WO2014185972A2

WO2014185972A2 - Variable control engine decompression brake

Info

Publication number: WO2014185972A2
Application number: PCT/US2014/000109
Authority: WO
Inventors: Joseph Reilly
Original assignee: Parker-Hannifin Corporation
Priority date: 2013-05-14
Filing date: 2014-05-14
Publication date: 2014-11-20
Also published as: WO2014185972A3

Abstract

A variable decompression engine brake system includes a master piston assembly that engages an engine prime drive mechanism, and a slave piston assembly including a piston that is moveable within a cylinder, the slave piston assembly being hydraulically connected to the master piston assembly and configured to operate at least one exhaust valve. A control valve has a differential area that sets a control pressure, and the slave piston assembly is configured to actuate the exhaust valves when a pressure within the cylinder reaches the control pressure. The brake system further includes a command solenoid coil that is energized to set a command pressure, and the differential converts the command pressure proportionally to the control pressure for opening of the exhaust valves. The slave piston may reach the control pressure when or after the piston has reached a top dead center position.

Description

TITLE: VARIABLE CONTROL ENGINE DECOMPRESSION BRAKE

Related Applications

This application claims the benefit of U.S. Provisional Application No.

61/823,320, filed May 14, 2013, which is incorporated herein by reference.

Field of the Invention

The present invention relates generally to engine brakes, and more

particularly to variable control engine decompression brakes.

Background of the invention

Engine decompression brakes commonly are utilized in large vehicle engines, such as in commercial trucking engines. The use of the decompression engine braking system reduces usage of the regular service brakes, which extends the life of the service brakes.

During decompression engine braking, the full engine cycle is turned off to cut fuel to the engine and an engine decompression braking system is activated. The engine decompression braking system may be activated by a driver initially applying force to the service brakes, with the driver force being converted into a proper force level of the engine brake. In conventional configurations of decompression engine braking systems, a master piston follows the prime drive mechanism of the engine. The master piston further is hydraulically connected to a working piston that drives decompression engine braking. In particular, the working piston (also referred to as a slave piston) telescopes outward to actuate a rocker arm that opens one or more exhaust valves to release the pressure inside the working piston. By releasing the pressure in the working piston with the full engine cycling off, the engine itself acts as a brake to reduce the usage of the service brakes.

Further in conventional systems, the exhaust valves are opened to release the pressure in the working piston cylinder when the working piston essentially is at top dead center (TDC). Full pressure release at TDC corresponds to the maximum or 100 percent braking force of the decompression engine brake. In most circumstances, however, full braking force is undesirable, so the decompression braking system employs a variable control mechanism. Conventional variable decompression engine brakes utilize a back pressure device in the exhaust manifold to induce a variation in the braking force. Such a conventional back pressure device always operates after the pressure is fully released while working piston is at TDC. The back pressure device returns force back to the engine crankshaft. This varies the net braking force of the decompression engine brake to a desirable level based on the force applied by the driver to the service brakes.

Accordingly, in conventional systems back pressure is applied by the exhaust manifold only after the pressure release at TDC of the working piston. The use of a back pressure device in the exhaust manifold separate from the working piston has disadvantages by virtue of the additional components.

Summary of the Invention

The present invention provides improved systems and methods for operating a variable decompression engine brake system. An enhanced variable

decompression engine brake utilizes a variable control pressure in the working piston cylinder to control the opening of the engine exhaust valve, varying the reactive braking force or effectiveness of the engine decompression brake. A differential area is built into a control valve of the decompression brake system. With such differential, the control valve allows a much lower solenoid command pressure to proportionally control the pressure in the piston chambers of the decompression brake at which the exhaust valves open, thus controlling the opening of the exhaust valves during the decompression brake cycle. With the ability to pressure control the opening of the exhaust valves after top dead center (TDC), the cylinder pressure of the engine is more directly controlled. This configuration obviates the use of a separate back control device in the exhaust manifold, as utilized in conventional configurations.

An aspect of the invention, therefore, is a variable decompression engine brake system. In exemplary embodiments, the brake system includes a master piston assembly configured to engage a prime drive mechanism, and a slave piston assembly including a piston that is moveable within a cylinder. The slave piston assembly may be hydraulically connected to the master piston assembly and configured to operate at least one exhaust valve. A control valve has a differential area built into the control valve that sets a control pressure, and the slave piston assembly is configured to actuate the at least one exhaust valve when a pressure within the cylinder reaches the control pressure to perform decompression engine braking.

