Note: Descriptions are shown in the official language in which they were submitted.
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ADJUSTMENT MECHANISM FOR VALVES
FIELD
The present invention relates to improvements in engines, such as
internal combustion engines, particularly to the actuation of valves and most
particularly, poppet valves for internal combustion engines.
The present invention also has application to engines or pumps which
uses valves.
BACKGROUND
The available torque from an internal combustion engine is largely
dependant on the volumetric efficiency of the engine. For reciprocating piston
engines, this efficiency is a measure of the volume of atmospheric air drawn
into
the cylinders during an inducaion stroke, relative to the swept volume of the
cylinderls.
The valve timing of the reciprocating internal combustion engine has a
significant effect on the volumetric efficiency of the engine at particular
engines
speeds. An engine having fixed valve timing, ie a fixed crankshaft angle for
valve opening before piston at top dead centre and a fixed crankshaft angle
for
valve closing beyond top dead centre, will have a particular engine speed
where it operates most efficiently. At this speed, the fixed syncnroncsation
of the
inlet and exhaust valves opf;ning and closing relative to the piston position
create the combination giving the most torque.
Obviously it is desirable to have the most torque possible available over a
wide range of engine speeds.. To achieve the maximum torque at high engine
speeds, it is desirable to have the valves (inlet and exhaust) open for as
much
piston travel as possible. This gives air more time to enter and exhaust gases
more time to exit the cylinder and therefore increases volumetric efficiency.
However, there are limiting factors for how much piston travel, or how much of
' an angle (of crankshaft rotation) the inlet and exhaust valves can be kept
open.
For example, increasing the angle that the valves are open increases the angle
that both the inlet and exhaust valves are open at the same time, which is
called
valve overlap. Valve overlap is desirable at high engine speeds as it
increases
torque output. However, the same amount of valve overlap that produces good
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torque at high engine speeds will cause the engine to run poorly and reduces
torque output at low engine speeds. Accordingly, in general, opening the
valves
earlier and closing them later improves volumetric efficiency at high engine
speed at the expense of torque at low engine speed. Conversely decreasing
valve overlap increases engine torque at low engine speeds but does not give
the best efficiency at hiigh engine speeds.
It is therefore desirable to have a mechanism whereby the timing of the
valve opening and closing can be adjusted according to parameters such as
engine speed, in order to optimise the torque across a range of engine speeds.
Further, other parameters, such as for example, throttle position and which
gear
is engaged, may be used to vary the timing of the opening and closing of the
valves.
Apart from valve timing, there are other factors which are important in the
operation of reciprocating engine poppet valves. Firstly, just before the
valve is
opened, the valve actuator should accelerate slowly towards the valve, in
order
to reduce then eliminate the ~~learance between the valve and the actuator or
between any intervening tappet arrangement and the actuator. This is to ensure
that the valve and actuator do not impact on each other with large velocities
or
forces. The valve then needs to be opened as quickly as possible in order to
facilitate the filling of the cylinder with fresh air and fuel in the case of
an intake
valve, or empty the cylinder of exhaust gas in the case of an exhaust valve.
Once opened, the valve should be held open for as long as possible before
closing rapidly. The valve should then reseat as gently as possible and then
stay closed until the cycle repeats. As there should be no radical changes in
motion, {excessive acceleration) of the valve, a substantially sinusoidal
motion
has been found to be acceptable in providing a path for valve movement.
The actuation of valves and the control of their motion has been
accomplished in the past by the use of camshafts. Camshafts have an eccentric
cam lobe that actuates a valve, wherein the profile of the cam lobe determines
the motion characteristic of the valve. A problem with this arrangement is
that
the camshafts spin rapidly and the valves rely on valve springs to keep them
in
contact with the outer surface of the cam lobe. As the camshafts spin more
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rapidly, the valves can leave the surface of the cam lobe due to inertia. This
problem has been addressed in part by increasing the strength of the valve
spring, however this makes opening the valve harder and increases wear on the
cam lobe surface.
Another major problem with camshafts is the inability to change the lobe
shape, making modification of t:he motion characteristics of the valve
difficult. To
modify the valve timing, the camshaft needs to be replaced or machined, with
the result that torque is only optimised over a narrow speed range for a
particular cam lobe profile. This is one of the reasons that engines that
perform
well at high speed usually lank torque in the lower range of engine speeds.
Further, as the valve spring:. push against the cam lobes as the camshaft
rotates, significant twi~;ting forces are generated along the camshaft, which
can
result in camshaft breakage.
