CH641526A5 - COMBUSTION ENGINE WITH AN ENGINE BRAKING SYSTEM. - Google Patents

Description

The invention relates generally to an internal combustion engine having the features of the preamble of claim 1. Such an engine braking system opens the exhaust valves of the engine near the top of the compression stroke of the engine so that the energy absorbed by the engine during the compression stroke does not increase during the expansion stroke the engine is returned. The invention was particularly concerned with a timer mechanism for an engine brake of this type.

It has long been recognized that the conventional wheel brake mechanisms, which are designed as drum or disc brakes in commercially available vehicles, can absorb a large amount of energy during short periods of time, but are unable to increase the smaller amounts of energy during an extended period of time absorb, such as is required when driving downhill on a long slope. Under these circumstances, the friction material used in the brake mechanism becomes overheated (causing the brake fading, i.e. the braking effect is reduced) and can be destroyed, causing the metal parts to warp or buckle. This problem has usually been solved either by using a lower gear ratio so that the motor can act more effectively as a brake due to its own friction, or by using some kind of auxiliary braking system. A number of such auxiliary braking systems, commonly known as engine retarders, have been developed in the prior art.

developed, for example hydrogenetic retarders, exhaust brakes, electric brakes and engine brakes. In each of these systems, a portion of the vehicle's kinetic energy is converted to heat as a result of gas compression, fluid friction, or electrical resistance and then released to the atmosphere directly or through the exhaust or cooling system. The common characteristic of such auxiliary braking systems is the ability to absorb and dissipate certain amounts of energy continuously or at least for an unlimited period of time. The types of engine retarders described above are described in detail in the publication "Retarders for Commercial Vehicles", 1975, Mechanical Engineering Publications Limited, London, England.

The hydroconetic and electrical retarders are usually quite heavy and large because they require a turbine or dynamomechanism; this is undesirable from the point of view of production costs and operating costs. While the exhaust brake is generally simple and compact, it necessarily increases the exhaust manifold pressure and can therefore cause the engine's exhaust valves to "float".

It has already been recognized that in the normal operation of an internal combustion engine with, for example, the Otto or Diesel operating mode, a considerable amount of work is done during the compression stroke of the air introduced into the cylinders or of the air / fuel mixture. During the expansion or the power stroke of the engine, the work of the compression stroke is recovered, so that neglecting friction losses, the net energy due to compression and expansion is zero and the net energy output results from the combustion of the fuel / air mixture. If the throttle valve is closed or the fuel supply is interrupted, the engine will of course act as a brake to the extent of the friction occurring in the engine mechanism.

Many attempts have been made to increase the braking performance of an engine by converting the engine into an air compressor and releasing the compressed air through the exhaust system. A simple and practical method for achieving this function is described in US Pat. No. 3,220,392. An auxiliary exhaust valve actuation device synchronized with the crankshaft of the engine is provided, which opens the exhaust valve near the end of the compression stroke without disturbing the normal actuation cam device for the exhaust valve; In addition, a suitable control device for the auxiliary outlet valve actuating device is provided. While the engine brake described in detail in U.S. Patent 3,220,392 can produce a deceleration power that approximates the driving force of the engine under normal operating conditions, it has been found in practice that with this mechanism the deceleration effect is significantly increased by the timing of the Opening the engine exhaust valve can be affected.

If the exhaust valve is opened too late, a significant portion of the retarding performance may be lost due to the expansion of the compressed air during the initial portion of the expansion stroke. On the other hand, if the exhaust valve is opened too early, insufficient compression may occur during the compression stroke, so that the amount of deceleration power developed is also reduced.

The timing of the exhaust valve opening is significantly affected by the temperature conditions in

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the engine, which vary as a result of changes in environmental conditions as well as changes in operating conditions. For example, the length of the engine exhaust valve increases with increasing temperature, thereby reducing the play in the valve actuation mechanism. Although it is known to provide adjustable elements in the valve actuation mechanism with which the play can be adjusted (US Pat. No. 3,220,392), this play, which is determined by the rocker arm adjusting screw (or an equivalent element), must at least then be large enough be when the engine is cold so that there is some play when the engine is hot. If there is insufficient play when the engine is hot, the exhaust valve can be held in a partially open position. In this case, the operation of the engine may be adversely affected and the exhaust valves tend to burn. To avoid these phenomena, it is common to provide a clearance of the order of 0.4572 mm in the actuation mechanism for the exhaust (and intake) valves of internal combustion engines to compensate for the dimensional changes in the mechanism that result from temperature changes. When using the exhaust valve actuation mechanism as part of an engine braking mechanism, it is highly advantageous to minimize the backlash or idle in the valve actuation mechanism to allow precise control of the valve timing pump, thereby maximizing the deceleration performance of the engine can be.

