EP0111232A1 - Kompressionsverringerungsmotorbremse für Vielzylinder-Brennkraftmaschinen - Google Patents

Kompressionsverringerungsmotorbremse für Vielzylinder-Brennkraftmaschinen Download PDF

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Publication number
EP0111232A1
EP0111232A1 EP83111861A EP83111861A EP0111232A1 EP 0111232 A1 EP0111232 A1 EP 0111232A1 EP 83111861 A EP83111861 A EP 83111861A EP 83111861 A EP83111861 A EP 83111861A EP 0111232 A1 EP0111232 A1 EP 0111232A1
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EP
European Patent Office
Prior art keywords
engine
cylinder
ducts
piston
valve
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Granted
Application number
EP83111861A
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English (en)
French (fr)
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EP0111232B1 (de
Inventor
Raymond Noel Quenneville
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Jacobs Vehicle Systems Inc
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Jacobs Manufacturing Co
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Filing date
Publication date
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Priority to AT83111861T priority Critical patent/ATE20268T1/de
Publication of EP0111232A1 publication Critical patent/EP0111232A1/de
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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01L—CYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L13/00—Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations
    • F01L13/06—Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations for braking
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02D—CONTROLLING COMBUSTION ENGINES
    • F02D13/00—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing
    • F02D13/02—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation
    • F02D13/04—Controlling the engine output power by varying inlet or exhaust valve operating characteristics, e.g. timing during engine operation using engine as brake
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01L—CYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L13/00—Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations
    • F01L13/06—Modifications of valve-gear to facilitate reversing, braking, starting, changing compression ratio, or other specific operations for braking
    • F01L13/065—Compression release engine retarders of the "Jacobs Manufacturing" type
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01L—CYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L9/00—Valve-gear or valve arrangements actuated non-mechanically
    • F01L9/10—Valve-gear or valve arrangements actuated non-mechanically by fluid means, e.g. hydraulic
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01L—CYCLICALLY OPERATING VALVES FOR MACHINES OR ENGINES
    • F01L9/00—Valve-gear or valve arrangements actuated non-mechanically
    • F01L9/10—Valve-gear or valve arrangements actuated non-mechanically by fluid means, e.g. hydraulic
    • F01L9/11—Valve-gear or valve arrangements actuated non-mechanically by fluid means, e.g. hydraulic in which the action of a cam is being transmitted to a valve by a liquid column
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F02—COMBUSTION ENGINES; HOT-GAS OR COMBUSTION-PRODUCT ENGINE PLANTS
    • F02B—INTERNAL-COMBUSTION PISTON ENGINES; COMBUSTION ENGINES IN GENERAL
    • F02B75/00—Other engines
    • F02B75/02—Engines characterised by their cycles, e.g. six-stroke
    • F02B2075/022—Engines characterised by their cycles, e.g. six-stroke having less than six strokes per cycle
    • F02B2075/027—Engines characterised by their cycles, e.g. six-stroke having less than six strokes per cycle four

