WO2021068284A1 - 一种多功率分配模式的机械液压复合传动装置及控制方法 - Google Patents

一种多功率分配模式的机械液压复合传动装置及控制方法 Download PDF

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Publication number
WO2021068284A1
WO2021068284A1 PCT/CN2019/112636 CN2019112636W WO2021068284A1 WO 2021068284 A1 WO2021068284 A1 WO 2021068284A1 CN 2019112636 W CN2019112636 W CN 2019112636W WO 2021068284 A1 WO2021068284 A1 WO 2021068284A1
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Prior art keywords
gear
transmission
clutch
mechanical
hydraulic
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PCT/CN2019/112636
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English (en)
French (fr)
Inventor
朱镇
蔡英凤
陈龙
夏长高
田翔
韩江义
孙晓东
施德华
王峰
袁朝春
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江苏大学
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Application filed by 江苏大学 filed Critical 江苏大学
Priority to GB2015405.0A priority Critical patent/GB2596367B/en
Priority to JP2021564182A priority patent/JP7094056B2/ja
Priority to DE112019004691.8T priority patent/DE112019004691T5/de
Priority to US17/042,955 priority patent/US10955038B1/en
Priority to CH00271/21A priority patent/CH716823B1/de
Publication of WO2021068284A1 publication Critical patent/WO2021068284A1/zh

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    • 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
    • F16H47/00—Combinations of mechanical gearing with fluid clutches or fluid gearing
    • F16H47/02—Combinations of mechanical gearing with fluid clutches or fluid gearing the fluid gearing being of the volumetric type
    • F16H47/04—Combinations of mechanical gearing with fluid clutches or fluid gearing the fluid gearing being of the volumetric type the mechanical gearing being of the type with members 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
    • F16H3/00—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion
    • F16H3/44—Toothed gearings for conveying rotary motion with variable gear ratio or for reversing rotary motion using 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
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H61/02—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing characterised by the signals used
    • F16H61/0202—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing characterised by the signals used the signals being electric
    • F16H61/0204—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing characterised by the signals used the signals being electric for gearshift control, e.g. control functions for performing shifting or generation of shift signal
    • F16H61/0213—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing characterised by the signals used the signals being electric for gearshift control, e.g. control functions for performing shifting or generation of shift signal characterised by the method for generating shift signals
    • 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
    • F16H37/00—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00
    • F16H37/02—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings
    • F16H37/06—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts
    • F16H37/08—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing
    • F16H37/0833—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing with arrangements for dividing torque between two or more intermediate shafts, i.e. with two or more internal power paths
    • F16H37/084—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings with a plurality of driving or driven shafts; with arrangements for dividing torque between two or more intermediate shafts with differential gearing with arrangements for dividing torque between two or more intermediate shafts, i.e. with two or more internal power paths at least one power path being a continuously variable transmission, i.e. CVT
    • F16H2037/0866—Power split variators with distributing differentials, with the output of the CVT connected or connectable to the output shaft
    • F16H2037/0873—Power split variators with distributing differentials, with the output of the CVT connected or connectable to the output shaft with switching, e.g. to change ranges
    • 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
    • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
    • F16H2061/0075—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing characterised by a particular control method
    • F16H2061/009—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing characterised by a particular control method using formulas or mathematic relations for calculating parameters
    • 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
    • F16H2200/00—Transmissions for multiple ratios
    • F16H2200/003—Transmissions for multiple ratios characterised by the number of forward speeds
    • F16H2200/0043—Transmissions for multiple ratios characterised by the number of forward speeds the gear ratios comprising four forward speeds
    • 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
    • F16H2200/00—Transmissions for multiple ratios
    • F16H2200/20—Transmissions using gears with orbital motion
    • F16H2200/2002—Transmissions using gears with orbital motion characterised by the number of sets of orbital gears
    • F16H2200/2007—Transmissions using gears with orbital motion characterised by the number of sets of orbital gears with two sets of orbital gears
    • 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
    • F16H2200/00—Transmissions for multiple ratios
    • F16H2200/20—Transmissions using gears with orbital motion
    • F16H2200/2002—Transmissions using gears with orbital motion characterised by the number of sets of orbital gears
    • F16H2200/201—Transmissions using gears with orbital motion characterised by the number of sets of orbital gears with three sets of orbital gears
    • 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
    • F16H2200/00—Transmissions for multiple ratios
    • F16H2200/20—Transmissions using gears with orbital motion
    • F16H2200/203—Transmissions using gears with orbital motion characterised by the engaging friction means not of the freewheel type, e.g. friction clutches or brakes
    • F16H2200/2058—Transmissions using gears with orbital motion characterised by the engaging friction means not of the freewheel type, e.g. friction clutches or brakes with eleven engaging means
    • 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
    • F16H2200/00—Transmissions for multiple ratios
    • F16H2200/20—Transmissions using gears with orbital motion
    • F16H2200/203—Transmissions using gears with orbital motion characterised by the engaging friction means not of the freewheel type, e.g. friction clutches or brakes
    • F16H2200/2069—Transmissions using gears with orbital motion characterised by the engaging friction means not of the freewheel type, e.g. friction clutches or brakes using two freewheel mechanism
    • 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
    • F16H2302/00—Determining the way or trajectory to new ratio, e.g. by determining speed, torque or time parameters for shift transition
    • F16H2302/02—Optimizing the way to the new ratio
    • 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
    • F16H2306/00—Shifting
    • F16H2306/30—Shifting characterised by the way or trajectory to a new ratio, e.g. by performing shift according to a particular algorithm or function

Definitions

  • the invention relates to a transmission device and a control method thereof, in particular to a mechanical-hydraulic composite transmission device and a control method in a multi-power distribution mode.
  • Engineering work equipment requires the transmission device to provide a transmission ratio of high torque and low speed when starting or moving operations, and the transmission device is required to provide a transmission ratio of small torque and high speed during the transition process. Therefore, the complexity of the operating conditions of the engineering operation equipment determines that its transmission device has higher requirements than the ordinary vehicle transmission device, and the structure is more complicated. The transmission device of the engineering operation equipment has higher requirements on the transmission ratio and torque at the same time.
  • the transmission device has an important impact on the fuel economy of the engineering operation equipment.
  • the transmission device is used to control the engine to run under economic conditions as much as possible, which is beneficial to improve the fuel economy.
  • the transmission device of traditional construction machinery consists of a hydraulic torque converter and a gearbox, commonly known as a dual-variable system; the role of the hydraulic torque converter is to form a non-rigid connection between the engine and the gearbox, and to reduce the speed and torque of the engine. It is transmitted to the gearbox; the transmission ratio of the final engineering work equipment is changed through the gear shift of the gearbox, which belongs to the traditional mechanical transmission mode and cannot meet the transmission ratio requirements of complex working conditions.
  • the transmission medium of this type of hydraulic transmission system is hydraulic oil, and high-pressure oil is generated during the transmission process to seal between the components.
  • the performance requirements are higher, the higher the pressure resistance requirements of the hydraulic components, the higher the component costs; and the transmission efficiency is not high compared with the mechanical transmission system.
  • the objective of the present invention is to solve the above-mentioned problems and provide a mechanical-hydraulic composite transmission device with multiple power distribution modes and a control method.
  • the invention can provide multiple power distribution modes according to complex operating conditions, make full use of the high-efficiency transmission performance of the mechanical transmission and the stepless speed regulation performance of the hydraulic transmission, and is beneficial to improving the operating efficiency, power and economy of the vehicle.
  • a mechanical-hydraulic composite transmission device with multiple power distribution modes including an input shaft, a shunt mechanism, a mechanical transmission assembly, a hydraulic transmission assembly, and an output shaft.
  • the input shaft is connected to each other in parallel with the mechanical transmission assembly and hydraulic through the shunt mechanism.
  • the transmission assembly is connected, the mechanical transmission assembly and the hydraulic transmission assembly are connected with the output shaft at the same time;
  • the mechanical transmission assembly includes a front planetary row assembly and a rear planetary row assembly connected in series;
  • the front planetary row assembly includes a first clutch, a second clutch, a third clutch, a front planetary row ring gear, a front planetary row planet carrier, a front planetary row sun gear, and a first one-way clutch.
  • the second clutch and the third clutch are connected in parallel with each other and are respectively connected to the front planetary row sun gear, and a first one-way is provided between the third clutch and the front planetary row sun gear.
  • Clutch the output shaft is connected with the front planetary planet carrier;
  • the rear planetary row assembly includes a fourth clutch, a rear planetary row sun gear, a rear planetary row planet carrier, a rear planetary row gear ring, a first brake, a second brake, and a second one-way clutch.
  • the fourth clutch is located in the shunt. Between the mechanism and the rear planetary row sun gear, the first brake is connected to the rear planetary row sun gear, and the rear planetary row planet carrier is connected to the front planetary row ring gear and is connected in parallel with the second brake and the second unidirectional The clutch is connected, and the rear planetary gear ring is connected with the output shaft.
  • the hydraulic transmission component includes a hydraulic transmission input clutch, a hydraulic transmission input gear pair, a hydraulic pump, a hydraulic oil pipe, a hydraulic motor, a hydraulic transmission output gear pair, and a hydraulic transmission output clutch; the hydraulic pump is connected to the shunt mechanism through the hydraulic transmission input gear pair, A hydraulic transmission input clutch is provided between the hydraulic transmission input gear pair and the hydraulic pump.
  • the hydraulic pump is connected to a hydraulic motor through a hydraulic oil pipe, and the hydraulic motor is connected to an output shaft through a hydraulic transmission output gear pair.
  • a hydraulic transmission output clutch is arranged between the output gear pairs through hydraulic transmission.
  • the diverting mechanism includes a diverging brake, a diverging mechanism sun gear, a diverging mechanism planet carrier, and a diverging mechanism gear ring; the input shaft is connected with the diverging mechanism gear ring, the diverging mechanism sun gear is connected with the hydraulic transmission assembly, and the diverging mechanism
  • the sun gear is provided with a shunt brake, and the planet carrier of the shunt mechanism is connected with the mechanical transmission assembly.
  • Pure hydraulic transmission the first brake, the hydraulic transmission input clutch, the hydraulic transmission output clutch, the fourth clutch are engaged, the other clutches and brakes are disengaged, the fourth clutch and the first brake are engaged, the planet carrier of the shunt mechanism is braked, and the hydraulic transmission input clutch It is engaged with the hydraulic transmission output clutch, and the power is driven by the input shaft, the split mechanism gear ring, the split mechanism sun gear, the hydraulic transmission input gear pair, and the input clutch to drive the hydraulic pump to work.
  • the hydraulic pump converts the mechanical power into high-pressure oil and passes through the hydraulic pressure.
  • the pipeline drives the hydraulic motor to work, and the mechanical power output by the hydraulic motor is transmitted to the output shaft through the hydraulic transmission output clutch and the hydraulic transmission output gear pair;
  • the shunt brake is engaged, the hydraulic transmission input clutch, the hydraulic transmission output clutch, the second brake and the fourth clutch are disengaged, and the combination of other clutches and brakes realizes the gears with different transmission ratios in the pure mechanical transmission mode Switching, the shunt brake is engaged, the hydraulic transmission input clutch and the hydraulic transmission output clutch are separated, the hydraulic transmission component has no power input, and the power is transmitted to the mechanical transmission component through the input shaft, the diverter gear ring, and the diverter mechanism planetary carrier. The mechanical transmission component is adjusted by the speed ratio. Then output from the output shaft;
  • Mechanical hydraulic compound transmission hydraulic transmission input clutch and hydraulic transmission output clutch are engaged, the shunt brake, the third clutch, the first one-way clutch, the second one-way clutch are separated, and the mechanical hydraulic compound is realized through the combination of other clutches and brakes.
  • the gears with different transmission ratios are switched.
  • the power flows through the input shaft and the ring gear of the splitter mechanism to the planetary carrier of the splitter mechanism.
  • the flow is split at the planetary carrier of the splitter mechanism and flows to the mechanical transmission assembly and the sun gear of the splitter mechanism respectively.
  • the sun gear of the mechanism is connected with the hydraulic transmission assembly, and finally merged to the output shaft for output.
  • the purely mechanical transmission modes include mechanical transmission I gear, mechanical transmission II gear, mechanical transmission III gear, and mechanical transmission IV gear.
  • the specific implementation methods are as follows:
  • Mechanical transmission gear I the first brake, the first clutch and the second clutch are disengaged, and the third clutch, the first one-way clutch and the second one-way clutch are engaged.
