Abstract:
An apparatus and method for varying a counter force to valve spring preload of a brake exhaust valve to undertake engine braking, includes a solenoid controlled hydraulic actuator. A control cylinder is arranged to move with a rocker arm and a control piston is arranged to slide within the control cylinder. During engine braking the control piston slides to press the valve stem to open the brake exhaust valve. An oil chamber is arranged above the control piston and is open into the control cylinder. A source of pressurized oil is selectably introduced into the oil chamber by the solenoid controlled hydraulic actuator to slide the control piston within the control cylinder to open and hold open the brake exhaust valve.

Description:
TECHNICAL FIELD 
       [0001]    This disclosure relates to vehicles, particularly large tractor trailer trucks, including but not limited to control and operation of an engine for engine braking. 
       BACKGROUND 
       [0002]    Adequate and reliable braking for vehicles, particularly for large tractor-trailer trucks, is desirable. While drum or disc wheel brakes are capable of absorbing a large amount of energy over a short period of time, the absorbed energy is transformed into heat in the braking mechanism. 
         [0003]    Braking systems are known which include exhaust brakes which inhibit the flow of exhaust gases through the exhaust system, and compression release systems wherein the energy required to compress the intake air during the compression stroke of the engine is dissipated by exhausting the compressed air through the exhaust system. 
         [0004]    In order to achieve a high engine-braking action, a brake valve in the exhaust line may be closed during braking, and excess pressure is built up in the exhaust line upstream of the brake valve. For turbocharged engines, the built-up exhaust gas flows at high velocity into the turbine of the turbocharger and acts on the turbine rotor, whereupon the driven compressor increases pressure in the air intake duct. The cylinders are subjected to an increased charging pressure. In the exhaust system, an excess pressure develops between the cylinder outlet and the brake valve and counteracts the discharge of the air compressed in the cylinder into the exhaust tract via the exhaust valves. During braking, the piston performs compression work against the high excess pressure in the exhaust tract, with the result that a strong braking action is achieved. 
         [0005]    Another engine braking method, as disclosed in U.S. Pat. No. 4,395,884, includes employing a turbocharged engine equipped with a double entry turbine and a compression release engine retarder in combination with a diverter valve. During engine braking, the diverter valve directs the flow of gas through one scroll of the divided volute of the turbine. When engine braking is employed, the turbine speed is increased, and the inlet manifold pressure is also increased, thereby increasing braking horsepower developed by the engine. 
         [0006]    Other methods employ a variable geometry turbocharger (VGT). When engine braking is commanded, the variable geometry turbocharger is “clamped down” which means the turbine vanes are closed and used to generate both high exhaust manifold pressure and high turbine speeds and high turbocharger compressor speeds. Increasing the turbocharger compressor speed in turn increases the engine airflow and available engine brake power. The method disclosed in U.S. Pat. No. 6,594,996 includes controlling the geometry of the turbocharger turbine for engine braking as a function of engine speed and pressure (exhaust or intake, preferably exhaust). 
         [0007]    U.S. Pat. No. 6,148,793 describes a brake control for an engine having a variable geometry turbocharger which is controllable to vary intake manifold pressure. The engine is operable in a braking mode using a turbocharger geometry actuator for varying turbocharger geometry, and using an exhaust valve actuator for opening an exhaust valve of the engine. 
         [0008]    In compression-release engine brakes, there is an exhaust valve event for engine braking operation. For example, in the “Jake” brake, such as disclosed in U.S. Pat. Nos. 4,423,712; 4,485,780; 4,706,625 and 4,572,114, during braking, a braking exhaust valve is closed during the compression stroke to accumulate the air mass in engine cylinders and is then opened at a selected valve timing somewhere before the top-dead-center (TDC) to suddenly release the in-cylinder pressure to produce negative shaft power or retarding power. 
         [0009]    In “Bleeder” brake systems, during engine braking, a braking exhaust valve is held constantly open during a large portion of the engine cycle to generate a compression-release effect. 
         [0010]    According to the “EVBec” engine braking system of Man Nutzfahrzeuge AG, there is an exhaust secondary valve lift event induced by high exhaust manifold pressure pulses during intake stroke or compression stroke. The secondary lift profile is generated in each engine cycle and it can be designed to last long enough to pass TDC and high enough near TDC to generate the compression-release braking effect. 
