WO2019137278A1 - 一种可实现米勒循环的曲柄连杆机构及控制方法 - Google Patents

一种可实现米勒循环的曲柄连杆机构及控制方法 Download PDF

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Publication number
WO2019137278A1
WO2019137278A1 PCT/CN2019/070088 CN2019070088W WO2019137278A1 WO 2019137278 A1 WO2019137278 A1 WO 2019137278A1 CN 2019070088 W CN2019070088 W CN 2019070088W WO 2019137278 A1 WO2019137278 A1 WO 2019137278A1
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WO
WIPO (PCT)
Prior art keywords
oil passage
gear
crankshaft
compression ratio
drive
Prior art date
Application number
PCT/CN2019/070088
Other languages
English (en)
French (fr)
Inventor
田维
楚云路
胡明艳
肖邦
吴学舜
韩志强
Original Assignee
西华大学
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Publication date
Application filed by 西华大学 filed Critical 西华大学
Priority to GB2010515.1A priority Critical patent/GB2583627B/en
Priority to US16/959,151 priority patent/US11008937B2/en
Publication of WO2019137278A1 publication Critical patent/WO2019137278A1/zh

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    • 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
    • F02B75/04—Engines with variable distances between pistons at top dead-centre positions and cylinder heads
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C9/00—Bearings for crankshafts or connecting-rods; Attachment of connecting-rods
    • F16C9/04—Connecting-rod bearings; Attachments thereof
    • 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
    • F02B75/04—Engines with variable distances between pistons at top dead-centre positions and cylinder heads
    • F02B75/048—Engines with variable distances between pistons at top dead-centre positions and cylinder heads by means of a variable crank stroke length
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C3/00—Shafts; Axles; Cranks; Eccentrics
    • F16C3/04—Crankshafts, eccentric-shafts; Cranks, eccentrics
    • F16C3/06—Crankshafts
    • F16C3/14—Features relating to lubrication
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C3/00—Shafts; Axles; Cranks; Eccentrics
    • F16C3/04—Crankshafts, eccentric-shafts; Cranks, eccentrics
    • F16C3/22—Cranks; Eccentrics
    • F16C3/28—Adjustable cranks or eccentrics
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C9/00—Bearings for crankshafts or connecting-rods; Attachment of connecting-rods
    • F16C9/02—Crankshaft bearings
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16F—SPRINGS; SHOCK-ABSORBERS; MEANS FOR DAMPING VIBRATION
    • F16F15/00—Suppression of vibrations in systems; Means or arrangements for avoiding or reducing out-of-balance forces, e.g. due to motion
    • F16F15/28—Counterweights, i.e. additional weights counterbalancing inertia forces induced by the reciprocating movement of masses in the system, e.g. of pistons attached to an engine crankshaft; Attaching or mounting same
    • F16F15/283—Counterweights, i.e. additional weights counterbalancing inertia forces induced by the reciprocating movement of masses in the system, e.g. of pistons attached to an engine crankshaft; Attaching or mounting same for engine crankshafts
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H1/00—Toothed gearings for conveying rotary motion
    • F16H1/28—Toothed gearings for conveying rotary motion with gears having orbital motion
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H57/00—General details of gearing
    • F16H57/04—Features relating to lubrication or cooling or heating
    • F16H57/0467—Elements of gearings to be lubricated, cooled or heated
    • F16H57/0479—Gears or bearings on planet carriers
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02—TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02T—CLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00—Road transport of goods or passengers
    • Y02T10/10—Internal combustion engine [ICE] based vehicles
    • Y02T10/12—Improving ICE efficiencies

Definitions

  • the invention relates to the technical field of a piston type internal combustion engine, in particular to a crank linkage mechanism and a control method capable of realizing a Miller cycle.
  • Piston internal combustion engines are widely used as power sources in vehicles, non-road mobile machinery, ships and some stationary generator sets, and have made important contributions to the development of the national economy. With the development of economy and industrial technology, the number of piston internal combustion engines continues to increase, which has brought great pressure on energy and the environment. Therefore, energy saving and emission reduction has always been the pursuit goal of piston engine workers.
  • variable compression ratio engine In general, when the compression ratio of a piston type internal combustion engine is increased, fuel economy and emission levels can be improved. However, when the load of the engine is relatively large, a large compression ratio may cause a rough explosion of the engine, deterioration of noise and emissions of the engine, and even mechanical damage of the engine. In order to solve a problem, the idea of a variable compression ratio engine has been proposed in which the engine dynamically adjusts its compression ratio according to the state in which the engine is located to optimize its performance.
  • variable compression ratio engines are generally divided into the following categories: 1 change the engine compression ratio by changing the position of the crankshaft, such as the patent CN201210494582.4; 2 change the compression of the engine by changing the shape of the piston For example, the patent CN201110253913.0; 3 change the clearance ratio by changing the clearance volume through a device on the cylinder head.
  • the patent CN201710106970.34 changes the compression ratio by changing the relative positional relationship between the connecting rod and the crankshaft or changing the length of the connecting rod.
  • patents CN200810067205.6, CN200710029634.X and CN201710229147.1 are generally divided into the following categories: 1 change the engine compression ratio by changing the position of the crankshaft, such as the patent CN201210494582.4; 2 change the compression of the engine by changing the shape of the piston
  • the patent CN201110253913.0 change the clearance ratio by changing the clearance volume through a device on the cylinder head.
  • the patent CN201710106970.34 changes the
  • the scheme of changing the relative position of the connecting rod and the crankshaft has high feasibility, such as the technical solution proposed by the patent CN201710229147.1, but the control scheme of the patent adopts a stepping motor and an external contact circuit, and the stepping motor
  • the response speed is difficult to synchronize the variable compression ratio control with the high engine speed, and the increase of the stepping motor quality on the high-speed rotating crankshaft also has a great influence on the mechanical properties such as the dynamic balance of the crankshaft, and the external contact circuit. The way it also reduces its control reliability.
  • an object of the present invention is to provide a hydraulic-mechanical variable compression ratio crankshaft linkage mechanism; the mechanism can be synchronized with the rotation of the engine to achieve dynamic compression ratio adjustment, and further can be realized in the engine The compression ratio is adjusted during the working stroke to achieve the Miller cycle.
  • a crank linkage mechanism capable of realizing a Miller cycle comprising a crankshaft, a crankshaft support cover, a crankshaft support seat, a drive gear assembly, a planetary gear assembly and a lock pin;
  • the crankshaft includes a main journal, a connecting rod neck and a crank balance weight fixedly coupled to each other;
  • the main journal is rotatably mounted in the cylindrical support hole and the crankshaft support to form a cylindrical bore;
  • the inner wall of the cylindrical bore is provided with a circumferentially concave crankshaft drive drive oil passage ring groove and a crankshaft support lock oil passage ring a slot
  • the crank bearing support is provided with a driving oil passage leading to the crankshaft supporting the driving oil passage ring groove, and a locking oil passage leading to the crankshaft supporting the locking oil passage ring groove;
  • the connecting rod neck and the main journal axis are offset and parallel, and the outer sleeve is sleeved with a cylindrical eccentric connecting rod.