  • Second gear of mechanical transmission the first brake, the first clutch, the second clutch and the second one-way clutch are disengaged, the third clutch and the first one-way clutch are engaged, and the power flows from the planet carrier of the splitter mechanism through the third clutch and the first one in order.
  • the front planetary row sun gear to the front planetary row planet carrier, the power at the front planetary row planet carrier is split to the output shaft and the front planetary row gear ring respectively, and the power of the front planetary row gear ring is sequentially transmitted to the rear planetary row planets The carrier and the rear planetary gear ring, finally converge to the output shaft;
  • Mechanical transmission third gear the first brake, the second clutch and the second one-way clutch are disengaged, the first clutch, the third clutch and the first one-way clutch are engaged, and the power is input from the planet carrier of the splitter mechanism to the first clutch.
  • the flow forward planetary row gear ring and the front planetary row sun gear converge on the front planetary row planet carrier, and finally output from the output shaft;
  • Mechanical transmission gear IV the second clutch, the third clutch, the first one-way clutch and the second one-way clutch are disengaged, the first clutch and the first brake are engaged, and the power flows from the planet carrier of the shunt mechanism through the first clutch and the front planetary row in turn
  • the gear ring, the rear planetary row planet carrier, and the rear planetary row gear ring are output to the output shaft.
  • the mechanical-hydraulic compound transmission modes include compound transmission I, compound transmission II, compound transmission III, and compound transmission IV.
  • the specific implementation methods are as follows:
  • Compound transmission gear I the first brake, the first clutch and the fourth clutch are disengaged, the second brake and the second clutch are engaged, the power is diverted from the gear ring of the splitter mechanism from the planet carrier of the splitter mechanism, and a part of the power is transferred from the planet carrier of the splitter mechanism.
  • the sun gear of the splitter mechanism flows to the hydraulic transmission assembly, and the other part of the power flows from the planet carrier of the splitter mechanism through the second clutch, the front planetary sun gear, the front planetary planet carrier, and finally converges to the output shaft for output;
  • Compound transmission II gear the second brake, the first clutch and the fourth clutch are disengaged, the first brake and the second clutch are engaged, the power is diverted from the gear ring of the splitter mechanism from the planet carrier of the splitter mechanism, and a part of the power is transferred from the planet carrier of the splitter mechanism.
  • the sun wheel of the splitter mechanism flows to the hydraulic transmission assembly, and the other part of the power flows from the planetary carrier of the splitter mechanism to the front planetary planet carrier through the second clutch, the front planetary sun gear and the front planetary planetary carrier.
  • the power at the front planetary planetary carrier is split to the output shaft and the front planet respectively.
  • Row gear ring, the power of the front planetary row gear ring is sequentially transmitted to the rear planetary row planet carrier and the rear planetary row gear ring, and the divided three sets of power are finally converged to the output shaft;
  • Compound transmission third gear the first brake, the second brake and the fourth clutch are disengaged, the first clutch and the second clutch are engaged, the power is diverted from the gear ring of the splitter mechanism from the planet carrier of the splitter mechanism, and a part of the power is transferred from the planet carrier of the splitter mechanism.
  • the sun gear of the splitter mechanism flows to the hydraulic transmission assembly, and the other part of the power is input from the planet carrier of the splitter mechanism to the second clutch.
  • the forward planetary gear ring and the front planetary sun gear respectively converge on the front planetary carrier, the mechanical transmission assembly and The hydraulic transmission components converge and output on the output shaft;
  • Compound transmission gear IV the first brake, the first clutch and the second clutch are disengaged, the second brake and the fourth clutch are engaged, the power is diverted from the gear ring of the splitter mechanism from the planetary carrier of the splitter mechanism, and a part of the power is transferred from the planetary carrier of the splitter mechanism.
  • the sun gear of the splitter mechanism flows to the hydraulic transmission assembly, and the other part of the power flows from the planet carrier of the split mechanism through the fourth clutch, the rear planetary row sun gear, and the rear planetary row gear ring to the output shaft.
  • the mechanical transmission assembly and the hydraulic transmission assembly are combined and output on the output shaft. .
  • B 1 is the shunt brake
  • B 2 is the first brake
  • B 3 is the second brake
  • C 1 is the hydraulic transmission input clutch
  • C 2 is the hydraulic transmission output clutch
  • C 3 is the fourth clutch
  • C 4 is the first clutch
  • C 6 is the third clutch
  • F 1 is a first one-way clutch
  • F 2 is a second one-way clutch.
  • compound transmission I ⁇ compound transmission II involves 2 shift elements
  • compound transmission II ⁇ compound transmission III involves 2 shift elements
  • compound transmission III ⁇ Compound transmission gear IV involves 4 shifting elements
  • compound transmission gear I ⁇ composite transmission gear III involves 2 shifting components
  • compound transmission gear I ⁇ composite transmission gear IV involves 2 shifting elements
  • compound transmission II Gear ⁇ Compound transmission gear IV involves 4 shifting elements
  • Gear shifts involving 2 or less shift elements can be optimized through no more than 3 experiments; gear shifts involving 3 or 4 shift elements without interaction are analyzed through orthogonal tables; The gear shift of 4 shift elements makes full use of the 4 columns of the orthogonal table, and the gear shift involving 3 shift elements selects any 3 columns of the orthogonal table;
  • the angular velocity change of the output shaft is defined as:
  • ⁇ is the change in angular velocity of the output shaft
  • I the steady-state angular velocity of the output shaft
  • ⁇ o min is the lowest angular velocity of the output shaft
  • the angular velocity change rate of the output shaft is defined as:
  • ⁇ is the rate of change of the angular velocity of the output shaft
  • the output shaft torque is:
  • T o is the torque of the output shaft
  • J o is the moment of inertia of the output shaft
  • the second derivative of the angular velocity of the output shaft is defined as:
  • ⁇ is the second differential of the angular velocity of the output shaft
  • the impact degree is the rate of change of the longitudinal acceleration of the vehicle:
  • j is the degree of impact
  • r d is the tire power radius
  • i g is the transmission ratio
  • i 0 is the drive axle transmission ratio
  • ⁇ is a comprehensive evaluation index
  • ⁇ k is a single evaluation index
  • ⁇ k min / ⁇ k max is the upper/lower limit of a single evaluation index
  • ⁇ k is the weight coefficient
  • ⁇ is a comprehensive evaluation index
  • ⁇ k is a single evaluation index
  • ⁇ k min / ⁇ k max is the upper/lower limit of a single evaluation index
  • ⁇ k is the weight coefficient
  • the "advance” and “delay” time can be increased or decreased, or different “advance” time and “delay” time can be selected until the requirements are met.
  • the switch from the mechanical I gear to the mechanical II gear involves one shift element
  • the switch from the mechanical II gear to the mechanical III gear involves one shift element
  • the mechanical III gear to switch to the mechanical IV Gear involves 3 shifting elements, switching from mechanical I to mechanical III involves 2 shifting elements, switching from mechanical I to mechanical IV involves 5 shifting elements, and switching from mechanical II to mechanical IV involves Up to 4 shift elements;
  • Gear shifts involving 2 or less gear shifting elements can be optimized through no more than 3 tests; gear shifting involving 3 shifting elements and 2 of the shifting elements have an interactive effect, involving 5 gears Shifting elements and three of the shifting elements have interactive gear switching, and involving 4 shifting elements and two of the shifting elements have interactive gear switching are all analyzed by orthogonal table. Choose the corresponding list;
  • the angular velocity change of the output shaft is defined as:
  • ⁇ is the change in angular velocity of the output shaft
  • Is the steady-state angular velocity of the output shaft
  • ⁇ o min is the lowest angular velocity of the output shaft
  • the rate of change of the output shaft angular velocity is defined as:
  • ⁇ is the rate of change of the angular velocity of the output shaft
  • the output shaft torque is determined by the formula:
  • T o is the torque of the output shaft
  • J o is the moment of inertia of the output shaft
  • the second derivative of the angular velocity of the output shaft is defined as:
  • ⁇ is the second differential of the angular velocity of the output shaft
  • the impact degree is the rate of change of the longitudinal acceleration of the vehicle, which is determined by the formula:
  • j is the degree of impact
  • r d is the tire power radius
  • i g is the transmission ratio
  • i 0 is the drive axle transmission ratio
  • the sources of variance are mainly in three parts: 1 “a”, “b”, “c”, “d” and “e”; 2 “e ⁇ c” (including e ⁇ c1 and e ⁇ c2), “e ⁇ d” (including two parts e ⁇ d1 and e ⁇ d2) and “c ⁇ d” (including two parts c ⁇ d1 and c ⁇ d2); 3error e * ;
  • the mechanical gear optimization scheme is determined by the following formula:
  • ⁇ is a comprehensive evaluation index
  • ⁇ k is a single evaluation index
  • ⁇ k min / ⁇ k max is the upper/lower limit of a single evaluation index
  • ⁇ k is a weighting coefficient
  • the present invention can switch between hydraulic transmission, mechanical-hydraulic composite transmission and mechanical transmission, and each transmission mode has multiple gears to choose from, and can provide multiple power distribution modes according to complex operating conditions. Making full use of the transmission performance of mechanical transmission and the stepless speed regulation performance of hydraulic transmission is conducive to improving the operating efficiency, power and economy of the vehicle.
  • the mechanical gear shifting device adopts a one-way clutch to not only realize engine braking, but also to use all shifting elements evenly to increase the service life.
  • Figure 1 is a schematic diagram of the structure of the present invention
  • Figure 2 is a schematic diagram of the power flow of the pure hydraulic transmission of the present invention.
  • Figure 3 is a schematic diagram of the power flow of the first gear of the mechanical transmission of the present invention.
  • Figure 4 is a schematic diagram of the power flow in the second gear of the mechanical transmission of the present invention.
  • Figure 5 is a schematic diagram of the power flow of the third gear of the mechanical transmission of the present invention.
  • Figure 6 is a schematic diagram of the power flow of gear IV of the mechanical transmission of the present invention.
  • Figure 7 is a schematic diagram of the power flow of the first gear of the compound transmission of the present invention.
  • Figure 8 is a schematic diagram of the power flow in the second gear of the compound transmission of the present invention.
  • Figure 9 is a schematic diagram of the power flow of the third gear of the compound transmission of the present invention.
  • Figure 10 is a schematic diagram of the power flow in the fourth gear of the compound transmission of the present invention.
  • a mechanical-hydraulic composite transmission device with multiple power distribution modes includes an input shaft 1, a shunt mechanism 2, a mechanical transmission assembly 3, a hydraulic transmission assembly 4, and an output shaft 5.
  • the input shaft 1 passes through the shunt mechanism 2 is connected to the mechanical transmission assembly 3 and the hydraulic transmission assembly 4 connected in parallel with each other.
  • the mechanical transmission assembly 3 and the hydraulic transmission assembly 4 are simultaneously connected with the output shaft 5;
  • the mechanical transmission assembly 3 includes a front planetary assembly 31 and Rear planetary row assembly 32;
  • the front planetary row assembly 31 includes a first clutch 311, a second clutch 312, a third clutch 313, a front planetary row ring gear 314, a front planetary row carrier 315, a front planetary row sun gear 316, and a first one-way clutch 317
  • the first clutch 311 is located between the splitter mechanism 2 and the front planetary row ring gear 314, the second clutch 312 and the third clutch 313 are connected in parallel with each other and are respectively connected to the front planetary row sun gear 316, and the third clutch
  • a first one-way clutch 317 is provided between the 313 and the front planetary row sun gear 316, and the output shaft 5 is connected to the front planetary row planet carrier 315;
  • the rear planetary row assembly 32 includes a fourth clutch 321, a rear planetary row sun gear 322, a rear planetary row carrier 323, a rear planetary row gear ring 324, a first brake 325, a second brake 326, and a second one-way clutch 327 ,
  • the fourth clutch 321 is located between the shunt mechanism 2 and the rear planetary row sun gear 322, the first brake 325 is connected to the rear planetary row sun gear 322, and the rear planetary row carrier 323 is connected to the front planetary row gear ring 314 is connected with the second brake 326 and the second one-way clutch 327 connected in parallel with each other, and the rear planetary gear ring 324 is connected with the output shaft 5.
