Patent Publication Number: US-6983725-B2

Title: Exhaust valve mechanism in internal combustion engines

Description:
FIELD OF THE INVENTION 
   An exhaust valve mechanism in an internal combustion engine, comprising at least one exhaust valve in each cylinder, a rocker arm shaft-mounted rocker arm for each cylinder for operating the exhaust valve, a cam shaft with a cam element for each rocker arm, said cam element cooperating with motion transmitting means at one end of the rocker arm, a first piston-cylinder device disposed between an opposite end of the rocker arm and the exhaust valve, said first piston-cylinder device having a first cylinder chamber in said opposite rocker arm end, a hydraulic circuit for supplying and draining off pressure fluid to and from said cylinder chamber, and a piston disposed in said cylinder chamber, said piston being biased towards the exhaust valve when pressure fluid is supplied to the cylinder chamber. 
   BACKROUND OF THE INVENTION 
   SE-A-468 132 describes an exhaust valve mechanism of the above mentioned type which, together with a special type of camshaft with exhaust cams with extra lobes can be used to increase the engine braking power. The extra cam lobes are dimensioned so that their lifting height corresponds to the normal valve play of the valve mechanism. By reducing, by means of the piston cylinder device, the valve plate to zero, one or more extra lifts of the exhaust valve corresponding to the normal valve play can be achieved during a suitable time interval. For example, an extra cam lobe can be placed in relation to the regular cam lobe so as to provide an extra exhaust valve lift during a later part of the compression stroke, resulting in a loss of a portion of the compression work during the compression stroke which will not be recovered during the expansion stroke. This increases the braking effect of the engine. 
   In an engine with such an arrangement, the maximum lift height of the exhaust valve during the compression when engine braking, is limited to the valve play. Furthermore, the overlap of the exhaust valve and the intake valve in braking mode increases by virtue of the fact that the maximum lift height of the exhaust valve increases by a distance corresponding to the valve play as compared to drive mode. Since the pressure in the exhaust manifold is much higher than the pressure in the intake manifold in braking mode (ca 5 bar on the exhaust side as opposed to ca 1 bar on the intake side), hot exhaust in an amount depending on the overlap will flow between the exhaust side and the intake side during braking mode, which will impair the engine cooling during braking mode as compared to driving mode, especially since fuel as a cooling medium for the injection nozzle is not available during braking mode. Finally, the exhaust rocker arm must be dimensioned more robustly for braking mode than for normal driving mode, since the opening force on the exhaust valve in braking mode must overcome the force from a high compression pressure in the cylinder, this force being substantially higher than the force on the valve required for normal opening during the exhaust stroke. 
   OBJECTS OF THE INVENTION 
   One purpose of the present invention is to achieve an exhaust valve mechanism of the type described by way of introduction which is constructed so that extra lifting of the exhaust valve during braking mode can be effected without affecting the regular lifting of the exhaust valve, to thereby avoid increasing the overlap between the exhaust valve and the intake valve with accompanying large back-flow and reduction of the mass-flow through the engine. 
   Another purpose of the invention is to achieve an exhaust valve mechanism, where the lifting height of the extra lift of the exhaust valve during braking mode is not limited to the valve play. 
   An additional purpose of the invention is to achieve an exhaust valve device, in which the exhaust rocker am does not need to be dimensioned for braking mode but only for driving mode. 
   SUMMARY OF THE INVENTION 
   This is achieved according to the invention by virtue of the fact that the rocker arm is provided with a second piston-cylinder device on the same side of the rocker arm shaft as the first piston-cylinder device, said second piston-cylinder device having a second cylinder chamber communicating with the first cylinder chamber and housing a second piston which, upon supply of pressure fluid to the second cylinder chamber, is biased in a direction from the exhaust valve, and that a second rocker arm mounted on a rocker arm shaft has an end acting against the second piston and an opposite end with motion-transmitting means, which cooperate with a cam element on a cam shaft. 
