NO834522L - PRESSURE-DELIVERED ENGINE DELAY DEVICE - Google Patents

Abstract

A compression release engine retarder for a multi-cylinder four-stroke cycle engine is disclosed. The retarder incorporates an hydraulic pulse generator (85) which in one embodiment comprises a multi-chamber positive displacement pump of the piston (146) and cylinder (144) type which is positively driven at engine speed or at half engine speed in synchronism with the engine crankshaft (86). Means (84,96,97) are provided to adjust the timing of the hydraulic pulses so as to control precisely the opening of the engine exhaust valves and to maximize the compression release retarding power developed by the engine. Additional means (182,184,186,188,190,192,194) are provided to control the timing of the hydraulic pulses in response to the boost pressure in the engine inlet manifold produced by the engine turbocharger.

Description

The present invention relates to an engine brake for use with a multi-cylinder four-stroke internal combustion engine. More specifically, the invention relates to a hydraulic pulse-generating system which is arranged to open the exhaust valves in an internal combustion engine in order to provide a compression-triggered engine braking function when braking is to be carried out.

It is known that drum brakes and disc brakes are capable of

to absorb a large amount of energy over a short time. The absorbed energy is converted into heat, which rapidly increases the temperature of the braking mechanism to a level where the friction pads and other parts of the mechanism can have reduced effectiveness. Where repeated use of brakes under these conditions is not practically feasible, auxiliary braking devices have had to be used.

Such auxiliary devices include hydraulic or electrodynamic braking systems where the vehicle's kinetic energy is converted by fluid friction or magnetic eddy currents into heat that can be carried away through appropriate heat exchangers. Other braking assist systems include (a) exhaust brakes that block the flow of exhaust gases or air through the exhaust system and (b) compression-actuated brake mechanisms where the energy required to compress the intake air during the compression stroke of a four-stroke engine is dissipated by the compressed air is released through the exhaust system under the engine's expansion layer. In the case of compression actuated brake systems,

part of the vehicle's kinetic energy is released through the engine's cooling system, while another part of the kinetic energy is led away through the engine's exhaust system. With exhaust braking, the vehicle's kinetic energy will only be carried away as heat through the cooling system.

An important advantage of compression-triggered braking and exhaust braking compared to hydraulic and electrodynamic braking systems is that the latter two braking systems require dynamos or turbine equipment which can be space-consuming and expensive, compared to the mechanism that must be used for normal exhaust braking or compression-triggered braking. A typical compression actuated braking mechanism for internal combustion engines is shown in US Patent No. 3,220,392, while an exhaust brake system is described in US Patent No. 4,054,156. Other mechanisms that have compression-triggered braking are described in US Patent Nos. 3,809,033, 3,786,792, 3,547,087, 3,859,970 and 3,357,312.

Another advantage of a compression actuated braking system, such as that described in US Patent No. 3,220,392, is that it uses the existing valve mechanism and only requires the addition of a master piston and a slave piston for each cylinder along with an associated control system.

The problem with using the engine's valve lifters or fuel injection equipment as a source of movement as is the case in US patent no. 3,220,392, is that forces can be introduced into the valve mechanism that are different from and perhaps greater than the forces assumed by the original engines -calculations, which can lead to damage to the injection equipment. With such a device, moreover, the timing of the compression release will necessarily be determined by valve functions that take place elsewhere in the engine, functions that may not be optimal in terms of the braking effect.

One way to solve the foregoing problem of excessive stress in the valve mechanism is shown in US patent no. 4,271,796 where a pressure relief device is built into the hydraulic system to limit the maximum pressure that can occur in this . The use of a safety valve for this purpose is described in US patent no. 4,150,640.

The inventors have found that (a) improved engine braking and

(b) complete non-elimination of the risk of damage to

the engine's valve mechanism can be achieved using hydraulic pulses produced by a hydraulic pulse generator when this is operated directly in synchronism with the engine's crankshaft. More specifically and in accordance with the invention, a compression actuated engine brake has been arrived at for use in a multi-cylinder, four-stroke internal combustion engine with a crankshaft, intake and exhaust manifolds and at least one exhaust valve for each cylinder and at least one slave piston which during a braking operation shall open the exhaust valve to which it is connected and the invention is characterized by a rotating hydraulic pulse generator operated in synchronism with the engine crankshaft to sequentially deliver pulses of hydraulic fluid under pressure to predetermined slave pistons by means of channels connecting the slave pistons and the hydraulic pulse generator and a common valve control which is connected to the channels for, when the valve control is out of operation, to create a low-pressure condition in the fluid in the channels that occurs during fuel supply and when the valve control is in operation during braking, to create high-pressure conditions in the channels for the fluid.

The hydraulic pulse generator produces hydraulic pulses that are timed to coincide with the beginning of the expansion stroke or power stroke for each cylinder in the four-stroke engine. The hydraulic pulse generator can be driven by a shaft that works at engine speed or at half this speed. The timing can be set over a wide range by changing the position of the pulse generator and its drive shaft in relation to the engine's crankshaft, while the size of the pulses can be determined by the construction of the pulse generator. Additionally, timing can be adjusted as a function of added pressure to more precisely control the timing of compression release.