There are existing devices that attempt to solve some of the above
problems, however, none are completely satisfactory. One device is a cam shaft
having two standard cram lobes for the two inlet valves, and a third cam lobe
between the two standard inlet lobes. When the engine is spinning below a
certain engine speed, the inlet valves are actuated by the standard cam lobes.
When the engine accelerat~a over a predetermined engine speed, a pin
engages with the valve's actuators, which allows both the valves to be
actuated
by the third cam lobe, which has a different profile suited to high engine
speeds,
wherein the inlet valves open earlier and stay open longer. A similar
mechanism operates in the exhaust valve camshaft. This system has the
disadvantage that it is. not po ssible to vary the valve opening and closing
times
between the two predetermined valve motion characteristics, ie there are only
two valve opening durations available. This results in a marked "step" in
torque
output from lower rpm to higher rpm and fails to achieve the maximum torque
output across the whole range of engine speeds, as effectively only two
specific
engine speeds are optimised.
Another method of varying the valve opening and closing angle is where
the camshaft speed is. not always half the crankshaft speed over parts of a
single
revolution, but varies according to the engine speed. For example, at low rpm
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the camshaft may spin at the standard rate of half the crankshaft speed. At
higher engine speeds, a mechanism mounted on the camshaft causes the
camshaft to spin at lower than half crank speed while opening the valves and
keeping them open, thus ensuring that the valves are open over a wider angle
than at lower speed. In order to make up lost time (as the crankshaft must
average one revolution for every two crankshaft revolutions), the camshaft
must
then spin faster than half crank speed for the remainder of the revolution to
ensure that it is in the correct position when it is time for the valves to
open
again. This system is obviously less than ideal as a complex mechanism is
used to vary the speE~d of th~~ camshaft with respect to the crankshaft over a
single revolution. Furi:her, valve lift cannot be modified as the cam lobe
profiles
cannot be modified.
Another disadvantage of most valve actuation means is that they
comprise a cam shaft which opens the valves. Camshafts are difficult to
manufacture, and are subject to wear and breakage.
It has also beE~n foune~ that the method of adjusting valve clearances
between the top of the valve and the valve actuator, for example rocker arm or
cam shaft lobe, has disadvantages, such as the need for the clearance
adjusting
mechanism to be on the rocker arm, thereby adding inertia to the rocker arm,
or
the use of shims which are difficult to get at under the cam lobes, and
require
buckets to locate them, which add to the overall length of the valve assembly
and therefore add to the dimensions of the engine.
SUMMARY OF INVENTION
It is an objet, of the present invention to alleviate at least one
disadvantage associa~led with the prior art.
To this end, the present invention provides a means for adjusting the
motion characteristics of a valve. The motion characteristics of the valve
include
timing, such as the crankshaft's angular location before the top dead centre
reference angle where the valve opens, duration, such as the angle of
crankshaft rotation for which lche valve will stay open, lift or travel the
amount of
lift of the valve for a given crankshaft angular location, rate of travel
and/or force.
In one form, the adjustmenil is actuated mechanically. In another form, the
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adjustment means is located between a valve actuation means and the valve.
Advantageously, adjusting t;he motion characteristics of the valve by way of
the
present invention, enables selection of engine performance criteria from a
range
of predetermined characteristics, together with a selection of the degree to
5 which the criteria is. to be performed. For example, the adjustment of valve
motion characteristics may be selected in a manner which accentuates engine
torque. Or, selection may be made to accentuate engine fuel economy.
It may also be desirable to produce a valve actuation means which
produces an approximate sinusoidal motion of valve lift in relation to
crankshaft
rotation and/or which also ;allows the motion characteristics of the valve to
be
varied.
Usually the valve actuation means includes a rotating member.
Typically the adjustment means varies the valve opening angle, and/or
the valve closing angle and/or the valve lift, either individually or
collectively. It
has been found that it is advantageous to vary the valve lift and duration,
and
that while these may be done separately; it has been found that it is
beneficial to
increase valve lift and valve opening duration as engine speed rises.
Accordingly, it is desirable that the adjustment means varies the valve
opening and closing angle and the valve lift collectively.
In another form, thE~ invention provides an apparatus for adjusting the
motion characteristi~~s of a valve, including adjustment means to adjust the
valve
motion in accordance with the adjustment means travel along a non-straight
path.
In another form, the invention provides an adjustment means for use in an
apparatus for adjusting a ~~~notion characteristics of a valve comprising a
plate
having a guide path.
