on face 23. The result of
this is that friction between the ring face 17 and the cylinder wall 18 is reduced
compared with a conventional ring. Furthermore, because of the groove 22 machined
in the back of the ring 16, the tangential tension or load applied by the ring
in the radial direction against the cylinder wall 18 is also reduced giving an
additional reduction in friction betwen ring and cylinder. The reduction in tangential
tension is of the order of 10 to 20% and the reduction in the radial force generated
by the gas pressure is of the order of 20 to 40%.
Figure 3 shows the configuration assumed by the ring components at
the top of an exhaust stroke, for example, where the piston 10 has just changed
direction to descend in the cylinder, but piston ring 16 inertia carries the ring
to the top of the ring groove 11 where face 19 and face 12 meet. If there is still
any gas pressure above the ring 16 it will tend to keep the auxiliary sealing ring
27 depressed with its end 30 sealing against the face 25, thus preventing blowby.
Otherwise the components may adopt the configuration shown in Figure 4 where the
end 30 of the sealing ring 27 seals against the top face 24 of the groove 22. In
this configuration, the sealing ring 27 prevents passage of oil into the space
above the sealing ring 27 where it could be burnt by the combustion gases. This
could cause higher oil consumption and possible carbonisation and clogging of
the various working clearances.
In general, the angular and axial positions adopted by the piston
ring and the sealing ring will depend upon the balance of forces and turning moments
arising from gas pressures, component inertia and interfacial friction.
It has been found in pistons according to the invention where two
conventional top compression rings have been replaced by one composite ring as
described above, that reductions of piston compression height of 10 to 15% may
be achieved. On a piston of about 68mm diameter a reduction of compression height
of 13% over the conventional piston was achieved. This reduction of compression
height on the 68mm piston also resulted in a reduction of piston weight of 8%
leading potentially to other benefits as described above. In other piston embodiments,
weight reductions of up to 10% may be achieved.
Pistons may be produced as stated above where two conventional top
compression rings are replaced by one composite ring, giving benefits directly
attributable to the composite piston ring per se plus the additional indirect
benefit in that friction is still further reduced by elimination of the second
conventional compression ring. However, where it is considered that reduction
of gas blowby levels and/or oil consumption may be more important considerations
in an existing engine, for example, than reduction of compression height etc.,
it is of course advantageous in terms of these parameters to employ a composite
ring as described above as the top ring of a three-ring piston. The second ring
may be another composite ring. Even where a three-ring piston is retained, composite
rings as described still confer benefits of reduced ring-to-cylinder friction in
addition to reduced blowby and oil consumption. The second or third ring of a
two or a three-ring piston will usually be an oil control-type ring.
Extensive engine testing on dynamometers has been undertaken on pistons
according to the invention with total testing in excess of 3500 hours of engine
running having been accumulated. Pistons utilising a single composite compression
ring (and a conventional conformable oil control ring) and pistons having a composite
top compression ring and a conventional one-piece second compression ring (plus
oil control ring) have been tested in engines. The results of these engine tests
have been compared with tests carried out on standard production engines having
conventional three-ring pistons. The results of the engine dynamometer tests are
shown in Table 1. TABLE 1
In the above table all engine pistons employed a conventional oil
It may be seen from Table 1 that the single composite compression
ring produces blowby levels which are comparable with those of a conventional
three-ring piston and entirely adequate in performance. The oil consumption figures,
however, are in all cases superior when using a single composite ring compared
with the standard conventional piston. In the case of the 1.6L Ford engine a single
composite ring has given an improvement of 25% over the standard engine and when
used in conjunction with a conventional second compression ring the improvement
is over 60%.
A vehicle test, covering in excess of 30,000 miles, of a Ford Escort
using 1.6L gasoline engine which had already completed 400 hours running on a
dynamometer and having a single composite compression ring and conventional oil
control ring resulted in an oil consumption of 35,000 mpg (Imperial) under normal
A second vehicle test of a British Leyland Montego using a 1.6L engine
which had also completed 400 hours dynamometer running and fitted with a single
composite compression ring and a conventional oil control ring resulted in an
oil consumption of 24,000 m.p.g. (Imperial) over 10,000 miles of normal driving
Alternative embodiments are envisaged for some applications. Figure
5 shows a piston embodiment similar to that shown in Figure 1 but where the depression
15 of Figure 1 has been omitted. The sealing ring 27 is still produced such that
it grips the piston groove axial face 14 with an inwardly directed radial force.