  • the inner cylindrical surface of the eccentric connecting rod and the outer cylindrical surface are offset and parallel, and the eccentric connecting rod.
  • the inner wall of the tile is matched with the gap of the connecting rod neck, the outer wall is matched with the big head of the connecting rod, and the outer edge of one end is provided with an outer meshing ring gear, and the outer meshing ring gear is coaxial with the inner cylindrical surface;
  • crankshaft balance block and the connecting rod neck are respectively arranged on two sides of the main journal axis, and the crankshaft balance block is provided with a drive gear hollow shaft mounting hole whose axis is parallel to the main journal axis, and the inner wall of the drive gear hollow shaft mounting hole is provided with a ring-shaped concave
  • the locking oil passage ring groove and the driving oil passage ring groove are arranged, and the bottom of the groove of the driving oil passage ring groove is evenly arranged with a plurality of locking pin holes with a deep hole in the radial direction;
  • crankshaft is further provided with a crank lock oil passage connecting the crankshaft support lock oil passage ring groove and the lock oil passage ring groove, and a crank drive for connecting the crankshaft support drive oil passage ring groove and the drive oil passage ring groove.
  • Oil passage connecting the crankshaft support lock oil passage ring groove and the lock oil passage ring groove, and a crank drive for connecting the crankshaft support drive oil passage ring groove and the drive oil passage ring groove.
  • the drive gear assembly comprises a drive gear hollow shaft and a drive gear coaxially fixed to one end of the drive gear hollow shaft;
  • the drive gear hollow shaft is rotatably mounted in the drive gear hollow shaft mounting hole, and a plurality of sidewalls thereof are opened a lock oil passage hole corresponding to the lock oil passage ring groove, and a plurality of pin holes corresponding to the drive oil passage ring groove;
  • the drive gear meshes with the outer mesh gear ring through an idle gear;
  • the shaft of the idle gear Fixed on the crankshaft balance block and parallel to the main journal;
  • the planetary gear assembly includes a planetary gear shaft mounted in a hollow shaft of a drive gear, and a planetary gear coaxially fixed to one end of the planetary gear shaft;
  • the planetary gear shaft is provided with a planetary gear shaft on a circumference corresponding to the locking oil through hole The oil ring groove is locked, and a plurality of driving pin holes are arranged on the circumference of the pin hole;
  • the planetary gear shaft is further provided with a planetary gear shaft locking oil passage which is locked by the planetary gear shaft to the driving pin hole
  • the planetary gear meshes with the inner meshing ring gear;
  • the inner meshing ring gear is fixed on a plane formed by the crankshaft support cover and the crankshaft support seat, and is coaxial with the main journal;
  • the locking pin is disposed in the pin hole and is movable under the driving of the hydraulic oil to the locking pin hole or the driving pin hole; the length of the locking pin is equal to or slightly smaller than the wall thickness of the hollow shaft of the driving gear, and is larger than the locking The depth of the pin hole and the drive pin hole.
  • the eccentric connecting rod is divided into a positively biased half-watt and a negatively-biased half-watt along the axis of the inner cylindrical surface, and the split surface and the eccentric moment are not parallel or perpendicular.
  • the number of the locking pin holes is six; the number of the pin holes is eight; the number of the driving pin holes is six, and the number of the locking pins is eight.
  • a control method for a crank-link mechanism capable of realizing a Miller cycle comprising:
  • the compression ratio When the piston moves up to a position before the top dead center of the compression stroke, the compression ratio is gradually increased from small to a certain position after compression top dead center, the compression ratio reaches a maximum value, and then the compression ratio is kept unchanged;
  • the speed at which the engine compression ratio is changed is determined by the total gear ratio of the eccentric link outer meshing ring gear, the idle gear, the drive gear, the planetary gear, and the inner meshing ring gear.
  • the time step of changing the compression ratio is determined by the number of locking pin holes, the number of pin holes, the number of driving pin holes, and the gear ratio of the internal meshing ring gear to the planetary gears.
  • the invention changes the compression ratio by changing the eccentricity of the connecting rod, the required driving torque is small, and the driving system can be simplified;
  • the split surface of the eccentric connecting rod provided by the invention is not parallel or perpendicular to the direction of the eccentric moment, and the force of the crankshaft and the connecting rod is prevented from acting on the gap under a certain compression ratio state, thereby ensuring full compression.
  • the force of the crankshaft and the connecting rod within the range does not cause damage to the eccentric connecting rod;
  • the driving force for changing the compression ratio of the invention adopts gear transmission, and the transmission is stable, stable and reliable;
  • the driving force for changing the compression ratio of the present invention comes from the rotation of the crankshaft, which simplifies the driving force system
  • the driving force system for changing the compression ratio of the present invention can adopt a completely symmetrical design, and the mass of the driving force system itself can also be used as the balance mass to balance the asymmetry of the engine without causing excessive expansion of the original crankshaft balance weight. Influence, reduce the workload of crankshaft dynamic balance development;
  • the invention can easily change the changing speed of the engine compression ratio by changing the total gear ratio of the outer meshing ring gear, the idler gear, the driving gear, the planetary gear, and the inner meshing ring gear of the eccentric connecting rod;
  • the present invention can easily change the compression ratio by controlling the number of the locking pin holes, the number of pin holes on the hollow shaft of the driving gear, the number of driving pin holes, and the transmission ratio of the inner ring gear to the planetary gear. Crank angle;
  • the invention adopts mechanical hydraulic control, and has large control flexibility and flexible control;
  • the invention can be extended to an engine of any number of cylinders
  • the present invention can be applied to engines of various configurations such as in-line and V-type.
  • FIG. 1 is a schematic cross-sectional view showing the overall structure of a crank-link mechanism capable of realizing a Miller cycle according to the present invention.
  • FIG. 2 is a cross-sectional view of an eccentric connecting rod in a crank-link mechanism capable of realizing a Miller cycle of the present invention.
  • FIG 3 is a schematic view of a crankshaft oil passage in a crank-link mechanism capable of realizing a Miller cycle of the present invention.
  • FIG. 4 is a cross-sectional view of the drive gear assembly of the crank linkage mechanism that can realize the Miller cycle of the present invention.
  • Fig. 5 is a cross-sectional view showing the planetary gear assembly of the crank-link mechanism capable of realizing the Miller cycle of the present invention.
  • Figure 6 is a schematic view of the arrangement of the four-cylinder machine.
  • Figure 7 is a schematic view of the control of the present invention.
