CA1156526A - Engine braking apparatus - Google Patents
Engine braking apparatusInfo
- Publication number
- CA1156526A CA1156526A CA000364087A CA364087A CA1156526A CA 1156526 A CA1156526 A CA 1156526A CA 000364087 A CA000364087 A CA 000364087A CA 364087 A CA364087 A CA 364087A CA 1156526 A CA1156526 A CA 1156526A
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- valve
- stable
- guide
- spring
- area
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Abstract
ENGINE BRAKING APPARATUS
ABSTRACT
Pressure relief apparatus for an internal combustion engine compression relief engine brake is dis-closed. Normally an existing pushrod is used to drive a master piston which in turn controls the motion of a slave piston which opens an engine exhaust valve. The problem of assigning such additional function to the pushrod is that an increased load may be experienced which may exceed the design capacity of the pushrod. The invention solves this problem by providing a bi-stable relief valve located in a high pressure hydraulic fluid circuit which interconnects the slave and master pistons together with damping means adapted to damp out rapidly the oscillations of the valve during the period of its opening so as to maximize the flow of hydraulic fluid through the bi-stable valve and minimize the time required to relieve the pressure in the high pressure hydraulic system of the compression relief engine brake. The damping means comprise a spring controlled valve guide which inhibits premature reseating of the bi-stable valve and maximizes the average opening of the valve during its operating period.
ABSTRACT
Pressure relief apparatus for an internal combustion engine compression relief engine brake is dis-closed. Normally an existing pushrod is used to drive a master piston which in turn controls the motion of a slave piston which opens an engine exhaust valve. The problem of assigning such additional function to the pushrod is that an increased load may be experienced which may exceed the design capacity of the pushrod. The invention solves this problem by providing a bi-stable relief valve located in a high pressure hydraulic fluid circuit which interconnects the slave and master pistons together with damping means adapted to damp out rapidly the oscillations of the valve during the period of its opening so as to maximize the flow of hydraulic fluid through the bi-stable valve and minimize the time required to relieve the pressure in the high pressure hydraulic system of the compression relief engine brake. The damping means comprise a spring controlled valve guide which inhibits premature reseating of the bi-stable valve and maximizes the average opening of the valve during its operating period.
Description
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 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;
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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 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 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 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~
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 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 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 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
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 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;
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 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 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 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~
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 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 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 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 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~
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
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 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~
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,
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,
Claims (11)
PROPERTY OR PRIVILEGE IS CLAIMED ARE DEFINED AS FOLLOWS:
1. Engine braking apparatus of a gas compression release type including an internal combustion engine having exhaust valve means and pushrod means, hydraulically actuated first piston means for opening said exhaust valve means at a predetermined time, and further piston means actuated by said pushrod means and hydraulically inter-connected with said exhaust valve opening piston means in a high pressure hydraulic fluid circuit, characterized by a bi-stable valve located in said high pressure hydraulic fluid circuit and having at least primary and secondary orifices, and damping means associated with said valve to rapidly damp out vibrations of said valve while it is moving from its closed position defining a high pressure condition, until it comes to rest in an open position defining a low pressure condition, said primary and secondary orifices and damping means maximizing the flow through said valve and minimizing the the required to attain said low pressure condition.
2. The apparatus of claim 1, wherein said damping means comprises a valve guide located within said further piston means, a spring located within said further piston means to bias said valve guide against said valve and urge said valve to a normally closed position, and a hydraulic fluid drainage passageway in said further piston means.
3. The apparatus of claim 2, wherein said valve guide has a guide seat portion the diameter of which is smaller than the inside diameter of said further piston means thereby defining a tertiary area at least equal to the area of the primary orifice of said bi-stable valve,
4. The apparatus of claim 3, wherein said tertiary area is at least equal to the area of the primary orifice of said bi-stable valve but less than about 150% of the area of said primary orifice.
5. The apparatus of claim 1 or 2, wherein said damping means comprises a ball valve guide having a ball valve seat skewed with respect to the axis of said ball valve guide whereby said ball valve may be displaced from the axis of said ball valve guide when said ball valve is opened.
6. The apparatus of any one of claims 2, 3 or 4 wherein the area of said hydraulic fluid drainage passageway is at least equal to the area of said primary valve orifice.
7. The apparatus of claim 1, wherein the position of the bi-stable valve may be varied relative to the bottom surface of said further piston means whereby the bias between the valve and said valve guide induced by said spring may be varied.
8. The apparatus of claim 2, wherein the position of the bi-stable valve may be varied relative to the bottom surface of said further piston means whereby the bias between the valve and said valve guide induced by said spring may be varied.
9. The apparatus of claim 3, wherein the position of the bi-stable valve may be varied relative to the bottom surface of said further piston means whereby the bias between the valve and guide induced by said spring may be varied.
10. The apparatus of claims 7, 8 or 9 wherein said spring is a coil spring and the maximum travel of said valve guide is less than the maximum compression of said coil spring.
11. A modification of the apparatus of claim 1, wherein said damping means comprises a valve guide and spring combined in the form of a leaf spring having a ball engaging surface which is skewed with respect to the axis of a bi-stable ball valve whereby the point of contact between said leaf spring member and said ball valve is displaced from the axis of said bi-stable ball valve.
Priority Applications (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CA000364087A CA1156526A (en) | 1980-11-06 | 1980-11-06 | Engine braking apparatus |
Applications Claiming Priority (1)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| CA000364087A CA1156526A (en) | 1980-11-06 | 1980-11-06 | Engine braking apparatus |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| CA1156526A true CA1156526A (en) | 1983-11-08 |
Family
ID=4118353
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA000364087A Expired CA1156526A (en) | 1980-11-06 | 1980-11-06 | Engine braking apparatus |
Country Status (1)
| Country | Link |
|---|---|
| CA (1) | CA1156526A (en) |
Cited By (1)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| CN107939472A (en) * | 2017-10-17 | 2018-04-20 | 浙江大学 | Two cycle compression release type brake device of integrated engine and its braking method |
-
1980
- 1980-11-06 CA CA000364087A patent/CA1156526A/en not_active Expired
Cited By (2)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| CN107939472A (en) * | 2017-10-17 | 2018-04-20 | 浙江大学 | Two cycle compression release type brake device of integrated engine and its braking method |
| CN107939472B (en) * | 2017-10-17 | 2023-10-27 | 浙江大学 | Two-stroke compression release type braking device of integrated engine and braking method thereof |
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