WO2023159407A1 - 一种集齿轮-液压-多盘为一体的多模式复合传动装置 - Google Patents

一种集齿轮-液压-多盘为一体的多模式复合传动装置 Download PDF

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Publication number
WO2023159407A1
WO2023159407A1 PCT/CN2022/077551 CN2022077551W WO2023159407A1 WO 2023159407 A1 WO2023159407 A1 WO 2023159407A1 CN 2022077551 W CN2022077551 W CN 2022077551W WO 2023159407 A1 WO2023159407 A1 WO 2023159407A1
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Prior art keywords
clutch
transmission
hydraulic
gear
assembly
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PCT/CN2022/077551
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English (en)
French (fr)
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朱镇
盛杰
陈龙
蔡英凤
夏长高
韩江义
孙晓东
邹荣
张奕涵
后睿
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江苏大学
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Publication of WO2023159407A1 publication Critical patent/WO2023159407A1/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
    • F16H47/06—Combinations of mechanical gearing with fluid clutches or fluid gearing the fluid gearing being of the hydrokinetic type
    • F16H47/08—Combinations of mechanical gearing with fluid clutches or fluid gearing the fluid gearing being of the hydrokinetic type the mechanical gearing being of the type with members having orbital motion
    • F16H47/085—Combinations of mechanical gearing with fluid clutches or fluid gearing the fluid gearing being of the hydrokinetic type the mechanical gearing being of the type with members having orbital motion with at least two mechanical connections between the hydraulic device and the mechanical transmissions
    • 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
    • F16H37/021—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings toothed gearing combined with continuous variable friction gearing
    • F16H37/022—Combinations of mechanical gearings, not provided for in groups F16H1/00 - F16H35/00 comprising essentially only toothed or friction gearings toothed gearing combined with continuous variable friction gearing the toothed gearing 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
    • F16H57/08—General details of gearing of gearings with members having orbital motion
    • F16H57/082—Planet carriers
    • 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
    • F16H61/04—Smoothing ratio shift
    • 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
    • F16H61/66—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 specially adapted for continuously variable gearings
    • F16H61/664—Friction gearings
    • 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
    • F16H2200/203—Transmissions using gears with orbital motion characterised by the engaging friction means not of the freewheel type, e.g. friction clutches or brakes
    • F16H2200/2053—Transmissions using gears with orbital motion characterised by the engaging friction means not of the freewheel type, e.g. friction clutches or brakes with nine engaging means

Definitions

  • the invention relates to the field of variable speed transmission devices, in particular to a multi-mode compound transmission device integrating gears, hydraulic pressure and multiple discs.
  • variable speed transmission methods currently used in construction machinery generally include gear single-flow transmission, hydraulic single-flow transmission, and hydraulic-gear compound transmission; gear single-flow transmission has high efficiency, but the transmission ratio is fixed, and frequent shifts are required during operation; hydraulic single-flow transmission Flow transmission can easily realize stepless speed regulation, and the transmission torque is large, but its transmission efficiency is low; hydraulic-gear compound transmission is a transmission method in which hydraulic power flow and mechanical power flow are connected in parallel, and it has the advantages of gear transmission. High efficiency and high torque of hydraulic transmission, but it has high requirements for variable hydraulic pump, quantitative hydraulic motor and hydraulic system. Multi-disk continuously variable transmission has the advantages of large transmission power, strong bearing capacity, long life and high transmission efficiency. It is also mostly used in construction machinery, but its transmission ratio has a limited range of variation.
  • the existing technology only involves the design of a single-flow transmission device and a compound transmission device in which two single-flow transmissions are connected in parallel, which fails to fully meet the design requirements for multi-mode transmission devices, especially multiple compound modes, under different working conditions of construction machinery.
  • the present invention provides a multi-mode compound transmission device integrating gear-hydraulic-multi-disk, by switching the clutch assembly and brake assembly, hydraulic transmission, gear transmission, multi-disk Stage variable transmission, gear-multi-disc compound transmission, hydraulic-gear series compound transmission, hydraulic-gear parallel compound transmission, hydraulic-multi-disc series compound transmission, hydraulic-multi-disk parallel compound transmission, hydraulic-gear-multi-disc series compound Switching of multiple modes of transmission, hydraulic-gear-multi-disc parallel compound transmission.
  • the present invention achieves the above-mentioned technical purpose through the following technical means.
  • a multi-mode compound transmission device integrating gear-hydraulic-multi-disc, including input assembly, hydraulic transmission mechanism, front planetary gear mechanism, multi-disk continuously variable transmission mechanism, rear planetary gear mechanism, output assembly, clutch assembly and brake assembly;
  • the clutch assembly connects the output end of the input assembly with the input end of the hydraulic transmission mechanism, the input end of the front planetary gear mechanism and the input end of the multi-disc continuously variable transmission transmission mechanism;
  • the clutch assembly connects the hydraulic transmission mechanism
  • the output ends of the output terminals are respectively connected with the front planetary gear mechanism, the rear planetary gear mechanism and the multi-disk continuously variable transmission mechanism;
  • the clutch assembly connects the front planetary gear mechanism with the multi-disk continuously variable transmission mechanism, and connects the front planetary gear mechanism Connected to the rear planetary gear mechanism to connect the rear planetary gear mechanism to the output assembly;
  • the clutch assembly and brake assembly provide a continuous transmission ratio between the input assembly and the output assembly.
  • the hydraulic transmission H between the input assembly and the output assembly is provided.
  • the clutch assembly includes a first clutch C 1 , a second clutch C 2 , a third clutch C 3 , a fourth clutch C 4 , a fifth clutch C 5 , a sixth clutch C 6 , a seventh clutch C 7 , a Eight clutches C 8 , ninth clutch C 9 , brake B;
  • the first clutch C 1 is used to selectively connect the input assembly with the front planetary gear mechanism and the input end of the multi-disk continuously variable transmission;
  • the second clutch C 2 is used to selectively connect the input assembly with the input end of the hydraulic transmission mechanism;
  • the third clutch C 3 is used to selectively connect the output end of the hydraulic transmission mechanism with the front planetary gear mechanism or the multi-disk continuously variable transmission mechanism series or parallel;
  • the fourth clutch C4 is used to selectively connect the hydraulic transmission mechanism or input assembly with the input end of the multi-disc continuously variable transmission, and the fifth clutch C5 is used to selectively connect the hydraulic transmission mechanism Or the output end of the input assembly is connected to the ring gear of the front planetary gear
  • the sixth clutch C 6 and the ninth clutch C 9 provide hydraulic-gear-multi-disc series compound transmission HGS1 between the input assembly and the output assembly.
