Note: Descriptions are shown in the official language in which they were submitted.
11~652~
ENGINE BRAKING APPARATUS
TECHNICAL FIELD OF INVENTION
This invention relates generally to engine
braking apparatus of a gas compression relief type.
The invention relates more particularly to a pressure
relief apparatus which automatically disables one or more
operating cylinders of the compression relief engine brake
whenever the forces in the hydraulic circuit of the engine
brake exceed a predetermined level.
BACKGROU~D ART
Engine brakes of the compression relief type are
well known in the art, Such engine brakes are designed to
convert, temporarily, an internal combustion engine of the
spark ignition or compression ignition type into an air
compressor so as to develop a retarding horsepower which
may be a substantial portion of the operating horsepower
normally developed by the engine.
As a general rule, so long as the retarding
horsepower developed during braking operations does not
exceed in absolute value the operating horsepower for
which the engine was designed, the stresses on the crank-
shaft, bearings and drive train, though opposite in direc-
tion will not exceed the allowable stresses for these
parts and the addition of the compression relief engine
brake will not adversely affect the operating life of
the drive train components of the engine and vehicle.
At the same time, the engine brake will supplement ~he
braking capacity of the primary vehicle wheel braking
system and extend, substantially, the life of the primary
.
1 ~56S26
- ~ -2
braking system. The basic design for an engine braking
system of the type here involved is disclosed in the
Cummins U.S. Patent 3,220,392.
The compression relief engine brake of the type
disclosed in Patent 3,220,392 employs a hydraulic system
wherein the motion of a master piston controls the motion
of a slave piston which opens the exhaust valve of the
internal combustion engine near the end of the compres-
sion stroke whereby the work done in compressing the
intake air is not recovered during the expansion or
"power" stroke but, instead, is dissipated through the
exhaust and radiator systems. The master piston is cus-
tomarily driven by a pushrod controlled by the engine
camshaft. It will be apparent that the force required
to open the exhaust valve will be transmitted back through
the hydraulic system to the pushrod and camshaft. In order
to minimize modification of the engine, it is common to
utilize an existing pushrod which moves at an appropriate
time to operate the engine brake hydraulic system. In some
cases,an e~st valve pushrod is selected while, in other
cases, it is convenient to use the fuel in;ector pushrod.
The problem of assigning a second function to an
existing pushrod is that an increased load may be exper-
ienced which may exceed the design capacity of the pushrod
or cam shaft. The present invention generally solves this
problem by providing an automatic means (a) to unload
the engine brake whenever an excessive loading condition
becomes im~inent and (b) to reactivate the engine brake
as soon a8 the temporary excess loading condition has
terminated so as not to interfere with the effectiveness
of the engine brake.
DISCLOSURE OF THE INVENTION
With the foregoing in mind, we provide in accord-
ance with the invention an engine braking apparatus of a
gas compression relief type including an internal combus-
tion engine having exhaust valve means and pushrod means,
hydraulically actuated piston means for opening said
exhaust valve means at a predetermined time, and further
565~6
-3-
piston means actuated by said pushrod means and hydraulical-
ly interconnected with said exhaust valve opening piston
means in a high pressure hydraulic fluid circuit, char-
acterized by a bi-stable valve located in said high pres-
sure hydraulic fluid circuit and having at least primaryand secondary orifices, and damping means associated with
said valve to rapidly damp out vibrations of said valve
while it is.moving from it8 closed position defining a high
pressure condition, until it comes.to rest in an open posi-
tion defining-a low pressure condition, said primary and
secondary ori.fices and damping means maximizing the flow
through said valve and minimizing the time required to attain
said low pressure condition.
The apparatus of the invention will rapidly
respond to a hydraulic pressure in the brake system above
a desired predetermined pressure and maintain the system
pressure for the balance of a cycle at a fraction of the
predetermined pressure whenever an excess pressure is
sensed. With our apparatus the pressure drop will occur
rapidly and with a minimum number of pressure oscillations
and the apparatus will automatically reset itself after
operation so as to restore it to the regular operating
mode.