The brake system may include a command solenoid coil that is energized to set a command pressure, and the differential converts the command pressure proportionally to the control pressure for opening of the at least one exhaust valve. The slave piston assembly may be configured to reach the control pressure when or after the piston has reached a top dead center position. The slave piston assembly may be configured to reach the control pressure when the piston has reached the top dead center position to achieve full braking force. In addition, the slave piston assembly may be configured to reach the control pressure after the piston has - rebounded from the top dead center position, and a reduction of braking force from full braking force is proportional to an amount of rebound from the top dead center position.

Another aspect of the invention is a control method for controlling a variable decompression engine brake system. In exemplary embodiments, the control method includes cutting off fuel to an engine; applying a duty cycle current to a command solenoid coil thereby allowing a pressure to act on a control valve to set a control pressure; shifting the control valve to provide an oil volume to activate a decompression brake piston within a cylinder; advancing the piston to a top dead center position within the cylinder, thereby building up pressure in a piston chamber; and opening at least one exhaust valve when a pressure within the cylinder reaches a control pressure to perform decompression engine braking. The control valve may have a differential area, and the differential converts the command pressure proportionally to the control pressure for opening of the at least one exhaust valve. The control pressure in the cylinder may be reached when or after the piston has reached a top dead center position.

These and further features of the present invention will be apparent with reference to the following description and attached drawings. In the description and drawings, particular embodiments of the invention have been disclosed in detail as being indicative of some of the ways in which the principles of the invention may be employed, but it is understood that the invention is not limited correspondingly in scope. Rather, the invention includes all changes, modifications and equivalents coming within the spirit and terms of the claims appended hereto. Features that are described and/or illustrated with respect to one embodiment may be used in the same way or in a similar way in one or more other embodiments and/or in

combination with or instead of the features of the other embodiments.

Brief Description of the Drawings

Fig. 1 is drawing depicting a perspective view of an exemplary variable control decompression brake system in accordance with embodiments of the present invention.

Fig. 2 is a schematic diagram depicting the variable control decompression brake system, with the brake system in a neutral condition.

Fig. 3 is another schematic diagram depicting the variable control

decompression brake system, with the brake system in a rebound mode for varying the decompression braking force.

Fig. 4 is a flowchart diagram depicting an exemplary method of controlling braking force in a variable decompression brake system in accordance with embodiments of the present invention.

Fig. 5 is a drawing depicting a partial cross-sectional view of an exemplary working piston assembly with a piston in a first retracted position prior to the compression stroke.

Fig. 6 is a drawing depicting a partial cross-sectional view of the exemplary working piston assembly with the piston in a second position essentially at top dead center in a full braking force mode.

Fig. 7 is a drawing depicting a partial cross-sectional view of the exemplary working piston assembly with the piston in the second position essentially at top dead center in a rebound mode for variably controlling the braking force. Fig. 8 is a drawing depicting a partial cross-sectional view of the exemplary working piston assembly with the piston in a third position passed top dead center in the rebound mode corresponding to eighty percent braking force.

Fig. 9 is a drawing depicting a partial cross-sectional view of the exemplary working piston assembly with the piston in a fourth position passed top dead center in the rebound mode corresponding to fifty percent braking force.

Fig. 10 is a graph depicting an exemplary pressure profile for an engine brake in accordance with embodiments of the present invention.

Detailed Description

Embodiments of the present invention will now be described with reference to the drawings, wherein like reference numerals are used to refer to like elements throughout. It will be understood that the figures are not necessarily to scale.

Fig. 1 is drawing depicting a perspective view of an exemplary variable control decompression brake system 10 in accordance with embodiments of the present invention. The variable control decompression brake system 10 includes a working piston 12 (also referred to as a slave piston) that telescopes out and actuates a rocker arm 14. The rocker arm 14 in turn actuates at least one exhaust valve (not shown in Fig. 1) for releasing the pressure from the working piston cylinder during decompression engine braking. The working piston 12 is hydraulically connected with a master piston 16 through a lobe 20. The master piston is configured to engage a prime drive mechanism. Specifically, the master piston 16 includes a cam follower 18 that interacts with a dedicated cam lobe that is driven by operation of the prime drive mechanism of the engine.