         [0011]    The EVBec engine brake does not require a mechanical braking cam or variable valve actuation (“VVA”) device to produce the exhaust valve braking lift events. The secondary valve lift is produced by closing an exhaust back pressure (“EBP”) valve located at the turbocharger turbine outlet and the exhaust valve held open by a “lock-in” hydraulic mechanism during the engine compression stroke. When the engine brake needs to be deactivated, the EBP valve is set back to its fully open position to reduce the exhaust manifold pressure pulses during each engine cycle so that the exhaust valve floating and secondary lift as well as the braking lift event at TDC do not occur. Such a system is described for example in U.S. Pat. No. 4,981,119. 
         [0012]    When operating the EVBec engine brake, when the turbine outlet EBP valve is very closed, turbine pressure ratio becomes very low, hence engine air flow rate becomes low. Also, engine delta P (i.e., exhaust manifold pressure minus intake manifold pressure) and exhaust manifold pressure may become undesirably high. As a result, the compression-release effect can be weakened, retarding power can be reduced, and in-cylinder component (e.g. fuel injector tip) temperature can become undesirably high. 
         [0013]    For the EVBec compression-release engine brake, the valve motion of the braking exhaust valve is determined passively by mainly the valve spring preload and exhaust manifold pressure pulses. The braking exhaust valve may open at an undesirable location (e.g., during the intake stroke), and it results in excessive gas leaking from the cylinder to intake manifold so that retarding power is reduced. Moreover, at low engine speed or when the turbine outlet exhaust back pressure (EBP) valve is opened, exhaust manifold pressure pulse is weaker (lower) than that at high speed or when the EBP valve is closed. In this situation, the braking valve is difficult to open due to the relatively strong spring preload and weak exhaust pressure pulse. For the purpose of increasing engine retarding power, it is desirable to open the EBP valve to increase turbine pressure ratio and engine air flow rate. 
         [0014]    The present inventors recognize the desirability of producing a variable counter force to exhaust valve spring preload to control the braking valve motion and timing at variable speeds and exhaust manifold pressure levels. 
         [0015]    The present inventors have recognized the desirability of providing a more effective engine braking system. 
       SUMMARY 
       [0016]    The exemplary embodiment of the invention provides an apparatus for varying a counter force to exhaust valve spring preload of a brake exhaust valve to undertake engine braking. The embodiment includes the brake exhaust valve having a first valve stem and a valve spring to urge the valve closed; a rocker arm for pressing the valve stem to open the valve by overcoming spring preload during engine firing operation; a control cylinder arranged to move with the rocker arm; a control piston arranged to slide within the control cylinder, during engine braking the control piston slidable to press the valve stem to open the valve; an oil chamber arranged above the control piston and open into the control cylinder; and a source of pressurized oil selectably introduced into the oil chamber to slide the control piston within the control cylinder. 
         [0017]    The component for selectively introducing pressurized oil can be a solenoid valve arranged to selectively open the oil chamber to the source of pressurized oil. Alternately, a first passage can be arranged between the source of pressurized oil and the oil chamber and a second passage can be arranged between the oil chamber and the crankcase and a solenoid valve can be arranged in the second passage to close in order to subject the oil chamber to the source of pressurized oil. 
         [0018]    More particularly, the embodiment can include a valve bridge and a further exhaust valve having a second valve stem, the valve bridge arranged between the rocker arm and the first and second valve stems of the brake exhaust valve and the further exhaust valve. The valve bridge is movable with the rocker arm to open the brake exhaust valve and the further exhaust valve. The control cylinder can be formed into the valve bridge. 
         [0019]    The source of pressurized oil can be oil pressurized by the engine oil circulation pump taken from the oil passage at the rocker arm shaft. The source of pressurized oil can also be a booster oil pump taking suction from engine oil pre-pressurized by the engine oil circulation pump, which delivers a higher oil pressure and can change the equivalent net spring load more significantly. 
         [0020]    An exemplary method of the invention includes the steps of: 
         [0021]    generating a source of pressurized oil; and 
         [0022]    during engine braking, using the source of pressurized oil to selectively press the first valve stem to overcome spring preload to open the brake exhaust valve. 