   The invention is based on the idea of using two separate rocker arms, one for exhaust valve lifting during regular driving mode and one for exhaust valve lifting in braking mode. The regular exhaust rocker arm can have a normal lever ratio on the order of 1:1,4–1,6 and need only be dimensioned for the forces occurring during driving mode. The exhaust valve rocker arm for braking mode transmits the valve movement from a separate cam element, whereby the extra cam lobes on the cam elements for regular drive mode can be eliminated. The rocker arm for braking mode acts on the second piston which functions as a pump piston and pumps fluid to the first cylinder chamber. The pressure in the first cylinder chamber presses the first piston towards the exhaust valve. The valve movement during braking mode is thus transmitted partially hydraulically. The second exhaust rocker arm can have another lever ratio than the first exhaust rocker arm, e.g. 1:0,7–1,1, which reduces the forces and the contact pressure in the mechanism. The cam element cooperating with the second rocker arm can have a greater base diameter than the cam element of the first rocker arm, which reduces the contact pressure and/or offers more rapid upward or downward movement. 
   Through the invention it is possible to eliminate the large valve overlap which is necessary when extra cam lobes are used in the regulate cam element for braking mode, because a high and long ramp is not required to conceal the extra lobes during driving mode. The return flow of exhaust into the cylinder and on through the inlet port, caused by overpressure in the exhaust manifold, is thereby reduced. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
     The invention will be described in more detail below with reference to examples shown in the accompanying drawings, where 
       FIG. 1  shows a side view of one embodiment of an exhaust valve mechanism according to the invention with a longitudinal section through the exhaust valve rocker arm for regular valve lifting during driving mode but without the rocker arm for braking mode, 
       FIG. 2  shows a side view, mirror reversed in relation to  FIG. 1 , of the valve mechanism according to the invention with the rocker arm for braking mode and with the rocker arm for regular valve lift partially in section, 
       FIG. 3  shows a section through the rocker arm in  FIG. 1  along the line III—III, 
       FIG. 4  shows a section through the rocker arm in  FIG. 1  along the line IV—IV, 
       FIG. 5  is a diagram illustrating the lifting curves of the exhaust valve and of the intake valve in normal driving mode, 
       FIG. 6  is a corresponding diagram during braking mode with the described previously known exhaust valve mechanism, and 
       FIG. 7  is a corresponding diagram during braking mode with the valve mechanism according to the present invention. 
   

   DETAILED DESCRIPTION OF THE INVENTION 
     FIG. 1  shows schematically a valve mechanism  1  in an internal combustion engine (not shown). The mechanism  1  comprises an exhaust valve rocker arm  2 , which is rockably mounted on a rocker arm shaft  3 . One end of the rocker arm  2  has a cam follower roller  4 , rotatably mounted thereon. The cam follower roller  4  is in contact with a schematically shown cam element  5  on the camshaft  6 . The designation “a” indicates the base circle of the cam element  5 , and “b” designates its top radius. At its end  7  opposite to the end with the cam follower roller  4 , the rocker arm  2  is provided with a piston cylinder device  8  consisting of a cylinder chamber  9  formed in the rocker arm end  7  and a piston  10  housed in the cylinder chamber. The piston  10  is provided with a piston pin  11  with a spherical end extending into a socket  12  on a yoke  13  which, during operation, applies pressure to two exhaust valve spindles  14 .  15  designates two valve springs for closing the valves. Beyond the springs  15  there is an additional spring  16 , which is designed to keep the yoke  13  in such a position that the play, which is always present in a valve mechanism of this type, is disposed between the ends  14  of the spindles and the underside of the yoke  13 . 