  • the multi-mode composite transmission device integrating gear-hydraulic-multi-disk described in the present invention realizes gear transmission, hydraulic transmission, multi-disc continuously variable transmission, gear-multi-disc Compound transmission, hydraulic-gear series compound transmission, hydraulic-gear parallel compound transmission, hydraulic-multi-disc series compound transmission, hydraulic-multi-disk parallel compound transmission, hydraulic-gear-multi-disc series compound transmission, hydraulic-gear-multi-disk
  • the multi-mode switching of parallel compound transmission can meet the requirements of multi-working conditions of construction machinery, improve the utilization rate of engine power, and improve fuel economy.
  • the multi-mode compound transmission device integrating gear-hydraulic-multi-disk of the present invention effectively reduces the impact of gear shifting and increases the range of speed ratio adjustment; the hydraulic transmission starts quickly and works smoothly, and is easy to realize fast.
  • the speed change and reversing without impact, the transmission ratio change process of the multi-disk continuously variable transmission has continuity, and the impact on the mechanism is extremely small during use.
  • the multi-mode composite transmission device integrating gear-hydraulic-multi-disk according to the present invention is equipped with hydraulic-multi-disc series, hydraulic-gear-multi-disc series and hydraulic-multi-disk parallel, hydraulic-gear-multi-disk
  • There are two transmission modes for parallel disks hydraulic-multi-disc series, hydraulic-gear-multi-disc series transmission modes effectively broaden the range of speed regulation, and can meet the requirements of large-scale non-linear stepless speed regulation, hydraulic-multi-disc parallel, hydraulic -The gear-multi-disc parallel transmission method improves the transmission efficiency of the system and can meet the requirements of high-efficiency stepless speed regulation in the region.
  • Fig. 1 is a structural principle diagram of a multi-mode compound transmission device integrating gears, hydraulic pressure and multiple discs according to the present invention.
  • Fig. 2 is a schematic diagram of power flow in H mode of hydraulic transmission according to the present invention.
  • Fig. 3 is a schematic diagram of the power flow in G mode of gear transmission according to the present invention.
  • Fig. 4 is a schematic diagram of power flow in S mode of the multi-disc continuously variable transmission according to the present invention.
  • Fig. 5 is a schematic diagram of the power flow of the gear-multi-disc compound transmission GS according to the present invention.
  • Fig. 6 is a schematic diagram of the power flow of the hydraulic-gear series compound transmission HG1 according to the present invention.
  • Fig. 7 is a schematic diagram of the power flow of the hydraulic-multi-disc series compound transmission HS1 according to the present invention.
  • Fig. 8 is a schematic diagram of the power flow of the hydraulic-gear-multi-disk series compound transmission HGS1 according to the present invention.
  • Fig. 9 is a schematic diagram of the power flow of the hydraulic-gear parallel compound transmission HG2 according to the present invention.
  • Fig. 10 is a schematic diagram of the power flow of the hydraulic-multi-disc parallel compound transmission HS2 according to the present invention.
  • Fig. 11 is a schematic diagram of the power flow of the hydraulic-gear-multi-disc parallel compound transmission HGS2 according to the present invention.
  • FIG. 12 is a diagram showing the relationship between the output speed and the input speed in the mode switching process (H1 ⁇ S ⁇ HS) according to the present invention.
  • Fig. 13 is a diagram showing the relationship between the output speed and the input speed during the mode switching process (H ⁇ G ⁇ HG or H ⁇ HGS) according to the present invention.
  • Fig. 14 is a diagram showing the relationship between the output speed and the input speed in the mode switching process (H ⁇ GS ⁇ HGS1) according to the present invention.
  • Fig. 15 is a diagram of the relationship between the output speed and the input speed in the mode switching process (HG ⁇ S ⁇ HS1) according to the present invention.
  • first and second are used for descriptive purposes only, and cannot be interpreted as indicating or implying relative importance or implicitly specifying the quantity of indicated technical features. Thus, a feature defined as “first” and “second” may explicitly or implicitly include one or more of these features.
  • “plurality” means two or more, unless otherwise specifically defined.
  • the multi-mode compound transmission device integrating gear-hydraulic-multi-disc includes an input assembly 1, a hydraulic transmission mechanism 2, a multi-disc continuously variable transmission mechanism 3, and a front planetary gear Mechanism 4, rear planetary gear mechanism 5, output assembly 6, clutch assembly and brake assembly; said clutch assembly connects the output end of the input assembly 1 with the input end of the hydraulic transmission mechanism 2 and the multi-disc continuously variable transmission mechanism 3 respectively.
  • the input end is connected; the clutch assembly connects the output end of the hydraulic transmission mechanism 2 with the multi-disc continuously variable transmission mechanism 3 and the rear planetary gear mechanism 5; the clutch assembly connects the multi-disc continuously variable transmission mechanism 3 with the rear planetary gear mechanism 5;
  • the front planetary gear mechanism 4 is connected, the front planetary gear mechanism 4 is connected with the rear planetary gear mechanism 5, and the rear planetary gear mechanism 5 is connected with the output assembly 6; the clutch assembly and the brake assembly are provided between the input assembly 1 and the output assembly 6 Continuous gear ratio.
  • Input assembly 1 includes 1-1-engine, 1-2-second gear pair; 1-3-engine input shaft, 1-4-first gear pair, 1-5-first clutch C 1 , 1-6- Engine output shaft; the input shaft 1-2 of the engine 1-1 is connected to the input end of the multi-disk continuously variable transmission mechanism 3 through the first gear pair 1-3.
  • the first clutch C 1 1-5 is used to selectively connect the input shaft 1-3 to the multi-disk continuously variable transmission mechanism 3 .
  • the input shaft 1-3 of the engine 1-1 is connected with the input end of the hydraulic transmission mechanism 2 through the second gear pair 1-2.
  • the hydraulic transmission mechanism 2 includes 2-1-hydraulic transmission input shaft, 2-2-second clutch C 2 , 2-3-variable hydraulic pump, 2-4-quantitative hydraulic motor, 2-5-hydraulic transmission output shaft, 2 -6-the third clutch C3 , 2-7-the third gear pair, 2-8-the eighth clutch C8 ;
  • the hydraulic transmission input shaft 2-1 is connected with the variable hydraulic pump 2-3, and the quantitative hydraulic motor 2-4 It is connected with the hydraulic transmission output shaft 2-5, the variable hydraulic pump 2-3 is used to drive the quantitative hydraulic motor 2-4, and the second clutch C 2 2-2 is used to selectively connect the input shaft 1-2 with the
  • the hydraulic transmission input shaft 2-1 is connected;
  • the third clutch C 3 2-6 is used to selectively connect the hydraulic transmission output shaft 2-5 to the multi-disk continuously variable transmission mechanism through the third gear pair 2-7 3.