  • the hydraulic transmission assembly 4 includes a hydraulic transmission input clutch 41, a hydraulic transmission input gear pair 42, a hydraulic pump 43, a hydraulic oil pipe 44, a hydraulic motor 45, a hydraulic transmission output gear pair 46, and a hydraulic transmission output clutch 47;
  • the transmission input gear pair 42 is connected to the shunt mechanism 2.
  • a hydraulic transmission input clutch 41 is provided between the hydraulic transmission input gear pair 42 and the hydraulic pump 43.
  • the hydraulic pump 43 is connected to the hydraulic motor 45 through a hydraulic oil pipe 44.
  • the hydraulic motor 45 is connected to the output shaft 5 through a hydraulic transmission output gear pair 46, and a hydraulic transmission output clutch 47 is provided between the hydraulic motor 45 through a hydraulic transmission output gear pair 46.
  • the split mechanism 2 includes a split brake 21, a split mechanism sun gear 22, a split mechanism planet carrier 23, and a split mechanism ring gear 24; the input shaft 1 is connected with a split mechanism ring gear 24, and the split mechanism sun gear 22 is connected to hydraulic
  • the transmission assembly 4 is connected, a shunt brake 21 is provided on the sun gear 22 of the shunt mechanism, and the planet carrier 23 of the shunt mechanism is connected with the mechanical transmission assembly 3.
  • the pure hydraulic transmission the first brake 325, the hydraulic transmission input clutch 41, the hydraulic transmission output clutch 47, the fourth clutch 321 are engaged, the other clutches and brakes are disengaged, the fourth clutch 321 and the first brake 325 are engaged,
  • the shunt mechanism planetary carrier 23 is braked, the hydraulic transmission input clutch 41 and the hydraulic transmission output clutch 47 are engaged, and the power passes through the input shaft 1, the shunt mechanism ring gear 24, the shunt mechanism sun gear 22, the hydraulic transmission input gear pair 42, and the input clutch 41
  • the hydraulic pump 43 is driven to work.
  • the hydraulic pump 43 converts mechanical power into high-pressure oil and drives the hydraulic motor 45 to work through the hydraulic pipe 44.
  • the mechanical power output by the hydraulic motor 45 is transmitted through the hydraulic transmission output clutch 47 and the hydraulic transmission output gear pair 46 Transfer to the output shaft 5;
  • the purely mechanical transmission the shunt brake 21 is engaged, the hydraulic transmission input clutch 41, the hydraulic transmission output clutch 47, the second brake 326, and the fourth clutch 321 are disengaged.
  • the combination between the two realizes switching between gears with different transmission ratios in the purely mechanical transmission mode, the shunt brake 21 is engaged, the hydraulic transmission input clutch 41 and the hydraulic transmission output clutch 47 are separated, the hydraulic transmission component 4 has no power input, and the power passes through the input shaft 1.
  • the splitter mechanism ring gear 24 and the splitter mechanism planet carrier 23 transmit the mechanical transmission assembly 3, and the mechanical transmission assembly 3 is output from the output shaft 5 after the speed ratio is adjusted;
  • the mechanical hydraulic compound transmission the hydraulic transmission input clutch 41 and the hydraulic transmission output clutch 47 are engaged, the shunt brake 21, the third clutch 313, the first one-way clutch 317, and the second one-way clutch are engaged.
  • the clutch 327 is disengaged, and through the combination of other clutches and brakes, the switching between different transmission ratio gears in the mechanical-hydraulic compound transmission mode is realized.
  • the power passes through the input shaft 1, the splitter mechanism ring gear 24 to the splitter mechanism planet carrier 23, and The flow is divided at the planet carrier 23 of the dividing mechanism, and flows to the mechanical transmission assembly 3 and the dividing mechanism sun gear 22 respectively.
  • the dividing mechanism sun gear 22 is connected with the hydraulic transmission assembly 4 and finally merges to the output shaft 5 for output.
  • the purely mechanical transmission modes include mechanical transmission I gear, mechanical transmission II gear, mechanical transmission III gear, and mechanical transmission IV gear.
  • the specific implementation methods are as follows:
  • the first gear of the mechanical transmission the first brake 325, the first clutch 311 and the second clutch 312 are disengaged, the third clutch 313, the first one-way clutch 317 and the second one-way clutch 327 are engaged, and the power is split from
  • the mechanism planet carrier 23 passes through the third clutch 313, the first one-way clutch 317, and the front planet carrier 315 of the front planetary row sun gear 316 to the output shaft 5 in sequence;
  • the second gear of the mechanical transmission the first brake 325, the first clutch 311, the second clutch 312 and the second one-way clutch 327 are disengaged, the third clutch 313 and the first one-way clutch 317 are engaged, and the power is split from
  • the mechanism planet carrier 23 sequentially passes through the third clutch 313, the first one-way clutch 317, and the front planetary row sun gear 316 to the front planetary row planet carrier 315.
  • the power at the front planetary row planet carrier 315 is divided to the output shaft 5 and the front planetary row respectively.
  • Ring gear 314, the power of the front planetary row gear ring 314 is sequentially transmitted to the rear planetary row planet carrier 323 and the rear planetary row gear ring 324, and finally converges to the output shaft 5;
  • the third gear of mechanical transmission the first brake 325, the second clutch 312 and the second one-way clutch 327 are disengaged, the first clutch 311, the third clutch 313 and the first one-way clutch 317 are engaged, and the power is split from
  • the mechanism planet carrier 23 is input to the first clutch 311 to split the flow, respectively flow to the front planetary row gear ring 314 and the front planetary sun gear 316 to converge in the front planetary row planet carrier 315, and finally output from the output shaft 5;
  • the mechanical transmission gear IV the second clutch 312, the third clutch 313, the first one-way clutch 317 and the second one-way clutch 327 are disengaged, the first clutch 311 and the first brake 325 are engaged, and the power is split from
  • the mechanism planet carrier 23 is outputted to the output shaft 5 through the first clutch 311, the front planetary gear ring 314, the rear planetary gear carrier 323, and the rear planetary gear ring 324 in sequence.
  • the mechanical-hydraulic compound transmission modes include compound transmission I, compound transmission II, compound transmission III, and compound transmission IV.
  • the specific implementation methods are as follows:
  • the first gear of the compound transmission the first brake 325, the first clutch 311 and the fourth clutch 321 are disengaged, the second brake 326 and the second clutch 312 are engaged, and the power is from the splitter mechanism ring gear 24 from the splitter mechanism planet carrier Part of the power flows from the planet carrier 23 of the split mechanism through the sun gear 22 of the split mechanism to the hydraulic transmission assembly 4, and the other part of the power flows from the planet carrier 23 of the split mechanism through the second clutch 312, the front planetary sun gear 316, and the front planet. Row planet carrier 315, and finally converge to output shaft 5 for output;
  • the second gear of the compound transmission the second brake 326, the first clutch 311 and the fourth clutch 321 are disengaged, the first brake 325 and the second clutch 312 are engaged, and the power is transferred from the split mechanism ring gear 24 from the split mechanism planet carrier Part of the power flows from the planet carrier 23 of the split mechanism to the hydraulic transmission assembly 4 through the sun gear 22 of the split mechanism, and the other part of the power flows from the planet carrier 23 of the split mechanism through the second clutch 312 and the front planetary sun gear 316 to the front planet.
  • Planet carrier 315 the power at the front planet carrier 315 is split to the output shaft 5 and the front planetary gear ring 314, and the power of the front planetary gear ring 314 is sequentially transmitted to the rear planetary carrier 323 and the rear planetary gear In the ring gear 324, the divided three sets of power finally converge to the output shaft 5;
  • the third gear of the compound transmission the first brake 325, the second brake 326 and the fourth clutch 321 are disengaged, the first clutch 311 and the second clutch 312 are engaged, and the power is from the splitter mechanism ring gear 24 to the splitter mechanism planet carrier Part of the power flows from the planetary carrier 23 of the splitter mechanism to the hydraulic transmission assembly 4 through the sun gear 22 of the splitter mechanism, and the other part of the power is input from the planetary carrier 23 of the splitter mechanism to the second clutch 312 to split, respectively flow to the forward planetary gear ring 314 and the front planetary row sun gear 316 converge on the front planetary carrier 315, and the mechanical transmission assembly 3 and hydraulic transmission assembly 4 converge and output on the output shaft 5;
  • the fourth gear of the compound transmission the first brake 325, the first clutch 311 and the second clutch 312 are disengaged, the second brake 326 and the fourth clutch 321 are engaged, and the power is transferred from the split mechanism ring gear 24 from the split mechanism planet carrier Part of the power flows from the planet carrier 23 of the split mechanism through the sun gear 22 of the split mechanism to the hydraulic transmission assembly 4, and the other part of the power flows from the planet carrier 23 of the split mechanism through the fourth clutch 321, the rear planetary sun gear 322, and the rear planet.
  • the gear ring 324 is connected to the output shaft 5, and the mechanical transmission assembly 3 and the hydraulic transmission assembly 4 converge and output on the output shaft 5.
  • compound transmission I ⁇ compound transmission II involves 2 shift elements
  • compound transmission II ⁇ compound transmission III involves 2 shift elements
  • compound transmission III ⁇ Compound transmission gear IV involves 4 shifting elements
  • compound transmission gear I ⁇ composite transmission gear III involves 2 shifting components
  • compound transmission gear I ⁇ composite transmission gear IV involves 2 shifting elements
  • compound transmission II Gear ⁇ Compound transmission gear IV involves 4 shifting elements
  • Gear shifts involving 2 or less shift elements can be optimized through no more than 3 experiments; gear shifts involving 3 or 4 shift elements without interaction are analyzed through orthogonal tables; The gear shift of 4 shift elements makes full use of the 4 columns of the orthogonal table, and the gear shift involving 3 shift elements selects any 3 columns of the orthogonal table;
  • the angular velocity change of the output shaft is defined as:
  • ⁇ is the change in angular velocity of the output shaft
  • I the steady-state angular velocity of the output shaft
  • ⁇ o min is the lowest angular velocity of the output shaft
  • the angular velocity change rate of the output shaft is defined as:
  • ⁇ is the rate of change of the angular velocity of the output shaft
  • T o is the torque of the output shaft
  • J o is the moment of inertia of the output shaft
  • the second derivative of the angular velocity of the output shaft is:
  • ⁇ is the second differential of the angular velocity of the output shaft
  • the impact degree is the rate of change of the longitudinal acceleration of the vehicle:
  • j is the degree of impact
  • r d is the tire power radius
  • i g is the transmission ratio
  • i 0 is the drive axle transmission ratio
  • the "four factors" are determined to be brake B2, brake B3, clutch C3, and clutch C5.
  • the switching timing of these four shifting elements is the four factors that affect the shifting quality of the transmission system.
  • the time of "advance” and “delay” can be selected according to the actual situation.
  • the time of “advance” and “delay” can be the same, or It can be different.
  • the time for selecting "advance” and “delay” is 0.3s.
  • ⁇ is a comprehensive evaluation index
  • ⁇ k is a single evaluation index
  • ⁇ k min / ⁇ k max is the upper/lower limit of a single evaluation index
  • ⁇ k is the weight coefficient
  • ⁇ is a comprehensive evaluation index
  • ⁇ k is a single evaluation index
  • ⁇ k min / ⁇ k max is the upper/lower limit of a single evaluation index
  • ⁇ k is a weight coefficient
  • the "advance” and “delay” time can be increased or decreased, or different “advance” time and “delay” time can be selected until the requirements are met.
  • the switch from the mechanical I gear to the mechanical II gear involves one shift element
  • the switch from the mechanical II gear to the mechanical III gear involves one shift element
  • the mechanical III gear to switch to the mechanical IV Gear involves 3 shifting elements, switching from mechanical I to mechanical III involves 2 shifting elements, switching from mechanical I to mechanical IV involves 5 shifting elements, and switching from mechanical II to mechanical IV involves Up to 4 shift elements;
  • Gear shifts involving 2 or less gear shifting elements can be optimized through no more than 3 tests; gear shifting involving 3 shifting elements and 2 of the shifting elements have an interactive effect, involving 5 gears Shifting elements and three of the shifting elements have interactive gear switching, and involving 4 shifting elements and two of the shifting elements have interactive gear switching are all analyzed by orthogonal table. Choose the corresponding list;
  • the angular velocity change of the output shaft is defined as:
  • ⁇ is the angular velocity change of the output shaft
  • Is the steady-state angular velocity of the output shaft
  • ⁇ o min is the minimum angular velocity of the output shaft
  • rate of change of the output shaft angular velocity is defined as:
  • ⁇ is the rate of change of the angular velocity of the output shaft
  • the output shaft torque is determined by formula (3):
  • T o is the torque of the output shaft
  • J o is the moment of inertia of the output shaft
  • the second derivative of the angular velocity of the output shaft is defined as:
  • ⁇ is the second differential of the angular velocity of the output shaft
  • the impact degree is the rate of change of the longitudinal acceleration of the vehicle, which is determined by equation (5):
  • j is the degree of impact
  • r d is the tire power radius
  • i g is the transmission ratio
  • i 0 is the drive axle transmission ratio
  • the time of "advance” and “delay” can be selected according to the actual situation.