   The valve mechanism  1  described is lubricated by pressurized oil which is supplied by the engine oil pump via channels in the engine block and the cylinder head (not shown) to a channel  17  in the rocker arm shaft  3 . The rocker arm  2  has journal bearings  18 , which are lubricated by a minor leakage flow between the shaft  3  and the bearing  18 . The excess oil is returned via a return line  19 , in a hydraulic circuit generally designated  20 , which contains a valve device  21  consisting of a valve housing  22  and a valve element  24  biased by a spring  23 . The housing  22  has an outlet  25  through which return oil flows back to the engine oil sump, when the valve element is in the position shown in  FIG. 1 . The housing  22  also has an inlet  26  for a pressure medium (compressed air or hydraulic fluid). When pressure medium is supplied through the inlet  26 , the valve element  24  is biased upwards in  FIG. 1 , thereby closing the outlet  25  and blocking the return flow through the line  19 . The result will be that the pressure in the channel  17  rises. The channel  17  communicates via a channel  27  with the cylinder chamber  9  above the piston  10 , which leads to the piston being loaded downwards towards the valve yoke  13  so that the play between the yoke and the upper end surfaces of the valve spindles is adjusted down to zero. In the piston  10  there is a relief valve, which limits the pressure to a predetermined level. If this level is exceeded, the valve  28 ,  29  opens so that oil can drain out through channels  30  in the piston. 
   In order to prevent the pumping of oil between the cylinder chamber  9  and the chamber  17  in the rocker arm shaft during operation with zero valve play, a one-way valve  31  ( FIG. 3 ) is arranged in the rocker arm channel  27 . The one-way valve  31  comprises a valve element  32  in the form of a ball which, when there is high pressure in the hydraulic circuit, is held in its closed position by the pressure in the cylinder chamber  9  and by a spring. The pressure in the hydraulic circuit acts also against the end of a piston  34  biased by a spring  33 . The piston  34  has a shaft  35  extending to the seat of the ball  32 . When there is high pressure in the circuit, i.e. when the valve  21  is closed, the pressure will keep the piston  34  in a position with the end of the shaft  35  at a distance from the ball  32 , thereby keeping the valve closed. When the valve  21  opens the return line  19 , the oil pressure drops and when the force on the piston caused by the oil pressure exceeds the force from the spring  33 , the shaft  35  will push the ball  32  away so that the valve opens and the cylinder chamber  9  is put in communication with the return line  19 . 
   The features hitherto described with reference to  FIG. 1  are background art. 
   According to the present invention, the exhaust rocker arm  2  is made with a second piston cylinder device  40  comprising a cylinder chamber  41  spaced from the rocker arm end  7  and a piston  42  disposed in the cylinder chamber. As can be seen in the figures, the cylinder chamber  41  is essentially directed opposite to the cylinder chamber  9 , i.e. it opens upwards as seen in  FIGS. 1 and 2  and communicates with the first cylinder chamber via a channel  48 . As is particularly evident from  FIG. 2 , the piston  42  is concave as is the piston  10 . Between the bottom  43  of the depression in the piston  42  and a lock ring  44 , a helical spring  45  is tensioned, thereby loading the piston  42  towards the bottom of the cylinder chamber  41 . A second exhaust rocker arm  46  is mounted on a laterally extending portion  47  of the bearing bushing  18  non-rotatably joined to the first exhaust rocker arm  2  (see  FIGS. 3 and 4 ). At one end of the second rocker arm  46  there is a cam follower roller  49  rotatably mounted. The cam follower roller  49  is in contact with a schematically shown cam element  50  on the camshaft  6 . “c” designates the base circle of the cam element and “d” its top radius. At its opposite end designated  51  an adjustable spindle  52  is screwed in, which extends into the depression of the piston  42  and has a spherical end  53  held in a corresponding depression in a guide  54 . 
   As is particularly evident from  FIG. 4 , in the example shown the cylinder chamber  41  has the same cross-sectional area as the cylinder chamber  9 , which means that a pump stroke with a certain stroke length of the piston  42  results in the same stroke length in the piston  10 . Other embodiments with different cross-sectional areas for the cylinder chambers  9  and  41  are conceivable, but the stroke lengths for the pistons  10  and  42  will then be inversely proportional to their cross-sectional areas. The reactive forces, which can be different, from the two cylinder chambers  9  and  41 , form together with the lever lengths L 1  and L 3  a resulting reactive torque in the rocker arm  2 . The mechanical advantage of the rocker arms  2  and  46  differ however, firstly, by virtue of the fact that the cylinder chambers  9 ,  41  are placed at different distances from the rocker arm shaft  2  and, secondly, by virtue of the fact that the cam follower rollers  4  and  49  are mounted on their respective rocker arms at different distances from the rotational axis of the rocker arm. In the example shown in  FIG. 2 , the ratio L 2 /L 1  of the exhaust rocker arm  2  is ca 1:1.6, while the ratio L 4 /L 3  of the exhaust rocker arm  46  is ca 1:0.7. A suitable interval for the mechanical advantage of the rocker arm  2  can be ca 1:1.1–1.6 and for the mechanical advantage of the rocker arm  46  ca 1:0.7–1.1. 