Definitions

  • This invention relates generally to an engine retarder for use in a multi-cylinder four cycle internal combustion engine.
  • the invention relates more particularly to an hydraulic pulse generating system adapted to open sequentially the exhaust valves of an internal combustion engine so as to provide a compression release engine retarding function during braking operations.
  • drum or disc type wheel brakes are capable of absorbing a large amount of energy over a short period of time.
  • the absorbed energy is transformed into heat which rapidly raises the temperature of the braking mechanism to a level which may render ineffective the friction surfaces and other parts of the mechanism.
  • Such auxiliary devices include hydraulic or electrodynamic retarding systems wherein the kinetic energy of the vehicle is transformed by fluid friction or magnetic eddy currents into heat which may be dissipated through appropriate heat exchangers.
  • Other auxiliary retarding systems include (a) exhaust brakes which inhibit the flow of exhaust gases or air through the exhaust system and (b) compression release retarder mechanisms wherein the energy required to compress the intake air during the compression stroke of a four cycle engine is dissipated by exhausting the compressed air through the exhaust system during the expansion stroke of the engine.
  • the engine compression release retarder With respect to the engine compression release retarder, a portion of the kinetic energy of the vehicle is dissipated through the engine cooling system while another portion of the kinetic energy is dissipated through the engine exhaust system.
  • the exhaust brake the kinetic energy of the vehicle is dissipated only as heat through the engine cooling system.
  • Another advantage of the engine compression release retarder such as that shown in the Cummins U.S. Patent 3,220,392 is that it employs the existing valve train mechanism and requires only the addition of a master piston and a slave piston for each cylinder together with an appropriate control system.
  • a compression release engine retarder for use in a multi-cylinder four cycle internal combustion engine having a crankshaft, intake and exhaust manifolds, at least one exhaust valve for each cylinder, and at least one slave piston for opening, during a braking operation, the exhaust valve with which it is associated, characterized by a rotary hydraulic pulse generator positively driven in synchronism with said engine crankshaft for sequentially supplying pulses of hydraulic fluid under pressure to predetermined slave pistons by means of ducts interconnecting said slave pistons and said hydraulic pulse generator, and a common valve control arranged in relation to said ducts to establish, when the valve control is inoperative, low pressure fluid conditions in said ducts which occur during a fueling mode, and to establish when the valve control
  • the hydraulic pulse generator produces hydraulic pulses timed to coincide with the beginning of the expansion or power stroke of each cylinder of the four cycle engine.
  • the hydraulic pulse generator may be driven by a shaft operating at or half of the engine speed. Timing may be adjustable over a broad range by changing the position of the pulse generator and its driving shaft relative to the engine crankshaft while the magnitude of the pulse may be determined by the design of the pulse generator. In addition, timing may be adjusted as a function of boost pressure to control more precisely the compression release event.
  • the compression release retarder employs the motion of an element in the valve train or fuel injector train, typically a push tube, which has close to the desired timing, and uses this motion to drive a master piston.
  • the master piston is hydraulically interconnected with a slave piston which acts, during braking, to open an exhaust valve (or dual exhaust valves) at or near the beginning of the expansion or "power" stroke of a four cycle engine.
  • the compression release engine retarder is in operation, as it is during braking, the fuel supply is automatically shut off so that the engine is pumping air only.
  • the exhaust valves By opening the exhaust valves at the beginning of the expansion or "power" stroke of the engine, the work done in compressing air during the compression stroke is not recovered during the expansion stroke, but, instead, is dissipated through the engine exhaust and cooling systems.
  • the energy so absorbed by the engine can be measured by an electric dynamometer in terms of the retarding horsepower developed by the engine.
  • the total retarding effect includes, of course, the energy loss represented by the internal friction of the engine and all other losses related to the engine auxiliaries.
  • the absolute value of the retarding horsepower developed during compression release retarding can be a substantial portion of the positive horsepower developed by the engine during normal fueling operation.
  • FIG. 1 shows various exhaust valve motions.
  • exhaust valve motion is plotted as the ordinate while crank angle, starting at top dead center (TDC), is the abscissa.
  • Curve 10 indicates the normal mode of operation of the exhaust valve at the beginning of the exhaust stroke of the engine.
  • Curve 12 illustrates a typical exhaust valve motion of an engine equipped with a compression release brake designed to open the exhaust valves at the beginning of the expansion stroke of the engine as close to top dead center as may be practicable. More specifically, curve 12 shows the motion of the exhaust valve imparted by the fuel injector pushtube.