BRIEF DESCRIPTION OF THE DRAWING
Objects and advantages of the invention, involving
a special design multi-stage pressure relief valve, which
may be accommodated within the master piston of the engine
brake, will be apparent from the following disclosure
taken in conjunction with the following drawings, in
which:
Figure l is a schematic drawing of a compression
relief engine brake incorporating the improved pressure
relief system in accordance with the present invention;
Figure 2 is an enlarged cross-sectional view
of an engine brake master cylinder incorporating a pres-
sure relief system according to the present invention;
1 156S26
r4~
Figure 3 is an enlarged cross-sectional view of
an engine brake master cylinder havlng a modified pressure
relief system aecording to the present invention;
Figure 4 is a diagram showing the variation
in the force exerted on the pushrod to open the exhaust
valve and to actuate the fuel injector as a function of
engine crank angle position;
Figure 5 is a diagram showing the variation
in the force exerted on the pushrod as a function of
the crank angle when the pressure relief system of the
present invention is activated;
Figure 6 is a graph of engine brake hydraulic
pressure as a function of the engine crank angle for
two configurations of the pressure relief device shown
ln Figure 2.
DETAILED DESCRIPTION OF THE INVENTI~N
Figure 1 is a schematic diagram of a compression
relief engine brake adapted for use in conjunction with
an internal combustion engine of the spark ignition or
compression ignition type. As noted above, the basic
design of the compression relief brake is disclosed in
the Cummins U.S. Patent 3,220,392. For purposes of sim
plicity and clarity, the present invention will be des.
cribed wi.th reference to an engine brake applied to a
Cummins compression ignition engine in which the master
piston of the engine brake is driven by the injector push-
rod. It will be understood that the invention may also
be applied to other applications where, for example, the
master piston is driven by an exhaust valve pushrod.
Moreover, as will be explained below, the pressure re-
lief device herein disclosed may be placed at any con-
venient point in the high pressure hydraulic circuit
although its combination with the master piston is par-
ticularly desirable.
Referring now to Figure 1, the numeral lQ
represents a housing fitted on an internal combustion
engine w.thin which the components of a compression
relief engine brake are contained. Oil 12 from a sump
S6~26
~5--
14 which.may ~e, for example, the eng~ne c~ankcase is
pumped thxough a duct 16 by a low press.ure pump 18 to
the inlet 20 of a solenoid valve 22 mounted in the hous-
ing 10. Low pressu~e oill2 is conducted f~o~ the sole~
noid valve 22 to a control cylinder 24 b~ a duct 26~ A
control valve 28 is fitted for reclprocating moYement
within the control cylinder and is urged into a closed
position by a compression spring 3Q. The control ~alve
28 contains an inlet passage 32 closed by a ball check
valve 34 which is biased into the closed position by a
com,pression spring 36 and an outlet passage 38. When
the control valve 28 is in the open position (~s shown
in Figure 1) the outlet passage 38 registers with the
control cylinder outlet duct 40 which communicates with
the inlet of a slave cylinder 42 also formed in the
housing. It will be understood that low pressure oil
12 passing through the solenoid valve 22 enters the con-
trol valve cylinder 24 and raises the control valve 28
to the open position, Thereafter, the ball check valve
34 opens against the bias of spring 36 to permi,t the
oil 12 to flow into the slave cylinder 42. From the
outl,et 44 of the slave cylinder 42 the oil 12 flows
through a duct 46 into the master cylinder 48 formed
in the housing 10.
A slave piston 50 is fitted for reciprocating
motion within the slave cylinder 42~ The slave piston 5Q
is biased in an upward direction (,as shown in Figure 1)
against an adjustable stop 52 by a compression spring
54 which is mounted within the slave piston 50 and acts
against a bracket 56 seated in the slave cylinder 42.