The degree of such force may be increased to limit substantial axial movement of
the end 31 relative to the face 14.
A second alternative embodiment is shown in Figure 6. In this embodiment,
a main compression ring shown generally at 40 does not employ a full groove in
the inner circumferential face but a rebate having an upper vertical face portion
41, a sloping face portion 42 and a lower vertical face portion 43. A sealing
ring is shown generally at 44 and having rounded ends 45 and 46 is seated in a
depression 15 in an axial face 14 or a circumferential piston ring groove shown
generally at 11. The sealing ring 44 is dished such that it permanently grips the
groove 15 with an inwardly directed radial force. The ring 44 is also formed such
that the end 45 is always resiliently biased towards the sloping face portion
42 and is, under most circumstances, in contact. The resilient biasing, however,
is of insufficient force to significantly restrict movement of the ring 16, either
inwardly or axially in the groove 11 the biassing force being merely sufficient
to maintain the end 45 in substantially constant contact with the face 42.
In a third alternative embodiment shown in Figure 7 a piston 50 having
a diameter of 74mm is provided with a piston ring groove 51. The main compression
ring 52 has a nominal axial height of 1.2mm and a radial thickness of 3.6mm. The
ring 52 has a rebate in the upper face 53 having an upper face portion 54 and
a sloping face 55. The ring 51 is slightly dished and the upper face 53 and lower
face 56 touch the piston ring groove 51 at the points 57 and 58 respectively.
The inner axial face of the ring groove 51 is profiled to provide in effect a
supplementary groove 59 which both axially locates and provides a seat for the
steel rail sealing ring 60 at one end 61. The sealing ring 60 in its free state
is dished and is fitted with its natural bias in a downwardly direction. The free
position of the ring 60 in the absence of the main ring 52 is shown as the dashed
line 62. The radial thickness of the sealing ring 60 is about 2.75mm and its axial
thickness about 0.5mm. With the main ring 52 and the sealing ring 60 in position
the sealing ring contacts the main ring at the point 63 on the sloping face 55.
In operation the composite ring behaves in the following manner: On combustion
the main ring 52 is forced in a downwardly and outwardly direction such that the
face 56 is held against the lower radial face of the groove 51 and the cylinder
wall 64 thus sealing the combustion gases. The sealing ring is also forced in
a downwardly direction increasing the sealing force at the position 63. The position
of the main ring 52 on the exhaust stroke will depend upon the balance of forces
operating but the sealing ring 60 is substantially always in contact with the
main ring 52 at the point 63 thus maintaining an efficient gas and oil seal. On
the induction stroke the main ring 52 will contact the groove 51 at the points
57 and 58 thus helping to minimise the passage of oil past the ring.
The main ring 52 shown in Figure 7 may be manufactured from either
cast iron or steel as indeed may be the main ring of any of the embodiments described.
The latter material may be in the form of an extruded wire section which may be
wrapped to shape with or without subsequent machining. The embodiment of Figure
7 may be easier and thus more economic to manufacture due to the less stringent
and easier to apply tolerances on the rebate of the ring 52 compared to the grooves
of the embodiments shown in Figures 1 to 5. It will also be appreciated that this
ring is of extremely low axial height and potentially further reduces the piston
Figure 8 shows a detail of a ring and piston assembly similar to
that shown in Figure 7. In this embodiment the inner axial wall 70 of the groove
51 is flat. The sealing ring 60 seats on the wall 70 at the point 61 and the composite
ring operates in essentially the same manner as the embodiment of Figure 7. It
is sufficient when assembling to ensure that the point of seating 61 of the sealing
ring 60 is towards the top of the wall 70 to allow easy fitting of the main ring
In Figure 9 the main ring 80 has a rebate in the upper face 81, the
rebate having an upper face portion 82, a generally horizontal portion 83 and
a sloping portion 84. The purpose of the sloping portion 84 in this embodiment
is to provide a lead-in during assembly for the sealing ring 85 which is dished
in a downwardly direction in its free state (similar to that described with reference
to Figure 7). The end 86 of the ring 85 rests on the face portion 83 in operation.
The end 87 of the ring 85 is located in a supplementary groove 88 of the piston
ring groove 89.
It may be seen in all of the embodiments described that the inner
sealing ring does not physically limit the extent of movement of the main compression
ring within its groove. The ability of the sealing ring to pivot on its seat on
the piston wall allows it to form dynamic seals with various faces of the groove
or recess in the compression ring, thus enhancing the performance of the piston.