         [0023]    More particularly, the method is further defined by arranging a control cylinder to move with the rocker arm, and a control piston arranged to slide within the control cylinder, the control piston operable to press the valve stem to open the valve, and an oil chamber arranged above the control piston and open into the control cylinder; and 
         [0024]    selectably introducing pressurized oil into the oil chamber to slide the control piston within the control cylinder. 
         [0025]    Furthermore, the step of selectively introducing pressurized oil can be further defined in that pressurized oil flowing though the oil chamber and into the crankcase is shut off downstream of the oil chamber, allowing the oil chamber to reach the elevated pressure of the source of pressurized oil. 
         [0026]    Alternately, the step of selectively introducing pressurized oil can be further defined in that the source of pressurized oil is first closed from the oil chamber is then opened to the oil chamber to reach the pressure of the source of pressurized oil. 
         [0027]    The exemplary apparatus and methods of the invention use solenoid valves and electro-hydraulic actuation designs to dynamically effect a counter force to exhaust valve spring preload. The electro-hydraulic actuation may occur once or multiple times during the engine cycle. When it occurs once during an engine cycle, it may produce a constant force acting on the braking valve. When it occurs multiple times, it may modulate to produce variable forces with certain higher resolution at the crank angle level. 
         [0028]    The exemplary apparatus and methods of the invention use an electro-hydraulic design to vary the net force acting on the exhaust braking valve(s) in compression-release engine brakes to control the braking valve timing and motion according to the needs at different engine speeds and levels of exhaust manifold pressure pulses. In addition, it reduces the need for high back pressure build up. As a result, engine retarding power can be increased. 
         [0029]    Engine retarding power may be increased through better braking motion control due to three reasons: less leakage of cylinder flow into the intake manifold; and more exhaust mass or energy can be harvested into the cylinder from the exhaust manifold to be further compressed by the engine piston motion to even hotter at the braking TDC (i.e., transferring more energy fed to the turbine); and more airflow mass from the intake manifold into the cylinder due to improved turbocharger efficiency from reduced back pressure. At low speed, it is possible to open the braking exhaust valve to activate the EVBec engine brake under a reduced net opening force across the valve. 
         [0030]    Numerous other advantages and features of the present invention will become readily apparent from the following detailed description of the invention and the embodiments thereof, from the claims and from the accompanying drawings. 
     
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         [0031]      FIG. 1  is a schematic diagram of an engine braking system according to an exemplary apparatus of the invention; 
           [0032]      FIG. 2  is a schematic side view of an exhaust valve system according to an exemplary apparatus of the invention; 
           [0033]      FIG. 3  is an enlarged fragmentary sectional view of a portion of a first embodiment of the exemplary apparatus shown in  FIG. 2 , as taken from  FIG. 4B ; 
           [0034]      FIG. 4A  is a fragmentary sectional view of an engine incorporating the exemplary apparatus shown in  FIG. 3 , shown in an “on” engine brake state; 
           [0035]      FIG. 4B  is a fragmentary sectional view of an engine incorporating the exemplary apparatus shown in  FIG. 3 , shown in an “off” engine brake state; 
           [0036]      FIG. 5  is an enlarged fragmentary sectional view of a portion of a second embodiment of the exemplary apparatus shown in  FIG. 2 , as taken from  FIG. 6A ; 
           [0037]      FIG. 6A  is a fragmentary sectional view of an engine incorporating the exemplary apparatus shown in  FIG. 5 , shown in an “on” engine brake state; and 
           [0038]      FIG. 6B  is a fragmentary sectional view of an engine incorporating the exemplary apparatus shown in  FIG. 5 , shown in an “off” engine brake state. 
       
    
    
     DETAILED DESCRIPTION 
       [0039]    While this invention is susceptible of embodiment in many different forms, there are shown in the drawings, and will be described herein in detail, specific embodiments thereof with the understanding that the present disclosure is to be considered as an exemplification of the principles of the invention and is not intended to limit the invention to the specific embodiments illustrated. 