  • the time of "advance” and “delay” can be the same, or It can be different. As far as this example is concerned, because there are many components involved and there are interactions among them, the time for selecting "advanced” and “delayed” is 0.5s.
  • the sources of variance are mainly in three parts: 1 “a”, “b”, “c”, “d” and “e”; 2 “e ⁇ c” (including e ⁇ c1 and e ⁇ c2), “e ⁇ d” (including two parts e ⁇ d1 and e ⁇ d2) and “c ⁇ d” (including two parts c ⁇ d1 and c ⁇ d2); 3error e * ;
  • the mechanical gear optimization scheme is determined by the following formula:
  • ⁇ is a comprehensive evaluation index
  • ⁇ k is a single evaluation index
  • ⁇ k min / ⁇ k max is the upper/lower limit of a single evaluation index
  • ⁇ k is a weighting coefficient
  • Switching from mechanical third gear to mechanical fourth gear involves three shifting elements. Among them, the third clutch 313 and the first one-way clutch 317 interact with each other.
  • the L 9 (3 4 ) orthogonal table can be used for analysis.
  • the fourth factor column needs to be replaced with an interaction column.
  • Switching from mechanical second gear to mechanical fourth gear involves 4 shifting elements.
  • the third clutch 313 and the first one-way clutch 317 interact with each other.
  • the L 27 (3 13 ) orthogonal table can be used for analysis.
  • the extra column is vacant.

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Abstract

一种多功率分配模式的机械液压复合传动装置及控制方法，包括输入轴（1）、分流机构（2）、机械传动组件（3）、液压传动组件（4）和输出轴（5），所述输入轴（1）通过分流机构（2）与相互并联的机械传动组件（3）和液压传动组件（4）连接，所述机械传动组件（3）和液压传动组件（4）同时与输出轴（5）连接；通过换挡元件之间的组合切换实现纯液压传动、机械液压复合传动和纯机械传动三种类型的传动。有多个档位可供选择，能够根据复杂的作业工况提供多种功率分配模式，充分利用机械传动的传动性能和液压传动的无级调速性能，有利于提高车辆的作业效率、动力性和经济性。机械档换挡装置采用单向离合器既能实现发动机制动，又能均匀使用各换挡元件以增加使用寿命。

Description

一种多功率分配模式的机械液压复合传动装置及控制方法 技术领域
本发明涉及一种传动装置及其控制方法，特别提供了一种多功率分配模式的机械液压复合传动装置及控制方法。
背景技术
工程作业设备在起步或移动作业时要求传动装置能够提供大扭矩低转速的传动比，而在转场过程中要求传动装置能够提供小扭矩高转速的传动比。因此工程作业设备使用工况的复杂性决定了其传动装置比普通车辆传动装置的要求更高，结构更复杂，工程作业设备的传动装置对传动比和扭矩同时有着较高的要求。
传动装置对工程作业设备的燃油经济性有着重要影响，通过传动装置将发动机尽量控制在经济工况下运行，有利于提高燃油经济性。但是传统工程机械的传动装置由液力变矩器和变速箱构成，俗称双变系统；液力变矩器的作用是使发动机和变速箱之间形成非刚性连接，并将发动机的转速和扭矩传递到变速箱；最终工程作业设备的传动比变换还是通过变速箱的挡位切换实现，属于传统的机械传动模式，无法满足复杂工况对传动比的要求。
目前工程作业设备出现了通过发动机驱动液压泵进而驱动液压马达实现行走的液压传动系统，但是这类液压传动系统的传动介质是液压油，传动过程中会产生高压油液，对元件之间的密封性能要求较高，液压元件的耐压性能要求越高、元件成本越高；并且与机械传动系统相比传动效率不高。
发明内容
发明目的：本发明的目的是为了解决上述问题，提供一种多功率分配模式的机械液压复合传动装置及控制方法。本发明能够根据复杂的作业工况提供多种功率分配模式，充分利用机械传动的高效传动性能和液压传动的无级调速性能，有利于提高车辆的作业效率、动力性和经济性。
技术方案：一种多功率分配模式的机械液压复合传动装置,包括输入轴、分流机构、机械传动组件、液压传动组件和输出轴，所述输入轴通过分流机构与相互并联的机械传动组件和液压传动组件连接，所述机械传动组件和液压传动组件同时与输出轴连接；所述机械传动组件包括相互串联的前行星排组件和后行星排组件；
所述前行星排组件包括第一离合器、第二离合器、第三离合器、前行星排齿圈、前行星排行星架、前行星排太阳轮和第一单向离合器，所述第一离合器位于分流机构与前行星排齿圈之间，所述第二离合器和第三离合器相互并联并且分别与前行星排太阳轮连接，所述第三离合器与前行星排太阳轮之间设有第一单向离合器，所述输出轴与前行星排行星架连接；
所述后行星排组件包括第四离合器、后行星排太阳轮，后行星排行星架，后行星排齿 圈、第一制动器、第二制动器和第二单向离合器，所述第四离合器位于分流机构与后行星排太阳轮之间，所述第一制动器与后行星排太阳轮连接，所述后行星排行星架与前行星排齿圈连接并且与相互并联的第二制动器和第二单向离合器连接，所述后行星排齿圈与输出轴连接。
液压传动组件包括液压传动输入离合器、液压传动输入齿轮副、液压泵、液压油管、液压马达、液压传动输出齿轮副、液压传动输出离合器；所述液压泵通过液压传动输入齿轮副与分流机构连接，所述液压传动输入齿轮副与液压泵之间设有液压传动输入离合器，所述液压泵通过液压油管与液压马达连接，所述液压马达通过液压传动输出齿轮副与输出轴连接，所述液压马达通过液压传动输出齿轮副之间设有液压传动输出离合器。
所述分流机构包括分流制动器、分流机构太阳轮、分流机构行星架、分流机构齿圈；所述输入轴与分流机构齿圈连接，所述分流机构太阳轮与液压传动组件连接，所述分流机构太阳轮上设有分流制动器，所述分流机构行星架与机械传动组件连接。
通过制动器和离合器之间的组合切换实现纯液压传动、机械液压复合传动和纯机械传动三种类型的传动，具体传动类型如下：
纯液压传动：第一制动器、液压传动输入离合器、液压传动输出离合器、第四离合器接合，其他离合器和制动器分离，第四离合器和第一制动器接合，分流机构行星架被制动，液压传动输入离合器和液压传动输出离合器接合，动力经输入轴、分流机构齿圈、分流机构太阳轮、液压传动输入齿轮副、输入离合器驱动液压泵工作，所述液压泵将机械动力转化为高压油液并经过液压管道驱动液压马达工作，液压马达输出的机械动力经液压传动输出离合器和液压传动输出齿轮副传递到输出轴；
纯机械传动：分流制动器接合，液压传动输入离合器、液压传动输出离合器、第二制动器和第四离合器分离，通过其他离合器和制动器之间的组合实现纯机械传动模式中不同传动比的挡位之间切换，分流制动器接合，液压传动输入离合器和液压传动输出离合器分离，液压传动组件无动力输入，动力经输入轴、分流机构齿圈、分流机构行星架传递机械传动组件，机械传动组件经过速比调节后从输出轴输出；
机械液压复合传动：液压传动输入离合器和液压传动输出离合器接合，分流制动器、第三离合器、第一单向离合器、第二单向离合器分离，通过其他离合器和和制动器之间的组合实现机械液压复合传动模式中不同传动比的挡位之间切换，动力经输入轴、分流机构齿圈至分流机构行星架，在分流机构行星架处分流，分别流向机械传动组件和分流机构太阳轮，所述分流机构太阳轮与液压传动组件连接，最后汇合至输出轴输出。
所述纯机械传动模式包括机械传动Ⅰ档、机械传动Ⅱ档、机械传动Ⅲ档和机械传动Ⅳ档，具体实现方法如下：
机械传动Ⅰ档：第一制动器、第一离合器和第二离合器分离，第三离合器、第一单向离合器和第二单向离合器接合，动力从分流机构行星架依次经第三离合器、第一单向离合 器、前行星排太阳轮前行星排行星架至输出轴；
机械传动Ⅱ档：第一制动器、第一离合器、第二离合器和第二单向离合器分离，第三离合器、第一单向离合器接合，动力从分流机构行星架依次经第三离合器、第一单向离合器、前行星排太阳轮至前行星排行星架，前行星排行星架处动力分别分流到输出轴和前行星排齿圈，所述前行星排齿圈的动力依次传递给后行星排行星架和后行星排齿圈，最后汇流至输出轴；
机械传动Ⅲ档：第一制动器、第二离合器和第二单向离合器分离，第一离合器、第三离合器和第一单向离合器接合，动力从分流机构行星架输入至第一离合器处分流，分别流向前行星排齿圈和前行星排太阳轮在前行星排行星架汇流，最后从输出轴输出；
机械传动Ⅳ档：第二离合器、第三离合器、第一单向离合器和第二单向离合器分离，第一离合器和第一制动器接合，动力从分流机构行星架依次经第一离合器、前行星排齿圈、后行星排行星架、后行星排齿圈至输出轴输出。
所述机械液压复合传动模式包括复合传动Ⅰ档、复合传动Ⅱ档、复合传动Ⅲ档和复合传动Ⅳ档，具体实现方法如下：
复合传动Ⅰ档：第一制动器、第一离合器和第四离合器分离，第二制动器和第二离合器接合，动力从分流机构齿圈从分流机构行星架处进行分流，一部分动力从分流机构行星架经分流机构太阳轮流向液压传动组件，另一部分动力从分流机构行星架依次经第二离合器、前行星排太阳轮、前行星排行星架，最后汇流至输出轴输出；
复合传动Ⅱ档：第二制动器、第一离合器和第四离合器分离，第一制动器和第二离合器接合，动力从分流机构齿圈从分流机构行星架处进行分流，一部分动力从分流机构行星架经分流机构太阳轮流向液压传动组件，另一部分动力从分流机构行星架依次经第二离合器、前行星排太阳轮至前行星排行星架，前行星排行星架处动力分别分流到输出轴和前行星排齿圈，所述前行星排齿圈的动力依次传递给后行星排行星架和后行星排齿圈，分流的三组动力最后汇流至输出轴；
复合传动Ⅲ档：第一制动器、第二制动器和第四离合器分离，第一离合器和第二离合器接合，动力从分流机构齿圈从分流机构行星架处进行分流，一部分动力从分流机构行星架经分流机构太阳轮流向液压传动组件，另一部分动力从分流机构行星架输入至第二离合器处分流，分别流向前行星排齿圈和前行星排太阳轮在前行星排行星架汇流，机械传动组件和液压传动组件在输出轴汇流输出；
复合传动Ⅳ档：第一制动器、第一离合器和第二离合器分离，第二制动器和第四离合器接合，动力从分流机构齿圈从分流机构行星架处进行分流，一部分动力从分流机构行星架经分流机构太阳轮流向液压传动组件，另一部分动力从分流机构行星架依次经第四离合器、后行星排太阳轮、后行星排齿圈至输出轴，机械传动组件和液压传动组件在输出轴汇流输出。
表1传动档位与换挡元件关系表
Figure PCTCN2019112636-appb-000001
注：“▲”代表元件处于接合状态；
注：B 1为分流制动器、B 2为第一制动器、B 3为第二制动器、C 1为液压传动输入离合器、C 2为液压传动输出离合器、C 3为第四离合器、C 4为第一离合器、C 5为第二离合器、C 6为第三离合器、F 1为第一单向离合器、F 2为第二单向离合器。