The invention is characterized by the features reproduced in the claims and will be explained in more detail in the following with reference to the drawings, where Figure 1 is a curve showing the exhaust valve movement during different forms of engine operation,

figure 2(A) is a curve showing the displacement in a pulse generator with a positive displacement as a function of time,

figure 2(B) schematically shows a pulse generator with positive displacement in the form of a cylinder and a piston driven by a crankshaft,

Figure 3(A) schematically shows the construction of a compression-triggered engine brake with a three-unit pulse generator operating at engine speed, each unit being arranged to alternately open the exhaust valves of two cylinders in a six-cylinder engine at the start of braking in cylinder No. 6 ,

Figure 3(B) shows the mechanism of Figure 3(A) at an early stage during the pulse when the slave piston for cylinder No. 1

has come into contact with a stop device,

figure 3(C) shows the mechanism in figure 3(A) at maximum pulse pressure when the slave piston for cylinder no. 6 has opened the exhaust valve for this cylinder,

Figure 3(D) shows the mechanism of Figure 3(A) when the pulse has ended and the slave pistons for cylinders No. 1 and 6 have returned to their initial position,

Figure 3(E) shows the mechanism of Figure 3(A) at an early stage during the formation of the pulse for cylinders No. 1 and 6 when the slave piston for cylinder No. 6 has come into contact with a sto<p>pin direction,

figure 3(F) shows the mechanism in figure 3(A) at the maximum pulse pressure when the slave piston for cylinder no. 1 has opened

the exhaust valve for this cylinder,

figure 4(A) shows on an enlarged scale the pulse generator with additional construction details and with the solenoid in the inactive position as the timing mechanism in a low auxiliary position,

figure 4(B) shows on an enlarged scale the pulse generator with solenoids in active position and the time control mechanism in high auxiliary position,

figure 5 shows on an enlarged scale an alternative design for a pulse generator with six units, intended to be driven at half the speed of the engine and the generator has massive pistons, a one-way valve and a mechanism to prevent excessive speeds,

figure 6 is a collection of curves showing the effect of boost pressure from a turbocharger on exhaust valve timing,

figure 7 shows, seen from the end and partially in section, a broken piece of an alternative pulse generator with a gear pump with positive displacement and

figure 8 shows a cross-section taken along the line 8-8 in figure 7.

The operation of compression-triggered engine braking is now well known among professionals and need not be described in detail here. Quite briefly, however, with compression-triggered braking, use is made of the movement of an element in the valve mechanism or in the mechanism for fuel injection, for example a pressure pipe that moves close to the desired timing and this movement is used to drive a main piston. The master piston is hydraulically connected to a slave piston which, during braking, will open an exhaust valve (or dual exhaust valves) at or near the beginning of the expansion stroke or the working stroke for a four-stroke engine.

When the compression actuated engine brake is in operation during braking, the fuel supply is automatically shut off so that the engine only pumps air. By opening the exhaust valves at the beginning of the expansion stroke or working stroke of the engine, the work that has gone into compressing the air during the compression stroke will not be recovered during the expansion stroke, but will be led away through the engine's exhaust system and cooling system.

The energy absorbed by the engine in this way can be measured with an electric dynamometer expressed as the brake horsepower the engine has developed. The total braking force naturally includes energy losses represented by the internal friction in the engine and all other losses linked to the engine's equipment. The absolute value of the brake horsepower developed during the compression-triggered braking can be a significant portion of the total horsepower developed by the engine when operating normally.

Reference should now be made to figure 1, which shows the different exhaust valve movements. In figure 1, the friction valve movement is shown as the ordinate, while the crankshaft angle starting at top dead center (TDC) is the abscissa. Curve 10

shows the normal operation of the exhaust valve at the beginning of the engine's exhaust stroke. Curve 12 shows a typical exhaust valve movement for an engine equipped with a compression actuated brake designed to open the exhaust valves at the beginning of the engine expansion stroke as close to top dead center as is practical. More specifically, curve 12 shows the movement of the exhaust valve

obtained from the fuel injection pressure pipe. The movement shown with curve 12 can be produced with the device described in US patent no. 4,271,796. Curve 16 shows the desired optimal movement of the exhaust valve which will optimize the dissipation of the energy stored in the engine during the compression stroke. It will be seen that the precise timing of the compression actuated exhaust valve

opening and is chosen so as to optimize the compression work and minimize the recovery of this work during the subsequent expansion stroke. This requires a rapid opening of the valve near the end of the compression stroke and close to the top dead center position. However, when the valve operating movement is derived from another part of the engine, e.g. the pressure pipe in the fuel injection device or the exhaust pressure pipe for another cylinder,

it may be impossible to achieve optimal operation of the compression-triggered mechanism.

Figure 2(A) schematically shows a curve for volume versus time for a displacement pulse generator, e.g. the piston pump 18 in figure 2(B). The pump 18 comprises a shaft 20

with a crank arm 22 which in turn drives a piston rod 24 and this is connected to a piston 26 which is forced to move in a cylinder 28. It will be seen that it; >positive part of the displacement curve 30 corresponds to the optimal valve lift curve 16 in Figure 1. Part of the displacement curve 30, e.g. the one above line 32 can therefore act as a valve actuation pulse if timed correctly.

Reference should now be made to figures 3A-3F which show the various stages of the operation of a compression-triggered engine brake where a three-part pulse generator is used according to the invention. The engine block of a six-cylinder four-stroke internal combustion engine is shown in broken pieces at 34. Cylinder No. 1 is indicated at 36 and contains piston 38. Cylinder No. 6 is shown at 40 and contains piston 42. In the embodiment of the engine shown, the pistons for cylinder no. 1 and cylinder no. 6 to be at top dead center (TDC) at the same time and they move in tandem. However, when piston No. 1 (38) moves upward on an exhaust stroke,

piston No. 6 (42) will move upwards in a compression stroke. It goes without saying that cylinders no. 2 and 5 and cylinders no. 3 and 4 work in pairs in the same way. The way it works

the present invention will therefore be limited to a description of the various events linked to cylinders no. 1 and no. 6, it being understood that the same will happen with the other cylinders, but at other times.