  • a crank linkage mechanism that can realize a Miller cycle includes a crankshaft 2, a crankshaft support cover 1, a crankshaft support base 3, a drive gear assembly 11, a planetary gear assembly 8, and a lock pin 10.
  • the crankshaft support base 3 is a part of the engine body and is mainly used for mounting the crankshaft 2.
  • the crankshaft support cover 1 and the crankshaft support base 3 are paired to form a cylindrical hole, and all the cylindrical holes in one engine are kept coaxial.
  • the crankshaft 2 includes a main journal 13 that is fixedly coupled to each other, a connecting rod neck 3, and a crank weighting block 9.
  • the main journal 13 is rotatably mounted in the cylindrical support hole of the crankshaft support cover 1 and the crank support base 3; the inner wall of the cylindrical bore is provided with a circumferentially concave crankshaft support drive oil passage ring groove f3 and a crankshaft support lock
  • the oil passage ring groove f4 is provided with a drive oil passage 6 leading to the crankshaft support drive oil passage ring groove f3, and a lock oil passage 7 leading to the crankshaft support lock oil passage ring groove f4.
  • the connecting rod neck 3 and the main journal 13 are axially offset and parallel, and the outer sleeve is sleeved with a cylindrical eccentric connecting rod 4, and the inner cylindrical surface of the eccentric connecting rod 4 is offset and parallel with the axis of the outer cylindrical surface.
  • the inner wall of the eccentric connecting rod 4 is matched with the connecting rod neck 3, the outer wall is matched with the big end of the connecting rod, and the outer edge of one end is provided with an outer meshing ring 15 which is coaxial with the inner cylindrical surface. As shown in FIG.
  • the eccentric connecting rod 4 is divided into a positively biased half watt a1 and a negatively biased half watt a3 along the inner cylindrical axis, and a positively biased half watt a1 and a negatively biased half watt a3 are formed.
  • the complete eccentric connecting rod, the split surface and the eccentric moment are not parallel or perpendicular.
  • the eccentric connecting rod 4 can freely rotate between the crankshaft connecting rod neck and the connecting rod big head, but cannot move axially;
  • the crank balance weight 9 and the connecting rod neck 3 are respectively arranged on both sides of the main shaft 13 axis, and the crank balance weight 9 is provided with a driving gear hollow shaft mounting hole f7 whose axis is parallel to the main journal 13 axis, and the driving gear hollow shaft mounting hole f7
  • the inner wall is provided with a circumferentially concave locking oil passage ring groove f0 and a driving oil passage ring groove f1, and the groove bottom of the driving oil passage ring groove f1 is evenly arranged with a plurality of hole locking pin holes f2 in the radial direction;
  • the lock pin holes f2 are six, and are uniformly machined on the circumference of the bottom of the drive oil passage ring groove f1.
  • the six locking pin holes f2 are in the same plane and are in the same plane as the driving oil passage ring groove f1, and the plane is perpendicular to the driving gear hollow shaft mounting hole axis.
  • crankshaft 2 is further provided with a crank lock oil passage f6 that connects the crankshaft support lock oil passage ring groove f4 and the lock oil passage ring groove f0, and supports the crankshaft to drive the oil passage ring groove.
  • F3 drives the oil passage f5 with the crankshaft that drives the oil passage ring groove f1.
  • the drive gear assembly 11 includes a drive gear hollow shaft b2 and a drive gear b3 coaxially fixed to one end of the drive gear hollow shaft b2; the drive gear hollow shaft b2 is rotatably mounted on the drive gear hollow shaft.
  • a plurality of lock oil passage holes b1 corresponding to the lock oil passage ring groove f0 and a plurality of pin holes b4 corresponding to the drive oil passage ring groove f1 are opened in the side wall.
  • the pin holes b4 are eight through holes distributed on the same circumference of the drive gear hollow shaft b2.
  • the lock oil through hole b1 and the pin hole b4 are not in the same plane, and the two do not interfere with each other on the circumference.
  • the drive gear assembly 11 is mounted in the drive gear hollow shaft mounting hole f7 of the crank balance weight 9 through the drive gear hollow shaft b2, and the drive gear hollow shaft b2 and the drive gear hollow shaft mounting hole f7 can be rotated. However, it is not movable in the direction of the axis.
  • the driving gear assembly is installed, the driving gear b3 and the outer meshing ring gear of the eccentric connecting rod 4 are on the same plane, and the driving gear b3 is meshed with the outer meshing ring 15 through an idler pulley 12;
  • the shaft of the wheel 12 is fixed to the crank weight block 9 and is parallel to the main journal 13.
  • the outer gear ring of the drive gear b3, the idler gear 12, and the eccentric link tile 4 are sequentially engaged in two.
  • the planetary gear assembly 8 includes a planetary gear shaft c4 mounted in a hollow shaft b2 of the drive gear, and a planetary gear c3 coaxially fixed to one end of the planetary gear shaft c4; the planetary gear assembly passes The planetary gear shaft c4 is mounted in the drive gear hollow shaft b2. After installation, the planetary gear shaft c4 can rotate in the drive gear hollow shaft b2 but cannot move axially.
  • the planetary gear shaft c4 is provided with a planetary gear shaft lock oil ring groove c1 corresponding to the circumference of the lock oil passage hole b1, and a plurality of drive pin holes c5 are provided on the circumference of the pin hole b4; the number of the drive pin holes c5 It is six and evenly distributed on the planetary gear shaft.
  • the planetary gear shaft c4 is further provided with a planetary gear shaft locking oil passage c2 which is connected to the driving pin hole c5 by the planetary gear shaft lock oil ring groove c1; the planetary gear shaft locking oil passage c2 is a closed oil passage, and the two ends are respectively c5 connected to the drive pin and the planetary gear shaft holes oil lock ring groove c1.
  • the planetary gear c3 meshes with the inner meshing ring gear 5; the inner meshing ring gear 5 is fixed to a plane formed by the crankshaft support cover 1 and the crankshaft support base 3, and is coaxial with the main journal 13 .
  • the number of the locking pins 10 is 8, disposed in the pin hole b4, and can be moved by the hydraulic oil to the locking pin hole f2 or the driving pin hole c5; the length of the locking pin 10 is equal to or slightly smaller than the driving gear hollow
  • the wall thickness of the shaft b2 is greater than the depth of the lock pin hole f2 and the drive pin hole c5.
  • the lock pin 10 is movable in the drive gear hollow shaft pin hole b4 and the crank weight block lock pin hole f2, and the lock pin 10 can also move in the drive gear hollow shaft pin hole b4 and the planetary gear shaft drive pin hole c5.
  • the hydraulic-mechanical variable compression ratio crankshaft linkage has one set per cylinder.
  • Figure 6 is a schematic view of the arrangement of the four-cylinder machine.