  • the eighth clutch C 8 2-8 is used to selectively connect the hydraulic transmission output shaft 2-5 to the rear planetary sun gear 5-4 through the eighth gear pair 5-5 for common rotation.
  • the multi-disc continuously variable transmission mechanism 3 includes 3-1-fourth gear pair, 3-2-fifth gear pair, 3-3-multi-disc continuously variable transmission mechanism power input shaft, 3-4-fourth Clutch C 4 , 3-5-multi-disc continuously variable transmission, 3-6-sixth gear pair, 3-7-sixth clutch C 6 , 3-8-multi-disc continuously variable transmission mechanism power output shaft, 3-9-Brake B; where the fourth clutch C 4 3-4 is used to selectively connect the engine output shaft 1-6 or the hydraulic transmission output shaft 2-5 to the power input shaft 3 of the multi-disc continuously variable transmission mechanism -3 for common rotation; the sixth clutch C 6 is used to selectively connect the power take-off shaft 3-8 of the multi-disc continuously variable transmission mechanism to the front planetary gear sun gear 4-4 for common rotation; the brake B 3 -9 is used to fix the sun gear of the front planetary gear mechanism.
  • the front planetary gear mechanism 4 includes 4-1-the seventh gear pair; 4-2-the fifth clutch C 5 ; 4-3-the front planetary gear ring gear; 4-4-the front planetary gear sun gear; Seven clutches C 7 ; 4-6-front planetary gear planet carrier; 4-7-front planetary gear mechanism power output shaft;; said front planetary gear ring gear 4-3, front planetary gear sun gear 4-4 and front planet
  • the gear carrier 4-6 constitutes a planetary gear train; the brake B 3-5 is used to selectively fix the front planetary gear sun gear 4-4; the fifth clutch C 5 4-5 is used to selectively fix the
  • the engine output shaft 1-6 is connected for common rotation to the front planetary gear ring gear 4-3; the seventh clutch C7 4-5 connects the front planetary gear sun gear 4-4 to the front planetary gear carrier 4-6 for Rotating together means that the front planetary gear mechanism 4 is fixedly connected as one; the power output shaft 4-7 of the front planetary gear mechanism is connected with the front planetary gear carrier 4-6.
  • the rear planetary gear mechanism 5 includes 5-1-the ninth clutch C 9 ; 5-2-the rear planetary gear carrier; 5-3-the rear planetary gear ring gear; 5-4-the rear planetary gear sun gear; 5-5- The eighth gear pair; the 5-2-rear planetary gear carrier, 5-3-rear planetary gear ring gear, and 5-4-rear planetary gear sun gear form a planetary gear train; the ninth clutch C 9 5- 1 is used to selectively connect the 5-2-rear planetary gear carrier to the 5-3-rear planetary gear ring gear for common rotation, that is, the rear planetary gear mechanism is fixedly connected as one; the output assembly 6 and the rear planetary gear planet Frame 5-2 is connected.
  • the gear transmission between the input assembly 1 and the output assembly 6 is provided.
  • the engagement elements of each transmission mode are shown in Table 1. details as follows:
  • ⁇ represents the actuator is in the engaged state
  • ⁇ represents the actuator is in the disengaged state
  • n o (H) is the output speed of the hydraulic transmission H mode
  • n o (G) is the output speed of the gear transmission G mode
  • n o (S) is The output speed of multi-disc continuously variable transmission in S mode
  • n o (HG1) is the output speed of hydraulic-gear series compound transmission HG1 mode
  • n o (HS1) is the output speed of hydraulic-multi-disc series compound transmission HS1 mode
  • n o ( HGS1) is the output speed of hydraulic-gear-multi-disc series compound transmission HGS1 mode
  • n o (GS) is the output speed of gear-multi-disk compound transmission GS mode
  • n o (HG2) is the output speed of hydraulic-gear parallel compound transmission HG2 mode
  • n o (HS2) is the output speed of the hydraulic-multi-disk parallel compound transmission HS2 mode
  • n o (HGS2)
  • the transmission mode of hydraulic transmission H is shown in Figure 2. Only the second clutch C 2 2-2, the eighth clutch C 8 2-8 and the ninth clutch C 9 5-1 are engaged, and other clutches and brakes are disengaged. At this time, the input shaft The engine power transmitted by 1-2 drives the variable hydraulic pump 2-3 to work through the second gear pair 12, and then drives the quantitative hydraulic motor 2-4 to rotate, and the power output by the quantitative hydraulic motor 2-4 is then transmitted to the multi- The disc-type continuously variable transmission mechanism 3 is transmitted to the rear planetary gear sun gear 5-4 through the eighth gear pair 5-5. At this time, the rear planetary gear mechanism 5 is firmly connected as one, and the power is output from the output assembly 6.
  • the gear transmission G transmission mode is shown in Figure 3, only the first clutch C 1 1-5, the fifth clutch C 5 4-2, the ninth clutch C 9 5-1 and the brake B 3-9 are engaged, and other clutch components are disengaged .
  • the engine power transmitted by the input shaft 1-2 drives the front planetary gear ring gear 4-3 through the first clutch C 1 1-5 and the fifth clutch C 5 4-2, because the front planetary gear sun gear 4-4 is The brake B 3-9 brakes, so the power is output from the front planetary gear carrier 4-6 to the power output shaft 4-7 of the front planetary gear mechanism. Since the ninth clutch C 9 5-1 is engaged, the rear planetary gear mechanism 5 is firmly connected As a whole, the power is output from the output assembly 6 .
  • the S mode of the multi-disc continuously variable transmission is shown in Figure 4, engaging the first clutch C 1 1-5, the fourth clutch C 4 3-4, the sixth clutch C 6 4-1, and the seventh clutch C 7 4- 5.
  • the ninth clutch C 9 5-1, other clutches and brakes are separated.
  • the engine power is transmitted to the sun gear 3-1 of the front planetary gear mechanism through the input shaft 1-2.
  • the front planetary gear mechanism 3 is solidly connected as a whole, and the engine power is transmitted to the power input shaft 3-3 of the multi-disc continuously variable transmission mechanism.
  • the power output by the multi-disk continuously variable transmission is transmitted to the sun gear 4-4 of the front planetary gear mechanism through the power output shaft 3-8 of the multi-disc continuously variable transmission mechanism.
  • the rear planetary gear ring gear 5-3, at this time, the rear planetary gear mechanism 5 is firmly connected as one, and the power is output from the output assembly 6.
  • the hydraulic-gear series compound transmission HG1 transmission mode is shown in Figure 5, only the second clutch C 2 2-2, the third clutch C 3 2-6, the fifth clutch C 5 4-2, and the ninth clutch C 9 5 are engaged -1. Brake B 3-9, other clutches are disengaged.