In another form, the invention provides an apparatus for adjusting the
motion characteristics of a valve including a first guide path and a second
guide
path wherein the motion characteristics of the valve are determined by
differences in shape and/or alignment between the first and second guide
paths.
In another form, the invention provides an apparatus for adjusting the
clearance of a valve aci:uated by a desmodromic valve actuation means
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including a valve having a threaded end portion.
In a preferred embodiment, the means for adjusting the motion
characteristics of thE~ valve include an adjustment member having a guide path
and a pivotally mounted valve actuation member having at least one guide
surface, wherein a pin moves along both the guide path and the guide surtace,
causing the pivotally mounted actuation member to pivot and move the valve.
Typically, the pin is driven in a substantially cyclic motion.
Desirably the guide path of the adjustment member and the guide surface
of the valve actuation member are not collateral over their entire length, ie
there
is a difference in thE~ paths such that they deviate from each other at least
over
part of their length. This difference in paths produces the movement of the
actuation member as the pin travels along both paths. Also, a kinematic
inversion of pin and guide i s contemplated as an alternative embodiment.
PREFERRED EMBODIMENT
One or more of the preferred embodiments of the present invention will
now be described, with reference to the accompanying drawings, wherein:
Figures 1 a-1 d show a schematic representation of an adjustment
mechanism in accordance with the present invention in various states of
assembly;
Figure 2a shows a schematic side view of the adjustment mechanism of
5 the present invention and a prior art valve actuation mechanism;
Figure 2b shows a schematic view of a non-desmodromic adjustment
mechanism of the present invention and a prior art valve actuation mechanism.
Figure 3 shows an isometric view of part of a first embodiment of the
adjustment mechanism of 'the present invention;
10 Figure 4 shows an isometric view of all of the first embodiment of the
adjustment mechanism of the present invention;
Figures 5a and 5b show end views of a second embodiment of the
adjustment mechanism of the present invention;
Figures 6a and 6b show end views of the adjustment mechanism shown
15 in figures 5a and 5b;
Figure 7 is a graph of typical extremes of variation in lift and duration of
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the valves compared with the position of the crankshaft, as varied by the
adjustment mechanism of the present invention;
Figure 8 is a first embodiment of a guide plate of the adjustment
mechanism of the present invention;
Figure 9 is a second embodiment of the guide plate of the adjustment
mechanism of the present invention;
Figures 10a-10d are embodiments of rocker arms of adjustment
mechanism of the present invention;
Figure 11 is a schematic side view of a first embodiment of a slot of a
guide plate of the adjustment mechanism in accordance with the present
invention;
Figure 12 is a schematic side view of a profiled surface of a guide plate of
the adjustment mechanism in accordance with the present invention;
Figure 13 is a schematic side view of a second embodiment of the slot of
the guide plate of the adjustment mechanism of the present invention;
Figure 14 is a schematic representation of the slot of the guide plate of
the adjustment mechanism of the present invention;
Figures 15a-15d are embodiments of a sliding pin of the adjustment
mechanism of the present invention;
Figures 16a-'I6d are embodiments of the guide plates of the adjustment
mechanism of the present invention;
Figure 17a is a first embodiment of a guide plate adjustment means of the
adjustment mechanism of the present invention;
Figure 17b is a second embodiment of the guide plate adjustment means
of the adjustment mechanism of the present invention;
Figure 18a is a perspective view of a valve clearance adjustment
mechanism in accordance with the present invention.
Figure 18b is an exploded perspective view of the valve clearance
adjustment mechanism shown in figure 18a.
Referring to I=figures 2a, 2b, 4, 5a and 5b, a mechanism 10 is shown for
adjusting the motion characteristics of a poppet valve 1. The mechanism 10
includes an actuation means, for example a valve crankshaft 12 having a crank
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pin 13, which is used to provide the cyclic displacement motion and base
timing
for actuation of the valve 1. The valve crankshaft 12 is normally driven by
the
crankshaft (not shown) by known means such as a belt, chain drive, or gears,
at
half the crankshaft rotation speed. The mechanism 10 is typically mounted in
the
head of a reciprocating four stroke engine (not shown) and further includes a
pivot point 14 which is fixed to the head of the engine to pivotally locate a
valve
actuator, such as rocker arm 16 which actuates the motion of the valve 1. An
adjustment member ouch as guide plate 18 is mounted to the head such that it
is
able to be moved within a range of positions, for example a first position 20
and
a second position 22 as shown in figures 1 b, 5a, 5b, 6a, 6b and 9.