The lower end of the slave piston 50 acts against an
exhaust valve cap or crosshead 58 fitted on the stem of
exhaust valve 60 which is, in turn, seated in the engine
cylinder head 62. As exhaust valve spring 64 normally
biases the exhaust valve 60 to the closed position as shown
1 1 S6526
--6--
in Figure 1, Nor~ally the ~dju~table st~p 52 is set to
proY~de a clearance o~ about 0,Q18 inch (i.e. "lash"~
~etween the slave piston 50 and the exhaust valve cap 58
when the ex~aust valve 60 is closed, the slave piston 50 is
seated against the adjustable stop 52 and the engine is cold.
This clearance is required and is normally suf~icient to
accommodate expansion of the parts comprising the exhaust
valve train when the engine is ho~ without opening the
exhaust valve 60,
A master piston 66 is fitted for recip~ocating
movement within the master cylinder 48 and ~iased in an
upward direction (:as viewed in Figure 1~ ~y a light leaf
spring 68, The lower end of the master piston 66 contacts
an adjusting screw mechanism 70 of a rocker arm 72 con-
trolled by a pushrod 74 driven from the engine camshaft
(not shown), As noted above, when applied to the Cummins
engine, the rocker arm 72 is conveniently the fuel injector
rocker arm and the pushrod 74 is the injector pushrod, In
this circumstance, the pushrod 74 and the exhaust valve
60 are associated with the same engine cylinder,
It will be understood that when the solenoid
valve 22 is opened, oil 12 will raise the control valve
28 and then fill both the slave cylinder 42 and the
master cylinder 48. Reverse flow of oil out of the slave
cylinder 42 and master cylinder 48 is prevented by the
action of the ball check valve 34. However, once the
system is filled with oil, upward movement of the push-
rod 74 will drive the master piston 66 upwardly and the
hydraulic pressure, in turn, will drive the slave pis-
ton 50 downwardly to open the exhaust valve 6Q, Thevalve timing is selected so that the exhaust Yalve 6Q
is opened near the end of the compression stroke of the
cylinder with which exhaust valve 6Q is associated,
Thus, the work done by the engine piston in compressing
air during the compression stroke is released to the
exhaust and radiator systems of the engine and not
~, ...
- 1~56526
--7--
recovered during the expansion str~ke of the engine~
When it is desired to deactivate the compres~ion
brake, the solenoid valve 22 is closed whereby the oil
12 in the control valve cylinder 24 passes through the
duct 26, t~e solenoid valve 22 and the return duct 76
to the sump 14, When the control valve 28 drops down~
wardly as viewed in Figure 1, a portion of the oil în
the slave cylinder 42 and master cylinder 48 is vented
past the control valve 28 and returned to the sump 14
by duct means (not shown~,
The electrical control system for the engine
brake includes the vehicle battery 7~ which is grounded
at 80. The hot terminal of the battery 78 is connected,
in series, to a fuse 82, a dash switch 84, a clutch
switch 86, a fuel pump switch 88 and, preferably, through
a diode 90 back to ground 8Q, The switches 84, 86 and
88 are provided to assure the safe operation of the sys-
tem. Switch 84 is a manual control to deact vate the
entire system. Switch 86 is an automatic switch connected
to the clutch to deactivate the system whenever the clutch
is disengaged so as to prevent engine stalling. Switch 88
is a second automatic switch connected to the fuel ~ystem
to prevent engine fueling when the engine brake is in
operation.
Reference is now made to Figure 2 which shows in
an enlarged cross-sectional view one form of a modified
master piston in accordance with the present invention.