  • the engine power transmitted by the input shaft 1-2 is transmitted to the hydraulic transmission input shaft 2-1 to drive the variable hydraulic pump 2-3 through the second gear pair 1-2, and then drives the quantitative hydraulic motor 2-4 to rotate, and the quantitative hydraulic motor 2
  • the power output by -4 drives the front planetary gear ring gear 4-3 through the hydraulic transmission output shaft 2-5, the third gear pair 2-7, and the fourth gear pair 3-1, because the front planetary gear sun gear 4-4 is The brake B 3-9 brakes, so the power is output from the front planetary gear carrier 4-6 to the power output shaft 4-7 of the front planetary gear mechanism. Since the ninth clutch C 9 5-1 is engaged, the rear planetary gear mechanism 5 is firmly connected As a whole, the power is output from the output assembly 6 .
  • the hydraulic-multi-disc series HS1 transmission mode is shown in Figure 6, only the second clutch C 2 2-2, the third clutch C 3 3-6, the fourth clutch C 4 3-4, and the sixth clutch C 6 3- 7.
  • the seventh clutch C 7 4-5, the ninth clutch C 9 5-1, other clutches and brakes are separated, at this time the engine power transmitted by the input shaft 1-2 is transmitted to the hydraulic transmission input through the first gear pair 1-4
  • the shaft 2-1 drives the variable hydraulic pump 2-3, and then drives the quantitative hydraulic motor 2-4 to rotate.
  • the power output by the quantitative hydraulic motor 2-4 passes through the hydraulic transmission output shaft 2-5, the third gear pair 2-7, the fourth The gear pair 3-1 and the fifth gear pair 3-2 are transmitted to the power input shaft 3-3 of the multi-disc continuously variable transmission mechanism to drive the multi-disk continuously variable transmission 3-5, and the power output by the multi-disc continuously variable transmission is passed through
  • the power output shaft 3-8 of the multi-disk continuously variable transmission mechanism is transmitted to the sun gear 4-4 of the front planetary gear mechanism.
  • the planetary gear mechanism 5 is solidly connected as a whole, and the power is output from the output assembly 6.
  • the hydraulic-gear-multi-disc compound transmission HGS1 mode is shown in Figure 7, only the second clutch C 2 2-2, the third clutch C 3 2-6, the fourth clutch C 4 3-4, and the sixth clutch C 6 are engaged 3-7, the ninth clutch C 9 5-1, other clutches and brakes are separated, at this time the engine power transmitted by the input shaft 1-2 is transmitted to the hydraulic transmission input shaft 2-1 through the first gear pair 1-4 to drive the variable hydraulic pressure
  • the pump 2-3 then drives the quantitative hydraulic motor 2-4 to rotate, and the power output by the quantitative hydraulic motor 2-4 passes through the hydraulic transmission output shaft 2-5, the third gear pair 2-7, and the fourth gear pair 3-1.
  • the hourly power is transmitted in two ways: one way drives the front planetary gear ring gear 4-3 through the fifth clutch C 5 4-2, and the other way drives the multi-disk continuously variable transmission through the fourth clutch C 4 3-4.
  • the power output by the multi-stage transmission is transmitted to the sun gear 4-4 of the front planetary gear mechanism through the power output shaft 3-8 of the multi-disk continuously variable transmission mechanism.
  • the planetary gear carrier 4-6 is output to the power output shaft 4-7 of the front planetary gear mechanism, and because the ninth clutch C 9 5-1 is engaged, the rear planetary gear mechanism 5 is solidly connected as one, and the power is output from the output assembly 6.
  • the gear-multi-disc compound transmission GS transmission mode is shown in Figure 8, only the first clutch C 1 1-5, the fourth clutch C 4 3-4, the fifth clutch C 5 4-2, and the sixth clutch C 6 3 are engaged -7 and the ninth clutch C 9 5-1, other clutches and brakes are separated.
  • the engine power transmitted by the input shaft 1-2 is divided into two roads after passing through the first gear pair 1-4: one road is transmitted to the front planetary gear ring gear 4-3 through the fifth clutch C 5 4-2; the other road is passed through the fifth gear Auxiliary 3-2, the fourth clutch C 4 3-4 is transmitted to the power input shaft 3-3 of the multi-disc continuously variable transmission mechanism to drive the multi-disc continuously variable transmission 3-5, and the power output by the multi-disc continuously variable transmission is passed through
  • the power output shaft 3-8 of the multi-disk continuously variable transmission mechanism is transmitted to the sun gear 4-4 of the front planetary gear mechanism; the two-way power converges at the front planetary gear carrier 4-6, and is output to the power output shaft of the front planetary gear mechanism 4-7. Since the ninth clutch C 9 5-1 is engaged, the rear planetary gear mechanism 5 is solidly connected as a whole, and the power is output from the output assembly 6 .
  • the hydraulic-gear parallel compound transmission HG2 transmission mode is shown in Figure 9, only the first clutch C 1 1-5, the second clutch C 2 2-2, the fifth clutch C 5 4-2, and the eighth clutch C 8 2 are engaged -8 and brake B 3-9, other clutches are disengaged.
  • the engine power transmitted by the input shaft 1-2 is divided into two paths through the first gear pair 1-4 and the second gear pair 1-2: one path is transmitted to the hydraulic transmission input shaft 2-1 through the first gear pair 1-4 for driving
  • the variable hydraulic pump 2-3 drives the quantitative hydraulic motor 2-4 to rotate, and the power output by the quantitative hydraulic motor 2-4 is transmitted to the rear planetary gear sun gear through the hydraulic transmission output shaft 2-5 and the eighth clutch C 8 2-8 5-4; another road drives the front planetary gear ring gear 4-3 through the first clutch C 1 1-5 and the fifth clutch C 5 4-2, because the front planetary gear sun gear 4-4 is restrained by the brake B 3-9 Therefore, the power is output from the front planetary gear carrier 4-6 to the power output shaft 4-7 of the front planetary gear mechanism, and the power output shaft 4-7 of the front planetary gear mechanism transmits the power to the rear planetary gear ring gear 5-3; Road power converges at the rear planetary gear carrier 5-2, and the power is output from the output assembly 6.
  • the hydraulic-multi-disc parallel HS2 transmission mode is shown in Figure 10, only the first clutch C 1 1-5, the second clutch C 2 2-2, the fourth clutch C 4 3-4, and the sixth clutch C 6 3- 7.