The brake system 10 further includes a hydraulic control valve 22 that utilizes a control valve spool 24 to control pressures within the brake system 10 as further detailed below. The hydraulic control valve 22 includes a relief rod 26 associated with a spring 28. A tank port 30 provides for an inlet of hydraulic fluid for operation of the hydraulic control valve 22. The working piston 12, master piston 16, and hydraulic control valve 22 may be mounted within a brake housing 32. A differential area is built into the hydraulic control valve 22. A command solenoid coil (not shown specifically in Fig. 1 ) is energized to generate a command pressure at a command port 34. The differential allows a relatively low solenoid command pressure at the command port 34 to proportionally control the pressure in the piston chambers as measured at a chamber pressure port 36 of the

decompression brake system. The pressure at the chamber pressure port 36 is determinative of the maximum pressure that the working piston cylinder may achieve, upon which the opening of the exhaust valves occurs during the

decompression brake cycle. In other words, the pressure for opening the exhaust valves of the decompression engine brake system is directly controlled.

Operation of engine braking generally occurs as follows. Once activated by the proportional solenoid valve, the slave piston with cam follower extends to make contact with the dedicated cam lobe. The cam profile begins to raise the applied hydraulic pressure to the slave piston via the cam follower-master piston. The control valve regulates the maximum pressure desired for the selected mode of braking operation. This applied pressure to the slave piston actuates the rocker arm that strategically opens and closes the exhaust valves during compression stroke at or after top dead center (TDC), depending upon the desired braking force selected. The exhaust valve opens when the slave piston's regulated pressure generates enough force to overcome the opposing force generated by the cylinder's

compression pressure acting on the exhaust valve's face or closing area. A functional relationship is created between the cylinder's pressure and the engine brake control pressure. Extra compression loading creates the engine braking effect, and maximum braking effect is achieved when the exhaust valves open when the slave position reaches TDC. When maximum braking force is not desired, the braking force can be proportionally reduced by reducing the control pressure at which the exhaust valve will open after the slave piston rebounds from TDC.

With the ability to directly control the pressure at which the exhaust valve opens, the cylinder pressure of the working cylinder and thus the decompression braking force are more directly controlled. The controlled decompression pressure in the working cylinder after top dead center is directly related to the magnitude of force returned back to the engine crankshaft. By allowing the decompression brake to proportionately return a portion of the energy back to the engine's crankshaft, the net variable engine decompression brake effect is increased with lower cylinder pressure or decreased with higher cylinder pressure after TDC.

The brake system may include a command solenoid coil that is energized to set a command pressure, and the differential converts the command pressure proportionally to the control pressure for opening of the at least one exhaust valve. The slave piston assembly may be configured to reach the control pressure when or after the piston has reached a top dead center position. The slave piston assembly may be configured to reach the control pressure when the piston has reached the top dead center position to achieve full braking force. In addition, the slave piston assembly may be configured to reach the control pressure after the piston has rebounded from the top dead center position, and a reduction of braking force from full braking force is proportional to an amount of rebound from the top dead center position.

Fig. 2 is a schematic illustration of the variable control decompression brake system, with the brake system in a neutral condition. Fig. 3 is another schematic illustration of the variable control decompression brake system, with the brake system in a rebound mode for varying the decompression braking force. Like components in Figs. 2 and 3 are identified with comparable reference numerals as in Fig. 1.

As seen in Figs. 2 and 3, a pressure control system 40 is operable to control pressures through the decompression brake system 10. The pressure control system has an inlet pressure at inlet 42. Such pressure at inlet 42 is derived from the engine lubrication system and is typically held constant. The pressure at inlet 42 aids in controlling the hydraulic fluid flow into the hydraulic control valve 22, and otherwise provides a base source for generating the pressure at the command port 34 of the control valve 22. In the example of Figs. 2 and 3, the pressure at inlet 42 is set to 120 psi, although this pressure may be varied and typically may range from approximately 100-150 psi. The pressure control system 40 further may include a command solenoid coil 44, which controls the pressure at the command port 34 of the hydraulic control valve 22. The control of such pressure results in variable control of the decompression brake force as described below.

As referenced above, Fig. 2 illustrates a neutral condition of the

decompression engine brake system 10. In such state, the command solenoid coil 44 is de-energized, which means none of pressure at inlet 42 is transferred from the pressure control system 40. As a result, there is no command pressure at the command port 34 of the hydraulic control valve 22. Accordingly, the control valve spool 24 is not actuated and there is no force to overcome the spring bias of the spring 28 to shift the control valve 22. The piston chamber pressure 36 of the decompression brake is therefore vented and inactive. In this mode, a normal running combustion cycle can be achieved, and the engine brake is off with both master and slave pistons being retracted and not active.