The master piston 66 comprises a hollow cylindrical body
92 Dp~n at the top and having a plurality of drainage
passageways 94 communicating between the interior and
exterior of the body 92, A cap 96 is threaded into the
top of the body 92 and contains ad~usting bores 98 adapted
to receive an appropriate wrench or spanner (not shown),
A central or primary orifice 100 is formed in the cap 96
and communicates with a larger valve bore or secondary
orifice 102, The intersection of the orifice 100 and the
valve bore 102 defines a valve seat lQ4 for a valve lQ6,
-`` 11S6526
-8 .
prefera~ly in the orm of a spheroid or ball, but which
may also, for exa~ple, be conical in shape~ The diameter
of the valve 106 is selected so as to Be slightl~ smaller
than the bore 102 while the cap 96 has a thîckness such
that the bottom surface 108 lies slightly below the center
of the valve 106, A spring 110 mounted ~ithln the hody 92
of the master piston 66 carries a valve guide 112 whlch
biases the valve 106 against the valve seat 104, The
valve guide includes a seat portion 114 and a plunger
portion 116 designed to limit the downward motion of the
valve guide 112 before the spring llQ becomes fully com-
pressed.
In operation, it will be understood that the
pressure in the high pressure side of the engine brake
hydraulic system which includes the slave cylinder 42
and the master cylinder 48 will be transmitted through
the master piston 66 and will appear as a force tending
to compress or buckle the pushrod 74. In addition, the
force required to operate the fuel injector will be
carried as a force moment by the rocker arm 72 and then
reflected as a compressive or buckling force on the push-
rod. However, the hydraulic pressure alone will act on
the valve 106 over an area defined by the orifice 100 to
produce a force tending to open the valve. If the
force due to the hydraulic pressure exceeds the force
due to the spring 110, the valve 106 will be displaced
slightly from the seat 104 whereupon the hydraulic pres-
sure will act on the full projected cross-section of
the valve 106, an area known as the "secondary" area.
3Q As a result, the valve 106 will be rapidly accelerated
to the fully displaced position as limited by contact be-
tween the plunger end 116 of the valve guide 112 with
the bottom of the piston body 92,
Applicants have found that in order to cause
the pressure to be dumped rapidly with a minimum of pres-
sure oscillation it is desirable accurately to define the
ratio of the annular area between the inside of the piston
56526
body 92 and the outer periphery of the sh~ulder portion 118
and the a~e~ af the ori~ice lQ0, The annular area may be
called the "tert~ary'' area as distinguished from the "pri-
mary" area of the ori~fîce lOa and the t'secondary" pro~ected
5 area of t~e valve lQ6, Applicants have discovered that
the ratio of the tertiar~ and primary areas s~ould be
at least 1,0 and preferably about 1,5, W~ere the ratio
is less than 1, a a t~rottling of the flow of ~draulic
fluid occurs which tends to decrease the rate at w~ich
hydraulic fluid is dumped through the piston, When the
area between the shoulder 118 of the ~alve guide 112 and
the inner wall of the master piston 92, the '~'tertiary'?
area, is controlled so as to be between about lQ0% and
150% of the size of the orifice 100, the resistance to
the flow of hydraulic fluid is sufficient so that the
pressure'acts on the upper surface of the valve guide 112
and quickly damps out the vibratory motion of the valve
guide 112 and the valve 106 resulting from the reaction
of the spring 110. As a result, the average opening and
the average time in the open position of the valve 106
are increased whereby the flow ~hrough the valve is
maximized. Tests have shown that when the rat~o of the
tertiary and primary areas exceeds about 150% the damping
effect on the normal vibratory motion of the valve 106
and the valve guide 112 is diminished and when the area
ratio is below 100.% secondary throttling occurs which
also restricts the flow of hydraulic fluid through
the piston 66,
In addition, applicants believe that the effect
of locating the bottom 108 of the cap 98 below the maxi-
mum diameter of the valYe 106 is that the valve 106, when-
fully contained in the valve bore 102, allows a pressure
to develop behind the valve lQ6 in the valve bore lQ2 and
this causes a greater acceleration and increased velocity
of the val~e. The effect is that the valve is Qpen for