  • the motion shown by curve 12 may be produced by the device described, for example, in Sickler et al U.S. Patent 4,271,796 referred to above.
  • Curve 16 indicates the desired optimum motion of the exhaust valve to maximize the dissipation of the power stored in the engine during the compression stroke. It will be appreciated that the precise timing of the compression release exhaust valve opening should be selected so as to maximize the work of compression and minimize the recovery of that work during the ensuing expansion stroke. This implies a rapid opening of the valve near the end of the compression stroke and close to the top dead center position. However, where the valve operating motion is derived from another part of the engine, such as the fuel injector pushtube or the exhaust pushtube for another cylinder, it may be impossible to attain an optimum operation of the compression release mechanism.
  • Figure 2(A) is a schematic plot of volume vs. time for a positive displacement pulse generator such as the piston pump 18 of Figure 2(B).
  • Pump 18 comprises a shaft 20 which drives a crank 22 which, in turn, drives a connecting rod 24 affixed to a piston 26 constrained to move within a cylinder 28.
  • the positive portion of the displacement curve 30 is similar to the optimum valve lift curve 16 of Figure 1.
  • a portion of the displacement curve 30, such as that above the line 32, may, therefore, function as a valve-operating pulse, if properly timed.
  • FIGS. 3A-3F illustrate sequentially the operation of a compression release engine retarder using a three-unit pulse generator in accordance with the present invention.
  • the engine block for a six-cylinder four-stroke cycle internal combustion engine is shown in fragmentary form at 34.
  • Cylinder No. 1 is indicated at 36 and contains piston 38.
  • Cylinder No. 6 is indicated at 40 and contains piston 42.
  • the pistons for Cylinder Nos. 1 and 6 reach top dead center (TDC) at the same time and move in tandem.
  • piston No. 1 (38) is moving upwards in an exhaust stroke
  • piston No. 6 (42) is moving upwards in a compression stroke.
  • Cylinder Nos. 2 and 5 and Cylinder Nos. 3 and 4 are similarly paired.
  • the operation of the present invention will be limited to a description of the sequential events related to Cylinder Nos. 1 and 6, it being understood that similar events will occur with the other cylinders but at a different time.
  • Intake valve 44 for Cylinder No. 1 (36) is normally biased toward a closed position by valve spring 46.
  • intake valve 48 for Cylinder No. 6 (40) is biased toward the closed position by valve spring 50.
  • the intake valves 44, 46 may be operated by rocker arms (not shown).
  • the exhaust valve 52 for Cylinder No. 1 (36) is biased toward its closed position by valve spring 54. Exhaust valve 52 is opened during the fueling mode of engine operation by rocker arm 56 driven by the engine cam shaft (not shown) and acts through a crosshead 58.
  • a slave piston 60 is mounted for reciprocating motion in a slave piston cylinder 62 formed in a housing 64 secured to the engine head 63 which is, in turn, secured to the engine block 34.
  • the slave piston 60 is biased away from the crosshead 58 by a spring 65.
  • exhaust valve 66 for Cylinder No. 6 (40) is biased toward its closed position by valve spring 68.
  • Exhaust valve 66 is opened during the fueling mode by rocker arm 70 which acts through crosshead 72.
  • Slave piston 74 is mounted for reciprocating motion in slave piston cylinder 76 formed in the housing 64.
  • the slave piston 74 is biased away from the crosshead 72 by a spring 78.
  • a duct 80 communicates between the slave piston cylinders 62, 76 and the master cylinder 82 formed in housing 84 of pulse generator 85.
  • Master cylinder 82 is disposed in the housing 84 radially with respect to pulse generator drive shaft 86 which may be any engine accessory or auxiliary drive shaft operating at engine speed and positively driven directly or indirectly from the engine crank shaft so as to maintain synchronism therewith.
  • Master cylinders 88 and 90 are also disposed radially with respect to the drive shaft 86 and separated from each other and master cylinder 82 by 120° of arc.
  • a duct 92 leads from master cylinder 88 to a pair of slave piston cylinders (not shown) identical to slave piston cylinders 62 and 76 but associated with engine Cylinder Nos. 2 and 5.
  • a duct 94 leads from master cylinder 90 to a third pair of slave piston cylinders (not shown) identical to slave piston cylinders 62 and 76 but associated with engine Cylinder Nos. 3 and 4.
  • An eccentric 96 is keyed to the drive shaft 86 which, as shown in Figures 3A-3F, is driven in a counterclockwise direction. It will be understood that rotation of the drive shaft 86 drives the eccentric 96 against a ring 97 which contacts master pistons 98, 100 and 102 respectively mounted for reciprocating motion in master cylinders 82, 88, and 90 thereby causing the pistons sequentially to move radially outwardly from the drive shaft 86.
  • piston 38 of Cylinder No. 1 is beginning to move upwardly on an exhaust stroke while piston 42 of Cylinder No. 6 is beginning to move upwardly on a compression stroke.
  • the ring 97 driven by the eccentric 96 is also beginning to move the master piston 98 in a radially outward direction.