The decompression brake system operates as follows. When the driver initiates braking, fuel is cut off to the engine. A duty cycle current is applied to the command solenoid coil 44. The energizing of the command solenoid 44 causes the pressure to rise at the command port 34 of the hydraulic control valve 22,

proportionately with the current duty cycle. For example, at a current level for maximum duty cycle, maximum pressure results at the command port 34, and at 80% duty cycle the pressure is 80% maximum, at 50% duty cycle the pressure is 50% maximum, and so forth. In this manner, the duty cycle current is proportional to the pressure established at the command port 34. The resultant pressure at the command port 34 operates to shift the control valve 22 fully and provide an oil volume to activate the master piston 16 and working piston 12 of the decompression engine brake system 10. The pistons are actuated by virtue of a rise in pressure as measured at the chamber pressure port 36. For example, one hundred percent current permits one hundred percent maximum pressure in the piston chambers which allows full actuation, and thus full pressure in the engine cylinder is absorbed resulting in braking force at one hundred percent. As the current applied to the command solenoid coil 44 is reduced, a proportionate reduction in braking force results.

As referenced above, a differential area, denoted in Figs. 2 and 3 by reference numeral 46, is built into the hydraulic control valve 22 allowing a relatively low command solenoid control pressure at the command port 34 to proportionally control the maximum pressure in the piston chambers after which the exhaust valves are opened, as measured at a chamber pressure port 36 of the decompression brake system. In exemplary embodiments, the pressure ratio of the differential is between forty and fifty.

The operation of the decompression brake system may be understood with reference to the following numerical example. It will be appreciated that the precise numerical values may be varied, and the following values are provided for illustrative purposes. In this example, the pressure at inlet 42 is set to 120 psi and held constant. The command solenoid coil 44 may be energized such that at maximum duty cycle, the pressure at the command port 34 is 50 psi. The ratio across the differential 46 is deemed in this example to be forty. Accordingly, at maximum duty cycle current, corresponding to the command port pressure of 50 psi, the maximum piston chamber pressure is 2000 psi.

In a case of maximum braking force, the pneumatic pressure of compression building in the master and working cylinders before TDC is enough to cause a hydraulic pressure to build in the piston chambers of the decompression brake to the maximum of 2000 psi. For maximum braking force, the pressure in the piston chamber acting on the differential area of the control valve must reach the maximum 2000 psi to shift the decompression brake into a release state in which the exhaust valves open, thereby releasing the pressure that has built up within the working piston chamber. Accordingly, in the case of maximum braking force, the exhaust valves are allowed to open essentially upon reaching TDC of the working piston, allowing all pneumatic pressure built during the compression stroke to be released into the exhaust manifold. No rebound energy will be returned to the engine's crankshaft, and maximum engine braking effect is achieved. The differential area 46 aids in regulating the maximum pressure 36 permitted in the working cylinder chamber upon which opening the exhaust valves is permitted. Fig. 3 particularly depicts the brake system in a rebound mode for varying the decompression braking, in which the exhausts valves are permitted to open after the working piston has rebounded from TDC. A proportionally controlled duty cycle current is now applied to the command solenoid coil 44, which reduces the pressure at the command port 34 sufficient to shift the control valve 22 fully and provide the oil volume to activate the decompression brake system. In this specific numerical example of Fig. 3, the reduced pressure at the command port 34 is now 40 psi (reduced from the 50 psi maximum), resulting in a proportionately reduced piston chamber pressure of 1600 psi based on the differential ratio of forty. As 1600 psi represents a reduction to eighty percent of the 2000 psi pressure associated with full braking force, Fig. 3 illustrates the decompression engine brake system as set to provide eighty percent braking force.

The pneumatic pressure of compression building in the engine's cylinders before TDC is high enough to cause a hydraulic pressure to build in the piston chambers of the decompression brake. That pressure again acts on the differential area of the control valve and shifts the control valve into a rebound mode, biased by the command solenoid pressure at command port 34 acting on the control valve spool 24. While the control valve 22 is in the relieving mode, the engine's exhaust valves cannot open against the pneumatic pressure of compression in the engine's cylinder. Any pneumatic pressure in the engine's cylinder after TDC will return energy back to the engine's crankshaft. The pneumatic pressure decreases due to the expanding volume in the working piston cylinder as the piston retracts after TDC. A point in time, measured for example in crankshaft degrees, will be reached in which force on the exhaust valves developed by the pneumatic pressure in the engine cylinder is insufficient to hold the exhaust valves closed against the hydraulic force produced by the decompression brake. When that force balance is reached, the exhaust valves will open relieving the pneumatic cylinder pressure in the working cylinder, and no further energy will be returned to the engine's crankshaft.