  • the external oil pressure fills the crankshaft supporting the lock oil passage ring groove f4 processed on the crankshaft support cover 1 and the support base 4 through the lock oil passage 7; and is processed on the crankshaft 2
  • the crankshaft lock oil passage f6 enters the lock oil passage ring groove f0 on the crank balance weight 9; the hydraulic oil in the lock oil passage ring groove f0 on the crank balance weight 9 passes through the lock oil on the drive gear hollow shaft b2
  • the through hole b1 enters the planetary gear shaft lock oil ring groove c1 machined on the planetary gear shaft c4; and then enters the drive pin hole c5 processed on the planetary gear shaft c4 via the planetary gear shaft lock oil passage c2, and drives the pin
  • the lock pin 10 in the hole c5 is pushed out, and the lock pin 10 is pushed out either in the pin hole b4 of the drive gear hollow shaft b2 or in the pin hole b4 of the drive gear hollow shaft b2 and the
  • the drive gear hollow shaft b2 is locked with the crank weight block 9; when the drive gear hollow shaft b2 is locked with the crank weight block 9, the drive gear b3 and the crank balance weight 9 move synchronously, the planetary gear C3 rotates with the crank weight block 9 in the inner ring gear 5, the planetary gear shaft C3 rotates in the driving gear hollow shaft b2, the idler gear 12 and the eccentric link outer meshing ring gear do not rotate, and the positional relationship of the eccentric connecting rod 4 and the crank connecting rod neck does not change, thereby keeping the compression ratio constant.
  • the external oil pressure is filled with the hydraulic oil through the driving oil passage 6 to fill the crankshaft supporting the driving oil passage ring groove f3 processed on the crank bearing cover 1 and the crank bearing support 3; the crankshaft is further processed on the crankshaft.
  • the drive oil passage f5 enters the drive oil passage ring groove f1 on the crank balance weight 9; the lock pin 10 in the lock pin hole f2 in the drive oil passage ring groove f1 on the crank balance weight 9 is pushed out by the hydraulic oil, and the lock pin 10 After being pushed out, either stay in the pin hole b4 of the drive gear hollow shaft b2 or between the pin hole b4 of the drive gear hollow shaft b2 and the drive pin hole c5 of the planetary gear shaft c4, thereby driving the drive gear hollow shaft b2 and
  • the planetary gear shaft c4 is locked together; the planetary gear c3 rotates with the crank balance weight 9 in the inner meshing ring gear 5, and when the drive gear hollow shaft b2 is locked with the planetary gear shaft c4, the planetary gear shaft c4 and the drive gear hollow shaft B2 synchronous rotation; thereby the planetary gear shaft c4 transmits the motion to the driving gear b3, and then transmits the eccentric connecting rod to the outer meshing ring
  • the compression ratio of the cylinder is controlled to be at the maximum state; when the piston continues to move, the control compression ratio is changed from large to small until the intake stroke At some point after the bottom dead center, the compression ratio is minimized, and then the compression ratio is kept constant; when the piston moves up to a position before the top dead center of the compression stroke, the control compression ratio becomes gradually larger from the small to the After the top dead center is compressed, the maximum value is reached, and then the compression ratio is kept constant; when the piston continues to move to the bottom of the power stroke, the control compression ratio is reduced from large to small, until after the power stroke ends.
  • the compression ratio is minimized, and then the compression ratio is kept constant; when the piston continues to move upward to a position before the top dead center of the exhaust stroke, the control compression ratio is gradually increased from small to compressed. After a dead point, a certain position reaches a maximum value, and then the compression ratio remains unchanged; then the engine enters the next intake stroke, and thus the cycle can be continuously operated.
  • This control allows the engine's intake compression ratio to be less than the work expansion ratio to achieve the Miller cycle.
  • the timing of changing the compression ratio can be flexibly controlled according to the requirements of the engine; the speed of changing the compression ratio of the engine is determined by the total gear ratio of the outer meshing ring gear, the idler gear, the driving gear, the planetary gear, and the inner meshing ring gear of the eccentric connecting rod
  • the time step for changing the compression ratio is determined by the crank angle, determined by the number of locking pin holes, the number of pin holes on the hollow shaft of the drive gear, the number of drive pin holes, and the gear ratio of the internal meshing ring gear to the planetary gear.
  • Figure 7 is a schematic diagram of control.
  • the speed of changing the compression ratio of the engine mainly depends on the transmission ratio. For example, when the total gear ratio of the internal gear ring to the external gear is 1:1, that is, the crankshaft rotates one turn, when it is always in the transmission state, the engine The compression ratio can be continuously changed from maximum to minimum to maximum once. If it is 2:1 or twice, then the speed of change is twice the speed.
  • the time step for changing the compression ratio mainly depends on the number of locking pin holes, the number of pin holes on the hollow shaft of the drive gear, and the number of drive pin holes.
  • the number of the locking pin holes provided in this embodiment is six, the number of pin holes on the hollow shaft of the driving gear is eight, the number of driving pin holes is six, and the time step is 15 degrees, and the crank angle can be changed once. If the total gear ratio is 1:1, the state can be changed every 15 degrees. If the total gear ratio is 2:1, the state changes every 7.5 degrees.
  • the number of the above "686" is set to a relatively optimized combination. Of course, other combinations can be designed.
  • any number that can be evenly distributed within 360 degrees can be changed, as long as the position of the hole corresponds to the state, for example, As long as the locking pin hole corresponds to the position of the hollow shaft pin hole, the locking state can be realized, and as long as the driving pin hole corresponds to the position of the hollow shaft pin hole, the driving state can be realized.