  • the exhaust valves 52 and 66 are still closed but the master piston 98 has moved sufficiently to isolate duct 80 and the hydraulic fluid contained therein from the low pressure hydraulic fluid supply within the eccentric chamber 104 of the pulse generator.
  • Figure 3B shows the compression release mechanism at a point slightly later in time when the rocker arm 56 has opened exhaust valve 52 in a normal valve opening sequence as required in conjunction with the normal exhaust stroke of Cylinder No. 1(36).
  • the exhaust valve 52 and its crosshead 58 are displaced downwardly, the motion of the slave piston 60 is restrained only by the bias of the spring 65. Consequently, as the pressure in the duct 80 builds up from the radially outward motion of the master piston 98, hydraulic fluid enters the slave cylinder 62 and drives the slave piston 60 downwardly until an abutment 108 on the slave piston 60 strikes a stop 110 mounted on the housing 64.
  • slave piston 60 which occurs while the hydraulic fluid in duct 80 is still at a relatively low pressure has no effect on the exhaust valve 52 which, during this interval, is controlled by rocker arm 56.
  • the pressure in duct 80 is high enough to overcome the bias of spring 65 in slave piston cylinder 62 it will also overcome the bias of spring 78 in slave piston 76 and bring the slave piston 74 into contact with crosshead 72, thereby taking up the lash or clearance normally existing between each slave piston and its associated exhaust valve crosshead whenever the duct 80 is in communication with the low pressure existing in the eccentric chamber 104.
  • slave piston 74 will not open exhaust valve 66 at this time because the force required to open valve 66 greatly exceeds that required to move the slave piston 74 against the bias of spring 78 and therefore only the latter motion will occur.
  • Figure 3C shows the compression release mechanism when the master piston 98 has attained its maximum stroke.
  • the hydraulic pressure in the duct 80 has reached a high level and is sufficient to overcome the bias of valve spring 68 and the forces on valve 66 due to the air compressed in Cylinder No. 6 (40) so as to open exhaust valve 66.
  • movement of the master piston 98 should be controlled by the position of the eccentric 96 so that the exhaust valve 66 is opened close to the TDC position of the piston 42 in order to maximize the retarding horsepower developed by the engine. This insures that the maximum amount of work has been done in compressing air during the compression stroke and that the minimum portion of that work is recovered during the ensuing expansion stroke.
  • Figure 3D illustrates a subsequent condition of the mechanism in which the eccentric 96 has moved to a point where the master piston 98 has retracted to essentially the position it had in Figure 3B.
  • the pressure in the slave cylinder 76 drops to a point where the valve spring 68 will close the valve 66.
  • rocker arm 56 oscillates back to its original position and valve spring 54 closes exhaust valve 52.
  • slave piston 60 is driven upwardly.
  • shaft 86 has moved through an angle of slightly less than 90°.
  • Figure 3E shows the condition of the compression release mechanism when the eccentric 96 has again reached master piston 98.
  • Cylinder No. 6 (40) is beginning its exhaust stroke while Cylinder No. 1 (36) is beginning its compression stroke.
  • rocker arm 70 oscillates downwardly against the crosshead 72 to open exhaust valve 66.
  • hydraulic fluid enters the slave cylinder 76 and drives slave piston 74 downwardly until the abutment 114 in slave piston 74 strikes stop 116 mounted on the housing 64.
  • hydraulic fluid will also enter slave cylinder 62 driving slave piston 60 downwardly until it contacts the crosshead 58.
  • Figure 3F corresponds essentially with Figure 3C described above and represents the point at which the master piston 90 has moved to its extreme outward position and has raised the pressure in duct 80 to a point where slave piston 60 opens exhaust valve 52. Again, this occurs close to top dead center (TDC) position.
  • TDC top dead center
  • each of the engine cylinders will have experienced a compression release event close to the top dead center position of the piston near the end of the compression stroke.
  • Figures 4A and 4B show, on an enlarged scale, the pulse generator illustrated schematically in Figures 3A and 3F.
  • Figures 4A and 4B also include a timing advance mechanism, a pressure relief mechanism, a solenoid switch to control the pulse generator and an alternative master piston construction.
  • the pulse generator 85 comprises a housing 84 containing an eccentric chamber 104. Master cylinders 82, 88 and 90 are located radially 120° apart around the eccentric chamber 104.
  • a duct 118 conducts low pressure hydrualic fluid, such as oil, to the eccentric chamber 104 and thence to the master cylinders 82, 88, and 90 to the duct 80 and slave cylinders 62 and 76.
  • hydraulic fluid is conducted to the ducts 92 and 94 and their corresponding slave cylinders.
  • a solenoid valve 120 which comprises a solenoid coil 122, an armature disc 124, a control rod 126 and a ball valve 128 which seats on a valve seat 130.
  • the ball chamber 132 communicates with a drain (not shown).
  • a disc valve chamber 134 Within the ball valve seat 130 is located a disc valve chamber 134. Passageways 136, 138, 140 lead from the disc valve chamber 134 respectively to master cylinders 82, 88, 90.
  • a disc valve 142 is moveably located within the disc valve chamber 134.