In other words, in the relieving mode the pressure at TDC does not cause the exhaust valves to open. Rather, the building pressure causes the working piston to retract after TDC, resulting in decreasing pressure in the working cylinder. When the working cylinder pressure reaches the pressure set by virtue of the energized solenoid command pressure multiplied through the differential, the exhaust valves then open. In this numerical example with a 2000 psi chamber pressure

representing maximum braking energy, Fig. 3 depicts an eighty percent braking force. The working piston will retract to eighty percent TDC corresponding to a proportionate chamber pressure of 1600 psi, at which point the exhaust values open to provide eighty percent braking power.

Other proportionate braking forces may be achieved in similar fashion. For example, if fifty percent braking force is desired, the command solenoid coil 44 may be energized at fifty percent duty cycle. This results in a command port pressure 34 of 25 psi, and a maximum chamber pressure of 1000 psi with the differential ratio of forty. The working piston will retract to fifty percent of TDC corresponding to a proportionate chamber pressure of 1000 psi, at which point the exhaust values open to provide 50 percent braking force. Other levels of braking force may be achieved in similar fashion. By proportionally controlling the maximum pressure in the working cylinder after TDC for opening the exhaust valves, the control of force returned back to the engine crankshaft is achieved. By allowing the decompression engine brake to proportionately return a portion of the energy back to the engine's crankshaft, the net variable engine decompression brake effect is realized. The longer the delay in opening the exhaust valves after TDC, the more energy is rebounded and the braking energy is proportionately reduced.

Fig. 4 is a flowchart diagram depicting an exemplary method of controlling braking force in a variable engine decompression brake system in accordance with embodiments of the present invention. Although the exemplary method is described as a specific order of executing functional logic steps, the order of executing the steps may be changed relative to the order described. Also, two or more steps described in succession may be executed concurrently or with partial concurrence. It is understood that all such variations are within the scope of the present invention.

The control method may begin at step 100, in which fuel is cut off to the engine and the decompression braking system is actuated. At step 1 10, a duty cycle current is applied to a command solenoid coil, thereby allowing a pressure to act on a control valve. At step 120, the control valve shifts to provide an oil volume to activate the decompression brake working piston. At step 130, the working piston advances to a top dead center position, thereby building up pressure in the cylinder. The opening of the exhaust valves is based on when the pressure in the working cylinder has reached the control pressure resulting from the current applied to the command solenoid. This is shown initially at step 140, which illustrates whether the condition is met that the control pressure is high enough to open the exhaust valves against the cylinder pressure. It will be appreciated that whether the condition is met is not an active decision being made by the system, but rather is a state the occurs passively by the action of the working cylinder.

If the control pressure of the working cylinder is high enough to overcome the cylinder pressure at step 140, the method proceeds to step 160 and the exhaust valves are opened and the cylinder pressure is relieved. For example, if the system is controlled for full braking force, the control pressure essentially corresponds to the TDC position of the working cylinder and the exhaust valves will open at TDC. If, however, control pressure of the working cylinder is not high enough to overcome the cylinder pressure at step 140, the method proceeds to step 150 at which the working cylinder rebounds. As illustrated by the loop of steps 140 and 150, the working piston will continue to rebound until the control pressure is reached, at which time the method now proceeds to step 160 and the exhaust valves are opened.

Accordingly, following this method, when or after the working piston reaches top dead center, at least one exhaust valve opens when the pressure in the cylinder reaches the control pressure for opening the exhaust valve. In exemplary

embodiments, the control valve has a differential area as described above, and the differential area allows the solenoid command pressure to proportionally generate the control pressure at which pressure in the piston cylinder permits opening the at least one exhaust valve.

In this manner, variable decompression engine braking may be achieved. For example, when 100 percent braking force is desired, the command solenoid coil may be energized at full duty cycle, resulting in the control pressure being achieved essentially when the working piston reaches top dead center. Braking force may be proportionately reduced by reducing the magnitude of the current supplied to the command solenoid coil. In such cases, the control pressure is proportionately reduced. After reaching top dead center, the working piston rebounds, which reduces the pressure in the cylinder. The at least one exhaust valve opens when the cylinder pressure reaches the control pressure based on the magnitude of the energizing current supplied to the command solenoid coil.