Abstract

一种可实现米勒循环的曲柄连杆机构及控制方法，曲轴（2）内部设有驱动油道（6）和锁止油道（7），曲轴（2）的连杆颈（14）上设有偏心连杆瓦（4），曲轴平衡块（9）内装有通过惰轮（12）与偏心连杆瓦外啮合齿圈（15）啮合的驱动齿轮（b3），驱动齿轮空心轴（b2）内安装有行星齿轮（c3），驱动齿轮空心轴（b2）设有可将其与曲轴平衡块（9）锁定或与行星齿轮轴（b2）锁定的锁销（10）；通过向驱动油道（6）施加液压油，锁定驱动齿轮空心轴（b2）和行星齿轮轴（c4），使行星齿轮轴（c4）将运动传递给驱动齿轮（b3），再通过惰轮（12）传递给偏心连杆外啮合齿圈（15），偏心连杆瓦（4）绕曲轴连杆颈（14）旋转，改变发动机的压缩比。曲柄连杆机构可以与发动机的旋转同步运动，还可实现在发动机一个工作冲程内的压缩比调节，进而实现米勒循环，控制柔性大，控制灵活。

Description

一种可实现米勒循环的曲柄连杆机构及控制方法 技术领域
本发明涉及活塞式内燃机技术领域，具体为一种可实现米勒循环的曲柄连杆机构及控制方法。
背景技术
活塞式内燃机在车辆、非道路移动机械、船只和一些固定式发电机组当中作为动力源而被广泛使用，为国民经济的发展做出了重要的贡献。随着经济和工业技术的发展，活塞式内燃机的保有量持续增大，为能源和环境带来了很大的压力。因此，节能减排一直以来都是活塞式内燃机工作者的追求目标。
通常情况下，当提高活塞式内燃机的压缩比时，可以提高其燃油经济性和排放水平。然而，当发动机的负荷比较大时，大的压缩比有可能造成发动机工作粗爆，发动机的噪声和排放恶化，甚至造成发动机的机械损坏。为了解决一问题，可变压缩比发动机的想法被提出来，这种可变压缩比发动机根据发动机所处的状态动态地调节其压缩比，以最优化其性能。
目前，已经公开的资料显示，可变压缩比发动机一般分为以下几大类：1通过改变曲轴的位置来改变发动机压缩比，如专利CN201210494582.4；2通过改变活塞的形状来改变发动机的压缩比如专利CN201110253913.0；3通过缸盖上的某个装置来改变余隙容积来改变压缩比，如专利CN201710106970.34通过改变连杆与曲轴的相对位置关系或者是改变连杆长度来改变压缩比，例如专利CN200810067205.6、CN200710029634.X和CN201710229147.1。上述方案虽然都能够实现改变压缩比，但是改变曲轴的位置涉及到与曲轴位置相关的装置都必须重新设计，例如输出端的飞轮等，造成发动机的结构十分复杂；改变活塞的形状会造成其动态压缩比控制很困难。改变连杆与曲轴的相对位置这种方案具有较高的可行性，例如专利CN201710229147.1所提出的技术方案，但是该专利的控制方案采用了步进电机和外部接触式电路，步进电机的响应速度很难与发动机的高转速进行同步可变压缩比控制，并且在高速旋转的曲轴上增加步进电机质量对于曲轴的动平衡等机械性能也会造成很大的影响，同时外部接触式电路的方式也会降低其控制可靠性。
发明内容
针对上述问题，本发明的目的在于提供一种液压-机械式可变压缩比的曲轴连杆机构；该机构可以与发动机的旋转同步运动，实现动态的压缩比调节，进一步的可以实现在发 动机一个工作冲程内的压缩比调节，进而实现米勒循环。技术方案如下：
一种可实现米勒循环的曲柄连杆机构，包括曲轴、曲轴支承盖、曲轴支承座、驱动齿轮总成、行星齿轮总成和锁销；
所述曲轴包括相互固定连接的主轴颈、连杆颈和曲轴平衡块；
所述主轴颈可旋转的安装于所述曲轴支承盖和曲轴支承座形成圆柱孔内；所述圆柱孔内壁设有环向内凹的曲轴支承驱动油道环槽和曲轴支承锁止油道环槽，曲轴支承座上设有通向曲轴支承驱动油道环槽的驱动油道，及通向曲轴支承锁止油道环槽的锁止油道；
所述连杆颈与主轴颈轴线相错且平行，其外部套设有圆筒状的偏心连杆瓦，偏心连杆瓦内柱面与外柱面的轴线相错且平行，且偏心连杆瓦内壁与连杆颈间隙配合，外壁与连杆大头间隙配合，且其一端的外沿设有外啮合齿圈，外啮合齿圈与内柱面同轴；
所述曲轴平衡块与连杆颈分局于主轴颈轴线两侧，曲轴平衡块上设有轴线与主轴颈轴线平行的驱动齿轮空心轴安装孔，驱动齿轮空心轴安装孔内壁设有环向内凹的锁止油道环槽和驱动油道环槽，驱动油道环槽的槽底均匀排布有多个孔深沿径向的锁止销孔；
所述曲轴内还设有将曲轴支承锁止油道环槽和锁止油道环槽连通的曲轴锁止油道，及将曲轴支承驱动油道环槽和驱动油道环槽连通的曲轴驱动油道；
所述驱动齿轮总成包括驱动齿轮空心轴及同轴固定于驱动齿轮空心轴一端的驱动齿轮；驱动齿轮空心轴可旋转的安装于驱动齿轮空心轴安装孔内，其侧壁上开设有多个与锁止油道环槽对应的锁止油通孔，及多个与驱动油道环槽对应的销孔；驱动齿轮通过一惰轮与所述外啮合齿圈啮合；所述惰轮的轴固定在曲轴平衡块上，且与主轴颈平行；
所述行星齿轮总成包括安装于驱动齿轮空心轴内的行星齿轮轴，及同轴固定于行星齿轮轴一端的行星齿轮；行星齿轮轴对应于锁止油通孔的圆周上设有行星齿轮轴锁止油环槽，对应于销孔的圆周上设有多个驱动销孔；行星齿轮轴上还设有由行星齿轮轴锁止油环槽通向驱动销孔的行星齿轮轴锁止油道；行星齿轮与内啮合齿圈啮合；内啮合齿圈固定在曲轴支承盖和曲轴支承座所形成的平面上，且与主轴颈同轴；
所述锁销设置于销孔内，并能够在液压油的驱动下向锁止销孔或驱动销孔内移动；锁销的长度等于或略小于驱动齿轮空心轴的壁厚，且大于锁止销孔和驱动销孔的深度。
进一步的，所述偏心连杆瓦沿内圆柱面轴线剖分为正偏置半瓦和负偏置半瓦，剖分面与偏心矩方向不平行也不垂直。
更进一步的，所述锁止销孔为6个；所述销孔为8个；所述驱动销孔为6个，所述锁销为8个。
一种可实现米勒循环的曲柄连杆机构的控制方法，包括：
当发动机某一缸的活塞处于进气行程下止点前某一时刻时，控制该缸的压缩比处于最大状态；