The intake valve 44 for cylinder no. 1 (36) is normally biased towards the closed position by a valve spring 46. Similarly, the intake valve 48 for cylinder no. 6 (40) is biased towards its closed position by a valve spring 50. The intake valves 44 and 48 are controlled with rocker arms (not shown).

The exhaust valve 52 for cylinder no. 1 (36) is biased towards

its closed position by a valve spring 54. The exhaust valve 52 is opened when the engine is running normally, by means of a rocker arm 56 which is controlled from the engine camshaft (not shown) and acts on a cross head 58. A slave piston 60 is mounted for reciprocating movement in a slave cylinder 62 formed in a housing 64 attached to the cylinder head 63 which, in turn, is attached to the engine block 34. The slave piston 60 is biased away from the cross head 58 by a spring 65.

Similarly, the exhaust valve 66 for cylinder No. 6 (40) is biased towards its closed position by a valve spring 68. The exhaust valve 66 is opened when the engine is operating in the normal manner by a rocker arm 70 which acts through the crosshead 72. A slave piston 74 is mounted forward -and reciprocating movement in a slave cylinder 76 which is formed in the housing 64. The slave piston 74 is also biased away from the crosshead 72 by a spring 78.

A line 80 forms a connection between the slave cylinders

62, 76 and the master cylinder 82 which is designed in the housing 84

for the pulse generator 85. The main cylinders 82 are arranged in the housing 84 radially in relation to the pulse generator's drive shaft 86, which can be one or another shaft for operating the engine's accessories or additional equipment and which rotates with the engine's

speed and is driven directly or indirectly from the engine's crankshaft, so that it moves synchronously with it. Master cylinders 88 and 90 are also arranged radially in relation to the drive shaft 86 and separated from each other and from master cylinder 82 by arcs of 120°. A line 92 leads from the master cylinder 88 to a pair of slave cylinders (not shown)

which are identical to slave cylinders 62 and 76, but are connected to engine cylinders No. 2 and No. 5. Similarly, a line 94 leads from master cylinder 90 to a third pair of slave cylinders (not shown) which are identical to slave cylinders 62 and 76, but is connected to the engine's cylinders No. 3 and No. 4. An eccentric disc 94 is locked onto the drive shaft 86 which, in Figures 3A-3F, is driven counter-clockwise. It can be seen that rotation of the drive shaft 86 moves the eccentric disk 96 towards a ring 97 which comes into contact with the main pistons 98, 100 and 102 which are stored for reciprocating movement in the main cylinders 82, 88 and 90, so that the pistons in turn move radially outwardly from the drive shaft 86. The movement imparted to the main pistons 98, 100 and 102 by the eccentric disc 96 is functionally equivalent to the movement produced by the cranking mechanism shown in Figures 2A and 2B. Other details relating to the construction of the pulse generator 85 will appear from the following description with reference to figures 4 and 5.

In Figure 3A, the piston 38 of cylinder #1 begins to move upward on an exhaust stroke, while the piston 42 of cylinder #6 begins to move upward on a compression stroke. The ring 97 driven by the eccentric disc 96 also begins to move the main piston 98 in a radially outward direction. At this point, the exhaust valves 52 and 66 are still closed, but the main piston 98 has moved sufficiently to isolate the line 80 and the hydraulic fluid it contains from the low pressure hydraulic supply source inside the eccentric chamber 104 of the pulse generator. This is shown schematically in figure 3A in that the L-shaped passage 106 in the main piston 98, which forms a connection between the line 80 and the eccentric chamber 104 when the main piston 98 is in its radially inner position, is now out of connection with the eccentric chamber 104 .-Figure 3B shows the compression-triggered mechanism at a somewhat later time when the rocker arm 56 has opened the exhaust valve 52 in a normal valve opening sequence as required for the normal exhaust stroke in cylinder No. 1 (36). When the exhaust valve 52 and its crosshead 58 are displaced downwards, the movement of the slave piston 60 will only be held back by the bias in the spring 65. The consequence of this is that when the pressure in the line 80 builds up due to the radially outward movement of the main piston 98, hydraulic fluid will enter the slave cylinder 62 and drive the slave piston 60 downwards until a stop 108 on the slave piston 60 strikes a stop device 110 found in the housing 64. It will be seen that the movement of the slave piston 60 which takes place while the hydraulic fluid in the line 80 still has a relatively low pressure, has no effect on the exhaust valve 52 which, during this time, is affected by the rocker arm 56. As soon as the pressure in the line 80 becomes sufficiently high to overcome the bias in the spring 65 in the slave cylinder 62, it will also overcome the bias from the spring 78 in the slave cylinder 76 and bring the slave piston 74 into contact with the cross head 72 and thereby take up the idle or clearance that normally fi between each slave piston and the associated valve junction head when the line 80 is in communication with the low pressure prevailing in the eccentric chamber 104. However, the slave piston 74 will not open the exhaust valve 66 at this time, because the force required to open the valve 66, greatly exceeds that required to move the slave piston 74 against the bias of the spring 78 and therefore only the latter movement will take place. Figure 3C shows the compression release mechanism when the main piston 98 has reached its maximum stroke. At this point, the hydraulic pressure in line 80 has reached a high level and is sufficient to overcome the bias from spring 68 and the forces on valve 66 due to the air compressed in cylinder No. 6 (40) so that the exhaust valve 66 opens. It should be pointed out that movement of the main piston 98 should be controlled by the position of the eccentric disk 96, so that the exhaust valve 66 opens close to the ØD position of the piston 42 in order to optimize the number of braking horsepower the engine develops. This ensures that the optimum amount of work has been done by compressing the air during the compression stroke and that the smallest possible part of this work is recovered during the following expansion stroke.