Figs. 5-9 are drawings depicting a partial cross-sectional view of an

exemplary working piston assembly 60 in various positions corresponding to different modes of the decompression engine braking system. The working piston assembly 60 includes a moveable piston formed of a piston head 62 attached to a piston rod 64, which are moveable within a cylinder 66. The working piston assembly may be positioned within a housing 68 comparable to the housing identified with respect to Fig. 1 that houses the components of the decompression engine braking system. The cylinder 66 may be in fluid communication with an exhaust flow path 70 that is part of an exhaust manifold 71. Flow from the cylinder 66 into the exhaust flow path 70 may be controlled with at least one exhaust valve 72 that opens to exhaust the pressure from the working piston assembly in the manner described above. In the exemplary embodiment of Figs. 5-9, two exhaust valves are shown, although the precise number of exhaust valves may be varied.

Fig. 5 is a drawing depicting a partial cross-sectional view of the exemplary working piston assembly 60 with the piston in a first retracted position prior to the compression stroke. Fig. 5 illustrates that at the beginning of a compression stroke, the exhaust valves 72 may open briefly to admit high-pressure gases from the exhaust manifold into the piston cylinder 66. This flow of gases is indicated by the arrows in Fig. 5. The piston translates towards TDC increasing the pressure in the cylinder 66 until the position of Fig. 6 is reached.

Fig. 6 is a drawing depicting a partial cross-sectional view of the exemplary working piston assembly with the piston in a second position essentially at top dead center (TDC) in a full braking force mode. Fig. 6 illustrates that extra compression loading during the compression stroke creates the engine braking effect. The exhaust valves 72 open essentially at the TDC position to release the pneumatic pressure in the cylinder 66, which flows into the exhaust manifold 71 via the flow path 70 as shown by the arrows in the figure. Because all the pressure is released from the cylinder 66 essentially at TDC, Fig. 6 represents the working piston cylinder assembly as would be operating to achieve 100 percent braking force. As described above with respect to Figs. 1-4, the pressure that permits opening of the exhaust valves may be controlled to vary the braking force of the decompressing engine braking system. In particular, pressure control is performed such that the exhaust valves will remain closed at TDC, resulting in rebound of the piston passed TDC due to the high pressure within the cylinder 66. As the piston rebounds, the cylinder pressure decreases until the pressure inside the cylinder 66 reaches the control pressure for permitting opening of the exhaust valves 72. In accordance with such operation, Fig. 7 is a drawing depicting a partial cross- sectional view of the exemplary working piston assembly with the piston in the second position essentially at top dead center in a rebound mode. In the rebound mode, in contrast to Fig. 6, Fig. 7 shows that the exhaust valves 72 have remained closed at TDC. Looking at Fig. 7, no flow arrows are present to illustrate the closed nature of the valves and the corresponding lack of flow. Accordingly, the pneumatic pressure in the cylinder 66 is not released at TDC, so 100 percent braking force is not realized.

Fig. 8 is a drawing depicting a partial cross-sectional view of the exemplary working piston assembly 60 with the piston in a third position passed top dead center in the rebound mode corresponding to eighty percent braking force. Looking at Figs. 7 and 8 in combination, the piston rebounds from the TDC position of Fig. 7 to eighty percent of TDC in Fig. 8. During such rebound, the pressure in the cylinder 66 falls proportionately as the piston reaches the position of Fig. 8. In the example of Fig. 8, the pressure control parameters have been set so as to permit the exhaust valves 72 to open at eighty percent of the pressure relative to TDC. As seen in Fig. 8, therefore, the exhaust valves 72 open when the piston has rebounded to the third position, as shown by the flow arrows. The example of Fig. 8 would correspond to the control parameters described with respect to the schematic diagram of Fig. 3, in which the command solenoid coil is energized to achieve a control pressure corresponding to eighty percent of full braking force.

Similarly, Fig. 9 is a drawing depicting a partial cross-sectional view of the exemplary working piston assembly 60 with the piston in a fourth position passed top dead center in the rebound mode corresponding to fifty percent braking energy. Looking at Figs. 7 and 9 in combination, the piston rebounds from the TDC position of Fig. 7 to fifty percent of TDC in Fig. 9. During such rebound, the pressure in the cylinder 66 falls proportionately as the piston reaches the position of Fig. 9. In the example of Fig. 9, the pressure control parameters have been set so as to permit the exhaust valves 72 to open at fifty percent of the pressure relative to TDC. As seen in Fig. 9, therefore, the exhaust valves 72 open when the piston has rebounded to the fourth position., as shown by the flow arrows. The example of Fig. 9 would

correspond to control parameters in which the command solenoid coil is energized to achieve a control pressure corresponding to fifty percent of full braking force.