当活塞继续运动时，控制压缩比由大变小，直到在进气行程下止点后的某一时刻其压缩比变到最小，然后保持压缩比不变；
当活塞向上运动到压缩行程上止点前某一位置时，控制压缩比由小逐渐变大直到在压缩上止点后某一位置，压缩比达到最大值，然后保持压缩比不变；
当活塞继续运动到做功行程下止点前某一时刻时，控制压缩比由大变小，直到在做功行程下止点后的某一时刻其压缩比变到最小，然后保持压缩比不变；
当活塞继续向上运动到排气行程上止点前某一位置时，控制压缩比由小逐渐变大，直到在压缩上止点后某一位置，压缩比达到最大值，然后保持压缩比不变。
进一步的，改变发动机压缩比的速度由偏心连杆瓦外啮合齿圈、惰轮、驱动齿轮、行星齿轮和内啮合齿圈的总传动比决定。
更进一步的，改变压缩比的时间步长由锁止销孔数目、销孔数目、驱动销孔数目，及内啮合齿圈与行星齿轮的传动比决定。
本发明的有益效果是：
1)本发明通过改变连杆瓦的偏心量来改变压缩比，所需要的驱动力矩小，驱动系统可以简化；
2)本发明所提供的偏心连杆瓦的剖分面不平行也不垂直于偏心矩方向，避免了某一特定压缩比状态下曲轴与连杆的力作用在缝隙上，保证了在全压缩比范围内的曲轴与连杆的力不会造成偏心连杆瓦的破坏；
3)本发明改变压缩比的驱动力采用了齿轮传动，传动平稳、稳定可靠；
4)本发明改变压缩比的驱动力来自曲轴的旋转，简化了驱动力系统，
5)本发明改变压缩比的驱动力系统可采用完全对称设计，并且驱动力系统本身的质量也可以作为平衡块的质量用于平衡发动机的不对称性，不会对原曲轴平衡块造成过大影响，减小曲轴动平衡开发工作量；
6)本发明通过改变偏心连杆瓦外啮合齿圈、惰轮、驱动齿轮、行星齿轮、内啮合齿圈的总传动比可以很容易地改变发动机压缩比的改变速度；
7)本发明通过改变锁止销孔数目、驱动齿轮空心轴上的销孔数目、驱动销孔数目和内啮合齿圈与行星齿轮的传动比可以很容易地改变压缩比的控制时间步长按曲轴转角记；
8)本发明采用机械液压控制，控制柔性大，控制灵活；
9)本发明可以扩展到任意缸数的发动机；
10)本发明可以用于直列和V型等多种结构形式的发动机。
附图说明
图1为本发明可实现米勒循环的曲柄连杆机构的整体结构剖面示意图。
图2为本发明可实现米勒循环的曲柄连杆机构中偏心连杆瓦剖视图。
图3为本发明可实现米勒循环的曲柄连杆机构中曲轴油道示意图。
图4为本发明可实现米勒循环的曲柄连杆机构中驱动齿轮总成剖视图。
图5为本发明可实现米勒循环的曲柄连杆机构中行星齿轮总成剖视图。
图6为四缸机布置示意图。
图7位本发明控制示意图。
图中：1-曲轴支承盖；2-曲轴；3-曲轴支承座；4-偏心连杆瓦；5-内啮合齿圈；6-驱动油道；7-锁止油道；8-行星齿轮总成；9-曲轴平衡块；10-锁销；11-驱动齿轮总成；12-惰轮；13-主轴颈；14-连杆颈-15-外啮合齿圈；a1-正偏置半瓦；a2-正偏置半瓦齿圈；a3-负偏置半瓦；a4-负偏置半瓦齿圈；a5-连杆瓦内圆圆心；a6-连杆瓦外圆圆心；b1-锁止油通孔；b2-驱动齿轮空心轴；b3-驱动齿轮；b4-销孔；c1-行星齿轮轴锁止油环槽；c2-行星齿轮轴锁止油道；c3-行星齿轮；c4-行星齿轮轴；c5-驱动销孔；f0-锁止油道环槽；f1-驱动油道环槽；f2-锁止销孔；f3-曲轴支承驱动油道环槽；f4-曲轴支承锁止油道环槽；f5-曲轴驱动油道；f6-曲轴锁止油道；f7-驱动齿轮空心轴安装孔；f20-锁止销孔中心圆；b40-销孔中心圆；c50-驱动销孔中心圆。
具体实施方式
下面结合附图和具体实施例对本实用新型做进一步详细说明。如图1所示，一种可实现米勒循环的曲柄连杆机构，包括曲轴2、曲轴支承盖1、曲轴支承座3、驱动齿轮总成11、行星齿轮总成8和锁销10。
所述曲轴支承座3为发动机机体的一部分，主要用于安装曲轴2，曲轴支承盖1与所述曲轴支承座3配对形成圆柱孔，一台发动机内所有的圆柱孔保持同轴。
所述曲轴2包括相互固定连接的主轴颈13、连杆颈3和曲轴平衡块9。
所述主轴颈13可旋转的安装于所述曲轴支承盖1和曲轴支承座3形成圆柱孔内；所述圆柱孔内壁设有环向内凹的曲轴支承驱动油道环槽f3和曲轴支承锁止油道环槽f4，曲轴支承座3上设有通向曲轴支承驱动油道环槽f3的驱动油道6，及通向曲轴支承锁止油道环槽 f4的锁止油道7。
所述连杆颈3与主轴颈13轴线相错且平行，其外部套设有圆筒状的偏心连杆瓦4，偏心连杆瓦4内柱面与外柱面的轴线相错且平行，且偏心连杆瓦4内壁与连杆颈3间隙配合，外壁与连杆大头间隙配合，且其一端的外沿设有外啮合齿圈15，外啮合齿圈15与内柱面同轴。如图2所示，所述偏心连杆瓦4沿内圆柱面轴线剖分为正偏置半瓦a1和负偏置半瓦a3，正偏置半瓦a1和负偏置半瓦a3配合形成完整的偏心连杆瓦，剖分面与偏心矩方向不平行也不垂直。偏心连杆瓦4可以在曲轴连杆颈与连杆大头之间自由转动，但不能轴向移动；
所述曲轴平衡块9与连杆颈3分局于主轴颈13轴线两侧，曲轴平衡块9上设有轴线与主轴颈13轴线平行的驱动齿轮空心轴安装孔f7，驱动齿轮空心轴安装孔f7内壁设有环向内凹的锁止油道环槽f0和驱动油道环槽f1，驱动油道环槽f1的槽底均匀排布有多个孔深沿径向的锁止销孔f2；所述锁止销孔f2为6个，且均匀加工在驱动油道环槽f1底部的圆周上。所述6个锁止销孔f2轴线在同一平面内，且与驱动油道环槽f1同平面，且平面与驱动齿轮空心轴安装孔轴线垂直。
如图3所示，所述曲轴2内还设有将曲轴支承锁止油道环槽f4和锁止油道环槽f0连通的曲轴锁止油道f6，及将曲轴支承驱动油道环槽f3和驱动油道环槽f1连通的曲轴驱动油道f5。