a longer time and the average opening is greater whereby
the flow through the ball valve is maximized,
-:-" 115652~
-10-
It will be understood that while a relatively
high hydraulic pressure is required initially to unseat
the valve 106, a much smaller pressure is required to
maintain the valve in the open position. The ratio of
these pressures is approximately equal to the inverse ratio
of the area of the orifice 100, the primary area, and the
cross-sectional area of the valve 106, the secondary area,
The total area of the drainage passageways 94 should be
greater than the area of the orifice 100 and the opening
between the valve 106 and the bore 102 to insure that the
drainage passageway ~4 do not throttle the flow of hydraulic
fluid. It will therefore be appreciated that whenever
the valve 106 opens, it will remain open in a stable
condition until a sufficient quantity of hydraulic fluid
has been dumped so as to establish a low pressure level
in the hydraulic system. The pressure at which the valve
106 will begin to open is controlled by the ~ias exerted
by the spring 110 which acts through the valve guide 112
to hold the valve 106 against the seat lQ4. Such bias
may be regulated by adjusting the cap 96 until the desired
load on the spring 110 is attained. Once adjusted, the
cap 96 may be staked or otherwise locked in the body 92
of the master piston 66 to maintain the adjustment,
It will be understood, that after the hydraulic
pressure in the high pressure system has been relieved,
the valve 106 will automatically reseat and the hydraulic
system will be restored to its normal operating mode, Thus,
the engine brake will again be in condition to oper-
ate.
Referring now to Figure 3, another form of a
pressure relief valve is shown, Parts which are common
to both Figures2 and 3 bear the same identification, The
principal difference in construction lies in the structure
of the valve guide 122 of Figure 1 which comprises a
assymmetric structure having a radially extending shoulder
portion 124, an axially extending plunger 125 and a skew
seat 127, By the term "skew seat", applicants mean that
,...
-- `` 1156526
-11-
the plane of the seat in the valve guide 122 against which
the valve 106 acts is not normal or perpendicular to the
axis of the guide 122 but, instead, is inclined with respect
to that axis as is clearly shown in Figure 3. In this
case the valve 106 should be a ball valve, If the force
due to hydraulicpxe~s~reexceeds the force due to the
spring 110, the ball 106 will be displaced slightly from
the seat 104 whereupon the hydraulic pressure will act
upon the full projected cross-section of the ball valve
106. As a result, the ball valve 106 will be rapidly
accelerated to the fully displaced position and will tend
to "ride down" the skew seat 127. The resulting skew move-
ment of the ball valve 106 in combination with the impact
of the plunger 125 against the bottom of the piston body
92 tends quickly to damp out vibrations and inhibit the
ball valve 106 from reseating itself before the hydraulic
pressure has been fully dissipated by the flow of hydraulic
fluid through the piston body 92 and then through the
drainage passageways 94. As in the case of the structure
shown in Figure 2, the bottom edge 108 of the cap 96
extends slightly below the center of the ball 106 whereby a
pressure of hydraulic fluid tends to be built up în the
valve bore 102 which accelerates the ball 106 to a high
velocity whereby the ball valve is opened more rapidly
to maximize the flow of hydraulic fluid therethrough.
Reference is now made to Figure 4 in which
pushrod force is plotted against engine crank position
in terms of the crank angle before and after top dead
center (TDC). Curve A represents the force required at
the in;ector pushrod to open an exhaust valve, This is
the force transmited through the high pressure hydraulic
system of the engine brake by the slave piston 50 and the
master piston 66. Curve A is in the form of a bell curve
es~sentially symmetric about the TDC point and reflects
the changing pressure within the cylinder. Curve A may
be displaced vertically depending upon the degree of
boost g;ven by the engine supercharger. In Figure 4,
curve A is shown with a typical normal boost of 15 inches
-- -" 1156526
-12-
of mercury. If a higher boost were used, the curve would
be raised while with a lower boost it would be lowered.
Curve B represents the force induced in the
pushrod to open the exhaust valve and hold it open.