Figure 3D shows the following situation for the mechanism where the eccentric disk 96 has moved to a point where the main piston 98 has returned to essentially the position it had in Figure 3B. As a result of the piston 98's return, the pressure in the slave cylinder 76 will drop to a point where the valve spring 68 will close the valve 66. Meanwhile, the rocker arm 56 will swing back to its initial position and the valve spring 54 will close the exhaust valve

52. As the exhaust valve 52 closes, the slave piston 60

be driven upwards. The curved movement shown by the dashed line 112 between Figures 3A and 3D, it is clear that the shaft 96 has moved through an angle somewhat less than 90°. During the next drive shaft movement of approximately 30°, the main piston 98 will be withdrawn so much that the line 80 comes into contact with the L-shaped passage 106 and thus with the eccentric chamber 104, so that the pressure in the line 80 and in the eccentric chamber 104 offset. Under these conditions, the bias from the springs 65, 78 will be sufficient to drive the slave pistons 60 and

74 away from the cross heads 58 and 72. It will be apparent from the stippling of the dotted line 112 in Figure 3D that continued crank shaft rotation, that is rotation of the pulse generator drive shaft 86 will cause the eccentric disc 96 to lead the main piston 100 radially outwards and thus begin a series of operations in cylinders no. 2 and no. 5, corresponding to those described above in connection with cylinders no. 1 and no. 6. Continued rotation of the pulse generator drive shaft 86 will then lead to a repetition of these operations as well

for cylinders No. 3 and No. 4 as a result of the movement of the main piston 102.

Figure 3E shows the conditions in the compression triggering mechanism when the eccentric disc 96 has again reached the main piston 98. At this time, cylinder No. 6 (40) begins its exhaust stroke, while cylinder No. 1 (36) begins its compression stroke. Under these conditions, the rocker arm 70 swings down towards the crosshead 72 to open the exhaust valve 66. As the pressure in the line 80 then begins to build, hydraulic fluid will be driven into the slave cylinder 76 and drive the slave piston 74 down until the facility 114 on the slave piston 74 hits the stop device 116 which is found in the housing 64. As noted above, hydraulic fluid will also enter the slave cylinder 62 and drive the slave piston 60 down until it comes into contact with the cross head 58. Figure 3F corresponds mainly to Figure 3C, which is described above, and reproduces that time when the main piston 98 has moved to its extreme position and has raised the pressure in the line 80 to a point where the slave piston 60 opens the exhaust valve 52. This again takes place close to the top dead center (TDC) position. Continued rotation of the pulse generator drive shaft 86 results in a drop in pressure in line 80 which enables the exhaust valve 52 to close and eventually the slave pistons 60 and 74 will return to their respective starting positions. This series of operations is repeated for cylinders No. 2 and No. 5 and then for cylinders No. 3 and No. 4. It will be seen that after two revolutions of the pulse generator drive shaft 86, which corresponds to two revolutions of the engine crankshaft, each of the engine's cylinders have undergone a compression release close to top dead center position for the piston at the end of the compression stroke.

Reference should now be made to Figures 4A and 4B, which reproduce on an enlarged scale the pulse generator shown schematically in Figures 3A to 3F. Figures 4A and 4B also include a timing mechanism, a pressure relief mechanism,

a solenoid switch for controlling the pulse generator and an alternative design for the main piston. The pulse generator 85 comprises a housing 84 with an eccentric chamber 104. The main cylinders 82, 88 and 90 are radially 120° apart around the eccentric chamber 104. A passage 118 carries low-pressure hydraulic fluid, e.g. oil to the eccentric chamber 104

and then to the master cylinders 82 , 88 and 90 , as well as to the line 80 and the slave cylinders 62 and 76 . Similarly, hydraulic fluid is directed to the passages 92<q>g 94 and their associated slave cylinders.

The action of the pulse generator 85 is controlled by a solenoid valve 120 which has a solenoid coil 122, an armature disc 124, an operating rod 126 and a ball valve 128 with valve seat 130. The ball chamber 122 is connected to a drain (not shown). On the outside of the ball valve seat 130 there is a disk valve chamber 134. Passengers 136, 138, 140 lead from the disk valve chamber 134 to the main cylinders 82, 88, 90 respectively. A valve disk 142 is movably arranged in the disk valve chamber 144. When the solenoid valve 120 is disengaged as shown in Figure 4A, the armature disc 124 is released from the coil 122

and the operating rod 126, as well as the valve ball 128 move to the right (seen in Figure 4A), so that hydraulic fluid from the disc valve chamber 134 can flow out of the housing 85. This causes the disc valve 142 to move to the right, whereby the disc valve chamber 134 acts as a distribution pipe as regards the passages 136, 138 and 140. It will be seen that when the disc valve 142 is open, hydraulic fluid pumped in sequence by the main pistons 98, 100, 102 will flow back to the eccentric chamber 104. It will also be seen that when one main piston pumps, the other two main pistons (and their associated passages) will empty.

However, when the solenoid is supplied with current (as shown in Figure 4B), the ball valve 128 will seal against the seat 130 upon movement of the operating rod 126 and the armature disc 124 and pressure will build up on the disc valve 142 and thus close the low pressure passages. If, as shown in Figure 4B, the eccentric disc 96 drives the main piston 98, high pressure from the passage 136 will cause the disc valve 142 to close the low pressure passages 138 and 140. This will cause hydraulic fluid at high pressure to enter the line 80 and perform the functions of is described above.