To solve the problems explained above, an internal combustion engine with the features of claim 1 is proposed according to the invention.

This design of the brake system reduces the play between the first piston and the valve stem to a value that maximizes the power when the engine brake is actuated. By reducing the play in this way, the exhaust valve is opened earlier and the timing of the valve opening coincides more closely with the actuation of the engine brake master piston, so that the deceleration power of the engine is maximized.

The invention has the additional advantage that the maximum load on the push rod can be reduced. The load on the push rod is caused by the force required to open the exhaust valve against the pressure of the air compressed during the compression cycle and against the force required to actuate the fuel injector. By effectively reducing backlash as discussed above, the exhaust valve opening timing is advanced so that the time interval between the brake actuation load and the injector actuation load is increased, thereby minimizing the combined impact of these two events. In addition, advancing the timing of the exhaust valve opening can reduce the engine cylinder peak pressure, which in turn reduces the pushrod load.

The invention thus provides an engine brake system of the gas decompression type in which the timing of the system can be controlled in a targeted manner. The hydromechanical mechanism opens the exhaust valve near the top of the compression stroke of the engine so that the energy absorbed by the engine during the compression stroke is not returned to the engine during the expansion stroke. The hydromechanical devices provided in the exhaust valve actuation mechanism reduce the play in this mechanism to a value which maximizes the deceleration power developed by the engine when the engine brake is applied.

The invention is explained in more detail below, for example with reference to the drawing; it shows:

Fig. 1 is a schematic view of an engine brake with a time setting advance mechanism according to the invention;

FIG. 2 is an enlarged partial cross section of a portion of the engine brake mechanism shown in FIG. 1, the timing mechanism being shown in more detail;

Fig. 3 is a bottom view taken along line 3-3 of Fig. 2; and

4 shows a graphical illustration of a comparison of the braking capacity or the deceleration power of two motors with the time setting pre-run mechanism according to the invention and the values of the two same motors without the device according to the invention.

In Fig. 1, reference numeral 10 designates various fragmentary portions of the engine brake housing, while at 12 is a schematic view of the engine pan containing oil 14. The oil 14 can be obtained from the tub 12 through a line 16 by means of an oil pump 18 and then directed into a solenoid valve 20 via a line 22. The solenoid valve 20 comprises a valve body 24, which is fastened to the engine brake housing 10 and has an inlet opening 26, an outlet opening 28 and an outlet opening 30. The inlet opening 26 and the outlet opening 30 are each connected to the valve cavity 32 at its upper and lower ends, respectively, while the outlet opening 28 is connected to an enlarged central part of the valve chamber 32. A valve stem 34 is supported for reciprocation within the valve body 24 and carries a cylindrical valve seat 36 which is formed to seat on the shoulders which are formed by the enlarged central portion of the valve chamber 32. A spring 38 normally biases the valve stem 34 such that oil flow from the inlet port 26 to the outlet port 28 of the solenoid valve is prevented.

A solenoid 40 surrounding the upper end of the valve stem 34 is constructed such that it opens the solenoid valve 20 against the bias of the spring 38 whenever an electrical current flows through the coil. The solenoid circuit in series includes a fuel pump switch 42, a clutch switch 44, an instrument panel switch 46, a fuse 48 and the vehicle battery 50. The purpose of each of the logic switches 42, 44 and 46 is described below in connection with an explanation of the operation of the engine brake .

The drain opening 30 is connected to the engine pan 12 via a line 52, while the outlet opening 28 is connected to a control valve 54 via a line 56. The control valve 54 is generally in the form of a piston which is reciprocably mounted in a control valve cylinder 58 which is formed in the engine brake housing 10. The control valve includes an inlet port 60 that communicates with an outlet port 62. The control valve inlet port 60 is normally closed by a ball check valve 64 which is biased by a valve spring 66.