Until the clearance or lash in the system is taken up,
essentially no force is induced in the pushrod. However,
once the clearances are taken up, the force in the push-
rod builds rapidly until the exhaust valve begins to
open. Once the exhaust valve begins to open as a result
of the coincidence of Curves B and A at point 126 (Figure
4~, Curve B will peak and then drop to a low level deter-
mined essentially by the force exerted by the exhaust
valve spring 64,
Curve C represents the force induced în the
pushrod due to the operation of the fuel injector train.
This force normally peaks shortly after TDC and the
peak, indicated at point 128, represents the crushing
load on the injector train as the injector is mechanically
seated in the injector body. The maximum force occurs at
point 128 and is considered in the normal design of the
engine.
Curve D represents the total force on the pushrod
due to the combined effect of the exhaust valve opening
load and the injector load and is determined as the sum
of the forces shown by Curves B and C. In general, it will
be noted that Curve D will have two pea~s -- the first occur-
ring approximately when the exhaust valve begins to open
and the second when the injector seats. These peaks are
indicated, respectively, at points 130 and 132, It will
be appreciated that if the time interval between the
peak loads indicated by points 13Q and 132 is decreased
for any reason, such as excessive lash in the system or
~ncreased supercharger boost, for example, the force
required to open the exhaust valve may not have decreased
to its minimum value before the maximum injector load
occurs with the result that the total load on the in;ec-
tor pushrod becomes excessive and buckling of the pushrod
may occur.
,
,
1 ~56526
~13~
Figure 5 illustrates a typical operation of
the present invention wherein the engine brake hydraulic
system is unloaded to prevent damage to the injector
pushrods when an overload condition occurs as a result
of excessive supercharger boost. Curve A is identical
to Curve A of Figure 4 and represents a normal 15 inch
supercharger boost while Curve E indicates the maximum
boost capable of being provided by the engine supercharger.
Curve C is also identical to Curve C of Figure 4 and
represents the force on the injector pushrod due to the
injector load alone.
The line 134 represents the set point for the
pressure relief system of the present invention This
is the predetermined pressure within the hydraulic system
of the engine brake at which the valve 106 will be dis-
placed from its seat 104 so as to dump hydraulic fluid
through the master piston 66 and therefore unload the
system. Curve B' in Figure 5 represents the force in
the pushrod due tothe hydraulic pressure in the engine
brake mechanism and the exhaust valve train, It is
similar to Curve B of Figure 4 but, beoause of the
increased supercharger boost, the curve reaches the set
point 134 before it coincides with the boost curve E,
As a result, the exhaust valve will not be opened, In-
stead, the pressure in the hydraulic system will be dumpedto a lower stable level determined by the characteristics
of the specific pressure relief system employed as des-
cribed above in connection with Figures 2 and 3~ The
total force on the pushrod is therefore shown by the
curve D' which, while necessarily in excess of the Curve C,
is still within the designed capacity of the pushrods, It
will be understood that the set point 134 is selected in
combination with the relative dimensions of the valve 106
and its bi-stable setting such that the resultant force
on the pushrods does not exceed a safe load.
From a consideration of Figure 5, it is apparent
that not only must the stable force on the pushrods be
limited, but the fugitive oscillations in this force should be
1~S652
-14-
dissipated before the injector load becomes operative.