As shown in Figures 4A and 4B, each of the main pistons 98, 100, 102 comprises three parts, namely an outer cylindrical shell 144, an inner sliding piston 146 and a snap ring 148 which limits the mutual sliding movement between the inner piston 146 and the cylindrical shell 144. It will be seen that the eccentric disk 96 and the ring 97 first drive the piston 146

in sealing contact with the cylindrical shell 144 and then both of these parts drive the main piston 98 (or 100,

102) radially outwards to create the necessary hydraulic pressure in line 80 (or 92, 94).

Figures 4A and 4B also show a pressure regulating mechanism 150 which is connected to the line 80. It should be pointed out that similar mechanisms can of course be placed in the corresponding line 92 leading to cylinders no. 2 and no. 5 and the line 94 leading to cylinder no. 3 and No. 4. The pressure regulation mechanism 150 comprises a housing 152 which has

a diametrically extending passage 154 in connection with the line 80 and, in connection with this, an axial passage 156 which ends in a cylindrical chamber 158. A piston 160 is arranged for reciprocating movement in the cylindrical chamber 158 in the housing 152 and is biased against the axial passage 156 by means of a relatively stiff spring 162. The spring 162 and the piston 160 are held in place by a cap 164 which is screwed to the housing 152. When in operation, the pressure regulating mechanism 150 will act

in a way like a shock absorber to -increase the volume of the hydraulic circuit when the pressure exceeds a predetermined level and thus limit the maximum pressure that can be produced by the pulse generator 85. Those skilled in the art will see that when a hydraulic circuit is relatively short , the hydraulic fluid mass will act so that the fluid can be regarded as essentially incompressible and some form of pressure regulating mechanism is then desirable to relieve excess pressure in the system. However, when the hydraulic lines are relatively long, the compressibility of the hydraulic fluid itself can be sufficient to form its own expansion chamber. Instead of the pressure regulating mechanism described above, another device can e.g. a Bourdon tube which expands with increasing pressure is used. In all cases, whether a specific pressure regulating mechanism is used or not, care must be taken to ensure that the lines 80, 92 and 94 are of approximately the same length and contain substantially the same amount of hydraulic fluid so that the effect on each slave cylinder is the same.

The pulse generator 85 can also include a timing mechanism by means of which the timing of the hydraulic pulses can be varied as a function of the auxiliary pressure developed by the engine's turbocharger (if the engine is equipped with this). It should be emphasized that the auxiliary pressure at the engine's inlet manifold varies with the speed of the turbocharger and that the mass of air introduced into the engine is also a function of the auxiliary pressure. Furthermore, the pressure developed in the engine cylinder during the compression stroke will vary with the mass of air drawn into the cylinder during the intake stroke, while the force required to open an exhaust valve is a function of the pressure in the cylinder to which the exhaust valve is connected to.

Figure 6 is a collection of curves where the force required to open an exhaust valve is plotted as the ordinate in relation to the crankshaft angle as the abscissa. Curve 166 shows how this force varies with crankshaft angle at low auxiliary pressure, while curve 168 is a corresponding curve for high auxiliary pressure. It can be seen that an engine can work with different auxiliary pressures within the range represented by curves 166 and 168.

If it is desired to start and open the exhaust valve at

a specific point, e.g. 15° before top dead center (TDC) to optimize the braking effect at the compression release, it is clear that the first movement of the main piston 98 (or 100 or 102) must lie before this point. The slope of line 170 represents the rate at which the pressure in the hydraulic system (eg lines 80, 92 and 94) builds up as a function of time (or crankshaft angle) at low auxiliary pressures to exert that force which is necessary to open the exhaust valve indicated at point 172. The point of intersection between line 170 and the axis (point 174) shows the crankshaft angle at which the movement of the main piston must begin in order to open the exhaust valve at point 172. Similarly, point 176 on curve 168 indicates for high auxiliary pressure the force required to open the exhaust valve 15° before top dead center when you have high auxiliary pressure. To achieve this force, the pressure in the hydraulic system rises along line 178 and intersects the axis at point 180. From figure 6 it is clear that a mechanism for adjusting the timing of the movement of the main piston automatically in relation to the auxiliary pressure would be desirable. Such a mechanism is built into figures 4A and 4B. A boss 182 integral with the housing 180 has a cylindrical bore 184 which communicates at one end with a channel 188 through a passage 186 and the channel 188

is connected to the engine's intake manifold (not shown). The intake manifold is at all times under the auxiliary pressure produced by the engine's turbocharger (not shown). A piston 190 is placed in the cylindrical bore

184 and is biased against the passage-186 with a spring 192.

A cam 194 with three waves is mounted on the drive shaft 86 in such a way that it can be rotated independently of the eccentric disc 96 and the ring 97. The cam 194 is axially offset from the eccentric disc 96 so that each wave of the cam 194 comes into contact with the cylindrical sliding sleeve 144 on one of the pistons 98, 100 or 102, but does not touch the internal sliding piston 146. A link 196 connects the cam 194 and the piston 190.

At low auxiliary pressure, the piston 190 is biased to the left (seen in figure 4A) and it will thereby rotate the cam 194 counter-clockwise to a position where the cylindrical shells 144 for the pistons 98, 100 and 102 are driven radially outwards from the drive shaft 86. As shown in Figure 4B, the piston 190 at high auxiliary pressure is driven to the right and the cam 194 is rotated clockwise. Such rotation of the cam 194 enables the cylindrical shells 144 for the pistons 98, 100 and 102 to move radially inward towards the shaft 86. It will be seen that movement of the cylindrical shells 144 inwards leads to movement of the main pistons 98, 100 and 102 relative to the drive shaft 86, a movement which is synchronous with the movement of the engine crankshaft. In this way, the timing of the movement of the main pistons will be a function of the auxiliary pressure. The result of this is that the exhaust valves will be opened during a compression-triggered braking at some point (e.g. 15° before ØD) regardless of changes that may occur in the auxiliary pressure as a result of e.g. changes in engine speed.