所述机械液压复合传动模式挡位切换时，复合传动Ⅰ档→复合传动Ⅱ档涉及到2个换挡元件，复合传动Ⅱ档→复合传动Ⅲ档涉及到2个换挡元件，复合传动Ⅲ档→复合传动Ⅳ档涉及到4个换挡元件，复合传动Ⅰ档→复合传动Ⅲ档涉及到2个换挡元件，复合传动Ⅰ档→复合传动Ⅳ档涉及到2个换挡元件，复合传动Ⅱ档→复合传动Ⅳ档涉及到4个换挡元件；
涉及到2个及以下换挡元件的档位切换可通过不多于3次试验给出优化方案；涉及3或4个换挡元件且无交互作用的挡位切换通过正交表进行分析；涉及4个换挡元件的档位切换充分利用正交表的4列，涉及3个换挡元件的档位切换则选择正交表的任意3列；
确定输出轴角速度的变化量、输出轴角速度的变化率、输出轴角速度的二次微分和时间作为评价指标，选择“提前”、“按时”和“延迟”切换为三水平，进行有交互作用的正交分析；通过试验获得试验数据；再根据方差分析表确定因素和误差的离差平方和及自由度，并与临界值进行比较，确定因素和误差的显著性；得到各评价指标的优选方案，根据 权重系数，确定优化方案；通过不同工况下得到的换挡机构切换时序数据成组控制各组换挡元件的切换时序。
所述机械液压复合传动模式挡位切换的控制方法具体步骤如下：
1)选择输出轴角速度的变化量α、输出轴角速度的变化率β、输出轴角速度的二次微分γ和换挡时间t作为评价指标。α、β、γ分别为转速的零阶、一阶、二阶微分，与时间t共同形成时空评价指标；
输出轴角速度变化量定义为：
Figure PCTCN2019112636-appb-000002
式中α为输出轴角速度变化量，
Figure PCTCN2019112636-appb-000003
为输出轴稳态角速度，ω o min为输出轴最低角速度；
输出轴角速度变化率定义为：
Figure PCTCN2019112636-appb-000004
式中β为输出轴角速度变化率；
输出轴转矩为：
T o＝β·J o
式中T o为输出轴转矩，J o为输出轴转动惯量；
输出轴角速度的二次微分定义为：
Figure PCTCN2019112636-appb-000005
式中γ为输出轴角速度二次微分；
冲击度是车辆纵向加速度的变化率：
Figure PCTCN2019112636-appb-000006
式中j为冲击度，r d为轮胎动力半径，i g为变速器传动比，i 0为驱动桥传动比；
2)建立如表2所示的L 9(3 4)正交表；表中，“1”、“2”和“3”为三水平，分别代表相关换挡元件“提前”、“按时”和“延迟”切换；“a”、“b”、“c”和“d”为四因素，分别代表无交互作用的换挡元件；n为总试验次数，n＝9，
Figure PCTCN2019112636-appb-000007
为与该因素第i个水平有关的试验结果之和(i∈(1，2,3)，F∈(a，b，c，d))；
表2机械液压复合传动模式元件切换时序正交表
  a b c d 试验结果x i
  1 2 3 4  
1 1 1 1 1  
2 1 2 2 2  
3 1 3 3 3  
4 2 1 2 3  
5 2 2 3 1  
6 2 3 1 2  
7 3 1 3 2  
8 3 2 1 3  
9 3 3 2 1  
表中各符号表达式如下：
Figure PCTCN2019112636-appb-000008
极差：
Figure PCTCN2019112636-appb-000009
通过极差数据，确定各因素的主次顺序，确定单个评价指标的优选方案，最终根据权重系数，确定最优方案；
最优方案决定：
Figure PCTCN2019112636-appb-000010
式中ξ为综合评价指标，ξ k为单个评价指标，ξ k min/ξ k max为单个评价指标上/下限，λ k为权重系数。
3)设计正交表头，明确试验方案，进行9次试验，得出试验结果；
4)根据试验结果计算极差，确定因素的主次顺序，得到各评价指标ξ k的优选方案，根据权重系数λ k，确定优化方案；
机液复合档位优化方案由下式决定：
Figure PCTCN2019112636-appb-000011
式中ξ为综合评价指标，ξ k为单个评价指标，ξ k min/ξ k max为单个评价指标上/下限，λ k为权重系数；
5)若优化方案尚不能满足要求，可增大或减小“提前”和“延迟”的时间，也可选择不相同的“提前”时间和“延迟”时间，直至满足要求为止。
所述纯机械传动模式挡位切换时，机械Ⅰ档切换到机械Ⅱ档涉及到1个换挡元件，机械Ⅱ档切换到机械Ⅲ档涉及到1个换挡元件，机械Ⅲ档切换到机械Ⅳ档涉及到3个换挡元件，机械Ⅰ档切换到机械Ⅲ档涉及到2个换挡元件，机械Ⅰ档切换到机械Ⅳ档涉及到5个换挡元件，机械Ⅱ档切换到机械Ⅳ档涉及到4个换挡元件；
涉及到2个及以下换挡元件的档位切换可通过不超过3次试验给出优化方案；涉及3个换挡元件且其中有2个换挡元件有交互作用的挡位切换、涉及5个换挡元件且其中有3个换挡元件有交互作用的挡位切换以及涉及4个换挡元件且其中有2个换挡元件有交互作用的挡位切换均采用正交表进行分析，分析时选用相应列表；
确定输出轴角速度的变化量、输出轴角速度的变化率、输出轴角速度的二次微分和时间作为评价指标，选择“提前”、“按时”和“延迟”切换为三水平，进行无交互作用的正交分析；通过试验获得试验数据；根据试验数据计算极差，确定因素的主次顺序，得到各评价指标的优选方案；根据权重系数，确定优化方案；将不同工况优化方案形成成对矩阵编列程序控制档位之间的切换。
所述纯机械传动模式挡位切换的控制方法具体步骤如下：
1)选择输出轴角速度的变化量α、输出轴角速度的变化率β、输出轴角速度的二次微分γ和时间t作为评价指标。α、β、γ分别为转速的零阶、一阶、二阶微分，与时间t共同形成时空评价指标；
输出轴角速度变化量定义为：
Figure PCTCN2019112636-appb-000012
式中α为输出轴角速度变化量，
Figure PCTCN2019112636-appb-000013
为输出轴稳态角速度，ω o min为输出轴最低角速度；输出轴角速度变化率定义为：
Figure PCTCN2019112636-appb-000014
式中β为输出轴角速度变化率；
输出轴转矩由式决定：
T o＝β·J o
式中T o为输出轴转矩，J o为输出轴转动惯量；
输出轴角速度的二次微分定义为：
Figure PCTCN2019112636-appb-000015
式中γ为输出轴角速度二次微分；
冲击度是车辆纵向加速度的变化率，由式决定：
Figure PCTCN2019112636-appb-000016
式中：j为冲击度，r d为轮胎动力半径，i g为变速器传动比，i 0为驱动桥传动比；
2)建立表3所示的L 27(3 13)正交表；表中，“1”、“2”和“3”分别代表相关换挡元件 “提前”、“按时”和“延迟”切换；“a”和“b”分别代表无交互作用的换挡元件；“c”、“d”和“e”分别有交互作用的换挡元件；
n为总试验次数，n＝27，
Figure PCTCN2019112636-appb-000017
为与该因素第i个水平有关的试验结果之和(i∈(1，2,3)，F∈(e，c，e×c1，e×c2，d，e×d1，e×d2，c×d1，a，b，c×d2))；
表3机械传动模式元件切换时序正交表
Figure PCTCN2019112636-appb-000018
Figure PCTCN2019112636-appb-000019
表中各符号表达式如下：
Figure PCTCN2019112636-appb-000020
Figure PCTCN2019112636-appb-000021
Figure PCTCN2019112636-appb-000022
Figure PCTCN2019112636-appb-000023
3)设计如表4正交表头，明确试验方案，进行27次试验，得出试验结果x i(i＝1,...,27)，计算出相关统计值；
4)再根据方差分析表确定因素和误差的离差平方和及自由度，确定F值，并与临界值进行比较，确定因素和误差的显著性；
表4机械传动模式元件切换时序方差分析表
Figure PCTCN2019112636-appb-000024
总离差平方和Q T，各因素离差平方和Q F和误差离差平方和Q e分别为：
Q T＝W-P＝ΣQ F+Q e
Q F＝U F-P
总自由度f T＝26，各因素自由度f F＝2，误差自由度f e＝4；
因素和误差的平均离差平方和为：
Figure PCTCN2019112636-appb-000025
Figure PCTCN2019112636-appb-000026
F值：
Figure PCTCN2019112636-appb-000027
机械传动模式元件切换时序方差分析表如表1所示：
方差来源主要在三部分：①“a”、“b”、“c”、“d”和“e”；②“e×c”(包括e×c1和e×c2两部分)、“e×d”(包括e×d1和e×d2两部分)和“c×d”(包括c×d1和c×d2两部分)；③误差e *；
根据计算出的方差来源各因素和误差的离差平方和、自由度、平均离差平方和，将F F值和F α(f F，f e)值进行比较；
若F F＞F α(f F，f e)，则该因素影响显著，否则不显著。选择显著性因素，并直观分析各因素的重要性主次，确定最佳换挡方案。
机械档位优化方案由下式决定：
Figure PCTCN2019112636-appb-000028
式中ζ为综合评价指标，ζ k为单个评价指标，ζ k min/ζ k max为单个评价指标上/下限，μ k为权重系数。
有益效果：1、本发明可在液压传动、机液复合传动和机械传动间进行切换，且各传动方式有多个档位可供选择，能够根据复杂的作业工况提供多种功率分配模式，充分利用机械传动的传动性能和液压传动的无级调速性能，有利于提高车辆的作业效率、动力性和经济性。2、机械档换挡装置采用单向离合器既能实现发动机制动，又能均匀使用各换挡元件以增加使用寿命。
附图说明
图1为本发明的结构原理图；
图2为本发明纯液压传动功率流向示意图；
图3为本发明机械传动Ⅰ档功率流向示意图；
图4为本发明机械传动Ⅱ档功率流向示意图；
图5为本发明机械传动Ⅲ档功率流向示意图；
图6为本发明机械传动Ⅳ档功率流向示意图；
图7为本发明复合传动Ⅰ档功率流向示意图；
图8为本发明复合传动Ⅱ档功率流向示意图；
图9为本发明复合传动Ⅲ档功率流向示意图；
图10为本发明复合传动Ⅳ档功率流向示意图。
具体实施方式
下面结合附图对本发明作进一步说明。
如图1所示，一种多功率分配模式的机械液压复合传动装置,包括输入轴1、分流机构2、机械传动组件3、液压传动组件4和输出轴5，所述输入轴1通过分流机构2与相互并联的机械传动组件3和液压传动组件4连接，所述机械传动组件3和液压传动组件4同时与输出轴5连接；所述机械传动组件3包括相互串联的前行星排组件31和后行星排组件32；
所述前行星排组件31包括第一离合器311、第二离合器312、第三离合器313、前行星排齿圈314、前行星排行星架315、前行星排太阳轮316和第一单向离合器317，所述第一离合器311位于分流机构2与前行星排齿圈314之间，所述第二离合器312和第三离合器313相互并联并且分别与前行星排太阳轮316连接，所述第三离合器313与前行星排太阳轮316之间设有第一单向离合器317，所述输出轴5与前行星排行星架315连接；
所述后行星排组件32包括第四离合器321、后行星排太阳轮322，后行星排行星架323，后行星排齿圈324、第一制动器325、第二制动器326和第二单向离合器327，所述第四离合器321位于分流机构2与后行星排太阳轮322之间，所述第一制动器325与后行星排太阳轮322连接，所述后行星排行星架323与前行星排齿圈314连接并且与相互并联的第二制动器326和第二单向离合器327连接，所述后行星排齿圈324与输出轴5连接。
液压传动组件4包括液压传动输入离合器41、液压传动输入齿轮副42、液压泵43、液压油管44、液压马达45、液压传动输出齿轮副46、液压传动输出离合器47；所述液压泵43通过液压传动输入齿轮副42与分流机构2连接，所述液压传动输入齿轮副42与液压泵43之间设有液压传动输入离合器41，所述液压泵43通过液压油管44与液压马达45连接，所述液压马达45通过液压传动输出齿轮副46与输出轴5连接，所述液压马达45通过液压传动输出齿轮副46之间设有液压传动输出离合器47。
所述分流机构2包括分流制动器21、分流机构太阳轮22、分流机构行星架23、分流机构齿圈24；所述输入轴1与分流机构齿圈24连接，所述分流机构太阳轮22与液压传动 组件4连接，所述分流机构太阳轮22上设有分流制动器21，所述分流机构行星架23与机械传动组件3连接。