  • the solenoid ball valve 128 When, however, the solenoid is energized (as shown in Figure 4B), the solenoid ball valve 128 will be sealed to the seat 130 by the movement of the control rod 126 and the armature disc 124 and pressure will build up on the disc valve 142 so as to seal off the low pressure passageways. If, as shown in Figure 4B the eccentric 96 is driving master piston 98, then the high pressure from passageway 136 will cause the disc valve 142 to seal off the low pressure passageways 138 and 140. This will then permit the high pressure hydraulic fluid to enter duct 80 and perform the functions described above.
  • the master pistons 98, 100, 102 each comprise three parts: an outer cylindrical shell 144, an inner sliding piston 146, and a snap ring 148 which limits the relative sliding motion between the inner piston 146 and the cylindrical shell 144. It will be understood that the eccentric 96 and the ring 97 first drive the piston 146 into sealing engagement with the cylindrical shell 144 and then drive both these parts of the master piston 98 (or 100, 102) radially outward to provide the requisite hydraulic pressure in the duct 80 (or 92, 94).
  • FIGS. 4A and 4B also illustrate a pressure control mechanism 150 inserted into the duct 80. It will be understood, of course, that similar mechanisms may be placed in the corresponding duct 92 leading to Cylinders No-. 2 and 5 and the duct 94 leading to Cylinder Nos. 3 and 4.-
  • the pressure control mechanism 150 comprises a housing 152 which contains a diametral passageway 154 communicating with the duct 80 and, communicating therewith, an axial passageway 156 which terminates in a cylindrical chamber 158.
  • a piston 160 is mounted for reciprocating motion within the cylindrical chamber 158 of the housing 152 and is biased toward the axial passageway 156 by a relatively stiff spring 162.
  • the spring 162 and piston 160 are held in place by a cap 164 which is threaded onto the housing 152.
  • the pressure control mechanism 150 functions as a sort of shock absorber to increase the volume of the hydraulic circuit whenever the pressure exceeds a predetermined level thereby limiting the maximum. pressure that can be produced by the pulse geneator mechanism 85.
  • the bulk modulus of the hydraulic fluid is such that the fluid may be regarded as essentially incompressible and some form of pressure control mechanism is desirable to relieve excess pressure in the system.
  • the compressibility of the hydraulic fluid itself may be sufficient to provide, in effect, its own expansion chamber.
  • another device such as a Bourdon-tube mechanism which expands with pressure, may be employed.
  • the pulse generator 85 may also incorporate a timing advance mechanism by which the timing of the hydraulic pulses may be varied as a function of the boost pressure developed by the engine turbocharger (if the engine is so equipped). It will be understood that the boost pressure in the inlet manifold of the engine varies with the speed of the turbocharger and that the mass of air introduced into the engine is also a function of the boost pressure. Moreover, the pressure developed in the engine cylinder during the compression stroke of the engine varies with the mass of air drawn into the cylinder on the intake stroke of the engine while the force required to open an exhaust valve is a function of the pressure in the engine cylinder with which that exhaust valve is associated.
  • Figure 6 is a family of curves in which the force required to open an engine exhaust valve is plotted as the ordinate against crank angle as the abscissa. Curve 166 shows how this force varies with the crank angle at low boost pressure while curve 168 is a similar curve for high boost pressure. It will be appreciated that an engine may operate at various boost pressures within the range represented by curves 166 and 168.
  • line 170 represents the rate at which the pressure in the hydraulic system, e.g. ducts 80, 92 and 94 builds up as a function of time (or crank angle) at low boost pressure so as to attain the force required to open the exhaust valve, indicated by point 172.
  • the intersection of line 170 with the axis (point 174) establishes the crank angle at which motion of the master piston must begin in order to open the exhaust valve at point 172.
  • point 176 on the high boost curve 168 indicates the force required to open the exhaust valve at 15° B.T.D.C. under high boost conditions. To attain this force, the pressure in the hydraulic system rises along curve 178 and intercepts the axis at point 180. From Fig. 6 it is apparent that a mechanism to adjust the timing of the master piston movement automatically in response to the boost pressure would be desirable.
  • a boss 182 formed integrally with the housing 84, contains a cylindrical bore 184 which communicates at one end through a passageway 186 to duct 188 which communicates with the engine intake manifold (not shown).
  • the intake manifold is at all times subject to the boost pressure produced by the engine turbocharger (not shown).
  • a piston 190 is disposed within the cylindrical bore 184 and is biased toward the passageway 186 by a spring 192.
  • a three- lobed cam 194 is mounted on the drive shaft 86 so as to be rotatable independently from the eccentric 96 and the ring 97.
  • the cam 194 is axially offset from the eccentric 96 so that each lobe of the cam 194 contacts the sliding cylindrical shell 144 of one of the pistons 98, 100 or 102 but does not contact the inner sliding piston 146.
  • a link 196 interconnects the cam 194 and the piston 190.