Reference should now be made to figure 5, which reproduces an alternative embodiment of the invention, intended for use together with a pulse generator with drive shaft 86' which operates at half the speed of the engine's crankshaft. In Figure 5, parts that are common to Figures 3 and 4 are designated with the same reference number, while parts that have been modified are designated with (') reference number. The 84' house has six

has six master cylinders 198, 200, 202, 204, 206 and 208

which during compression-triggered braking regulates the opening of the exhaust valves i:f or'respectively cylinder no. 6, 2, 4, 1, 5

and 3. The master cylinders 198, 200, 202, 204, 206 and 208 are in the same order as the engine firing order. Passages 210, 212, 214, 216, 218 and 220 form a connection between the master cylinders 198, 200, 202, 204, 206 and 208 and the chambers for six one-way ball valves, of which only three are visible, namely chambers 222, 226 and 230. The other three ball valve chambers are behind the chambers 222, 226 and 230 and are therefore not visible in Figure 5. The chamber for each one-way ball valve is equipped with a seat 234, a valve ball 236

and a spring 238 which normally biases the valve ball 236.

Each of the six ball valve chambers is connected to

a manifold chamber 240 through six passages of which only three are visible, namely the passages 242, 246 and 250. The other three passages lie behind the passages 242, 246 and 250 and are therefore not visible in Figure 5. An anchor 254 is placed in the manifold chamber 240.

A hydraulic fluid passage 256 is connected between the armature chamber 240 and the passage 258 supplies hydraulic fluid to the pulse generator 85'. The solenoid coil is shown at 260

and the wires to the solenoid coil are 262 and 264. A relatively stiff spring 266 sits in a bore 268 in the solenoid core and will normally bias the armature 254 away from the coil 260. Control rods 270 are placed between the armature 254 and the valve balls 236 in passages 242, 246 and 250 and in the three passages not shown.

The main pistons 272, 273, 276, 278, 280 and 282 are in the main cylinders 198, 200, 202, 204, 206 and 208 and are driven by a ring 97 and the eccentric disc 96 from the drive shaft 86'. A channel 284 forms a connection between the master cylinder 198 and a slave cylinder (not shown) which is connected to the engine's cylinder no. cylinders No. 2, 4, 1, 5 and 3.

A pressure regulating mechanism 150 similar to that which is . described with reference to figures 4A and 4B, can be placed in each of the channels 284, 286, 288, 290, 292 and 294.

A circular cam 296 is placed on the shaft 86' at the side

of the eccentric disk 96 and can either be fixed to the shaft 86' or be freely rotatable in relation to this. In all cases, the cam 296 will form a restriction to the inward movement of each of the main pistons 272, 274, 276, 278, 280 and 282.

The operation of the mechanism shown in Figure 5 is as follows: when the solenoid coil 260 is out of action (as shown in Figure 5) the one-way ball valves 236 are held in the open position. The result of this is that movement of the pistons 272, 274, 276, 278, 280 and 282 in sequence in their respective cylinders 198, 200, 202, 204, 206 and 208 causes hydraulic fluid to "circulate between the master cylinders via passages 210, 212 . 294 and the master cylinders associated with these passages. The system will remain filled with hydraulic fluid because any leakage past the massive master pistons 272, 274, 276, 278, 280 and 282 into the eccentric chamber 104 will be replenished from the supply passage 258. a drainage path (not shown) from the eccentric chamber 104 to a reservoir for hydraulic fluid, not shown, is also arranged.

However, when the solenoid coil 260 is energized, the armature 254 will be attracted to the coil 260 against the action of the spring 268

and will thereby secure the one-way valves 236 in a position to close against the seats 234. As the passages 210, 212, 214, 216, 218 and 220 are closed, the movement of the pistons 272, 274, 276, 278, 280 and 282 in sequence will create a corresponding catch up with

pressure pulses in the respective channels 284, 286, 288, 290,

292 and 294 which are timed so that they affect the slave pistons which are connected to cylinders no. 6, 2, 4 respectively,

1, 5 and 3 in the engine. It will be seen that the exact timing of the pressure pulses depends on the mutual position between the drive shaft 86' and the eccentric disc 96. As the shaft 86' is driven at half the rotational speed of the engine, one revolution of the shaft 86' will produce a pulse at each cylinder in the engine when the piston in this cylinder is near the end of its compression stroke. If it is desired to vary the timing of the pressure pulses as a function of the auxiliary pressure, the cam 296 can be provided with six waves of the same shape as the three waves on the cam 194 shown in Figures 4A and 4B. The setting of the cam 296 can be regulated by a piston and a link mechanism which is affected by the auxiliary pull as shown in Figures 4A and 4B. Under these conditions, it will naturally be necessary for the cam 296 to be freely rotatable in relation to the shaft 86'.

Modified cross heads can be used to open the exhaust valves during braking, for example as tin torn in the US

patent no. 4,399,787 and South African patent 80/7495.