Applicants have discovered that rapid damping of these
force oscillations can be accomplished by the special de-
signs of the pressure relief systems disclosed herein. In
Figure 3, the skewed seat 127 prevents the ball valve 108
from reseating prematurely as a result of such fugitive
oscillations. Similarly, a control of the ratio of the
area of the clearance space between the shoulder 118
and the piston ~ody ~2 and the area of the orîfice 100
as shown in Figure 2 is also effective to prevent pre-
mature reseating of the valve 106,
Figure 6 i5 a graph showing the effect o~ the
variations in the size of the fluid flow passages of the
configuration of Figure 2 on the performance of t~e pres-
sure relief system. For the tests represented by Figure6, a pressure relief system of the type shown in Figure
2 was employed wherein the area of the orifice laa (the
"primary" area) was 0.0102 square inches, Curve 136 shows
the performance of a pressure relief system whereîn the
area between the inner surface of the master piston body 92
and the shoulder 118 of the valve guide 112 ¢the "tertiary"
area) was 0.0676 square inches while Curve 138 shows the
improved performance resulting from a decrease in the size
of the orifice between the shoulder 118 and the inner sur-
face of the master piston body 92 ~the "tertiary" area)to 0.0146 square inches, Curve C of Figure 6 is identical
to Curve C of figures 4 and 5 and is reproduced for refer-
ence in the following discussion. Figure 6 shows two im^
provements which result from the change exemplified by
Curve 138; First, the pressure maintained in the system
after TDC was substantially lower and, second, the time
measured in crank angle degrees required to dump the
pressure, was substantially decreased. ~oth effects are
important: The first reduces the total maximum load on
the injector pushrod while the second tends to separate
the effect of the peak load required to open the exhaust
valve from the injector seating load.
~6~2
-15-
Applicants believe that when the area between the
shoulder 118 of the valve guide 112 and the inner wall of
the master piston 92, the "tertiary" area, is controlled
so as to be between about 100% and 150% of the size of the
orifice 100, the resistance to the flow of hydraulic fluid
is sufficient so that the pressure acts on the upper
surface of the valve guide 112 and quickly damps out the
vibratory motion of the valve guide 112 and the valve 106
resulting from the reaction of the spring 110. As a result,
the average opening and the average time in the open
position of the valve 106 are increased whereby the flow
through the valve 106 is maximized. Tests have shown that
when the ratio of the tertiary and primary areas exceeds
about 150% the damping effect on the normal vibratory
motion of the valye lQ6 and the valve guide 112 is diminished
and when the area ratio is below 100% secondary throttling
occurs which also restricts the flow of hydraulic fluid
through the piston 66.
Applicants believe that a similar damping pheno-
mena occurs in the pressure relief system shown in Figure3 although in that case it is believed that the damping
is a result of mechanical contact between the seat 127,
the ball 106 and the lower edge lQ8 of the cap 96.
By incorporating the pressure relief system
into the master piston as shown in Figures 2 and 3, appli-
cants provide a convenient mechanism whereby existing com-
pression relief engine brakes may be retrofitted to gain
the advantages of the present system at minimum cost.
It will also be noted that the hydraulic fluid which is
vented from the system is returned to the system without
the need for additional ducts or pumps siRce it is de-
livered to the pushrod area, an area where hydraulic
fluid is normally present,
However, the pressure relief system herein
contemplated may be placed at any point in the high pres~
sure hydraulic fluid circuit, for example, in ducts 40
or 46, While in such locations the dimensional limita-
1 ~5652~
-16-
tions presented by the master piston 66 are not present,
hydraulic fluid return ducts would be required, It will
be understood that if the pressure relief system of the
present invention were placed elsewhere in the high
pressure circuit, the body 92 or its equivalent would
be threaded or otherwise connected to the high pressure
circuit and the drainage passageways would be connected
to a hydraulic fluid return duct. In such a system, it
is apparent that either the three area pressure relief
lQ valve as shown in Figure 2 or the equivalent two area and
skew seat pressure relief valve of Figure 3 could be em-
ployed. However, because of the elimination of the dimen-
sional constraints, determined by the shape and size of
the master piston in such a modification, the coil spring
110 and the valve guide 122 of Figure 3 may be combined
in the form of an equivalent leaf spring having a ball
engaging surface of the shape and orientation of the ball
guide seat 127 and a spring rate equal to that of the
coil spring 110. Such a modification would operate in
a manner similar to the pressure relief system of Figure
3, as described herebefore but, as also noted, could
be located at any convenient point in the high pressure
hydraulic system.
The terms and expressions which have been em-
ployed are used as terms of description and not of
limitation and there is no intention in the use of
such terms and expressions of excluding any equivalents
of the features shown and described or portions thereof,
but it is recognized that various modifications are
possible within the scope of the invention claimed,