Crank pin 13 is attached to a conrod 24 through aperture 15 at one end
and has a member such as a guide member, for example sliding pin 26, at the
other end as shown in figures 1 a-1 d, 2 and 3. As the valve crankshaft 12
turns,
the conrod 24 moves the sliding pin 26 along a guide, such as path 25 in the
guide plate 18. The guide plate 18 does not move in response to movement of
the sliding pin 26, and the sliding pin 26. is constrained to move along path
25.
The sliding pin 26 also travels along a guide path 28 in rocker arm 16. There
is
typically one rocker' arm 16 per valve 1, and accordingly there may be two
rocker arms 16 if two inlet (or exhaust) valves are used per cylinder. The two
rocker arms 16 and guide plates 18 can be served by a single conrod 24 and
sliding pin 26 as shown in figure 4, thus actuating two valves (inlet or
exhaust)
simultaneously as in a four valve per cylinder engine head. Obviously the
number of valves that can be actuated by a single conrod and/or sliding pin is
not limited to two per inlet / exhaust. The path 28 may be in the form of a
slot
having an upper and lower profiled surface, in the case of desmodromic valve
actuation, as shown in figures 1 a, 1 b, 2a, 4, 5a, 5b, 6a, 6b,1 Oa, 1 Oc, 13
and 14
or it may be a single profiled surface in the case of conventional valve
actuation
with a spring provicling the valve closing force as shown in figures 2b 10b
and
1 Od.
In either method of valve actuation described above, the path 25 in the
guide plate 18 causes the sliding pin 26 to move in a way constrained by the
profile of the path 25, which causes the rocker arm 16 to pivot about pivot
point
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14. The rocker arm 16, when pivoting, has a forked actuator arm 32 distal from
the pivot point 14 which pushes on the valve 1 via two nuts tightened on the
valve stem, as shown in figures 5, 18a and 18b and described later, thus
causing valve 1 to open and close according to the differences in the profile
of
path 25 and path 28.. It is the difference in path profiles as shown in
figures 13
and 14 that causes the rocker arm 16 to pivot, and the profiles of the paths
28
and 25 may be varied according to the motion characteristics desired from the
valve, which may vary from engine to engine, or with the purpose of the
engine.
Thus the individual path profiles of the guide plate 18 and rocker arm 16 are
not
intended to be limited to the various embodiments shown in this specification.
Further, the valve 1 can be replaced by valve system 100 as shown in
figures 2a and 2b, wherein valve system 100 includes a known shim and bucket
arrangement that allows the valve clearance to be adjusted, and a valve spring
101 ensures that the valve stays in contact with the actuator 32 when the
valve
is closing, which rnay be used in non-desmodromic or conventional valve
actuation. The mechanism 10 may simply replace camshaft 102 as a valve
actuation means.
The assembled parts of the mechanism 10 can be seen in figures 1 a to
1 d, wherein the assembly c~f the conrod 24 to the crankshaft 12 by the crank
pin
13, and the attachment of the guide plate 18 and rocker arm 16 to the sliding
pin 26 is shown schematically. A partially assembled adjustment mechanism is
shown in figure 3, wherein the assembled valve crankshaft 12, conrod 24 and
sliding pin 26 are sE~en in relationship to the pivot point 14 which is
normally in a
fixed position, but can be rotated and includes in this embodiment eccentric
section 11. The assembled mechanism 10 is shown in fig 4 wherein two guide
plates 18 are slidinc~ly attached to the eccentric section 11 of the rocker
shaft 14
and two rocker arms 16 arE~ pivotally attached to the pivot point 14, which
would
enable the mechanism 10 to operate two inlet or exhaust valves. The guide
plates, in this case, are constrained to slide linearly by pins (not shown)
which fit
into guideways 9.
In operation, the valve crankshaft 12 rotates at half engine crankshaft
speed. The conrod 24 is connected at one end to a crank pin 13 on the valve
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crankshaft 12 and at the other end to the sliding pin 26. The pin 26 is
located in
a guide path 25 of the guide plate 18. As the valve crankshaft 12 rotates, the
pin
26 is constrained to move along the path 25 of the guide plate 18, however the
guide plate 18 can move from a first position 20 to a second position 22, and
5 any number of positions therebetween. The profile of the guide path 25, as
shown in the figures, defines the trajectory of pin 26. The pin 26 also slides
along path 28 of~~he rocker arm 16, and the different profile between the path
28
and path 25 causes the rocker arm 16 to pivot back and forth about pivot point
14. The actuator 32 attached to rocker arm 16 moves with the arm 16 and
10 contacts the end of valve 1, pushing the valve open and pulling the valve
closed. Where non-desmodromic valve actuation is desired, a valve spring may
close the valve 1.