如表1所示，通过制动器和离合器之间的组合切换实现纯液压传动、机械液压复合传动和纯机械传动三种类型的传动，具体传动类型如下：
如图2所示，纯液压传动：第一制动器325、液压传动输入离合器41、液压传动输出离合器47、第四离合器321接合，其他离合器和制动器分离，第四离合器321和第一制动器325接合，分流机构行星架23被制动，液压传动输入离合器41和液压传动输出离合器47接合，动力经输入轴1、分流机构齿圈24、分流机构太阳轮22、液压传动输入齿轮副42、输入离合器41驱动液压泵43工作，所述液压泵43将机械动力转化为高压油液并经过液压管道44驱动液压马达45工作，液压马达45输出的机械动力经液压传动输出离合器47和液压传动输出齿轮副46传递到输出轴5；
如图3、4、5和6所示，纯机械传动：分流制动器21接合，液压传动输入离合器41、液压传动输出离合器47、第二制动器326和第四离合器321分离，通过其他离合器和制动器之间的组合实现纯机械传动模式中不同传动比的挡位之间切换，分流制动器21接合，液压传动输入离合器41和液压传动输出离合器47分离，液压传动组件4无动力输入，动力经输入轴1、分流机构齿圈24、分流机构行星架23传递机械传动组件3，机械传动组件3经过速比调节后从输出轴5输出；
如图7、8、9和10所示，机械液压复合传动：液压传动输入离合器41和液压传动输出离合器47接合，分流制动器21、第三离合器313、第一单向离合器317、第二单向离合器327分离，通过其他离合器和和制动器之间的组合实现机械液压复合传动模式中不同传动比挡位之间的切换，动力经输入轴1、分流机构齿圈24至分流机构行星架23，在分流机构行星架23处分流，分别流向机械传动组件3和分流机构太阳轮22，所述分流机构太阳轮22与液压传动组件4连接，最后汇合至输出轴5输出。
所述纯机械传动模式包括机械传动Ⅰ档、机械传动Ⅱ档、机械传动Ⅲ档和机械传动Ⅳ档，具体实现方法如下：
如图3所示，机械传动Ⅰ档：第一制动器325、第一离合器311和第二离合器312分离，第三离合器313、第一单向离合器317和第二单向离合器327接合，动力从分流机构行星架23依次经第三离合器313、第一单向离合器317、前行星排太阳轮316前行星排行星架315至输出轴5；
如图4所示，机械传动Ⅱ档：第一制动器325、第一离合器311、第二离合器312和第二单向离合器327分离，第三离合器313、第一单向离合器317接合，动力从分流机构行星架23依次经第三离合器313、第一单向离合器317、前行星排太阳轮316至前行星排行星架315，前行星排行星架315处动力分别分流到输出轴5和前行星排齿圈314，所述前行星排齿圈314的动力依次传递给后行星排行星架323和后行星排齿圈324，最后汇流至 输出轴5；
如图5所示，机械传动Ⅲ档：第一制动器325、第二离合器312和第二单向离合器327分离，第一离合器311、第三离合器313和第一单向离合器317接合，动力从分流机构行星架23输入至第一离合器311处分流，分别流向前行星排齿圈314和前行星排太阳轮316在前行星排行星架315汇流，最后从输出轴5输出；
如图6所示，机械传动Ⅳ档：第二离合器312、第三离合器313、第一单向离合器317和第二单向离合器327分离，第一离合器311和第一制动器325接合，动力从分流机构行星架23依次经第一离合器311、前行星排齿圈314、后行星排行星架323、后行星排齿圈324至输出轴5输出。
所述机械液压复合传动模式包括复合传动Ⅰ档、复合传动Ⅱ档、复合传动Ⅲ档和复合传动Ⅳ档，具体实现方法如下：
如图7所示，复合传动Ⅰ档：第一制动器325、第一离合器311和第四离合器321分离，第二制动器326和第二离合器312接合，动力从分流机构齿圈24从分流机构行星架23处进行分流，一部分动力从分流机构行星架23经分流机构太阳轮22流向液压传动组件4，另一部分动力从分流机构行星架23依次经第二离合器312、前行星排太阳轮316、前行星排行星架315，最后汇流至输出轴5输出；
如图8所示，复合传动Ⅱ档：第二制动器326、第一离合器311和第四离合器321分离，第一制动器325和第二离合器312接合，动力从分流机构齿圈24从分流机构行星架23处进行分流，一部分动力从分流机构行星架23经分流机构太阳轮22流向液压传动组件4，另一部分动力从分流机构行星架23依次经第二离合器312、前行星排太阳轮316至前行星排行星架315，前行星排行星架315处动力分别分流到输出轴5和前行星排齿圈314，所述前行星排齿圈314的动力依次传递给后行星排行星架323和后行星排齿圈324，分流的三组动力最后汇流至输出轴5；
如图9所示，复合传动Ⅲ档：第一制动器325、第二制动器326和第四离合器321分离，第一离合器311和第二离合器312接合，动力从分流机构齿圈24到分流机构行星架23处进行分流，一部分动力从分流机构行星架23经分流机构太阳轮22流向液压传动组件4，另一部分动力从分流机构行星架23输入至第二离合器312处分流，分别流向前行星排齿圈314和前行星排太阳轮316在前行星排行星架315汇流，机械传动组件3和液压传动组件4在输出轴5汇流输出；
如图10所示，复合传动Ⅳ档：第一制动器325、第一离合器311和第二离合器312分离，第二制动器326和第四离合器321接合，动力从分流机构齿圈24从分流机构行星架23处进行分流，一部分动力从分流机构行星架23经分流机构太阳轮22流向液压传动组件4，另一部分动力从分流机构行星架23依次经第四离合器321、后行星排太阳轮322、后行星排齿圈324至输出轴5，机械传动组件3和液压传动组件4在输出轴5汇流输出。
所述机械液压复合传动模式挡位切换时，复合传动Ⅰ档→复合传动Ⅱ档涉及到2个换挡元件，复合传动Ⅱ档→复合传动Ⅲ档涉及到2个换挡元件，复合传动Ⅲ档→复合传动Ⅳ档涉及到4个换挡元件，复合传动Ⅰ档→复合传动Ⅲ档涉及到2个换挡元件，复合传动Ⅰ档→复合传动Ⅳ档涉及到2个换挡元件，复合传动Ⅱ档→复合传动Ⅳ档涉及到4个换挡元件；
涉及到2个及以下换挡元件的档位切换可通过不多于3次试验给出优化方案；涉及3或4个换挡元件且无交互作用的挡位切换通过正交表进行分析；涉及4个换挡元件的档位切换充分利用正交表的4列，涉及3个换挡元件的档位切换则选择正交表的任意3列；
确定输出轴角速度的变化量、输出轴角速度的变化率、输出轴角速度的二次微分和时间作为评价指标，选择“提前”、“按时”和“延迟”切换为三水平，进行有交互作用的正交分析；通过试验获得试验数据；再根据方差分析表确定因素和误差的离差平方和及自由度，并与临界值进行比较，确定因素和误差的显著性；得到各评价指标的优选方案，根据权重系数，确定优化方案；通过不同工况下得到的换挡机构切换时序数据成组控制各组换挡元件的切换时序。
所述机械液压复合传动模式挡位切换的控制方法具体步骤如下：
1)选择输出轴角速度的变化量α、输出轴角速度的变化率β、输出轴角速度的二次微分γ和时间t作为评价指标。α、β、γ分别为转速的零阶、一阶、二阶微分，与时间t共同形成时空评价指标；
输出轴角速度变化量定义为：
Figure PCTCN2019112636-appb-000029
式中α为输出轴角速度变化量，
Figure PCTCN2019112636-appb-000030
为输出轴稳态角速度，ω o min为输出轴最低角速度；
输出轴角速度变化率定义为：
Figure PCTCN2019112636-appb-000031
式中β为输出轴角速度变化率；
输出轴转矩：
T o＝β·J o
式中T o为输出轴转矩，J o为输出轴转动惯量；
输出轴角速度的二次微分为：
Figure PCTCN2019112636-appb-000032
式中γ为输出轴角速度二次微分；
冲击度是车辆纵向加速度的变化率：
Figure PCTCN2019112636-appb-000033
式中j为冲击度，r d为轮胎动力半径，i g为变速器传动比，i 0为驱动桥传动比；
2)建立如表2所示的L 9(3 4)正交表；表中，“1”、“2”和“3”为三水平，分别代表相关换挡元件“提前”、“按时”和“延迟”切换；“a”、“b”、“c”和“d”为四因素，分别代表无交互作用的换挡元件；n为总试验次数，n＝9，
Figure PCTCN2019112636-appb-000034
为与该因素第i个水平有关的试验结果之和(i∈(1，2,3)，F∈(a，b，c，d))；
以复合传动Ⅱ档切换到复合传动Ⅳ为例：
确定“四因素”为制动器B2、制动器B3、离合器C3和离合器C5，此4个换挡元件的切换时序为影响传动系统换挡品质的四因素。
选择“三水平”为换挡元件“提前”、“按时”和“延迟”切换，可根据实际情况选择“提前”和“延迟”的时间，“提前”和“延迟”的时间可以相同，也可以不同。就本例而言，选择“提前”和“延迟”的时间为0.3s。
表2机械液压复合传动模式元件切换时序正交表
  a b c d 试验结果x i
  1 2 3 4  
1 1 1 1 1  
2 1 2 2 2  
3 1 3 3 3  
4 2 1 2 3  
5 2 2 3 1  
6 2 3 1 2  
7 3 1 3 2  
8 3 2 1 3  
9 3 3 2 1  
表中各符号表达式如下：
Figure PCTCN2019112636-appb-000035
极差：
Figure PCTCN2019112636-appb-000036
通过极差数据，确定各因素的主次顺序，确定单个评价指标的优选方案，最终根据权重系数，确定最优方案；
最优方案决定：
Figure PCTCN2019112636-appb-000037
ξ为综合评价指标，ξ k为单个评价指标，ξ k min/ξ k max为单个评价指标上/下限，λ k为权重系数。
3)设计正交表头，明确试验方案，进行9次试验，得出试验结果；
4)根据试验结果计算极差，确定因素的主次顺序，得到各评价指标ξ k的优选方案，根据权重系数λ k，确定优化方案；
机液复合档位优化方案由下式决定：
Figure PCTCN2019112636-appb-000038
ξ为综合评价指标，ξ k为单个评价指标，ξ k min/ξ k max为单个评价指标上/下限，λ k为权重系数；
5)若优化方案尚不能满足要求，可增大或减小“提前”和“延迟”的时间，也可选择不相同的“提前”时间和“延迟”时间，直至满足要求为止。
将不同工况下得到的换挡机构切换时序数据成组输入到换挡控制器，进而通过控制各组换挡装置的切换时序，保证同种传动模式在各档位切换过程中都有良好的换挡品质。
复合传动Ⅲ档过渡到复合传动Ⅳ档同理可得，只需改变相应的“四因素”和“三水平”。
所述纯机械传动模式挡位切换时，机械Ⅰ档切换到机械Ⅱ档涉及到1个换挡元件，机械Ⅱ档切换到机械Ⅲ档涉及到1个换挡元件，机械Ⅲ档切换到机械Ⅳ档涉及到3个换挡元件，机械Ⅰ档切换到机械Ⅲ档涉及到2个换挡元件，机械Ⅰ档切换到机械Ⅳ档涉及到5个换挡元件，机械Ⅱ档切换到机械Ⅳ档涉及到4个换挡元件；
涉及到2个及以下换挡元件的档位切换可通过不超过3次试验给出优化方案；涉及3个换挡元件且其中有2个换挡元件有交互作用的挡位切换、涉及5个换挡元件且其中有3个换挡元件有交互作用的挡位切换以及涉及4个换挡元件且其中有2个换挡元件有交互作用的挡位切换均采用正交表进行分析，分析时选用相应列表；
确定输出轴角速度的变化量、输出轴角速度的变化率、输出轴角速度的二次微分和时间作为评价指标，选择“提前”、“按时”和“延迟”切换为三水平，进行无交互作用的正交分析；通过试验获得试验数据；根据试验数据计算极差，确定因素的主次顺序，得到各评价指标的优选方案；根据权重系数，确定优化方案；将不同工况优化方案形成成对矩阵编列程序控制档位之间的切换。
所述纯机械传动模式挡位切换的控制方法具体步骤如下：
1)选择输出轴角速度的变化量α、输出轴角速度的变化率β、输出轴角速度的二次微分γ和时间t作为评价指标。α、β、γ分别为转速的零阶、一阶、二阶微分，与时间t共同形成时空评价指标；
输出轴角速度的变化量α
输出轴角速度变化量定义为：
Figure PCTCN2019112636-appb-000039
式中：α为输出轴角速度变化量，
Figure PCTCN2019112636-appb-000040
为输出轴稳态角速度，ω o min为输出轴最低角速度输出轴角速度变化率定义为：
Figure PCTCN2019112636-appb-000041
式中：β为输出轴角速度变化率
输出轴转矩由式(3)决定：
T o＝β·J o