   In normal drive mode operation, the valve  21  is open and the pistons  10  and  42  lie in their end positions shown in  FIGS. 1 and 2 . The transition to braking mode is effected by closing the valve  21  so that the pressure is built up in the hydraulic circuit  20 . The piston  10  is thereby displaced downwards to adjust the valve play to zero at the same time as the piston  42  is displaced upwards to an upper end position abutting against the lock ring  44 . The brake cam element  50  can be provided with, for example, one or two (not shown) cam lobes with the top radius “d” shown in  FIG. 2 , either only one for opening the exhaust valve  14  at the end of the compression stroke (the decompression) or one for opening the exhaust valve  14  at the last portion of the intake stroke (the charging) and one for opening the exhaust valve  14  at the end of the compression stroke (the decompression). During the angular interval, when first the first and then the second of these brake cam lobes strikes the cam follower roller  49  of the rocker arm  46  and the rocker arm  46  thereby presses against the piston  42  so that oil is pumped into the cylinder chamber  9  behind the piston  10  to press it down and open the exhaust valve, the cam follower roller  4  of the regular exhaust rocker arm  2  lies on the base circle “a” of the cam element  5 . By virtue of the above described difference in the leverage of the two rocker arms  2  and  46 , there will be a limited reactive torque in the regular rocker arm  2 , which is continually taken up by its cam follower roller  4  on the base circle “a” of the cam element  5  during charging and decompression. The regular exhaust rocker arm  2  thus does not move on its own during the charging and decompression, which is advantageous for the bearing bushing  18  since it cannot be subjected to load on one edge. The design results in the two rocker arms  2  and  46  together taking up the loads during the charging and decompression sequence, even if the extra exhaust valve rocker arm  46 , for brake mode operation, has to absorb the major part of the load and perform the work of opening the exhaust valves. 
   The diagram of  FIG. 5  shows the lift curve A of the exhaust valve and the lift curve B of the intake valve during normal drive mode operation. As can be seen by the shaded area C, the valve overlap is relatively small. The dashed line D illustrates the increase in exhaust valve lift when going from driving mode to braking mode by adjusting down the valve play to zero and using the described previously known technology with extra cam lobes on the regular cam. As is evident from the diagram in  FIG. 6 , showing the lift curves A and B during braking mode while using the described known technology, the valve overlap C increases markedly as compared to driving mode. This in turn leads to, as mentioned above, a relatively significant back-flow from the exhaust side to the intake side. 
   The diagram in  FIG. 7  shows the lift curve A of the exhaust valve and the lift curve B of the intake valve during braking mode, using a valve mechanism  1  according to the present invention. As can be seen by a comparison with  FIG. 5 , in this case there is no change in the regular lift curve A of the exhaust valve when changing from drive mode to brake mode and, consequently, the valve overlap C does not change, as can be seen by comparison. 
   The diagrams in  FIGS. 6 and 7  reveal, when compared, that the extra lifts A 1 , A 2  during brake mode are of equal height. The lifting height when using the described known technology is limited to the valve play, in practice at most ca 1 mm. The lift height when using the valve mechanism according to the invention is limited to what the space between the valve disc and the top of the piston permit, when the piston is in its uppermost position, and can be appreciably higher than that shown. Furthermore, the valve mechanism according to the invention can absorb greater forces than the previously known valve mechanism, which means that a higher differential pressure can be permitted over the exhaust valve, ca 70 bar as compared to ca 45 bar previously. With 5 bar of counter-pressure in the exhaust manifold, this means that the compression pressure can be allowed to be raised from ca 50 bar to ca 75 bar, which corresponds to an increase in the braking power by ca 30%.