         [0040]      FIG. 1  illustrates a simplified schematic of an engine braking control system  100 . The system acts on an exhaust valve  114  that opens a cylinder  116  to an exhaust manifold  118 . A piston  117 , operatively connected to an engine crankshaft (not shown), reciprocates within the cylinder  116 . An engine braking electronic control is signal-connected to a downstream EBP valve  126  which, by closing, can increase backpressure through a turbocharger turbine  128  and back through the exhaust gas manifold  118 . Although the EBP valve  126  is shown downstream of the turbine  128 , it is poossible that the EBP valve could be located upstream of the turbine  128 . The control is also signal-connected to a counter-preload device  150  to allow the exhaust valve  114  to be opened by differential pressure between the exhaust manifold  118  and pressure within the cylinder  116 . The control  120  can initiate exhaust-manifold-pressure-pulse-induced valve motion by commanding the EBP valve  128  to close to a specified degree and also increasing the counter-preload force on the valve  114  by commanding an increase in counter-preload force by the device  150 . 
         [0041]      FIG. 2  shows a counter-preload device (either on/off type or variable type) for achieving an ultra-low required opening force across a spring loaded exhaust valve used in the engine brake with exhaust-manifold-pressure-pulse-induced valve motion. The device reduces the required opening force across a valve by countering the valve spring preload to enable high retarding power at very low engine speed because with very low required opening force, the exhaust braking valve may float easily to generate a high secondary valve lift to recover more exhaust gas mass from exhaust manifold to cylinder to enable the high-temperature-flow operation of the engine brake through a faster spinning turbine. The variable counter-preload device can also adjust retarding power continously by regulating the size of exhaust secondary valve lift event. Moreover, the variable counter-preload device, if designed with electro-magnetic means, may be used to totally or partially deactivate the engine brake by applying an attractive magnetic force on the top of the braking valve to increase the closing force on the valve to stop the secondary lift event. 
         [0042]      FIG. 2  shows an exemplary preload system  200  for ultra-low required valve opening force, either an on/off type or variable type, used in engine braking operation. Identical devices can be used at all cylinders or some of the cylinders, of the engine, although only the system  200  at the cylinder  116  is shown. The system  200  includes a rocker arm  212 , a valve bridge  216 , the counter-preload device  150 , a normally operated exhaust valve  220  and an braking exhaust valve  114 . The valves  220  and  114  open the cylinder  116  to the exhaust manifold via exhaust gas passages  224 ,  226  provided in a cylinder head  230 . 
         [0043]    Each valve includes a stem  234  having a stem end  237 , a head  235 , and a spring keeper  236 . A valve spring  238  surrounds the stem  234  and is fit between the keeper  236  and the cylinder head  230 . To move the heads  235  away from valve seats  240 ,  242  during normal engine operation, at the selected crankshaft angle, the rocker arm  212  presses the valve bridge  216  down to move the valve stems  234  down via force on the ends  237  against the expansion force of the springs  238  as the springs are being compressed between the keepers  236  and the cylinder head  230 , and against the cylinder pressure force on the valve. 
         [0044]    During an engine braking operation, differential pressure across the head  235  of the valve  114  moves the head  235  down and away from the valve seat  242  and exhaust gas can enter the cylinder  116 . In this regard the valve is a “floating exhaust valve” in that differential pressure across the valve is sufficient to push the valve downward away from its seat. The differential pressure force is due to the pressure difference between exhaust gas backpressure within the passage  226  and the pressure within the cylinder  116 . The differential pressure must also be sufficient to overcome the expansion force of the spring  238  as the opening of the valve  114  compresses the spring  238 . 
         [0045]    The counter-preload device  150  includes an actuator portion  244  shown installed on top of the valve bridge  216 . Alternatively, the actuator portion  244  can be installed within the valve bridge (shown dashed). The device  150  also includes a rod  250 . The rod  250  is moved by force from the actuator portion  244  to press down the end  237  of the stem. The required opening force across the valve refers to the net force on the valve of the normal spring preload and the opposing force exerted by the counter-preload device. The counter-preload device  150  can provide engine brake activation and deactivation controls and the ability of achieving variable required opening force across the valve to obtain variable or higher retarding power during engine braking operation. The device  150  can be variable or can be strictly on/off. 