Reference should now be made to figures 7 and 8, where one sees an alternative pulse generator 376 which has a gear pump with positive displacement. When used on a six cylinder engine, the pulse generator 376 includes an internal gear element 378 with six teeth and is designed to engage an eccentrically mounted gear 380 with five teeth. The gear wheel 380 is stored on an eccentric disc 382 which is locked to a shaft 384 and this is driven directly at half the engine's speed. More specifically, a splined ring 386 (Fig. 8) is attached to the shaft 384 and locked thereto by a washer 388 and a cap screw 390. The splined ring 386 mates with internal splines in an auxiliary drive shaft 392 which is driven at half the engine crankshaft speed . If other equipment is also to be operated from

the auxiliary drive shaft 392 can have a further key ring 394

attached to the opposite end of the shaft 384.

The main part of the pulse generator comprises a fastening plate 396 and

a main housing that attaches to the engine block (not shown) by cap screws 466 and a rear housing 400. The rear housing 400,

the internal gear element 378 and the main housing 398 are fixed together with six head screws 402 which stand between the roots of the six waves or teeth of the internal gear 378. An annular groove 404 is formed in the rear housing 400 and is in communication with a supply channel 408

for hydraulic fluid through a passage 406. The annular groove 404 has a smaller maximum diameter than a circle tangent to the waves in the internal gear 378, so that when the external gear 380 rotates in the internal gear 378, it sweeps over the annular groove 404 and encloses hydraulic fluid in the chamber formed by the teeth of the gears 378 and 380 and the surfaces of the housings 398 400. It should be noted that the diameter of the annular groove 404 determines the volume of the chamber 410 and the point where the hydraulic fluid can be pressurized. Thus, the diameter of the annular groove 404 will determine the position of line 32 in the diagram of Figure 2A. An annular groove 412 corresponding in size to the groove 404 is formed in the housing 398. O-rings 414, 416 can be placed between the internal gear 378 and the housings 400, 398 to prevent leakage of hydraulic fluid.

Passages 418 are formed in the housing 398 next to the root portions of each of the six teeth in the internal gear 378. The passages 418 are connected to an annular chamber 420 between the main housing 398 and the attachment plate 396 which acts as a branch pipe opposite the passages 418 and the various chambers 410 in the positive reinforcement pump. The annular manifold chamber 420 can advantageously be tightly closed by means of O-rings 422 and 424 which are placed between the surface 396 and the main housing 398. An annular valve 426 is freely movable in the annular chamber 420 and is biased against the housing 398 by a series of springs 428 which sits in blind hole 430 in surface 396.

A passage 432 forms a connection between the annular chamber 420 and a ball valve chamber 434 (Fig. 7) containing a ball valve 436. The passage 438 connects the ball valve chamber 434 and a supply line 440. A solenoid 442

is placed in the surface 396 and comprises a solenoid coil 444,

a core 446, a disk-shaped armature 448 which lies loosely in an armature chamber 450 and a pin which is placed in the core 446 between the armature 448 and the ball valve 436.

When solenoid coil 444 is energized, armature 448 will drive pin 452 against valve ball 436 to prevent flow of hydraulic fluid from passage 432 to passage 438 and from there to supply line 440.

Each passage 418 is connected to an outlet channel 454 which leads to an armature 456 in the main housing 398. The armatures 456 are arranged to occupy respectively the channels 284, 286, 288, 290, 292 and 294 (Figure 5) and the channels are connected to the slave cylinders which are linked respectively to cylinder no. 6,

2, 4, 1, 5 and 3 in the engine.

A passage 458 in the bottom of the housing 400 is connected

with a pipe 460 and a passage 464 found in the plate 398. The pipe 460 can advantageously be tightly connected to the housing 400

and plate 396 with O-rings 462. It is seen that during use, rotation of auxiliary drive shaft 392 will drive shaft 384 and eccentric disk 382, thereby rotating the five-tooth external gear 380 inside the six-tooth internal gear 378. Pulses of hydraulic fluid will be fed in order

from each of the chambers 410 through the passages 418 and into the annular chamber 420. Some of the hydraulic fluid will be circulated between the chambers 410 through the passages 418 and the annular manifold chamber 420, while part of

the hydraulic fluid will flow through the passages 432

and 438 to the supply line 440 (see figure 7). Fluid that may leak out of chambers 410 and along shaft 384 will fall into passage 458 and will be returned to the engine's supply of hydraulic fluid through tube 460 and passage 464. Because passage 432 is open when the solenoid is inoperative, it not possible to develop any particular pressure in the outlet cannulas 454.

However, when the solenoid 442 is activated, the ball valve 436 will close the passage 432 and allow a pressure build-up in the annular manifold chamber 420. This pressure will cause the annular valve 426 to seal against the passages 418 which are under low pressure, so that the pulse of hydraulic fluid from one chamber 410 will create pressure in a passage 418, an outlet channel 454 and be led to one of the channels 284, 286, 288, 290, 292 or 294 and the associated slave cylinder. It should be appreciated that continued rotation of the auxiliary drive shaft 392 will result in a series of pulses of hydraulic fluid in each of the channels 284, 286, 288, 290, 292 and 294 and in their associated slave cylinders. As the auxiliary drive shaft 392 rotates at one-half of the engine crankshaft revolutions, it will be seen that during two revolutions of the crankshaft, a hydraulic pulse will be fed to each of the channels 284, 286, 288, 290, 292 and 294

and to the associated impact cylinder. By correctly adjusting the key ring 386 in relation to the auxiliary drive shaft 392,

the pulses can be timed so that each exhaust valve in the engine opens for a predetermined time close to the top dead center position of the piston at the end of the compression stroke for the cylinder to which the exhaust valve belongs.

The preceding description has assumed that the engine is six cylinders, but it is of course that the pulse generator shown in Figures 7 and 8 can be modified so that it can be used for engines with any number of cylinders by ensuring that the generator has an internal gear 378 with the same number of teeth as the number of cylinders in the engine, while the external gear 380 has one tooth less than the number of cylinders in the engine.