式中：T o为输出轴转矩，J o为输出轴转动惯量
输出轴角速度的二次微分定义为：
Figure PCTCN2019112636-appb-000042
式中：γ为输出轴角速度二次微分
冲击度是车辆纵向加速度的变化率，由式(5)决定：
Figure PCTCN2019112636-appb-000043
式中：j为冲击度，r d为轮胎动力半径，i g为变速器传动比，i 0为驱动桥传动比；
2)建立表3所示的L 27(3 13)正交表；表中，“1”、“2”和“3”分别代表相关换挡元件“提前”、“按时”和“延迟”切换；“a”和“b”分别代表无交互作用的换挡元件；“c”、“d”和“e”分别有交互作用的换挡元件；
以机械Ⅰ档切换到机械Ⅳ档为例：
确定“五因素”为制动器B2、离合器C4、离合器C6、单向离合器F1和单向离合器F2，此5个换挡元件的切换时序为影响传动系统换挡品质的五因素。“五因素”中制动器B2和离合器C4无交互作用，离合器C6、单向离合器F1和单向离合器F2有交互作用。
选择“三水平”为换挡元件“提前”、“按时”和“延迟”切换，可根据实际情况选择“提前”和“延迟”的时间，“提前”和“延迟”的时间可以相同，也可以不同。就本例而言，因为涉及元件较多，其中还有交互作用，故选择“提前”和“延迟”的时间为0.5s。
表3机械传动模式元件切换时序正交表
Figure PCTCN2019112636-appb-000044
n为总试验次数，n＝27，
Figure PCTCN2019112636-appb-000045
为与该因素第i个水平有关的试验结果之和(i∈(1，2,3)，F∈(e，c，e×c1，e×c2，d，e×d1，e×d2，c×d1，a，b，c×d2))；
表中各符号表达式如下：
Figure PCTCN2019112636-appb-000046
Figure PCTCN2019112636-appb-000047
Figure PCTCN2019112636-appb-000048
Figure PCTCN2019112636-appb-000049
3)设计如表3正交表头，明确试验方案，进行27次试验，得出试验结果x i(i＝1,...,27)，计算出相关统计值；
4)再根据方差分析表确定因素和误差的离差平方和及自由度，确定F值，并与临界值进行比较，确定因素和误差的显著性；
表4机械传动模式元件切换时序方差分析表
Figure PCTCN2019112636-appb-000050
总离差平方和Q T，各因素离差平方和Q F和误差离差平方和Q e分别为：
Q T＝W-P＝ΣQ F+Q e
Q F＝U F-P
总自由度f T＝26，各因素自由度f F＝2，误差自由度f e＝4；
因素和误差的平均离差平方和为：
Figure PCTCN2019112636-appb-000051
Figure PCTCN2019112636-appb-000052
F值：
Figure PCTCN2019112636-appb-000053
机械传动模式元件切换时序方差分析表如表1所示：
方差来源主要在三部分：①“a”、“b”、“c”、“d”和“e”；②“e×c”(包括e×c1和e×c2两部分)、“e×d”(包括e×d1和e×d2两部分)和“c×d”(包括c×d1和c×d2两部分)；③误差e *；
根据计算出的方差来源各因素和误差的离差平方和、自由度、平均离差平方和，将F F值和F α(f F，f e)值进行比较；
若F F＞F α(f F，f e)，则该因素影响显著，否则不显著。选择显著性因素，并直观分析各因素的重要性主次，确定最佳换挡方案。
机械档位优化方案由下式决定：
Figure PCTCN2019112636-appb-000054
式中ζ为综合评价指标，ζ k为单个评价指标，ζ k min/ζ k max为单个评价指标上/下限，μ k为权重系数。
将不同工况下得到的换挡机构切换时序数据成组输入到换挡控制器，进而通过控制各组换挡装置的切换时序，保证同种传动模式在各档位切换过程中都有良好的换挡品质。
机械Ⅲ档切换到机械Ⅳ档涉及到3个换挡元件，其中第三离合器313和第一单向离合器317两个元件有交互作用，可采用L 9(3 4)正交表进行分析，只需将第四因素列替换为交互列。
机械Ⅱ档切换到机械Ⅳ档涉及到4个换挡元件，其中第三离合器313和第一单向离合器317两个元件有交互作用，可采用L 27(3 13)正交表进行分析，相关多余列空置。

Claims (10)

  1. 一种多功率分配模式的机械液压复合传动装置,包括输入轴(1)、分流机构(2)、机械传动组件(3)、液压传动组件(4)和输出轴(5)，所述输入轴(1)通过分流机构(2)与相互并联的机械传动组件(3)和液压传动组件(4)连接，所述机械传动组件(3)和液压传动组件(4)同时与输出轴(5)连接；其特征在于：所述机械传动组件(3)包括相互串联的前行星排组件(31)和后行星排组件(32)；
    所述前行星排组件(31)包括第一离合器(311)、第二离合器(312)、第三离合器(313)、前行星排齿圈(314)、前行星排行星架(315)、前行星排太阳轮(316)和第一单向离合器(317)，所述第一离合器(311)位于分流机构(2)与前行星排齿圈(314)之间，所述第二离合器(312)和第三离合器(313)相互并联并且分别与前行星排太阳轮(316)连接，所述第三离合器(313)与前行星排太阳轮(316)之间设有第一单向离合器(317)，所述输出轴(5)与前行星排行星架(315)连接；
    所述后行星排组件(32)包括第四离合器(321)、后行星排太阳轮(322)，后行星排行星架(323)，后行星排齿圈(324)、第一制动器(325)、第二制动器(326)和第二单向离合器(327)，所述第四离合器(321)位于分流机构(2)与后行星排太阳轮(322)之间，所述第一制动器(325)与后行星排太阳轮(322)连接，所述后行星排行星架(323)与前行星排齿圈(314)连接并且与相互并联的第二制动器(326)和第二单向离合器(327)连接，所述后行星排齿圈(324)与输出轴(5)连接。
  2. 根据权利要求1所述的多功率分配模式的机械液压复合传动装置，其特征在于：液压传动组件(4)包括液压传动输入离合器(41)、液压传动输入齿轮副(42)、液压泵(43)、液压油管(44)、液压马达(45)、液压传动输出齿轮副(46)、液压传动输出离合器(47)；所述液压泵(43)通过液压传动输入齿轮副(42)与分流机构(2)连接，所述液压传动输入齿轮副(42)与液压泵(43)之间设有液压传动输入离合器(41)，所述液压泵(43)通过液压油管(44)与液压马达(45)连接，所述液压马达(45)通过液压传动输出齿轮副(46)与输出轴(5)连接，所述液压马达(45)通过液压传动输出齿轮副(46)之间设有液压传动输出离合器(47)。
  3. 根据权利要求1或2所述的多功率分配模式的机械液压复合传动装置，其特征在于：所述分流机构(2)包括分流制动器(21)、分流机构太阳轮(22)、分流机构行星架(23)、分流机构齿圈(24)；所述输入轴(1)与分流机构齿圈(24)连接，所述分流机构太阳轮(22)与液压传动组件(4)连接，所述分流机构太阳轮(22)上设有分流制动器(21)，所述分流机构行星架(23)与机械传动组件(3)连接。
  4. 根据权利要求3所述的多功率分配模式的机械液压复合传动装置的控制方法，其特征在于：通过制动器和离合器之间的组合切换实现纯液压传动、机械液压复合传动和纯 机械传动三种类型的传动，具体传动类型如下：
    纯液压传动：第一制动器(325)、液压传动输入离合器(41)、液压传动输出离合器(47)、第四离合器(321)接合，其他离合器和制动器分离，第四离合器(321)和第一制动器(325)接合，分流机构行星架(23)被制动，液压传动输入离合器(41)和液压传动输出离合器(47)接合，动力经输入轴(1)、分流机构齿圈(24)、分流机构太阳轮(22)、液压传动输入齿轮副(42)、输入离合器(41)驱动液压泵(43)工作，所述液压泵(43)将机械动力转化为高压油液并经过液压管道(44)驱动液压马达(45)工作，液压马达(45)输出的机械动力经液压传动输出离合器(47)和液压传动输出齿轮副(46)传递到输出轴(5)；
    纯机械传动：分流制动器(21)接合，液压传动输入离合器(41)、液压传动输出离合器(47)、第二制动器(326)和第四离合器(321)分离，通过其他离合器和制动器之间的组合实现纯机械传动模式中不同传动速比的挡位之间切换，分流制动器(21)接合，液压传动输入离合器(41)和液压传动输出离合器(47)分离，液压传动组件(4)无动力输入，动力经输入轴(1)、分流机构齿圈(24)、分流机构行星架(23)传递机械传动组件(3)，机械传动组件(3)经过速比调节后从输出轴(5)输出；
    机械液压复合传动：液压传动输入离合器(41)和液压传动输出离合器(47)接合，分流制动器(21)、第三离合器(313)、第一单向离合器(317)、第二单向离合器(327)分离，通过其他离合器和和制动器之间的组合实现机械液压复合传动模式中不同传动速比的挡位之间切换，动力经输入轴(1)、分流机构齿圈(24)至分流机构行星架(23)，在分流机构行星架(23)处分流，分别流向机械传动组件(3)和分流机构太阳轮(22)，所述分流机构太阳轮(22)与液压传动组件(4)连接，最后汇合至输出轴(5)输出。
  5. 根据权利要求4所述的多功率分配模式的机械液压复合传动装置的控制方法，其特征在于：所述纯机械传动模式包括机械传动Ⅰ档、机械传动Ⅱ档、机械传动Ⅲ档和机械传动Ⅳ档，具体实现方法如下：
    机械传动Ⅰ档：第一制动器(325)、第一离合器(311)和第二离合器(312)分离，第三离合器(313)、第一单向离合器(317)和第二单向离合器(327)接合，动力从分流机构行星架(23)依次经第三离合器(313)、第一单向离合器(317)、前行星排太阳轮(316)前行星排行星架(315)至输出轴(5)；
    机械传动Ⅱ档：第一制动器(325)、第一离合器(311)、第二离合器(312)和第二单向离合器(327)分离，第三离合器(313)、第一单向离合器(317)接合，动力从分流机构行星架(23)依次经第三离合器(313)、第一单向离合器(317)、前行星排太阳轮(316)至前行星排行星架(315)，前行星排行星架(315)处动力分别分流到输出轴(5)和前行星排齿圈(314)，所述前行星排齿圈(314)的动力依次传递给后行星排行星架(323)和后行星排齿圈(324)，最后汇流至输出轴(5)；
    机械传动Ⅲ档：第一制动器(325)、第二离合器(312)和第二单向离合器(327)分离，第一离合器(311)、第三离合器(313)和第一单向离合器(317)接合，动力从分流机构行星架(23)输入至第一离合器(311)处分流，分别流向前行星排齿圈(314)和前行星排太阳轮(316)在前行星排行星架(315)汇流，最后从输出轴(5)输出；
    机械传动Ⅳ档：第二离合器(312)、第三离合器(313)、第一单向离合器(317)和第二单向离合器(327)分离，第一离合器(311)和第一制动器(325)接合，动力从分流机构行星架(23)依次经第一离合器(311)、前行星排齿圈(314)、后行星排行星架(323)、后行星排齿圈(324)至输出轴(5)输出。
  6. 根据权利要求4所述的多功率分配模式的机械液压复合传动装置的控制方法，其特征在于：所述机械液压复合传动模式包括复合传动Ⅰ档、复合传动Ⅱ档、复合传动Ⅲ档和复合传动Ⅳ档，具体实现方法如下：
    复合传动Ⅰ档：第一制动器(325)、第一离合器(311)和第四离合器(321)分离，第二制动器(326)和第二离合器(312)接合，动力从分流机构齿圈(24)从分流机构行星架(23)处进行分流，一部分动力从分流机构行星架(23)经分流机构太阳轮(22)流向液压传动组件(4)，另一部分动力从分流机构行星架(23)依次经第二离合器(312)、前行星排太阳轮(316)、前行星排行星架(315)，最后汇流至输出轴(5)输出；
    复合传动Ⅱ档：第二制动器(326)、第一离合器(311)和第四离合器(321)分离，第一制动器(325)和第二离合器(312)接合，动力从分流机构齿圈(24)从分流机构行星架(23)处进行分流，一部分动力从分流机构行星架(23)经分流机构太阳轮(22)流向液压传动组件(4)，另一部分动力从分流机构行星架(23)依次经第二离合器(312)、前行星排太阳轮(316)至前行星排行星架(315)，前行星排行星架(315)处动力分别分流到输出轴(5)和前行星排齿圈(314)，所述前行星排齿圈(314)的动力依次传递给后行星排行星架(323)和后行星排齿圈(324)，分流的三组动力最后汇流至输出轴(5)；