When the solenoid valve 20 is in the open position as shown in FIG. 1, oil 14 flows from the pan 12 through the solenoid valve 20 and the outlet line 56 to the inlet opening 60 of the control valve 54. The oil then lifts the control valve 54 against the bias a control valve spring 68 until the annular outlet opening 62 is aligned with the control valve cylinder outlet opening 70. Thereafter, the pressure of the oil opens the check valve 64 so that

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Oil can pass through the control valve 54 and into the line 72, which communicates with the outlet opening 70 of the control valve and the inlet opening 74 of the auxiliary cylinder or servo cylinder 76.

The auxiliary cylinder 76 is shaped in the engine brake housing 10 such that it is aligned with an exhaust valve 78 that is biased into a closed position by an exhaust valve spring 80. An auxiliary piston 82 is arranged to reciprocate within the cylinder 76. One end of the auxiliary piston 82 is formed to be in contact with the exhaust valve butt plate 84 while the opposite end of the auxiliary piston is in contact with an adjustable timer mechanism 86 which is in the engine brake housing 10 in alignment with the auxiliary piston 82 is screwed in and locked in position by a lock nut 88. An auxiliary piston return spring 77 is disposed within the auxiliary piston 82 such that one end of the spring biases the auxiliary piston up against the timer mechanism 86. The opposite end of spring 77 is carried by a bracket 79 which is inserted into the engine brake housing. The auxiliary cylinder 76 contains an outlet opening 90, which is connected via a line 96 to the inlet opening 92 of a master cylinder 94 formed in the engine brake housing 10. A master piston 98 is reciprocally supported in the master cylinder 94 and its outer end is adapted to abut the rocker arm adjusting screw 100 of the corresponding fuel injection or intake valve rocker arm 102, which in turn is actuated by the push rod 104. The main piston 98 is held in the housing bore by a light leaf spring 106.

An auxiliary piston 82 will usually be assigned to each exhaust valve, so that a six-cylinder engine will also have six auxiliary pistons, while a four-cylinder engine has four auxiliary pistons. In addition, each auxiliary piston is connected to a main piston associated with an intake or fuel injection rocker arm and a push rod. Of course, the main piston and its corresponding auxiliary piston can be assigned to different engine cylinders. An example linkage for this alternative is shown in Table 1 below for a six-cylinder engine:

Table 1

Location of the main piston

Position of the auxiliary piston

Push rod No. 1

Exhaust valve No. 3

Push rod No. 5

Exhaust valve No. 6

Push rod no.3

Exhaust valve No. 2

Push rod No. 6

Exhaust valve No. 4

Push rod no.2

Exhaust valve No. 1

Push rod no.4

Exhaust valve No. 5

When the solenoid valve 20 is open, oil 14 flows through the solenoid valve and the control valve 54 and fills the lines 72 and 96 as well as the auxiliary cylinder 76 and the master cylinder 94. The check valve 64 prevents the oil 14 from flowing backwards from the auxiliary and master cylinders. Thereafter, actuation of the push rod 104 moves the master piston 98 upward in the master cylinder 94 and causes the oil pressure to rise rapidly. The hydraulic pressure in the auxiliary cylinder 76 pushes the auxiliary piston 82 down so that the exhaust valve 78 is opened.

If the auxiliary piston 82 is not seated on the exhaust valve butt plate 84 when the main piston 98 begins to move, then the opening of the exhaust valve 78 will

Delayed by the time required to overcome the game or the gap in the system. However, it is necessary to accommodate dimensional changes in the mechanism such as the exhaust valve stem due to the temperature changes. In conventional devices, this play was controlled by an adjusting screw which is located at the position of the timing mechanism 86 and which has been set to a play of for example 0.46 mm when the machine is cold. In accordance with the invention, a timer mechanism 86 is provided which effectively maintains zero clearance in the exhaust valve actuation mechanism so that the exhaust valve begins to move as soon as the master piston begins to move when the brake mechanism is actuated.