The present invention, therefore, achieves highly precise and efficient variable control of an engine decompression brake system. Such control is achieved without utilizing a separate back pressure device within the exhaust manifold, as required in conventional configurations.

Fig. 10 is a graph depicting an exemplary pressure profile for an engine brake in accordance with embodiments of the present invention. The graph of Fig. 10 relates the control pressure to the crankshaft angle of the cam lobe that is

associated with the engine brake. As seen in Fig. 10, for maximum braking force (control pressure 100%), the pressure curve tracks somewhat that of the crankshaft angle. In the example of Fig. 10, the control pressure associated with the maximum braking force is approximately 3700 psi, which differs from the control pressure of maximum braking force in the above examples. Generally, the control pressure at maximum braking force may vary and are engine design driven. The control pressures may vary based on such parameters as exhaust valve face area and the mechanical advantage of the rocker arm.

As seen in Fig. 10, the pressure rises with the crankshaft angle until the pressure reaches a maximum corresponding to the control pressure. Engine braking is then achieved by release of the control pressure, and the pressure decreases with further progression of the crankshaft angle. As referenced above, the maximum braking force in this example is achieved at approximately 3700 psi. Fig. 10 further depicts the pressure profiles for control pressures at 75% and 50% braking force, corresponding to control pressures of approximately 2800 psi and 1850 psi respectively.

An aspect of the invention, therefore, is a variable decompression engine brake system. In exemplary embodiments, the brake system includes a master piston assembly configured to engage a prime drive mechanism, a slave piston assembly including a piston that is moveable within a cylinder, the slave piston assembly being hydraulically connected to the master piston assembly and configured to operate at least one exhaust valve, and a control valve having a differential area built into the control valve that sets a control pressure. The slave piston assembly is configured to actuate the at least one exhaust valve when a pressure within the cylinder reaches the control pressure to perform decompression engine braking.

In an exemplary embodiment of the brake system, the brake system further includes a command solenoid coil that is energized to set a command pressure, and the differential converts the command pressure proportionally to the control pressure for opening of the at least one exhaust valve.

In an exemplary embodiment of the brake system, the ratio of the control pressure to the command pressure across the differential is between 40:1 and 50:1.

In an exemplary embodiment of the brake system, the slave piston assembly is configured to reach the control pressure when or after the piston has reached a top dead center position.

In an exemplary embodiment of the brake system, the slave piston assembly is configured to reach the control pressure when the piston has reached the top dead center position to achieve full braking force.

In an exemplary embodiment of the brake system, the slave piston assembly is configured to reach the control pressure after the piston has rebounded from the top dead center position, and a reduction of braking force from full braking force is proportional to an amount of rebound from the top dead center position.

In an exemplary embodiment of the brake system, the piston of the slave piston assembly is configured to telescope to actuate a rocker arm that in turn actuates the at least one exhaust valve.

In an exemplary embodiment of the brake system, the master piston assembly is configured to engage a decompression brake lobe on the prime drive mechanism. In an exemplary embodiment of the brake system, the brake system further includes a manifold that houses the at least one exhaust valve, and has a flow path for exhausting the pressure within the cylinder of the working piston.

Another aspect of the invention is a control method for controlling a variable decompression engine brake system. In exemplary embodiments, the control method includes the steps of cutting off fuel to an engine; applying a duty cycle current to a command solenoid coil thereby allowing a pressure to act on a control valve to set a control pressure; shifting the control valve to provide an oil volume to activate a decompression brake piston within a cylinder; advancing the piston to a top dead center position within the cylinder, thereby building up pressure in a piston chamber; and opening at least one exhaust valve when a pressure within the cylinder reaches a control pressure to perform decompression engine braking.

In an exemplary embodiment of the control method, the control valve has a differential area, and the differential converts the command pressure proportionally to the control pressure for opening of the at least one exhaust valve.

In an exemplary embodiment of the control method, a ratio of the control pressure to the command pressure across the differential is between 40:1 and 50:1.

In an exemplary embodiment of the control method, the control pressure in the cylinder is reached when or after the piston has reached a top dead center position.

In an exemplary embodiment of the control method, the control pressure in the cylinder is reached when the piston has reached the top dead center position to achieve full braking force.

In an exemplary embodiment of the control method, the control method further includes rebounding the piston after top dead center to reach the control pressure, and a reduction of braking force from full braking force is proportional to an amount of rebound from the top dead center position.