The position ~of the guide plate 18 can be varied, in the case of Figures
2a, 2b, 3 and 4 by rotation of the rocker shaft 14, in a second embodiment
adjustment is due to eccentric adjusting shaft 30 as can be seen in figures
5a,
5b, 6a, 6b 17a and 17b. The shaft 30 has an eccentric off-centre lobe 31 which
can be turned within aperture 34, thus causing the guide plate 18 to move from
a first position 20 wherein the motion characteristics as shown by line 40a
and
40b of figure 7 suit: low engine speed, to a second position 22, wherein the
motion characteristics of the valve suit high engine speeds shown by line 42a
and 42b, also of in figure 7. The movement of the guide plate 18 can be seen
in
the comparison of open valve positions shown in figures 5a and 5b. In figure
5a, the adjusting shaft 30 and lobe 31 position the guide plate 18 in the
first
position 20. In figure 5b, the adjusting shaft 30 and lobe 31 position the
guide
plate 18 in the second position 22 and thus the maximum valve opening, as
seen in figure 5b is greater than the maximum valve position seen in figure
5a.
The operation of thE; shaft 30 and lobe 31 in the aperture 34 in the guide
plate is
shown in figures 17a and 17b and will be described in more detail below.
The motion characteristics of the valves can be seen in figure 7, wherein
line 40a represents the valve lift of an inlet valve (vertical axis) versus
crankshaft
rotation angle (horizontal axis) for the valve 1 actuated by valve crankshaft
12
while the guide plate 18 is in the first position 20. The exhaust valve motion
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characteristics when the guide plate 18 is in the first position 20 are shown
by
line 40b. Line 42a represents the valve motion characteristics when the guide
of an inlet valve when the guide plate 18 is in the second position 22. The
exhaust valve motion characteristics when the guide plate 18 is in the second
position 22, can be seen in line 42b. As can be seen from figure 7, there is a
significant difference between valve lift, valve opening duration and valve
overlap when the guide plate 18 moves from a first position 20 to a second
position 22.
The reason for the difference in motion characteristics is that when guide
plate 18 is in the first position 20, the profile of path 28 is positioned
such that
the differences in the profile between path 25 and path 28 are minimised, as
can be seen in figures 5a, Era and 8 and discussed below. This provides a
lower
valve lift as the pin 26 deflects less, as shown in the position of pin 26a in
figure
8.
The first position 20 of guide plate 18 opens the valve the least amount,
and over the shortest angle, and is therefore normally used for low engine
speeds where excessive v<~Ive overlap is undesirable and increased turbulence
is desirable. The second position 22 of the guide plate 18 is used to generate
larger valve overlap and higher lift in the valves, as seen in the position of
the
pin 26b in figure 8, and the extension of the valve in figure 5b compared to
the
extension of the valve in figure 5a. This arrangement is used during high
engine speeds whE~re maximum gas flow is required. Figures 6a and 6b show
the guide plate 18 in the first position 20 and second position 22
respectively,
but the valves in both cases are closed fully, i.e. regardless of the position
of
guide plate 18, the valves still close effectively as shown by the equal
positions
of the valves in figures 6a and 6b. The difference in the positions of the
plate 18
is clearly seen by 'the gap between pivot point 14 and guide plate 18 in
figure
6b, whereas in figure 6a there is no gap. The mechanism 10 allows the guide
plate 18 to be positioned at any point between the first position 20 and the
second position 22, thus allowing the amount of valve overlap and/or the angle
of valve opening to be adjusted to any point within the predetermined limits
of
the valve motion characteristics. This also provides the advantage of being
able
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to modify the valve opening angle and/or lift according to any change in the
conditions in order to maximise volumetric efficiency.
The differences between the profiles of path 25 in guide plate 18 and the
path 28 in rocker arm 16 .are designed to impart the desired valve motion
characteristics to thc~ valve. For example, when used with a rocker arm 16
having a straight pal:h 28, the path 25 shown in figure 11, is made up of four
portions, each with a specific; function. Portion A is a portion whereby, when
the
sliding pin 26 is in this portion, the valve will be closed. As the pin 26
travels
along the path 25, it moves to portion B, which is a ramp section designed to
allow the pin 26 to begin to move at an angle to the direction of motion in
portion
A. This allows the acauator ',32 to be brought into contact with the valve (or
valve
shim) relatively slowly, as there is usually a small gap between the top of
the
valve assembly and i:he valve actuator. Once the valve has contacted the top
of
the valve, sliding pin 26 entE~rs portion C of path 25, where the slope of the
path
increases greatly and thus causes the actuator to push open the valve quickly.