如图4所示，所述驱动齿轮总成11包括驱动齿轮空心轴b2及同轴固定于驱动齿轮空心轴b2一端的驱动齿轮b3；驱动齿轮空心轴b2可旋转的安装于驱动齿轮空心轴安装孔f7内，其侧壁上开设有多个与锁止油道环槽f0对应的锁止油通孔b1，及多个与驱动油道环槽f1对应的销孔b4。销孔b4为分布在驱动齿轮空心轴b2同一圆周上的8个通孔。所述锁止油通孔b1与所述销孔b4不在同一平面上，且两者在圆周上不相互干涉。所述驱动齿轮总成11通过所述驱动齿轮空心轴b2安装在所述曲轴平衡块9的驱动齿轮空心轴安装孔f7内，且驱动齿轮空心轴b2与驱动齿轮空心轴安装孔f7可以旋转，但在轴线方向上不可移动。所述驱动齿轮总成安装好后，驱动齿轮b3与偏心连杆瓦4的外啮合齿圈在同一平面上，驱动齿轮b3通过一惰轮12与所述外啮合齿圈15啮合；所述惰轮12的轴固定在曲轴平衡块9上，且与主轴颈13平行。所述驱动齿轮b3、惰轮12、偏心连杆瓦4的外啮合齿圈依次两两啮合。
如图5所示，所述行星齿轮总成8包括安装于驱动齿轮空心轴b2内的行星齿轮轴c4，及同轴固定于行星齿轮轴c4一端的行星齿轮c3；所述行星齿轮总成通过行星齿轮轴c4 安装在驱动齿轮空心轴b2内，安装好后行星齿轮轴c4在驱动齿轮空心轴b2内可以旋转但不能轴向运动。行星齿轮轴c4对应于锁止油通孔b1的圆周上设有行星齿轮轴锁止油环槽c1，对应于销孔b4的圆周上设有多个驱动销孔c5；驱动销孔c5的数量为6个，且均匀分布在行星齿轮轴上。行星齿轮轴c4上还设有由行星齿轮轴锁止油环槽c1通向驱动销孔c5的行星齿轮轴锁止油道c2；行星齿轮轴锁止油道c2为封闭油道，两端分别连接 驱动销孔c5和行星齿轮轴锁止油环槽c1。行星齿轮c3与内啮合齿圈5啮合；内啮合齿圈5固定在曲轴支承盖1和曲轴支承座3所形成的平面上，且与主轴颈13同轴。
所述锁销10数量为8，设置于销孔b4内，并能够在液压油的驱动下向锁止销孔f2或驱动销孔c5内移动；锁销10的长度等于或略小于驱动齿轮空心轴b2的壁厚，且大于锁止销孔f2和驱动销孔c5的深度。锁销10可以在驱动齿轮空心轴销孔b4与曲轴平衡块锁止销孔f2内移动，锁销10也可以在驱动齿轮空心轴销孔b4与行星齿轮轴驱动销孔c5内移动。
所述液压-机械式可变压缩比的曲轴连杆机构每一缸均有一套。图6为四缸机布置示意图。
在发动机运行过程中，所述液压-机械式可变压缩比的曲轴连杆机构的控制是这样实现的：
压缩比不需要改变时，外部的油压通过锁止油道7使液压油充满加工在曲轴支承盖1和支承座4上的曲轴支承锁止油道环槽f4；再经加工在曲轴2上的曲轴锁止油道f6进入曲轴平衡块9上的锁止油道环槽f0；曲轴平衡块9上的锁止油道环槽f0内的液压油经驱动齿轮空心轴b2上的锁止油通孔b1进入加工在行星齿轮轴c4上的行星齿轮轴锁止油环槽c1；再经行星齿轮轴锁止油道c2进入到加工在行星齿轮轴c4上的驱动销孔c5，将驱动销孔c5内的锁销10推出，锁销10被推出后要么停留在驱动齿轮空心轴b2的销孔b4内，要么处于驱动齿轮空心轴b2的销孔b4与曲轴平衡块9的锁止销孔f2之间，由此将驱动齿轮空心轴b2与曲轴平衡块9锁定在一起；当驱动齿轮空心轴b2与曲轴平衡块9锁定在一起后，驱动齿轮b3与曲轴平衡块9同步运动，行星齿轮c3随曲轴平衡块9在内啮合齿圈5内转动，行星齿轮轴c3在驱动齿轮空心轴b2内转动，惰轮12和偏心连杆外啮合齿圈不转动，偏心连杆瓦4与曲轴连杆颈的位置关系不变，由此保持压缩比不变。
当压缩比需要改变时，外部的油压通过驱动油道6使液压油充满加工在曲轴支承盖1和曲轴支承座3上的曲轴支承驱动油道环槽f3；再经加工在曲轴上的曲轴驱动油道f5进入曲轴平衡块9上的驱动油道环槽f1；曲轴平衡块9上的驱动油道环槽f1内的锁止销孔f2内 的锁销10被液压油推出，锁销10被推出后要么停留在驱动齿轮空心轴b2的销孔b4内，要么处于驱动齿轮空心轴b2的销孔b4与行星齿轮轴c4的驱动销孔c5之间，由此将驱动齿轮空心轴b2与行星齿轮轴c4锁定在一起；行星齿轮c3随曲轴平衡块9在内啮合齿圈5内转动，当驱动齿轮空心轴b2与行星齿轮轴c4锁定在一起后，行星齿轮轴c4与驱动齿轮空心轴b2同步转动；由此行星齿轮轴c4将运动传递给驱动齿轮b3，再通过惰轮12传递给偏心连杆外啮合齿圈，偏心连杆瓦4绕曲轴连杆颈旋转，由此改变发动机的压缩比；当在连杆颈最外侧由负偏置半瓦a3转向正偏置半瓦a1时发动机压缩比变大，当由正偏置半瓦a1转向负偏置半瓦a3时发动机压缩比变小。
为了实现米勒循环，在发动机运行过程中，所述液压-机械式可变压缩比的曲轴连杆机构的控制是这样实现的：
当发动机某一缸的活塞处于进气行程下止点前某一时刻，此时控制该缸的压缩比处于最大状态；当活塞继续运动时，控制压缩比由大变小，直到在进气行程下止点后的某一时刻其压缩比变到最小，然后保持压缩比不变；当活塞向上运动到压缩行程上止点前某一位置时，此时控制压缩比由小逐渐变大直到在压缩上止点后某一位置达到最大值，然后保持压缩比不变；当活塞继续运动到做功行程下止点前某一时刻，控制压缩比由大变小，直到在做功行程下止点后的某一时刻其压缩比变到最小，然后保持压缩比不变；当活塞继续向上运动到排气行程上止点前某一位置时，此时控制压缩比由小逐渐变大直到在压缩上止点后某一位置达到最大值，然后保持压缩比不变；然后发动机进入下一个进气行程，如此循环往复即可连续运行。
如此控制可以让发动机的进气压缩比小于做功膨胀比，实现米勒循环。改变压缩比的时刻可以根据发动机的需求进行灵活控制；改变发动机压缩比的速度由所述偏心连杆瓦外啮合齿圈、惰轮、驱动齿轮、行星齿轮、内啮合齿圈的总传动比决定；改变压缩比的时间步长按曲轴转角记，由锁止销孔数目、驱动齿轮空心轴上的销孔数目、驱动销孔数目和内啮合齿圈与行星齿轮的传动比决定。图7为控制示意图。
改变发动机压缩比的速度主要取决与传动比，如：当内啮合齿圈到外啮合齿轮的总传动比为1：1时，也就是说曲轴转一圈，当一直处于传动状态下，发动机的压缩比可以连续地由最大变到最小再变到最大一次，如果是2：1就是两次，那么改变速度就是两倍速。