Although the sealing ring has been described as being made of steel it may be
made from aluminium or copper-based alloys or even some of the later engineering
plastics materials an example of which is polyetheretherketone (PEEK).
In all cases the relative dimensions must be such that the main compression
ring is able to fully retract into its groove under thrust conditions, resulting
from piston lateral movements, without being impeded by the sealing ring. Conversely,
under conditions where the main compression ring is at the furthest outward extent
from its groove, the penetration of the sealing ring into the groove in the inner
axial face of the main ring should preferably not be less than 1mm.
A benefit conferred by the embodiment shown in Figure 1 etc., is
that the groove 22 is significantly easier and thus cheaper to produce by having
a frusto-conical form rather than parallel radial faces.
The main ring shown in Figures 1 to 5 may be used in conjunction
with a dished sealing ring of the type shown in Figures 7 to 8.
The main compression ring (16,40,52,80) may, of course, be surface
treated in any known manner such as, for example, nitro-carburising, chromium
or molybdenum facing to improve wear characteristics, etc.
Where the main ring is used in conjunction with a dished sealing
ring it has been found that wear of the lower face of the piston ring groove in
the piston body is reduced.
The invention disclosed in this specification may be further combined
with piston bodies of the type disclosed in GB 2,104,188 where the piston skirt
is provided with raised lands. The provision of such lands having appropriate
contours allows hydrodynamic lubrication of the piston in the cylinder bore, giving
reductions in friction between the piston and cylinder wall.
1. A piston for an internal combustion engine and comprising:
at least one annular piston ring groove (11,51,89) in the piston the groove having
two radially extending walls and an axially extending wall;
a compression ring (16,40,52,89) received in the piston ring groove for sealing
engagement with an associated cylinder wall or liner (18,64) and characterised
in that the compression ring has a recess extending in an outwardly direction;
an auxiliary sealing ring (16,44,60,85) in constant but deflectable contact with
the axially extending wall (14,59,70,88) of the annular piston ring groove and
penetrating the recess in the compression ring;
the arrangement being such that the auxiliary sealing ring does not physically
limit the extent of movement of the compression ring in the annular piston ring
2. A piston according to Claim 1 and characterised in that the recess in the
compression ring is in the form of a groove (22) in the inner axial face thereof.
3. A piston according to Claim 2 and characterised in that the groove in the
inner axial face of the compression ring is frustoconical in section.
4. A piston according to Claim 3 and characterised in that the upper (24) and
lower (25) radial faces of the groove in the compression ring extend at an angle
of between 0° and 4° to plane (26) passing through the groove, the upper face extending
upwardly and to the outside of the groove and the lower face extending downwardly
and to the outside of the groove.
5. A piston according to any one of Claims 2 to 4 and characterised in that
the radial depth of the groove is greater than 0.5mm.
6. A piston according to any one preceding claim and characterised in that
the compression ring is treated to improve wear performance.
7. A piston according to any one preceding claim and characterised in that
the axially extending wall of the piston ring groove in the piston has a circumferentially
extending depression (15) therein.
8. A piston according to Claim 7 and characterised in that the depression has
a radial depth of between 0.05 and 1.0mm.
9. A piston according to any one preceding claim and characterised in that
the auxiliary sealing ring is a steel rail.
10. A piston according to Claim 9 and characterised in that at least the radially
inner end of the steel rail section is rounded (31,46,87).
11. A piston according to Claim 9 and characterised in that both ends of the
steel rail section are rounded.
12. A piston according to any one claim from 2 to 11 and characterised in that
the radially outer end (30) of the auxiliary sealing ring penetrates the groove
in the inner axial face of the compression ring by at least 0.5mm.
13. A piston according to Claim 1 characterised in that the recess is in the
form of a rebate (41,42) (51,55)(82,83,84) in the top face (53,81) of the compression
14. A piston according to Claim 13 characterised in that the auxiliary sealing
ring is dished (62) and resiliently biassed so as to be in substantially constant
contact with the compression ring.
15. A piston according to Claim 13 characterised in that the compression ring
is dished (52).
16. A piston according to Claim 13 characterised in that the rebate in the
compression ring comprises a substantially horizontal portion (83) and a sloping
17. A piston according to Claim 1 and characterised by having a second compression
ring received in a second annular piston ring groove in the piston.
18. A piston according to Claim 17 and characterised in that the second compression
ring is a composite ring.
19. A piston according to any one preceding claim and characterised by having
an oil control ring.