Once the maximum valve opening is approached the sliding pin 26 enters
portion D whereby the velocity of the valve while opening is reduced, and the
valve starts to decelE:rate. In portion D, the sliding pin 26 reaches the end
of its
travel and the valve crankshaft 12 begins to pull the sliding pin 26 back
along
the portion D in the reverse direction, thus starting to close the valve
again.
Figure 12 shows the portions A-D of a profiled surface of a guide plate
which uses a spring to return the valve to the closed position, and therefore
does not require the lower portion of the path. The shape of the paths 50 and
51
in figures 11 and 1;? respectively are designed to be used with a rocker arm
having a substantially straight path 28.
If the path of the rocker arm had a shape the same as the shape of the
path of the guide plate, then the rocker arm would not move relative to the
guide
plate and accordingly there would be no motion of the valve. Therefore there
are numerous shapes that either the path of the guide plate, or the path of
the
rocker arm can taH;e in order to produce the required motion of the valve
providing that the other of the rocker arm or guide plate has a profile that
is
different. As an example, the shape of the path 52 of the guide plate shown in
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figure 13 can be used, provided the shape of the rocker arm path 53 differs in
the correct areas to provide the motion in the rocker arm. This difference in
path
shapes' is shown in figure 14 wherein the paths 25 and 28, have been
overlapped in order to highlight the differences in the profiles which then
cause
the rocker arm to deflect and actuate the valve.
From figures '13 and 14, the differences in the paths in the guide plates
and rocker arms can be seen, and the differences relate to valve lift. Figure
13
relates to path 52 in a guide plate that is adjusted rotatably, for example as
shown in figure 9. It can be seen that this arrangement allows a far greater
difference between paths 5.? and 53, and accordingly, a far higher valve lift
is
achieved than in thE~ linearly adjustable guide plate shown in figure 14. This
increased valve lift shown in figure 13 is accomplished without a radical
increase in path deviation, which would be necessary in a linearly adjustable
guide plate, such as that shown in figure 8. It is undesirable to have too
large a
deviation in any of i:he paths as this may lead to increased wear on the path
surfaces which will cause the valve motion characteristics to change.
An advantagE~ of the present system is that by altering the differences in
the profiles of the paths 25 and 28, it is possible to produce a valve motion
with,
for example, a more square top than that shown in figure 7.
In order to overcome or reduce wear due to high contact pressure
between the path 28 and the sliding pin 26, it has been found that the sliding
pin
can be made with a non-circular cross section, called a wear portion, in the
region where it travels alone the path 28.
In figure 15a,~ a wear portion 160 is shown having a flat upper and lower
surface where the pin conl:acts a straight path 28. The profile of the
surfaces
varies to match the 'facing surfaces of the paths. For example, the wear
surfaces
can be flat as in wear surtace 160 when used with path 128 in rocker arm 116
as shown in figure 10a. A,Iternatively the wear surfaces can be curved with a
common centre of curvature as shown by wear surfaces 360 in figure 15c to suit
a similarly curved path 328 in figure 10c of constant radius. If conventional
or
non-desmodromic valve actuation mechanisms are used, then the sliding pin
only needs one wear surface 260 or 460 as shown in figures 15b and 15d, as
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the valve spring willl ensure continuous contact of the wear surtace with the
opposing path 28 surface.
It should be noted that the embodiments shown in the figures 10a-10d,
15a-15d, and 16a-16d, woirk together in respective sets. A rocker arm 116
having a path 128 as shown in figure 10a is used with a gudgeon pin 126
shown in figure 15a and a guide plate 118 having a path 125 shown in figure
16a. This art'angement forms an adjustment mechanism employing
desmodromic valve .actuation wherein the guide plate 118 is adjusted linearly.
Similarly, a rocker arm 216 having a path 228 (figure 10b) works with a
gudgeon pin 226 (figure 15b) and a guide plate 218 having a path 225 (figure
16b) to form an adjustment mechanism employing a valve to close the valve,
wherein the guide plate 218 is adjusted linearly.