  • the seventh clutch C 7 4-5 and the eighth clutch C 8 2-8 are separated from other clutches and brakes, and the engine power transmitted by the input shaft 1-2 passes through the first gear pair 1-4 and the second gear pair 1-2
  • the latter is divided into two routes: one route is transmitted to the hydraulic transmission input shaft 2-1 through the first gear pair 1-4 to drive the variable hydraulic pump 2-3, and then drives the quantitative hydraulic motor 2-4 to rotate, and the quantitative hydraulic motor 2-4 outputs
  • the power is transmitted to the rear planetary gear sun gear 5-4 through the hydraulic transmission output shaft 2-5 and the eighth clutch C 8 2-8; the other way is transmitted to the rear planetary gear sun gear 5-4 through the fifth gear pair 3-2 and the fourth clutch C 4 3-4
  • the power input shaft 3-3 of the multi-disc continuously variable transmission mechanism drives the multi-disc continuously variable transmission 3-5, and
  • the hydraulic-gear-multi-disc parallel compound transmission HGS2 mode is shown in Figure 11, only the first clutch C 1 1-5, the second clutch C 2 2-2, the fourth clutch C 4 3-4, and the fifth clutch C are engaged 5 4-2, the sixth clutch C 6 3-7 and the eighth clutch C 8 2-8, other clutches and brakes are separated, the engine power transmitted by the input shaft 1-2 passes through the first gear pair 1-4 and the second gear After the pair 1-2, it is divided into two roads: one road is transmitted to the hydraulic transmission input shaft 2-1 to drive the variable hydraulic pump 2-3 through the first gear pair 1-4, and then drives the quantitative hydraulic motor 2-4 to rotate, and the quantitative hydraulic motor 2
  • the power output by -4 is transmitted to the rear planetary gear sun gear 5-4 through the hydraulic transmission output shaft 2-5 and the eighth clutch C 8 2-8; the other way is divided into two ways through the first clutch C 1 1-5: One path is transmitted to the front planetary gear ring gear 4-3 through the fifth clutch C 5 4-2; the other path is transmitted to the multi-disk continuously variable transmission
  • the hydraulic-gear series compound transmission mode HG1 is used to start, and the output speed increases linearly with the increase of the displacement ratio e of the hydraulic transmission mechanism.
  • Mode switching process 1 hydraulic transmission H ⁇ multi-disc continuously variable transmission S ⁇ hydraulic-multi-disc compound transmission HS
  • Hydraulic-multi-disk parallel compound transmission HS2 output-input speed relationship is:
  • the gear transmission G output-input speed relationship is:
  • the output-input speed relationship of the hydraulic-gear parallel compound transmission HG2 is:
  • Hydraulic-gear-multi-disk parallel compound transmission HGS2 output-input speed relationship is:
  • the hydraulic transmission mode H is used to start, and the output speed increases linearly with the increase of the displacement ratio e of the hydraulic transmission mechanism.
  • n o (HGS ) increases non-linearly, and the transmission is continuously variable in the range of [0.82, 1.79] n I by changing e and i S.
  • Mode switching process three hydraulic transmission H ⁇ gear-multi-disc compound transmission GS ⁇ hydraulic-gear-multi-disc series compound transmission HGS1
  • the gear-multi-disc continuously variable transmission GS output-input speed relationship is:
  • Hydraulic-gear-multi-disc series compound transmission HGS1 output-input speed relationship is:
  • the output speed increases linearly with the increase of the displacement ratio e of the hydraulic transmission mechanism.
  • n o GS
  • GS mode can be synchronously switched to HGS1 mode, The GS mode is e-fold positively correlated with the HGS1 mode.
  • n o increases nonlinearly from 0 to 2.12n I ; the switching position is different, resulting in gear-multi-disk compound transmission or hydraulic-gear-
  • the output values of the multi-disc series compound transmission mode are different, but the output speed decreases linearly with the reduction of the displacement ratio e of the hydraulic transmission mechanism.
  • Mode switching process four hydraulic-gear series compound transmission HG1 ⁇ multi-disc continuously variable transmission S ⁇ hydraulic-multi-disk series compound transmission HS1
  • the hydraulic-gear series transmission HG1 is used to start, and the output speed increases linearly with the increase of the displacement ratio e of the hydraulic transmission mechanism.
  • the GS mode is e-fold positively correlated with the HGS1 mode, and when switching to the HS1 mode, n o (HS1) increases nonlinearly from 0 to 3.79n I ; the position of switching is different, resulting in hydraulic-gear series compound transmission or hydraulic-multi-disc
  • the output values of the series compound transmission mode are different, but the output speed decreases linearly with the reduction of the displacement ratio e of the hydraulic transmission mechanism.
  • the interval [0, 2.98] n I changes, which is suitable for efficient stepless speed regulation in the region.
  • the transmission efficiency of the hydraulic-multi-disc series transmission HS1 is low, its speed regulation range is wide, and the output speed no(HS1) is in [ 0, 3.9] n I range changes, suitable for large-scale nonlinear stepless speed regulation.