         [0046]    The device may reduce the required opening force across the valve to enable the brake to operate at very low engine speed because with very low required opening force across the valve the exhaust braking valve may float easily off its valve seat to generate a secondary valve lift for braking. Moreover, the device can make the secondary lift very high to recover more exhaust gas mass from exhaust manifold to cylinder to enable the high-flow-temperature operation of the engine brake through a faster spinning turbine. 
         [0047]    Alternately, the rod  250  can be operatively connected to the valve stem  234  so that the actuator can exert a selectable two way force (up or down) on the valve  114 . In this way the device  150  can act to assist the spring  238  in closing the valve in addition to acting as a counter-preload to open the valve. It is also possible that the device, configured as a two way force acting device, can eliminte the need for the spring altogether. 
         [0048]    The variable counter-preload device can also adjust retarding power continously by regulating the size of exhaust secondary valve lift event. 
         [0049]      FIGS. 3-4B  illustrates one embodiment of the invention. Referring to  FIG. 4A , a rocker arm shaft  270  pivotally supports a plurality of rocker arms  212  (one shown). The rocker arms  212  pivot about the shaft  270  by reciprocating vertical movement of push rods  274  which are moved by a camshaft (not shown). In this configuration, oil is supplied through an oil passage  275  from the existing engine pressurized oil supply in the rocker arm shaft  270 , through the rocker arm  212 , through the valve bridge  216  and to an oil chamber  280  above a control piston  290  overlying the valve  114 . The control piston  290  is sealingly slidable within a control cylinder  292  formed in the valve bridge  216 . 
         [0050]    An end portion of the valve stem  234 , including the valve stem end  237 , fits within a socket portion  293  of the piston  290 . A spring  294  braced against the valve bridge  216  and the piston  290  maintains a pressing contact between the piston  290  and the valve stem end  237 . 
         [0051]    A solenoid valve  310  is normally open ( FIG. 4A ). The oil from the rocker arm shaft  270  and in the oil chamber  280  bleeds out through a channel  315 , through a channel  316 , through a valve passage  320  in a valve element  322  of the solenoid valve  310  that is in registry with side holes  324 ,  325  in a surrounding body  340  of the solenoid valve  310 , and through a channel  326  into the crankcase  330 . The hydraulic force acting down upon the top of the valve  114 , via the piston  290  is insignificant. 
         [0052]    As shown in  FIGS. 3 and 4B , when a solenoid coil  336  of the solenoid valve  310  is energized, the solenoid valve element  322  is raised by magnetic force and the valve passage  320  is closed with respect to the surrounding body  340  of the solenoid valve  310 . The oil pressure in the channels  316 ,  315 , in the oil chamber  280 , and in the passage  275  is raised to that of the oil pressure in the rocker arm shaft  270 . The elevated oil pressure in the oil chamber  280  acting on the piston  290  generates a step-change hydraulic force acting on the end  237  of the valve  114  and pushes the valve  114  downward and open. The amplitude of the hydraulic force is determined by the oil supply pressure and the area of the piston  290  at the top of the valve. 
         [0053]    During the compression stroke, when the air pressure within the cylinder  116  increases as the piston  117  ( FIG. 1 ) moves up, the pressure inside the oil chamber  280  pushes closed a check valve  350 , represented as a ball check valve, to reverse flow into the oil supply from the passage  275 . A ball  351  closed against a seat  352  effectively seals an inlet side of the oil chamber  280 . The valve  114  is therefore locked in the open position. 
         [0054]    The oil in the chamber  280  is eventually released during the exhaust stroke when the valve bridge  216  is pushed down by cam on the camshaft (not shown) via the pushrod  274  and the rocker arm  212 , and opens the channel  315  on top of the oil chamber  280  to the crankcase  330 . 
         [0055]    The operation of the solenoid valve  310  is controlled by control  120  which can be controlled by, or be part of, the electronic control unit (ECU) of the engine. This configuration requires no additional oil pump. 
         [0056]    To return the solenoid valve element  322  to the original position, the solenoid coil  310  is de-energized. A return spring  360  between a top of the element  322  and the body  340  forces the solenoid valve element  322  back to the original position with the passage  320  open with respect to side holes  324 ,  325  in the body  340 . Alternatively, another close solenoid may be mounted on the opposite side of the solenoid coil  336  to pull the valve element  322  to the original position. A seating spring  366  between the element  322  and a bottom surface of the body  340  reduces the amplitude of the impact noise. 