As shown in FIGS. 2 and 3, the timing mechanism 86 includes an adjustable body 108, on the outer cylindrical surface of which a thread is formed which is screwed into the engine brake housing 10 in alignment with the auxiliary cylinder 76. Body 108 may be provided with a slot 110 or other suitable recess for ease of adjustment and may be locked in the desired position using a lock nut 88. Although the timer mechanism 86 may be mounted in other locations, such as between the auxiliary piston 82 and the exhaust valve butt plate 84 or within the auxiliary piston, it is preferred for convenience of adjustment to locate one end of the mechanism outside the exhaust valve mechanism.

A series of three coaxial bores 112, 114 and 116 are formed in the body 108 and extend partially through the body 108 from the end opposite the end containing the adjustment slot 110. The first and largest bore 112 extends approximately halfway through the body 108 and is configured to receive a timer piston 118. The central bore 114 extends approximately halfway through the remaining length of the body 108 and contains a compression spring 120. The third and smallest bore 116 extends slightly lower than the central bore 114 and forms a seat for the check valve spring 122. Of course, a single bore can with the diameter of bore 112 and the depth of bore 116 may be used in place of the three bores shown and described.

The timer piston 118 is formed with an axial bore 124 that extends completely through the piston 118. An enlarged bore 126 is formed at the inner or upper end of the piston to form a seat 128 for a ball check valve 130. Ball check valve 130 is normally pressed against seat 128 by compression spring 122. A transverse bore or slot 132 is formed across a diameter of the piston 118. The body 108 contains a transverse bore 134 into which a pin 136 is pressed. The bore or slot 132 is considerably wider than the pin 136 so that the piston 118 can move axially with respect to the body 108 within a limited range. In the present exemplary embodiment, the range of motion of the piston 118 extends from a first position somewhat within the body 108 to a second position in which the piston 118 extends slightly beyond the end of the body 108, for example by 0.4572 to 0.7112 mm . It should be noted that the compression spring 120 normally urges the piston 118 to its extended position while the lighter, i.e. weaker compression spring 122 urges the ball check valve into a closed position.

Although it is advantageous to use a ball check valve 130 and s

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Using a compression spring 122, other check valve devices can also be used. For example, a leaf spring or other type of check valve can be located either in the timing piston 118 or in a separate conduit that connects the auxiliary cylinder 76 and the area of the bores 112, 114, 116 above the timing mechanism piston 118. In the same way, devices other than the pin 136 and the slot 132 can be used to allow limited axial movement of the piston 118 within the adjustable body 108. Such alternative devices may include a reduced diameter at the lower end of piston 118 and a mating inward flange or lip at the lower end of adjustable body 108.

The operation of the mechanism will now be described.

First, the adjustable body 108, as with the known device discussed above, can be adjusted to set any desired play, for example 0.4572 mm, between the auxiliary piston 82 and the exhaust valve butt plate 84, thus ensuring that there is sufficient play under all operating conditions is present to prevent inadvertent partial opening or lifting of the exhaust valve 78. Under these conditions, the timing mechanism piston 118 according to the invention does not extend away from the body 108 and the body 108 is in direct contact with the upper part of the auxiliary piston 82. When it is desired to apply the engine brake, the solenoid valves 20 and the control valve 54 is actuated. This results in a flow of oil 14 through lines 72 and 76 and into auxiliary cylinder 76 and master cylinder 94. When the main piston 98 begins to move, pressure is immediately built up in the hydraulic circuit since this circuit has been completely filled with oil. Movement of the master piston 98 therefore immediately results in movement of the auxiliary piston 82 and, as in conventional devices, after overcoming the play in the mechanism, opens the exhaust valve to which the auxiliary piston is assigned.

As the auxiliary piston 82 moves away from the timing mechanism body 108, according to the invention, the timing mechanism piston 118, controlled by the pin 136, extends a predetermined distance away from the timing mechanism body due to the force of the compression spring 120. This causes a pressure difference generated that is large enough to lift the ball check valve 130 from its seat and let oil into the area of the bores 112, 114 and 116. Upon return of the auxiliary piston 82 to its initial position under the action of the spring 77, it strikes the timing mechanism piston 118. The oil that has penetrated into the timing mechanism through the check valve 130 and is relatively incompressible acts on the auxiliary piston 82 the auxiliary return spring 77 counteracted force. In this way, the auxiliary piston 82 assumes a new starting position for all subsequent operating cycles, controlled by the predetermined stroke of the timer mechanism piston 118 and only variable by the movement of the timer mechanism piston 118 due to leakage due to the play between the piston and the piston Drilling between operating cycles. This leakage loss is replaced by ball check valve 130 during each cycle. When the engine brake solenoid valve 20 and the control valve 54 are turned off, the hydraulic circuit for drainage is vented. While the cyclical movement of the auxiliary piston 82 ends and it comes to rest on the timer mechanism piston 118, leakage by the timer enables

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Mechanism piston, due to the bore play, a full retraction of the timer mechanism piston and a return of the auxiliary piston to its original position on the adjusting screw body 108.