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Abstract

本发明提供了一种集齿轮-液压-多盘为一体的多模式复合传动装置，包括输入组件、液压传动机构、前行星齿轮机构、多盘式无级变速传动机构、后行星齿轮机构、输出组件、离合器组件和制动器组件；所述离合器组件将输入组件的输出端分别与液压传动机构的输入端和前行星齿轮机构的输入端连接；所述离合器组件将液压传动机构的输出端分别与前行星齿轮机构和后行星齿轮机构连接；所述离合器组件将前行星齿轮机构、多盘式无级变速传动机构、后行星齿轮机构和输出组件依次连接；所述离合器组件和制动器组件提供输入组件与输出组件之间连续的传动比。本发明可满足工程机械多工况作业要求，提高发动机功率利用率，改善燃油经济性。

Description

一种集齿轮-液压-多盘为一体的多模式复合传动装置 技术领域
本发明涉及变速传动装置领域，特别涉及一种集齿轮-液压-多盘为一体的多模式复合传动装置。
背景技术
我国是一个石油资源紧缺的能源消耗大国，其中大部分石油被车辆所消耗，随着国家对节能减排越来越重视，农机的节油也备受关注。农机在工作过程中运行的条件比道路车辆恶劣，经常遇到阻力激增的情况。小功率的农机往往要牺牲一定的作业效率才能克服这些极端情况，随之而来的还有油耗增加；大功率的农机则存在成本高，体积大等问题，且在相对良好的作业条件下存在功率过剩的问题。
目前应用在工程机械上的变速传动方式一般有齿轮单流传动、液压单流传动、液压-齿轮复合传动；齿轮单流传动效率高，但传动比固定，操作过程中需要频繁换挡；液压单流传动能够方便地实现无级调速，且传递转矩大，但其传动效率低；液压-齿轮复合传动是一种液压功率流和机械功率流并联的一种传动方式，兼具了齿轮传动的高效率和液压传动的大转矩，但其对变量液压泵和定量液压马达及液压系统要求高，多盘式无级变速传动具有传动功率大、承载能力强、寿命高，传递效率高的特点，也多用于工程机械，但其传动比变化范围有限。
现有技术仅涉及到单流传动装置和两种单流传动并联的复合传动装置的设计，未能完全满足工程机械不同工况下对传动装置多模式，尤其是多种复合模式的设计要求。
发明内容
针对现有技术中存在的不足，本发明提供了一种集齿轮-液压-多盘为一体的多模式复合传动装置，通过切换离合器组件和制动器组件，实现液压传动、齿轮传动、多盘式无级变速传动、齿轮-多盘复合传动、液压-齿轮串联复合传动、液压-齿轮并联复合传动、液压-多盘串联复合传动、液压-多盘并联复合传动、液压-齿轮-多盘式串联复合传动、液压-齿轮-多盘式并联复合传动多种模式的切换。
本发明是通过以下技术手段实现上述技术目的的。
一种集齿轮-液压-多盘为一体的多模式复合传动装置，包括输入组件、液压传动机构、前行星齿轮机构、多盘式无级变速传动机构、后行星齿轮机构、输出组件、离合器组件和制动器组件；所述离合器组件将输入组件的输出端分别与液压传动机构的输入端、前行星齿轮机构的输入端和多盘式无极变速器传动机构输入端连接；所述离合器组件将液压传动机构的输出端 分别与前行星齿轮机构、后行星齿轮机构和多盘式无极变速器传动机构连接；所述离合器组件将前行星齿轮机构与多盘式无级变速传动机构连接，将前行星齿轮传动机构与后行星齿轮机构连接，将后行星齿轮机构与输出组件连接；所述离合器组件和制动器组件提供输入组件与输出组件之间连续的传动比。
进一步，通过调节液压传动机构的排量比、调节多盘式无级变速传动机构的传动比和选择性控制所述离合器组件与制动器组件的接合，提供输入组件与输出组件之间的液压传动H、齿轮传动G和多盘式无级变速传动S中的任意一种或任意两种组合或三种组合的传动方式。
进一步，所述离合器组件包括第一离合器C 1、第二离合器C 2、第三离合器C 3、第四离合器C 4、第五离合器C 5、第六离合器C 6、第七离合器C 7、第八离合器C 8、第九离合器C 9、制动器B；所述第一离合器C 1用于选择性地将输入组件与前行星齿轮机构和多盘式无级变速器输入端连接；所述第二离合器C 2用于选择性地将输入组件与液压传动机构输入端连接；所述第三离合器C 3用于选择性地将液压传动机构输出端与前行星齿轮机构或多盘式无级变速器机构进行串联或者并联；所述第四离合器C 4用于选择性的将液压传动机构或输入组件与多盘式无级变速器输入端连接，所述第五离合器C 5用于选择性地将液压传动机构或输入组件输出端与前行星齿轮机构的齿圈连接以共同旋转；所述第六离合器C 6用于选择性的将多盘式无级变速器输出端与前齿轮机构输入端连接；所述第七离合器C 7用于选择性地将前行星齿轮机构的太阳轮连接到前行星齿轮机构的行星架以共同旋转；；所述第八离合器C 8用于选择性地将液压传动机构的输出端与后行星齿轮机构的太阳轮连接以共同旋转；第九离合器C 9选择性地将后行星齿轮机构的行星架连接到后行星齿轮机构的齿圈以共同旋转；所述制动器B用于选择性地将前行星齿轮机构的太阳轮连接到固定件；通过调节液压传动机构的排量比和选择性控制第二离合器C 2、第八离合器C 8和第九离合器C 9的接合，提供输入组件和输出组件之间的液压传动H。
进一步，通过控制所述第一离合器C 1、第五离合器C 5、第九离合器C 9和制动器B的接合，提供输入组件和输出组件之间的齿轮传动G。
进一步，通过调节多盘式无级变速传动机构的传动比和选择性控制所述第一离合器C 1、第四离合器C 4、第六离合器C 6、第七离合器C 7、第九离合器C 9的接合，提供输入组件和输出组件之间的多盘式无级变速传动S。
进一步，通过调节多盘式无级变速传动机构的传动比和选择性控制所述第一离合器C 1、第四离合器C 4、第五离合器C 5、第六离合器C 6、第九离合器C 9的接合，提供输入组件和输出组件之间的齿轮-多盘式复合传动GS。
通过调节液压传动机构的排量比和选择性控制第二离合器C 2、第三离合器C 3、第五离合 器C 5、第九离合器C 9和制动器B的接合，提供输入组件和输出组件之间的液压-齿轮串联复合传动HG1。
通过调节液压传动机构的排量比和选择性控制第二离合器C 2、第三离合器C 3、第四离合器C 4、第六离合器C 6、第七离合器C 7、和第九离合器C 9的接合，提供输入组件和输出组件之间的液压-多盘式串联复合传动HS1。
进一步，通过调节液压传动机构的排量比、调节多盘式无级变速传动机构的传动比和选择性控制第二离合器C 2、第三离合器C 3、第四离合器C 4、第五离合器C 5、第六离合器C 6和第九离合器C 9、提供输入组件和输出组件之间的液压-齿轮-多盘串联复合传动HGS1。