In an exemplary embodiment of the control method, the control pressure is proportional to a magnitude of the duty cycle current applied to the command solenoid. In an exemplary embodiment of the control method, a magnitude of braking force is proportional to a magnitude of the duty cycle current applied to the command solenoid.

In an exemplary embodiment of the control method, a magnitude of braking force is proportional to a magnitude of the control pressure.

Although the invention has been shown and described with respect to a certain embodiment or embodiments, it is obvious that equivalent alterations and modifications will occur to others skilled in the art upon the reading and

understanding of this specification and the annexed drawings. In particular regard to the various functions performed by the above described elements (components, assemblies, devices, compositions, etc.), the terms (including a reference to a "means") used to describe such elements are intended to correspond, unless otherwise indicated, to any element which performs the specified function of the described element (i.e., that is functionally equivalent), even though not structurally equivalent to the disclosed structure which performs the function in the herein illustrated exemplary embodiment or embodiments of the invention. In addition, while a particular feature of the invention may have been described above with respect to only one or more of several illustrated embodiments, such feature may be combined with one or more other features of the other embodiments, as may be desired and advantageous for any given or particular application.

Claims

Claims What is claimed is:

1. A variable decompression engine brake system comprising:

a master piston assembly configured to engage a prime drive mechanism; a slave piston assembly including a piston that is moveable within a cylinder, the slave piston assembly being hydraulically connected to the master piston assembly and configured to operate at least one exhaust valve; and

a control valve having a differential area built into the control valve that sets a control pressure;

wherein the slave piston assembly is configured to actuate the at least one exhaust valve when a pressure within the cylinder reaches the control pressure to perform decompression engine braking.

2. The brake system according to claim 1 , further comprising a command solenoid coil that is energized to set a command pressure, and the differential converts the command pressure proportionally to the control pressure for opening of the at least one exhaust valve.

3. The brake system according to claim 2, wherein the ratio of the control pressure to the command pressure across the differential is between 40:1 and 50:1.

4. The brake system according to any of claims 1-3, wherein the slave piston assembly is configured to reach the control pressure when or after the piston has reached a top dead center position.

5. The brake system according to claim 4, wherein the slave piston assembly is configured to reach the control pressure when the piston has reached the top dead center position to achieve full braking force.

6. The brake system according to claim 5, wherein the slave piston assembly is configured to reach the control pressure after the piston has rebounded from the top dead center position, and a reduction of braking force from full braking force is proportional to an amount of rebound from the top dead center position.

7. The brake system according to any of claims 1-6, wherein the piston of the slave piston assembly is configured to telescope to actuate a rocker arm that in turn actuates the at least one exhaust valve.

8. The brake system according to any of claims 1-7, wherein the master piston assembly is configured to engage a decompression brake lobe on the prime drive mechanism.

9. The brake system according to any of claims 1-8, further comprising a manifold that houses the at least one exhaust valve, and has a flow path for exhausting the pressure within the cylinder of the working piston.

10. A control method for controlling a variable decompression engine brake system comprising the steps of:

cutting off fuel to an engine;

applying a duty cycle current to a command solenoid coil thereby allowing a pressure to act on a control valve to set a control pressure;

shifting the control valve to provide an oil volume to activate a decompression brake piston within a cylinder;

advancing the piston to a top dead center position within the cylinder, thereby building up pressure in a piston chamber; and

opening at least one exhaust valve when a pressure within the cylinder reaches a control pressure to perform decompression engine braking.

11. The control method according to claim 10, wherein the control valve has a differential area, and the differential converts the command pressure

proportionally to the control pressure for opening of the at least one exhaust valve.

12. The control method according to claim 11 , wherein a ratio of the control pressure to the command pressure across the differential is between 40:1 and 50:1.

13. The control method according to any of claims 10-12, wherein the control pressure in the cylinder is reached when or after the piston has reached a top dead center position.

14. The control method according to claim 13, wherein the control pressure in the cylinder is reached when the piston has reached the top dead center position to achieve full braking force.

15. The control method according to claim 14, further comprising rebounding the piston after top dead center to reach the control pressure, and a reduction of braking force from full braking force is proportional to an amount of rebound from the top dead center position.

16. The control method according to any of claims 10-15, wherein the control pressure is proportional to a magnitude of the duty cycle current applied to the command solenoid.

17. The control method according to any of claims 10-16, wherein a magnitude of braking force is proportional to a magnitude of the duty cycle current applied to the command solenoid.

18. The control method according to any of claims 10-17, wherein a magnitude of braking force is proportional to a magnitude of the control pressure.