The result of effectively eliminating the slack or the valve actuator mechanism is shown in FIG. 4. 4 shows a graph of the relationship between engine speed and absorbed or braking power for a six-cylinder and a four-cylinder engine with and without the timer mechanism of the invention. Curve 138 represents the braking power in horsepower which is achieved with a six-cylinder diesel engine which is equipped with a Jacobs engine brake and in which the play setting has been adjusted to 0.4572 mm in accordance with the prior art. Curve 140 shows the horsepower in horsepower that was included from the same engine but with the lash adjuster replaced with a timing mechanism in accordance with the invention. Similarly, curve 142 shows the horsepower braking power developed by a four-cylinder diesel engine equipped with a standard Jacobs engine brake, such as described in U.S. Patent 3,220,992, during the curve 144 shows the effect which is achieved when a conventional adjusting screw set with a clearance of 0.4572 mm is replaced by a timer advance mechanism according to the invention. It can be seen that at normal engine operating speeds in the 2000 rpm range, a very significant increase in decelerating horsepower is achieved when the invention is applied.

As indicated schematically in Fig. 1, the engine brake according to the invention is actuated by a solenoid valve which is wired in series with three switches, namely with a fuel pump switch 42, a clutch switch 44 and a dashboard switch 46. If one or if several of these switches are in the open position, the brake cannot be actuated. The fuel pump switch 42 turns off the braking system whenever the engine is fueled, i.e. whenever the accelerator pedal is pressed. The clutch switch 44 opens whenever the clutch is disengaged to prevent the engine stalling. The dashboard switch 46 provides manual control to allow the operator to turn off the braking system if desired. The dashboard switch 46 may also be a multi-position switch that shuts down part of the system so that the operator can select a partial or full braking power as desired under certain operating conditions.

In addition to the basic advantage of considerably increasing the braking power of the motor, as shown in FIG. 4, the timer advance mechanism according to the invention can also be retrofitted in motors which are equipped with motor brakes of the type described without requiring any modification of the motor would. A second advantage of the timer advance mechanism described is a decrease in the push rod load when the device in an engine is provided with mechanical fuel splashes which are actuated with push rods. This advantage results from the increase in the time period between the opening of the exhaust valve and the actuation of the fuel injector. It was found that for a decrease in the play by 0.0254 mm, the push rod load in engines of the type tested for FIG. 4 is reduced by approximately 22.68 kg.

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2 sheets of drawings

Claims (4)

641 526 PATENT CLAIMS 1. internal combustion engine having a gas decompression type engine braking system having exhaust valve means and an adjustable body engageable with a first piston for positioning it to move the exhaust valve means at a selected predetermined time after the hydraulic pressure is placed on the first piston to open, characterized in that a hollow second piston (118) is reciprocable within the body (108) between a retracted and extended position, that the second piston in the absence of hydraulic fluid pressure is forced into its retracted position by the first piston (82) against the counterforce of a spring (120), and that a check valve device (130) is provided which is operatively associated with the second piston (118) around the second piston hold in its extended position so that the return of the first piston to engage the high len body is prevented while the hydraulic fluid pressure is applied to the first piston. 2. Internal combustion engine according to claim 1, characterized in that locking devices (88) are provided for the hollow body. 3. Internal combustion engine according to claim 1 or 2, characterized in that the check valve device comprises a ball check valve that an axial line (124) is formed through the second piston, that at one end of this line a valve seat (128) for the check valve device is formed, and that spring means (122) are provided which bias the ball check valve against the valve seat. 4. Internal combustion engine according to one of the preceding claims, characterized in that an adjustable play is formed between the second piston and the adjacent walls of the hollow body.