进一步，通过调节液压传动机构的排量比和选择性的控制第一离合器C 1、第二离合器C 2、第五离合器C 5、第八离合器C 8和制动器B的接合，提供输入组件和输出组件之间的液压-齿轮并联复合传动HG2；
通过调节液压传动机构的排量比、调节多盘式无级变速传动机构的传动比和选择性的控制第一离合器C 1、第二离合器C 2、第四离合器C 4、第六离合器C 6、第七离合器C 7、第八离合器C 8的接合，提供输入组件和输出组件之间的液压-多盘式并联复合传动HS2；
进一步，通过调节液压传动机构的排量比、调节多盘式无级变速传动机构的传动比和选择性控制第一离合器C 1、第二离合器C 2、第四离合器C 4、第五离合器C 5、第六离合器C 6和第八离合器C 8的接合，提供输入组件和输出组件之间的液压-齿轮-多盘并联复合传动HGS2。
进一步，通过调节液压传动机构的排量比、调节多盘式无级变速传动机构的传动比和控制离合器组件的接合，提供“液压传动H→多盘式无级变速传动S→液压-多盘复合传动HS”传动方式之间的无级调速切换。
进一步，通过调节液压传动机构的排量比、调节多盘式无级变速传动机构的传动比和选择性控制离合器与制动器组件的接合，提供“液压传动H→齿轮传动G→液压-齿轮复合传动HG→液压-齿轮-多盘复合传动HGS”传动方式之间的无级调速切换。
进一步，通过调节多盘式无级变速传动机构的传动比和选择性控制离合器与制动器组件的接合，提供“液压传动H→齿轮-多盘式复合传动GS→液压-齿轮-多盘复合传动HGS”传动方式之间的无级调速切换。
进一步，通过调节多盘式无级变速传动机构的传动比和选择性控制离合器与制动器组件的接合，提供“液压-齿轮复合传动HG→多盘式无级变速传动S→液压多盘复合传动HS”传动方式之间的无级调速切换。
本发明的有益效果在于：
1.本发明所述的集齿轮-液压-多盘为一体的多模式复合传动装置，通过切换离合器组件和制动器组件，实现齿轮传动、液压传动、多盘式无级变速传动、齿轮-多盘复合传动、液压-齿轮串联复合传动、液压-齿轮并联复合传动、液压-多盘串联复合传动、液压-多盘并联复合传动、液压-齿轮-多盘式串联复合传动、液压-齿轮-多盘式并联复合传动多种模式的切换，可满足工程机械多工况作业要求，提高发动机功率利用率，改善燃油经济性。
2.本发明所述的集齿轮-液压-多盘为一体的多模式复合传动装置，有效减小了换挡冲击，增大了速比调节范围；液压传动起步快、工作平稳，易于实现快速而无冲击地变速与换向，多盘式无级变速器传动比变化过程具有连续性，在使用过程中对机构的冲击极小。
3.本发明所述的集齿轮-液压-多盘为一体的多模式复合传动装置，设置了液压-多盘串联、液压-齿轮-多盘串联和液压-多盘并联、液压-齿轮-多盘并联各两种传动模式，液压-多盘串联、液压-齿轮-多盘串联传动方式有效拓宽了调速范围，能够满足大范围非线性无级调速的要求，液压-多盘并联、液压-齿轮-多盘并联传动方式提高了系统传动效率，能够满足区域内高效无级调速的要求。
附图说明
图1为本发明所述的集齿轮-液压-多盘为一体的多模式复合传动装置结构原理图。
图2为本发明所述的液压传动H模式功率流示意图。
图3为本发明所述的齿轮传动G模式功率流示意图。
图4为本发明所述的多盘式无级变速传动S模式功率流示意图。
图5为本发明所述的齿轮-多盘复合传动GS功率流示意图。
图6为本发明所述的液压-齿轮串联复合传动HG1功率流示意图。
图7为本发明所述的液压-多盘串联复合传动HS1功率流示意图。
图8为本发明所述的液压-齿轮-多盘串联复合传动HGS1功率流示意图。
图9为本发明所述的液压-齿轮并联复合传动HG2功率流示意图。
图10为本发明所述的液压-多盘并联复合传动HS2功率流示意图。
图11为本发明所述的液压-齿轮-多盘并联复合传动HGS2功率流示意图。
图12为本发明所述的模式切换过程(H1→S→HS)输出转速与输入转速关系图。
图13为本发明所述的模式切换过程(H→G→HG或H→HGS)输出转速与输入转速关系图。
图14为本发明所述的模式切换过程(H→GS→HGS1)输出转速与输入转速关系图。
图15为本发明所述的模式切换过程(HG→S→HS1)输出转速与输入转速关系图。
图中：
1-输入组件；1-1-发动机；1-2-第二齿轮副；1-3-输入轴；1-4-第一齿轮副；1-5-第一离合器C 1；1-6-发动机输出轴；2-液压传动机构；2-1-液压传动输入轴；2-2-第二离合器C 2；2-3-变量液压泵；2-4-定量液压马达；2-5-液压传动输出轴；2-6-第三离合器C 3；2-7-第三齿轮副；2-8-第八离合器C 8；3-多盘式无级变速传动机构；3-1-第四齿轮副；3-2-第五齿轮副；3-3-多盘式无级变速传动机构动力输入轴；3-4-第四离合器C 4；3-5-多盘式无级变速器；3-6-第六齿轮副；3-7-第六离合器C 6；3-8-多盘式无级变速传动机构动力输出轴；3-9-制动器B；4-前行星齿轮机构；4-1-第七齿轮副；4-2-第五离合器C 5；4-3-前行星齿轮齿圈；4-4-前行星齿轮太阳轮；4-5-第七离合器C 7；4-6-前行星齿轮行星架；4-7-前行星齿轮机构动力输出轴；5-后行星齿轮机构；5-1-第九离合器C 9；5-2-后行星齿轮行星架；5-3-后行星齿轮齿圈；5-4-后行星齿轮太阳轮；5-5-第八齿轮副；6-输出组件。
具体实施方式
下面结合附图以及具体实施例对本发明作进一步的说明，但本发明的保护范围并不限于此。
下面详细描述本发明的实施例，所述实施例的示例在附图中示出，其中自始至终相同或类似的标号表示相同或类似的元件或具有相同或类似功能的元件。下面通过参考附图描述的实施例是示例性的，旨在用于解释本发明，而不能理解为对本发明的限制。
在本发明的描述中，需要理解的是，术语“中心”、“纵向”、“横向”、“长度”、“宽度”、“厚度”、“上”、“下”、“轴向”、“径向”、“竖直”、“水平”、“内”、“外”等指示的方位或位置关系为基于附图所示的方位或位置关系，仅是为了便于描述本发明和简化描述，而不是指示或暗示所指的装置或元件必须具有特定的方位、以特定的方位构造和操作，因此不能理解为对本发明的限制。此外，术语“第一”、“第二”仅用于描述目的，而不能理解为指示或暗示相对重要性或者隐含指明所指示的技术特征的数量。由此，限定有“第一”、“第二”的特征可以明示或者隐含地包括一个或者更多个该特征。在本发明的描述中，“多个”的含义是两个或两个以上，除非另有明确具体的限定。