  • the piston 190 At low boost pressure, the piston 190 is biased toward the left (as viewed in Fig. 4A) thereby rotating the cam 194 counterclockwise to a position where the cylindrical shells 144 of the pistons 98, 100 and 102 are driven radially outward from the drive shaft 86. As shown in Fig. 4B, at high boost pressure, the piston 190 is driven to the right and the cam 194 is rotated in a clockwise direction. Such rotation of the cam 194 allows the cylindrical shells 144 of the pistons 98, 100 and 102 to move radially inward toward the shaft 86.
  • Fig. 5 illustrates an alternative form of the invention adapted for use in conjunction with a pulse generator drive shaft 86' which operates at half the speed of the engine crank shaft.
  • the housing 84' contains six master cylinders 198, 200, 202, 204, 206 and 208 which, during compression release retarding, respectively control the opening of the exhaust valves for Cylinders Nos. 6, 2, 4, 1, 5 and 3.
  • the master cylinders 198, 200, 202, 204, 206 and 208 are arranged in the same sequence as the firing order of the engine.
  • Passageways 210, 212, 216, 218 and 220 communicate respectively between master cylinders 198, 200, 202, 204,206 and 208 and six ball check valve chambers only three of which are visible, namely, chambers 222, 226 and 230.
  • the other three ball check valve members are behind chambers 222, 226 and 230 respectively, and thus are not visible in Fig. 5.
  • Each ball check valve chamber is equipped with a seat 234, a ball check valve 236 and a spring 238 which normally biases the ball check valve 234.
  • Each of the six ball check valve chambers communicate with an armature manifold chamber 240 respectively through six passageways only three of which are visible, namely passageways 242, 246 and 250.
  • the other three passageways are respectively located behind passageways 242, 246 and 250 and are therefore not visible in Fig. 5.
  • An armature 254 is disposed in armature manifold chamber 240.
  • a hydraulic fluid passageway 256 communicates between the armature manifold chamber 240 and duct 258 which supplies hydraulic fluid to the pulse generator 85'.
  • the solenoid coil is shown at 260 and the leads to the solenoid coil at 262, 264.
  • a relatively stiff spring 266 is seated in a bore 268 formed in the solenoid core and normally biases the armature 254 away from the coil 260.
  • Control rods 270 are disposed between the armature 254 and the ball check valves 236 in passageways 242, 246 and 250 in the three non-visible passageways.
  • Master pistons 272, 274, 276, 278, 280 and 282 are located respectively in master cylinders 198, 200, 202, 204, 206 and 208 and are driven by ring 97 and eccentric 96 from drive shaft 86'.
  • a duct 284 communicates between master cylinder 198 and a slave cylinder (not shown) associated with engine Cylinder No. 6.
  • Ducts 286, 288, 290, 292, 294 communicate, respectively, with slave cylinders (not shown) associated, respectively, with engine Cylinder Nos. 2, 4, 1, 5 and 3.
  • a pressure control mechanism 150 like that described in connection with Figs. 4A and 4B may be located in each of the ducts 284, 286, 290, 292 and 294.
  • a circular cam 296 is positioned on the shaft 86' adjacent to the cam 96 and may either be keyed to the shaft 86' or freely rotatable with respect thereto. In either event, the cam 296 provides a limit for the inward travel of each of the master pistons 272. 274, 276, 278, 280 and 282.
  • the chamber 240 will be seen to function as a manifold with respect to the ducts 284, 286, 290, 292 and 294 and the master cylinders respectively associated with those ducts.
  • the system will remain full of hydraulic fluid since any leakage of fluid past the solid master pistons 272, 274, 276, 278, 280 and 282 into the eccentric chamber 104 will be made up from the supply duct 258. It will be understood that a drainage path (not shown) is provided from the eccentric chamber 104 to an hydraulic fluid sump (not shown).
  • the cam 296 may be provided with six lobes similar in shape to the three lobes of cam 194 shown in Figs. 4A and 4B.
  • the position of the cam 296 may be controlled by a piston and linkage responsive to boost pressure as shown in Figs. 4A and 4B. In this circumstance it will, of course, be necessary for the cam 296 to be freely rotatable with respect to the shaft 86'.
  • Modified crossheads may be used for opening the exhaust valves during braking such as disclosed in U.S. Patent 4,399,787 and South African Patent 80/7495.
  • Figs. 7 and 8 illustrate an alternative pulse generator 376 employing a positive displacement gear pump.
  • the pulse generator 376 comprises an internal gear element 378 having six teeth and designed to mesh with an eccentrically mounted five toothed gear 380.
  • the gear 380 is journalled on an eccentric 382 keyed to a shaft 384 positively driven at half the speed of the engine.
  • a spline ring 386 (Fig. 8) is keyed to the shaft 384 and locked to the shaft by a washer 388 and a capscrew 390.
  • the spline ring 386 mates with internal splines formed on an auxiliary drive shaft 392 positively driven at half the speed of the engine crankshaft.
  • a second spline ring 394 may be affixed on the opposite end of shaft 384.
  • the body of the pulse generator comprises an adapter plate 396 and a main housing affixed to the engine block (not shown) by capscrews 466 and a rear housing 400.
  • the rear housing 400, internal gear element 378 and main housing 398 are fixed together with six capscrews 402 positioned intermediate the roots of the six lobes or teeth of the internal gear 378.