    复合传动Ⅲ档：第一制动器(325)、第二制动器(326)和第四离合器(321)分离，第一离合器(311)和第二离合器(312)接合，动力从分流机构齿圈(24)从分流机构行星架(23)处进行分流，一部分动力从分流机构行星架(23)经分流机构太阳轮(22)流向液压传动组件(4)，另一部分动力从分流机构行星架(23)输入至第二离合器(312)处分流，分别流向前行星排齿圈(314)和前行星排太阳轮(316)在前行星排行星架(315)汇流，机械传动组件(3)和液压传动组件(4)在输出轴(5)汇流输出；
    复合传动Ⅳ档：第一制动器(325)、第一离合器(311)和第二离合器(312)分离，第二制动器(326)和第四离合器(321)接合，动力从分流机构齿圈(24)从分流机构行星架(23)处进行分流，一部分动力从分流机构行星架(23)经分流机构太阳轮(22)流向液压传动组件(4)，另一部分动力从分流机构行星架(23)依次经第四离合器(321)、后行星排太阳轮(322)、后行星排齿圈(324)至输出轴(5)，机械传动组件(3)和液压 传动组件(4)在输出轴(5)汇流输出。
  7. 根据权利要求6所述的多功率分配模式的机械液压复合传动装置的控制方法，其特征在于：所述机械液压复合传动模式挡位切换时，复合传动Ⅰ档→复合传动Ⅱ档涉及到2个换挡元件，复合传动Ⅱ档→复合传动Ⅲ档涉及到2个换挡元件，复合传动Ⅲ档→复合传动Ⅳ档涉及到4个换挡元件，复合传动Ⅰ档→复合传动Ⅲ档涉及到2个换挡元件，复合传动Ⅰ档→复合传动Ⅳ档涉及到2个换挡元件，复合传动Ⅱ档→复合传动Ⅳ档涉及到4个换挡元件；
    涉及到2个及以下换挡元件的档位切换可通过不多于3次试验给出优化方案；涉及3或4个换挡元件且无交互作用的挡位切换通过正交表进行分析；涉及4个换挡元件的档位切换充分利用正交表的4列，涉及3个换挡元件的档位切换则选择正交表的任意3列；
    确定输出轴角速度的变化量、输出轴角速度的变化率、输出轴角速度的二次微分和时间作为评价指标，选择“提前”、“按时”和“延迟”切换为三水平，进行有交互作用的正交分析；通过试验获得试验数据；再根据方差分析表确定因素和误差的离差平方和及自由度，并与临界值进行比较，确定因素和误差的显著性；得到各评价指标的优选方案，根据权重系数，确定优化方案；通过不同工况下得到的换挡机构切换时序数据成组控制各组换挡元件的切换时序。
  8. 根据权利要求7所述的多功率分配模式的机械液压复合传动装置的控制方法，其特征在于：所述机械液压复合传动模式挡位切换的控制方法具体步骤如下：
    1)选择输出轴角速度的变化量α、输出轴角速度的变化率β、输出轴角速度的二次微分γ和时间t作为评价指标。α、β、γ分别为转速的零阶、一阶、二阶微分，与时间t共同形成时空评价指标；
    输出轴角速度变化量定义为：
    Figure PCTCN2019112636-appb-100001
    式中α为输出轴角速度变化量，
    Figure PCTCN2019112636-appb-100002
    为输出轴稳态角速度，ω o min为输出轴最低角速度；
    输出轴角速度变化率定义为：
    Figure PCTCN2019112636-appb-100003
    式中β为输出轴角速度变化率；
    输出轴转矩为：
    T o＝β·J o
    式中T o为输出轴转矩，J o为输出轴转动惯量；
    输出轴角速度的二次微分为：
    Figure PCTCN2019112636-appb-100004
    式中γ为输出轴角速度二次微分；
    冲击度是车辆纵向加速度的变化率：
    Figure PCTCN2019112636-appb-100005
    式中j为冲击度，r d为轮胎动力半径，i g为变速器传动比，i 0为驱动桥传动比；
    2)建立如表2所示的L 9(3 4)正交表；表中，“1”、“2”和“3”为三水平，分别代表相关换挡元件“提前”、“按时”和“延迟”切换；“a”、“b”、“c”和“d”为四因素，分别代表无交互作用的换挡元件；n为总试验次数，n＝9，
    Figure PCTCN2019112636-appb-100006
    为与该因素第i个水平有关的试验结果之和(i∈(1，2,3)，F∈(a，b，c，d))；
    表2机械液压复合传动模式元件切换时序正交表
      a b c d 试验结果x i   1 2 3 4   1 1 1 1 1   2 1 2 2 2   3 1 3 3 3   4 2 1 2 3   5 2 2 3 1   6 2 3 1 2   7 3 1 3 2   8 3 2 1 3   9 3 3 2 1  
    表中各符号表达式如下：
    Figure PCTCN2019112636-appb-100007
    极差：
    Figure PCTCN2019112636-appb-100008
    通过极差数据，确定各因素的主次顺序，确定单个评价指标的优选方案，最终根据权重系数，确定最优方案；
    最优方案决定：
    Figure PCTCN2019112636-appb-100009
    式中ξ为综合评价指标，ξ k为单个评价指标，ξ k min/ξ k max为单个评价指标上/下限，λ k为权重系数。
    3)设计正交表头，明确试验方案，进行9次试验，得出试验结果；
    4)根据试验结果计算极差，确定因素的主次顺序，得到各评价指标ξ k的优选方案，根据权重系数λ k，确定优化方案；
    机液复合档位优化方案由下式决定：
    Figure PCTCN2019112636-appb-100010
    式中ξ为综合评价指标，ξ k为单个评价指标，ξ k min/ξ k max为单个评价指标上/下限，λ k为权重系数；
    5)若优化方案尚不能满足要求，可增大或减小“提前”和“延迟”的时间，也可选择不相同的“提前”时间和“延迟”时间，直至满足要求为止；
    6)通过不同工况下得到的换挡机构切换时序数据，控制各组换挡元件的切换时序。
  9. 根据权利要求5所述的多功率分配模式的机械液压复合传动装置的控制方法，其特征在于：所述纯机械传动模式挡位切换时，机械Ⅰ档切换到机械Ⅱ档涉及到1个换挡元件，机械Ⅱ档切换到机械Ⅲ档涉及到1个换挡元件，机械Ⅲ档切换到机械Ⅳ档涉及到3个换挡元件，机械Ⅰ档切换到机械Ⅲ档涉及到2个换挡元件，机械Ⅰ档切换到机械Ⅳ档涉及到5个换挡元件，机械Ⅱ档切换到机械Ⅳ档涉及到4个换挡元件；
    涉及到2个及以下换挡元件的档位切换可通过不超过3次试验给出优化方案；涉及3个换挡元件且其中有2个换挡元件有交互作用的挡位切换、涉及5个换挡元件且其中有3个换挡元件有交互作用的挡位切换以及涉及4个换挡元件且其中有2个换挡元件有交互作用的挡位切换均采用正交表进行分析，分析时选用相应列表；
    确定输出轴角速度的变化量、输出轴角速度的变化率、输出轴角速度的二次微分和时间作为评价指标，选择“提前”、“按时”和“延迟”切换为三水平，进行无交互作用的正交分析；通过试验获得试验数据；根据试验数据计算极差，确定因素的主次顺序，得到各评价指标的优选方案；根据权重系数，确定优化方案；将不同工况优化方案形成成对矩阵编列程序控制档位之间的切换。
  10. 根据权利要求9所述的多功率分配模式的机械液压复合传动装置的控制方法，其特征在于：所述纯机械传动模式挡位切换的控制方法具体步骤如下：
    1)选择输出轴角速度的变化量α、输出轴角速度的变化率β、输出轴角速度的二次微分γ和时间t作为评价指标。α、β、γ分别为转速的零阶、一阶、二阶微分，与时间t共同形成时空评价指标；
    输出轴角速度变化量为：
    Figure PCTCN2019112636-appb-100011
    式中α为输出轴角速度变化量，
    Figure PCTCN2019112636-appb-100012
    为输出轴稳态角速度，ω o min为输出轴最低角速度；
    输出轴角速度变化率定义为：
    Figure PCTCN2019112636-appb-100013
    式中β为输出轴角速度变化率；
    输出轴转矩：
    T o＝β·J o
    式中T o为输出轴转矩，J o为输出轴转动惯量；
    输出轴角速度的二次微分为：
    Figure PCTCN2019112636-appb-100014
    式中γ为输出轴角速度二次微分
    冲击度是车辆纵向加速度的变化率：
    Figure PCTCN2019112636-appb-100015
    式中j为冲击度，r d为轮胎动力半径，i g为变速器传动比，i 0为驱动桥传动比；
    2)建立表2所示的L 27(3 13)正交表；表中，“1”、“2”和“3”分别代表相关换挡元件“提前”、“按时”和“延迟”切换；“a”和“b”分别代表无交互作用的换挡元件；“c”、“d”和“e”分别有交互作用的换挡元件；
    n为总试验次数，n＝27，
    Figure PCTCN2019112636-appb-100016
    为与该因素第i个水平有关的试验结果之和(i∈(1，2,3)，F∈(e，c，e×c1，e×c2，d，e×d1，e×d2，c×d1，a，b，c×d2))；
    表3机械传动模式元件切换时序正交表
    Figure PCTCN2019112636-appb-100017
    Figure PCTCN2019112636-appb-100018
    表中各符号表达式如下：
    Figure PCTCN2019112636-appb-100019
    Figure PCTCN2019112636-appb-100020
    Figure PCTCN2019112636-appb-100021
    Figure PCTCN2019112636-appb-100022
    3)设计如表3正交表头，明确试验方案，进行27次试验，得出试验结果x i(i＝1,...,27)，计算出相关统计值；
    4)再根据方差分析表确定因素和误差的离差平方和及自由度，确定F值，并与临界值进行比较，确定因素和误差的显著性；
    表4机械传动模式元件切换时序方差分析表
    Figure PCTCN2019112636-appb-100023
    总离差平方和Q T，各因素离差平方和Q F和误差离差平方和Q e分别为：
    Q T＝W-P＝∑Q F+Q e
    Q F＝U F-P
    总自由度f T＝26，各因素自由度f F＝2，误差自由度f e＝4；
    因素和误差的平均离差平方和为：
    Figure PCTCN2019112636-appb-100024
    Figure PCTCN2019112636-appb-100025
    F值：
    Figure PCTCN2019112636-appb-100026
    机械传动模式元件切换时序方差分析表如表4所示：
    方差来源主要在三部分：①“a”、“b”、“c”、“d”和“e”；②“e×c”(包括e×c1和e×c2两部分)、“e×d”(包括e×d1和e×d2两部分)和“c×d”(包括c×d1和c×d2两部分)；③误差e *；
    根据计算出的方差来源各因素和误差的离差平方和、自由度、平均离差平方和，将F F
    值和F α(f F，f e)值进行比较；
    若F F＞F α(f F，f e)，则该因素影响显著，否则不显著。选择显著性因素，并直观分析各因素的重要性主次，确定最佳换挡方案。
    机械档位优化方案由下式决定：
    Figure PCTCN2019112636-appb-100027
    式中ζ为综合评价指标，ζ k为单个评价指标，ζ k min/ζ k max为单个评价指标上/下限，μ k为权重系数；
    通过不同工况下得到的换挡机构切换时序数据，控制各组换挡元件的切换时序。
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