改变压缩比的时间步长主要取决于锁止销孔数目、驱动齿轮空心轴上的销孔数目、驱动销孔数目。本实施例设置的锁止销孔数目为6个，驱动齿轮空心轴上的销孔数目为8个，驱动销孔数目为6个，时间步长为15度曲轴转角就可以改变一次状态。如果总传动比为1：1，则每15度就可以改变一次状态，如果总传动比为2：1，则每7.5度改变一次状 态。上述“686”的数目设置为一种比较优化的组合方式，当然也可以设计其他的组合方式，任意能够均布在360度内的数目均可以，只要有孔位置相对应就可以改变状态，比如：只要有锁止销孔与空心轴销孔位置相对应，就可以实现锁止状态，只要有驱动销孔与空心轴销孔位置相对应，那么就可以实现驱动状态。

Claims (6)

  1. 一种可实现米勒循环的曲柄连杆机构，其特征在于，包括曲轴(2)、曲轴支承盖(1)、曲轴支承座(3)、驱动齿轮总成(11)、行星齿轮总成(8)和锁销(10)；
    所述曲轴(2)包括相互固定连接的主轴颈(13)、连杆颈(3)和曲轴平衡块(9)；
    所述主轴颈(13)可旋转的安装于所述曲轴支承盖(1)和曲轴支承座(3)形成圆柱孔内；
    所述圆柱孔内壁设有环向内凹的曲轴支承驱动油道环槽(f3)和曲轴支承锁止油道环槽(f4)，曲轴支承座(3)上设有通向曲轴支承驱动油道环槽(f3)的驱动油道(6)，及通向曲轴支承锁止油道环槽(f4)的锁止油道(7)；
    所述连杆颈(3)与主轴颈(13)轴线相错且平行，其外部套设有圆筒状的偏心连杆瓦(4)，偏心连杆瓦(4)内柱面与外柱面的轴线相错且平行，且偏心连杆瓦(4)内壁与连杆颈(3)间隙配合，外壁与连杆大头间隙配合，且其一端的外沿设有外啮合齿圈(15)，外啮合齿圈(15)与内柱面同轴；
    所述曲轴平衡块(9)与连杆颈(3)分局于主轴颈(13)轴线两侧，曲轴平衡块(9)上设有轴线与主轴颈(13)轴线平行的驱动齿轮空心轴安装孔(f7)，驱动齿轮空心轴安装孔(f7)内壁设有环向内凹的锁止油道环槽(f0)和驱动油道环槽(f1)，驱动油道环槽(f1)的槽底均匀排布有多个孔深沿径向的锁止销孔(f2)；
    所述曲轴(2)内还设有将曲轴支承锁止油道环槽(f4)和锁止油道环槽(f0)连通的曲轴锁止油道(f6)，及将曲轴支承驱动油道环槽(f3)和驱动油道环槽(f1)连通的曲轴驱动油道(f5)；
    所述驱动齿轮总成(11)包括驱动齿轮空心轴(b2)及同轴固定于驱动齿轮空心轴(b2)一端的驱动齿轮(b3)；驱动齿轮空心轴(b2)可旋转的安装于驱动齿轮空心轴安装孔(f7)内，其侧壁上开设有多个与锁止油道环槽(f0)对应的锁止油通孔(b1)，及多个与驱动油道环槽(f1)对应的销孔(b4)；驱动齿轮(b3)通过一惰轮(12)与所述外啮合齿圈(15)啮合；所述惰轮(12)的轴固定在曲轴平衡块(9)上，且与主轴颈(13)平行；
    所述行星齿轮总成(8)包括安装于驱动齿轮空心轴(b2)内的行星齿轮轴(c4)，及同轴固定于行星齿轮轴(c4)一端的行星齿轮(c3)；行星齿轮轴(c4)对应于锁止油通孔(b1)的圆周上设有行星齿轮轴锁止油环槽(c1)，对应于销孔(b4)的圆周上设有多个驱动销孔(c5)；行星齿轮轴(c4)上还设有由行星齿轮轴锁止油环槽(c1)通向驱动销孔(c5)的行星齿轮轴锁止油道(c2)；行星齿轮(c3)与内啮合齿圈(5)啮合；内啮合齿圈(5)固定在曲轴支承盖(1)和曲轴支承座(3)所形成的平面上，且与主轴颈(13)同轴；
    所述锁销(10)设置于销孔(b4)内，并能够在液压油的驱动下向锁止销孔(f2)或驱动销 孔(c5)内移动；锁销(10)的长度等于或略小于驱动齿轮空心轴(b2)的壁厚，且大于锁止销孔(f2)和驱动销孔(c5)的深度。
  2. 根据权利要1所述的可实现米勒循环的曲柄连杆机构，其特征在于，所述偏心连杆瓦(4)沿内圆柱面轴线剖分为正偏置半瓦(a1)和负偏置半瓦(a3)，剖分面与偏心矩方向不平行也不垂直。
  3. 根据权利要1所述的可实现米勒循环的曲柄连杆机构，其特征在于，所述锁止销孔(f2)为6个；所述销孔(b4)为8个；所述驱动销孔(c5)为6个，所述锁销(10)为8个。
  4. 一种如权利要求1所述的可实现米勒循环的曲柄连杆机构的控制方法，其特征在于，包括：
    当发动机某一缸的活塞处于进气行程下止点前某一时刻时，控制该缸的压缩比处于最大状态；
    当活塞继续运动时，控制压缩比由大变小，直到在进气行程下止点后的某一时刻其压缩比变到最小，然后保持压缩比不变；
    当活塞向上运动到压缩行程上止点前某一位置时，控制压缩比由小逐渐变大直到在压缩上止点后某一位置，压缩比达到最大值，然后保持压缩比不变；
    当活塞继续运动到做功行程下止点前某一时刻时，控制压缩比由大变小，直到在做功行程下止点后的某一时刻其压缩比变到最小，然后保持压缩比不变；
    当活塞继续向上运动到排气行程上止点前某一位置时，控制压缩比由小逐渐变大，直到在压缩上止点后某一位置，压缩比达到最大值，然后保持压缩比不变。
  5. 根据权利要求4所述的可实现米勒循环的曲柄连杆机构的控制方法，其特征在于，改变发动机压缩比的速度由偏心连杆瓦外啮合齿圈、惰轮、驱动齿轮、行星齿轮和内啮合齿圈的总传动比决定。
  6. 根据权利要求4所述的可实现米勒循环的曲柄连杆机构的控制方法，其特征在于，改变压缩比的时间步长由锁止销孔数目、销孔数目、驱动销孔数目，及内啮合齿圈与行星齿轮的传动比决定。
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