The pulse generator in Figures 7 and 8 has a chamber 410 for

each engine cylinder. If desired, one need only use three of the chambers with the slave cylinder design shown in Figure 3. In this case, the outlets 456 from the remaining chambers 410 can be fed back to the supply line 408 or 440 or the outlet from the remaining chambers 410

can be used for other purposes. Naturally, the pulse generator in figures 7 and 8 can also be made with a chamber 410 for each pair of engine cylinders. In the same way, the pulse generator in figure 5 can be modified so that it has a main piston for each pair of cylinders in the engine, while the pulse generator in figures 3 and 4 can be modified so that it has a main piston for each engine cylinder.

Although neither a pressure regulating mechanism nor a time control mechanism which is affected by the auxiliary pressure is shown specifically in connection with Figures 7 and 8, it is clear to those skilled in the art that one or both mechanisms can also be combined with the embodiment of Figures 7 and 8.

A pressure regulating mechanism 150 as shown in Figures 4A

and 4B, can be placed in the lines 80, 92 or 94 or in the channels 284, 296, 288, 290 or 294 which form a connection between the outlets 454 and the slave cylinders. In order to achieve a timing mechanism which is affected by the auxiliary pressure, in the embodiment of figures 7 and 8, it is necessary to provide for an oscillation of the internal gear 378 in relation to the drive shaft 384. This can be achieved by designing the holes for the head screws 402 as curved slots and by providing a hydraulic cylinder 184, 190 and linkage mechanism 196, as shown in Figures 4A and 4B for pivoting the gear 378

as a result of variations in the heel pressure in the manifold.

Claims (11)

1. Compression actuated engine brake for use in a multi-cylinder four-stroke internal combustion engine with crankshaft 86, intake and exhaust manifolds, at least one exhaust valve for each cylinder and at least one slave piston to open during braking the exhaust valve the slave piston is connected to, characterized by a rotating hydraulic pulse generator (85 . when the valve control is inactive, to create low-pressure conditions in the lines during normal operation of the engine and to create high-pressure conditions in the fluid in the lines when the valve control is in operation during braking. 2. Engine brake as stated in claim 1, characterized in that the lines have branches (136, 138, 140) or (210, 212, 214, 216, 218, 220) which, when the valve control is out of operation, connect the lines to empty , which branches are closed by the valve control during braking to prevent the discharge of hydraulic fluid, which ensures the transmission of hydraulic fluid under high pressure from the pulse generator through the lines to the slave pistons. 3. Engine brake as stated in claim 1 or 2, characterized by pressure regulating devices (150) which are connected to the lines and comprise a chamber which can expand under pressure to limit the maximum pressure which can be produced by the pulse generator to a predetermined level. 4. Engine brake as stated in claim 3, characterized in that the pressure regulating device comprises a cylinder (152) which is connected to the lines (80; 92,94), a piston (160) which is placed in the cylinder for reciprocating movement therein and a spring (162) which biases the piston against the hydraulic fluid pressure in the cylinder and the lines. 5. Engine brake as specified in claim 3, characterized in that the pressure regulating device comprises a Bourdon tube. 6. Engine brake as stated in claim 1, characterized in that the rotating hydraulic pulse generator is operated at engine speed and comprises a positive displacement pump (85) for each pair of engine cylinders with pistons that reach the upper dead center position simultaneously. 7. Engine brake as specified in claim 1, characterized in that the rotating, hydraulic pulse generator is operated at half the speed of the engine and comprises a positive displacement pump for each of the engine's cylinders and in that the wires connect each positive displacement pump and a slave stem <p> el for control of the exhaust valve piston is attached to. 8. Engine brake as stated in claim 6 or 7, characterized in that each positive displacement pump comprises a first piston (272) which is movable back and forth between first and second positions in a first cylinder (198) where the first position of the first piston is adjustable with pressure adjusting means which are affected by the pressure in the engine's intake manifold, which pressure adjusting means comprise a rotating cam (194) which cooperates with the first mentioned piston to determine the mentioned first position of the first piston, a second cylinder (182) which stands in connection with the intake manifold through a second channel, a second piston (190) movable back and forth in the second cylinder (182) affected by the pressure in the second cylinder and the intake manifold, with a spring (192) in the second cylinder for biasing of the second piston against the pressure in the second cylinder and a link (196) which connects the second piston and the said rotating cam. 9. Engine brake as stated in claim 8, characterized in that the valve regulating device comprises a first disc valve (142) which is movable in a manifold (134) between first and second position and solenoid-actuated one-way valves (128) which are connected to the manifold , which manifold connects each of the channels as the disc valves in one of their positions prevent connection of the channels but in another of the positions enable connection of each of the channels through the manifold to drainage. 10. Engine brake as stated in claim 1, characterized in that the rotating hydraulic pulse generator is operated at half the speed of the engine and comprises a gear pump with positive displacement and with an internal gear (378) with a number of teeth or waves corresponding to the number of cylinders the engine has and which meshes with an external gear (380) with one tooth smaller than the internal gear to thereby form a number of positive displacement pumping chambers (410) corresponding to the number of cylinders in the multi-cylinder engine and in that said lines connect each positive displacement pump chamber and slave piston . 11. Engine brake as stated in claim 10, characterized in that the control valve (442) comprises a ring valve (448) which is movable in a manifold (450) between first and second positions and a solenoid-actuated one-way valve (436) which is in connection with the manifold, which manifold forms a connection between the lines while the ring valve (448) in its first position prevents connection of the lines, but in its second position allows connection of each of the lines through the manifold to drain.