  • An annular groove 404 is formed in the rear housing 400 which groove communicates through a passageway 406 to an hydraulic fluid supply duct 408.
  • the annular groove 404 is smaller in maximum diameter than a circle tangent to the lobes of the internal gear 378 so that as the external gear 380 rotates within the internal gear 378, it sweeps across the annular groove 404 and traps hydrualic fluid in the chamber defined by the teeth of the gears 378, 380 and the surfaces of the housings 398, 400. It will be understood that the diameter of the annular groove 404 defines the volume of the chamber 410 and the point at which the hydraulic fluid may be pressurized. Thus the diameter of the annular groove 404 defines the position of the line 32 in the schematic diagram of Fig. 2A.
  • An annular groove 412 equal in size to the groove 404 is formed in the housing 398. 0-rings 414, 416 may be positioned between the internal gear 378 and the housings 400, 398 to inhibit leakage of hydraulic fluid.
  • Passageways 418 are formed in housing 398 adjacent the root portion of each of the six teeth of the internal gear 378.
  • the passageways 418 communicate with an annular chamber 420 formed between the main housing 398 and the adapter plate 398 which functions as a manifold with respect to the passageways 418 and the several chambers 410 of the positive displacement gear pump.
  • the annular manifold chamber 420 may conveniently be sealed by the use of 0-rings 422 and 424 located between the adapter plate 396 and the main housing 398.
  • An annular ring valve 426 is freely positioned in the annular chamber 420 and biased toward the housing 398 by a plurality of springs 428 seated in blind holes 430 formed in the adapter plate 396.
  • a passageway 432 communicates between the annular chamber 420 and a ball valve chamber 434 (Fig. 7) containing a ball valve 436.
  • Passageway 438 connects the ball valve chamber 434 and a supply line 440.
  • a solenoid 442 is mounted in the adapter plate 396 and comprises a solenoid coil 444, a core 446, a disc type armature 448 loosely fitted in an armature chamber 450 and a pin 452 positioned in the core 446 between the armature 448 and the ball valve 436.
  • the armature 448 drives the pin 452 against the ball valve 436 to prevent the flow of hydraulic fluid from the passageway 432 to the passageway 438 and thence to the supply line 440.
  • each passageway 418 Also communicating with each passageway 418 is an outlet duct 454 which leads to a fitting 456 in the main housing 398.
  • the fittings 456 are adapted to receive, respectively, ducts 284, 286, 288, 290, 292 and 294 (Fig. 5) which ducts are connected to slave cylinders associated respectively, with Cylinder Nos. 6, 2, 4, 1, 5 and 3 of the engine.
  • a passageway 458 formed at the bottom of the housing 400 communicates with a tube 460 and a passageway 464 formed in the adapter plate 398.
  • the tube 460 may conveniently be sealed into the housing 400 and adapter plate 396 by 0-rings 462.
  • the ball valve 436 will seal the passageway 432 and allow a build-up of pressure in the annular manifold chamber 420. Such pressure will cause the ring valve-426 to seal against the passageways 418 that are under low pressure so that the pulse of hydraulic fluid from one chamber 410 will pressurize one passageway 418, one outlet duct 454 and be delivered to one of the ducts 284, 286, 288, 290, 292 or 294 and its associated slave cylinder. It will be understood that continued rotation of the auxiliary drive shaft 392 will result in sequential pulses of hydraulic fluid in each of the ducts 284, 286, 288, 290, 292 and 294 and their associated slave cylinders.
  • auxiliary drive shaft 392 rotates at half the speed of the engine crankshaft it will be appreciated that during the course of two revolutions of the engine crankshaft one hydraulic pulse will be delivered to each of the ducts 284, 286, 288, 290, 292 and 294 and its associated slave cylinder.
  • the pulses may be timed so that each engine exhaust valve is opened at a predetermined time close to the T.D.C. position of the engine piston at the end of the compression stroke of the cylinder with which the exhaust valve is associated.
  • Figs. 7 and 8 may be modified so as to be applicable to engines having any number of cylinders by providing an internal gear 378 having the same number of teeth as the engine has cylinders and an external gear 380 having one tooth less than the number of cylinders in the engine.
  • the pulse generator of Figs. 7 and 8 contains a chamber 410 for each engine cylinder. If desired, only three of the chambers need be used in accordance with the slave cylinder arrangement shown in Fig. 3. In this event, the outlets 456 of the remaining chambers 410 may be directed back to the supply line 408 or 440 or the output of the remaining chambers 410 may be used for another purpose.
  • the pulse generator of Figs. 7 and 8 may also be designed with one chamber 410 for each pair of engine cylinders.
  • the pulse generator of Fig. 5 may be modified to provide one master piston for each pair of cylinders on an engine while the pulse generator of Figs. 3 and 4 may be modified to provide a master piston for each engine cylinder.
  • a pressure control mechanism 150 as shown in Figs. 4A and 4B may be placed in the ducts 80, 92, or 94 or the ducts 284, 286, 288, 290 or 294 which communicate between the outlets 456 and the slave cylinders.
  • boost pressure timing mechanism it is necessary to provide for oscillation of the internal gear 378 with respect to the drive shaft 384.