A rocker arm 316 having a curved path 328 (figure 10c) works with a
gudgeon pin 326 (figure 15c) and a guide plate 318 having a path 325 (figure
16c) to form an adjustment mechanism employing desmodromic valve
actuation, wherein the guidE; plate 318 is adjusted pivotally.
A rocker arm 416 having a curved path 428 (figure 10d) works with a
gudgeon pin 426 (figure 15d) and a guide plate 418 having a path 425 (figure
16d) to form an adjustment mechanism employing a valve to close the valve,
wherein the guide plate 418 is adjusted pivotally.
In figure 17a embodiments of a mechanism for adjusting the position of
the guide plate 18 is shown. The adjusting shaft 30 is situated in the
aperture
34 in the guide plate 18. E3y rotating the adjusting shaft 30, the eccentric
cam
lobe 31 on the adju:;ting shaft 30 causes the guide plate 18 to move linearly,
for
example, as shown in figure 8. The amount of linear movement of the guide
plate 18 is determined by the amount of rotation of the shaft 30. This allows
the
guide plate 18 to be adjusted to any point between and including the two
extreme positions, being the first position 20 and the second position 22.
In figure 17b, the shaft 30 is rotatably received in to an aperture 134 in a
guide plate 618 mounted :~o as to be pivotally adjustable about point 135. As
the shaft is rotated.. eccentric lobes 31 force the guide plate 618 to move.
The
guide plate is constrained 'to move pivotally and therefore, twisting the
shaft 30
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causes the guide plate 618 1:o move. As above, the amount of movement of the
guide plate 618 can be controlled by the rotation of the shaft 30.
A control means (not shown) is used to control the rotation of the shaft 30
for each mechanism 10 which enables the guide plate to be positioned
5 anywhere between the first (position 20 and the second position 22. The
control
means may be a simple device for advancing the valve opening by twisting the
shaft, or any other suitable means for moving the guide plate. Such
mechanisms are commonly used to advance the ignition timing as engine
speed rises. The valve timing in this case may be adjusted either with or
10 independent of the ignition timing.
A further embodiment of a guide plate 518 is shown in figure 9 wherein
the guide plate 518 is mounted to a rotatable pivot point 535, so that
adjustment
of the motion characteristics of the valve can be made by rotating pivot point
535
to which the guide plate !518 is attached, to any position between the two
15 positions as shown by the arrow and dotted line, rather than linear motion
as
shown by the arrow in figurE~ 8.
It should also be 'understood that the guide plates in any of the
embodiments disclosed may be positioned in discrete locations between the
first position 20 and the second position 22, for example by the use of a
stepper
motor. This would allow the position of the guide plates to be varied in steps
according to data from various parameters such as engine speed, rate of
change of engine apeed, throttle position and gear position. Accordingly, a
fuzzy logic table could be set up to position the guide plates in the optimum
position for a set of predefined parameters.
Figures 16a to 16d show further alternative arrangements for the guide
plates. Each guide plate is arranged to be mounted in such a way that its
position is able to be controlled in order for the position of the path for
the sliding
pin to be controlled. In a non-desmodromic arrangement as shown in figure 16b
and 16d, there is no requirement for the path to be a slot, and as such
profiles
125 and 325 can tie used, as a spring acting on the valve can be used in a
conventional manner to c4ose the valve and accordingly there will always be
pressure on either profile 125 or profile 325 and the underside of the
respective
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rocker arms 118 or 328. This arrangement has the advantage that there is a
large body of knowledge regarding the use of valve springs to close a valve.
Also, the reciprocating rocker arms may be made lighter.
Embodiments of the means for adjusting the position of the guide
plate is shown in figure 17a and 17b. Figure 17a relates to a method of
producing linear adjustment in the guide plate using a shaft 30 in an aperture
34
ur
in the guide plate 18. The :>haft has a lobe 31 which moves the guide plate to
the desired position when the shaft 30 is turned. Aperture 34 is designed to
move the guide plat~a 18 linearly, and therefore has substantially straight
side
walls. As many engines of the type that use poppet valves have numerous
valves in alignment, a single shaft with multiple lobes 31 can be used to move
all the guide plates 18 simultaneously.
A further embodiment is shown in figure 17b, wherein the shaft 30 is used
to cause a rotational motion in the guide plate 18. The twisting of the shaft
30
with eccentric lobe 31 in aperture 134 causes the guide plate to pivot about
fixed point 135. If the guide plate is mounted about a pivot point, as shown
in
figure 9, then the rotation of the shaft 30 will cause the guide plate to
rotate, and
thus increase or decrease the difference between the paths in the guide plate
and rocker arm, which will effect the motion characteristics of the valve. As
the
aperture 134 is designed to move the guide plate 18 pivotally, side wall 136
is
longer than side wall 137.