         [0057]    A cover  370  can be applied over the body  340  to retain the body into a wall  372  of the crankcase  330 . The cover  370  and/or the body  340  can have external threads to be threaded into internal threads in the wall  372  to retain the body into the wall  372 . An o-ring seal  376  can be applied between the body  340  and the wall  372 . 
         [0058]    The channel  316  can be formed through a fitting  380  having external threads that can engage inside threads of the wall  372 . A pair of o-ring seals  384 ,  386  seal the channel  316  between the fitting  380  and the wall  372 . An end surface  390  of the fitting  380  forms a seat between the fitting  380  and the bridge  216 , to form a substantially sealed connection between the channel  316  and the channel  315 . 
         [0059]    The solenoid valve  310  may include one coil, one preloaded spring, one seating spring, and one moving piston; or one actuation coil, one returning coil, one moving piston (not shown), or the like. 
         [0060]      FIGS. 5-6B  illustrate another embodiment of the invention. In this configuration, oil under higher pressure is supplied from a booster oil pump  392  (shown schematically) to a passage  394 . The booster pump can take suction from pressurized oil from the engine oil circulation pump and raises the oil pressure further. The solenoid valve  310  is normally in the closed position ( FIG. 6B ). The passage  394  at the hole  325  is blocked by the element  322 . The hydraulic force acting upon the top of the valve  114  via the control piston  290  is insignificant. 
         [0061]    When the solenoid valve  310  is energized, the solenoid valve element  322  is pulled up by the coil  336  and the passage  320  registers with the holes  324 ,  325  in the surrounding body  340  ( FIGS. 5 and 6A ). The passage  394  is connected with the passage  320  and the channel  316  that passes through the wall  372  and through the fitting  380 . The channel  316  is connected to the channel  315  and to the oil chamber  280  on top of the control piston  290 . Oil pressure builds up in the oil chamber  280 , which generates a step-change hydraulic force acting on top of the valve  114 , via the control piston  290 , and pushes the valve  114  open. The amplitude of the hydraulic force is determined by the oil supply pressure and the area of the control piston  290  at the top of the valve  114 . 
         [0062]    The solenoid valve  310  is then closed by the coil  336  lowering the element  322 , which locks in the oil in the oil chamber  280  and effectively seals the chamber  280 , and the valve  114  is locked in the open position. 
         [0063]    The oil in the chamber  280  is released at the exhaust stroke when the valve bridge  216  is pushed down by the cam and opens the hole on top of the oil chamber. 
         [0064]    The solenoid valve operation can be controlled by, or be part of, the ECU of the engine. This configuration may use an accumulator  420  which receives pressurized oil from the pump  392 . The oil pressure delivered from the booster oil pump can be made higher than the oil pressure from the rocker arm shaft ( FIG. 4A ), and a greater step change hydraulic force can be generated. The booster pump  392  takes suction from the oil lubrication system that is elevated in pressure by the engine oil circulation pump  410  taking suction from the oil sump  414  of the engine (shown schematically in  FIG. 5 ). This elevated oil pressure allows the valve  114  to open more swiftly, which leads to more precise control of the valve  114 . 
         [0065]    The solenoid valve  310  may include one coil  336 , one preloaded spring  360 , one seating spring  366 , and one moving valve element; or one actuation coil  336 , one returning coil (not shown), one moving valve element  322 , or the like. 
         [0066]    When the actuation solenoid coil is energized, it pulls the moving valve element towards the coil, and opens the valve  310 . To return the element  322  to the original position, the actuation solenoid coil is de-energized. The spring  360  forces the element  322  back to the original position. Alternatively, another close solenoid may be mounted on the opposite side of the solenoid coil  336  to pull the valve element  322  to the original position. The seating spring  366  reduces the amplitude of the impact noise. 
         [0067]    From the foregoing, it will be observed that numerous variations and modifications may be effected without departing from the spirit and scope of the invention. It is to be understood that no limitation with respect to the specific apparatus illustrated herein is intended or should be inferred.