Landscapes

  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Output Control And Ontrol Of Special Type Engine (AREA)
  • Valve Device For Special Equipments (AREA)
  • Transition And Organic Metals Composition Catalysts For Addition Polymerization (AREA)
EP83111861A 1982-12-09 1983-11-26 Kompressionsverringerungsmotorbremse für Vielzylinder-Brennkraftmaschinen Expired EP0111232B1 (de)

Priority Applications (1)

Application Number Priority Date Filing Date Title
AT83111861T ATE20268T1 (de) 1982-12-09 1983-11-26 Kompressionsverringerungsmotorbremse fuer vielzylinder-brennkraftmaschinen.

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
US44835082A 1982-12-09 1982-12-09
US448350 1982-12-09

Publications (2)

Publication Number Publication Date
EP0111232A1 true EP0111232A1 (de) 1984-06-20
EP0111232B1 EP0111232B1 (de) 1986-06-04

Family

ID=23779958

Family Applications (1)

Application Number Title Priority Date Filing Date
EP83111861A Expired EP0111232B1 (de) 1982-12-09 1983-11-26 Kompressionsverringerungsmotorbremse für Vielzylinder-Brennkraftmaschinen

Country Status (15)

Country Link
EP (1) EP0111232B1 (de)
JP (1) JPS59110820A (de)
KR (1) KR890002917B1 (de)
AT (1) ATE20268T1 (de)
AU (1) AU565286B2 (de)
BR (1) BR8306765A (de)
CA (1) CA1247483A (de)
DE (1) DE3363963D1 (de)
ES (1) ES527909A0 (de)
IE (1) IE54866B1 (de)
IN (1) IN160156B (de)
MX (1) MX160121A (de)
NO (1) NO834522L (de)
NZ (1) NZ206324A (de)
ZA (1) ZA838679B (de)

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4836162A (en) * 1986-10-30 1989-06-06 AVL Gesellschaft fur Verbrennungskraftmaschinen und Messtechnik m.b.H. Prof.Dr.Dr.h.c. Hans List Engine brake of an internal combustion engine
DE4038334C1 (de) * 1990-12-01 1991-11-28 Mercedes-Benz Aktiengesellschaft, 7000 Stuttgart, De
EP0608520A1 (de) * 1993-01-25 1994-08-03 Steyr Nutzfahrzeuge Ag Motorbremse bei einer 4-Takt-Brennkraftmaschine eines Nutzfahrzeuges

Families Citing this family (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS61110814U (de) * 1984-12-25 1986-07-14
US4592319A (en) * 1985-08-09 1986-06-03 The Jacobs Manufacturing Company Engine retarding method and apparatus

Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2635544A (en) * 1948-03-06 1953-04-21 Lossau Earl Hydraulic valve lifting mechanism
US2829628A (en) * 1954-08-30 1958-04-08 Nordberg Manufacturing Co Hydraulic valve actuating mechanism
FR1255370A (fr) * 1960-01-18 1961-03-10 Pompe à pistons sans clapet à débit et pression linéaires
US4206728A (en) * 1978-05-01 1980-06-10 General Motors Corporation Hydraulic valve actuator system
US4271796A (en) * 1979-06-11 1981-06-09 The Jacobs Manufacturing Company Pressure relief system for engine brake

Patent Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2635544A (en) * 1948-03-06 1953-04-21 Lossau Earl Hydraulic valve lifting mechanism
US2829628A (en) * 1954-08-30 1958-04-08 Nordberg Manufacturing Co Hydraulic valve actuating mechanism
FR1255370A (fr) * 1960-01-18 1961-03-10 Pompe à pistons sans clapet à débit et pression linéaires
US4206728A (en) * 1978-05-01 1980-06-10 General Motors Corporation Hydraulic valve actuator system
US4271796A (en) * 1979-06-11 1981-06-09 The Jacobs Manufacturing Company Pressure relief system for engine brake

Cited By (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4836162A (en) * 1986-10-30 1989-06-06 AVL Gesellschaft fur Verbrennungskraftmaschinen und Messtechnik m.b.H. Prof.Dr.Dr.h.c. Hans List Engine brake of an internal combustion engine
DE4038334C1 (de) * 1990-12-01 1991-11-28 Mercedes-Benz Aktiengesellschaft, 7000 Stuttgart, De
FR2669963A1 (fr) * 1990-12-01 1992-06-05 Daimler Benz Ag Frein moteur pour un moteur a combustion interne a plusieurs cylindres.
US5168848A (en) * 1990-12-01 1992-12-08 Mercedes-Benz Ag Engine brake for a multi-cylinder internal combustion engine
EP0608520A1 (de) * 1993-01-25 1994-08-03 Steyr Nutzfahrzeuge Ag Motorbremse bei einer 4-Takt-Brennkraftmaschine eines Nutzfahrzeuges

Also Published As

Publication number Publication date
ES8502211A1 (es) 1984-12-16
AU565286B2 (en) 1987-09-10
IE54866B1 (en) 1990-02-28
KR890002917B1 (ko) 1989-08-11
JPH0233850B2 (de) 1990-07-31
ZA838679B (en) 1984-07-25
CA1247483A (en) 1988-12-28
AU2149183A (en) 1984-06-14
EP0111232B1 (de) 1986-06-04
JPS59110820A (ja) 1984-06-26
IN160156B (de) 1987-06-27
MX160121A (es) 1989-12-01
BR8306765A (pt) 1984-07-17
DE3363963D1 (en) 1986-07-10
IE832882L (en) 1984-06-09
ES527909A0 (es) 1984-12-16
NO834522L (no) 1984-06-12
ATE20268T1 (de) 1986-06-15
NZ206324A (en) 1986-02-21
KR840007133A (ko) 1984-12-05

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