In the above embodiments, the rocker arm has pivoted while the guide
plate has moved either linearly or pivotally. It can be readily determined
that the
rocker arm could also move linearly in response to the movement of the pin in
the path of the guids~ plate. Further, the guide plate may be fixed in place,
and
all the adjustment movements can take place on the rocker arm, eg the rocker
arm could have its pivot point moveable with respect to the guide plate. This
arrangement has thE: advantage that the guide plate is then fixed, and all the
movement is undertaken by the rocker arm, making the mounting of the guide
plate greatly simplified.
It can be seen from the embodiments disclosed that the movement of the
guide plate 18 from its first position to the second position causes the
sliding pin
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26 travelling along path 25 to not only increase the crank rotation angle
across
which the valves open, but also increases valve lift at the same time. These
aspects in combination produce a result that is very desirable, as two of the
valve characteristics change with only a change in one parameter, that being
the movement of the guide plate. It is desirable to have the valves increase
their
lift at high engine speeds to ensure that the maximum amount of air enters the
cylinder or exhaust gas exits from the cylinder in the time provided. However,
at
low engine speeds, it has been found that increased turbulence in the air
entering the cylinder is desirable as it assists in the atomisation of the
fuel in the
air. When engines operate at low speeds, the velocity of the air entering the
cylinder is also low, and therefore there is not as much turbulence in the air
as it
passes the inlet valves into the cylinder. It has been found that decreasing
the
valve lift and duration increases turbulence and therefore increases fuel
atomisation, which increases torque. At higher engine speeds, the turbulence
from the faster air flow provides sufficient energy for fuel atomisation, and
the
limiting factor becomes the amount of air able to be squeezed into the
cylinder.
The present invention allows for the adjustment of not only the valve opening
duration, but also valve lift with only one parameter being adjusted.
The motion characteristics of the valve may also be varied in accordance
with factors such as throttle position and also which gear is selected.
It should be noted that it is not essential to increase valve lift and
duration
with engine speed, and that it may be desirable under certain circumstances to
decrease valve lift a.nd/or durations of the inlet and/or exhaust valve as
engine
speed increases which the present invention is also able to accommodate.
In figure 18a there is shown a guide plate 418 used in desmodromic
valve actuation, having two branches 420, each branch having an actuator 32.
The actuators 32 sit between an upper flange member 422 and a lower flange
member 424 at the upper end of a valve 1. The valve 1 includes a threaded
portion 426, which has a lower nut 425 including the lower flange member 424
threadedly attached thereto, as shown in figure 18b. An upper nut 428, which
is
threadedly attached to the threaded portion 426 of the valve 1, includes the
upper flange member 422. The gap between the upper flange member 422 and
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the lower flange member 424 may be set by an intermediate shim member (not
shown) which would fit between the upper nut 428 and lower nut 425, whereby
the size of the shim determines the gap between the upper flange member 422
and lower flange member 424.
In the embodiment shown, the upper flange member 422 includes a
spacer 423 which contacts a corresponding spacer 427 on the lower flange
r
member 424, thus. providing the appropriate gap between the flanges.
Typically, the size of the gap is slightly larger than the diameter of the
portions of
the actuator 32 that contact the upper and lower flanges, thereby allowing a
clearance between the flanges and the actuators 32. The upper and/or lower
nuts may be held in position by lock nuts (not shown). The valve clearance may
then be adjusted by removing the upper lock nut (if provided), removing the
upper nut 428 having upper flange member 422 and spacer 423, and replacing
the upper flange member 422 with spacer with another flange member and
spacer of suitable size, then reattaching the lock nut onto the threaded
portion of
the valve 1. In this way, the valve clearance can be adjusted to take into
account any wear in the system, without having to replace the guide plate 418.
The spacer 423 may be integral with or separate to the upper flange member
422, and the lower flange member 424 may also be replaced if desired.
Alternatively, the upper nut 428 and upper flange member 422 may be
locked into position by the upper nut 428, and the gap between the upper and
lower flanges can be set by the position in which the upper nut 428 and flange
member 422 are set.
It is important that the contact surfaces of actuator 32 be of constant
radius, as in this way the valve clearance will be constant as the valve is
opened
and closed, and the rocker ~~rm 16 moves about pivot point 14.
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