在本发明中，除非另有明确的规定和限定，术语“安装”、“相连”、“连接”、“固定”等术语应做广义理解，例如，可以是固定连接，也可以是可拆卸连接，或一体地连接；可以是机械连接，也可以是电连接；可以是直接相连，也可以通过中间媒介间接相连，可以是两个元件内部的连通。对于本领域的普通技术人员而言，可以根据具体情况理解上述术语在本发明中的具体含义。
如图1所示，本发明所述的集齿轮-液压-多盘为一体的多模式复合传动装置，包括输入组件1、液压传动机构2、多盘式无级变速传动机构3、前行星齿轮机构4、后行星齿轮机构5、 输出组件6、离合器组件和制动器组件；所述离合器组件将输入组件1的输出端分别与液压传动机构2的输入端和多盘式无级变速传动机构3的输入端连接；所述离合器组件将液压传动机构2的输出端分别与多盘式无级变速传动机构3和后行星齿轮机构5连接；所述离合器组件将多盘式无级变速传动机构3与前行星齿轮机构4连接、将前行星齿轮机构4与后行星齿轮机构5连接、将后行星齿轮机构5与输出组件6连接；所述离合器组件和制动器组件提供输入组件1与输出组件6之间连续的传动比。
输入组件1包括1-1-发动机、1-2-第二齿轮副；1-3-发动机输入轴、1-4-第一齿轮副、1-5-第一离合器C 1、1-6-发动机输出轴；所述发动机1-1的输入轴1-2通过第一齿轮副1-3和多盘式无级变速传动机构3输入端连接。所述第一离合器C 1 1-5用于选择性地将输入轴1-3连接到多盘式无级变速传动机构3。所述发动机1-1的输入轴1-3通过第二齿轮副1-2和液压传动机构2输入端连接。
液压传动机构2包括2-1-液压传动输入轴、2-2-第二离合器C 2、2-3-变量液压泵、2-4-定量液压马达、2-5-液压传动输出轴、2-6-第三离合器C 3、2-7-第三齿轮副、2-8-第八离合器C 8；液压传动输入轴2-1与变量液压泵2-3连接，定量液压马达2-4与液压传动输出轴2-5连接，所述变量液压泵2-3用于驱动定量液压马达2-4，所述第二离合器C 2 2-2用于选择性地将输入轴1-2与液压传动输入轴2-1连接；所述第三离合器C 3 2-6用于选择性地将液压传动输出轴2-5通过第三齿轮副2-7连接到多盘式无级变速传动机构3；所述第八离合器C 82-8用于选择性地将液压传动输出轴2-5通过第八齿轮副5-5连接到后行星齿轮太阳轮5-4以共同旋转。
多盘式无级变速传动机构3包括3-1-第四齿轮副、3-2-第五齿轮副、3-3-多盘式无级变速传动机构动力输入轴、3-4-第四离合器C 4、3-5-多盘式无级变速器、3-6-第六齿轮副、3-7-第六离合器C 6、3-8-多盘式无级变速传动机构动力输出轴、3-9-制动器B；其中第四离合器C 43-4用于选择性的将发动机输出轴1-6或液压传动输出轴2-5连接到多盘式无级变速传动机构动力输入轴3-3以共同旋转；第六离合器C 6用于选择性的将多盘式无级变速传动机构动力输出轴3-8连接到前行星齿轮太阳轮4-4以共同旋转；所述制动器B 3-9用于固定前行星齿轮机构太阳轮。
前行星齿轮机构4包括4-1-第七齿轮副；4-2-第五离合器C 5；4-3-前行星齿轮齿圈；4-4-前行星齿轮太阳轮；4-5-第七离合器C 7；4-6-前行星齿轮行星架；4-7-前行星齿轮机构动力输出轴；；所述前行星齿轮齿圈4-3、前行星齿轮太阳轮4-4和前行星齿轮行星架4-6构成行星轮系；所述制动器B 3-5用于选择性地将前行星齿轮太阳轮4-4固定；所述第五离合器C 54-5用于选择性地将发动机输出轴1-6连接到前行星齿轮齿圈4-3以共同旋转；所述第七离合 器C 7 4-5将前行星齿轮太阳轮4-4连接到前行星齿轮行星架4-6以共同旋转，即将前行星齿轮机构4固连为一体；前行星齿轮机构动力输出轴4-7与前行星齿轮行星架4-6连接。
后行星齿轮机构5包括5-1-第九离合器C 9；5-2-后行星齿轮行星架；5-3-后行星齿轮齿圈；5-4-后行星齿轮太阳轮；5-5-第八齿轮副；所述5-2-后行星齿轮行星架、5-3-后行星齿轮齿圈、5-4-后行星齿轮太阳轮构成行星轮系；所述第九离合器C 9 5-1用于选择性地将5-2-后行星齿轮行星架连接到5-3-后行星齿轮齿圈以共同旋转，即将后行星齿轮机构固连一体；所述输出组件6与后行星齿轮行星架5-2相连接。
通过调节液压传动机构2的排量比,调节多盘式无级变速传动机构3的传动比和选择性控制所述离合器组件与制动器组件的接合，提供输入组件1与输出组件6之间齿轮传动、液压传动、多盘式无级变速传动、齿轮-多盘复合传动、液压-齿轮串联复合传动、液压-齿轮并联复合传动、液压-多盘串联复合传动、液压-多盘并联复合传动、液压-齿轮-多盘式串联复合传动、液压-齿轮-多盘式并联复合传动多种模式的切换。各个传动模式的接合元件如表1所示。具体如下：
表1模式切换元件接合状态
Figure PCTCN2022077551-appb-000001
其中：▲代表执行元件处于接合状态，△代表执行元件处于分离状态；n o(H)为液压传 动H模式输出转速，n o(G)为齿轮传动G模式输出转速，n o(S)为多盘式无级变速传动S模式输出转速，n o(HG1)为液压-齿轮串联复合传动HG1模式输出转速，n o(HS1)为液压-多盘串联复合传动HS1模式输出转速，n o(HGS1)为液压-齿轮-多盘串联复合传动HGS1模式输出转速，n o(GS)为齿轮-多盘复合传动GS模式输出转速，n o(HG2)为液压-齿轮并联复合传动HG2模式输出转速，n o(HS2)为液压-多盘并联复合传动HS2模式输出转速，n o(HGS2)为液压-齿轮-多盘复合传动HGS2模式的输出转速，n I为发动机转速，k 1为前行星齿轮机构的行星齿轮特性参数，k 2为后行星齿轮机构的行星齿轮特性参数，e为液压传动机构排量比，i S为多盘式无级变速机构传动比，i 1为第一齿轮副1-4的传动比，i 2为第二齿轮副1-2的传动比，i 3为第三齿轮副2-7的传动比，i 4为第四齿轮副3-1的传动比，i 5为第五齿轮副3-2的传动比,i 6为第六齿轮副3-6的传动比,i 7为第七齿轮副4-1的传动比，i 8为第八齿轮副5-5的传动比。
液压传动H传动模式如图2所示，仅接合第二离合器C 22-2、第八离合器C 82-8和第九离合器C 9 5-1，其它离合器和制动器分离，此时输入轴1-2传递的发动机动力经第二齿轮副12驱动变量液压泵2-3工作，进而带动定量液压马达2-4转动，定量液压马达2-4输出的动力再经第三齿轮副传递到多盘式无级变速传动机构3或经过第八齿轮副5-5传递到后行星齿轮太阳轮5-4，此时后行星齿轮机构5固连为一体，动力从输出组件6输出。
齿轮传动G传动模式如图3所示，仅接合第一离合器C 11-5、第五离合器C 5 4-2、第九离合器C 9 5-1和制动器B 3-9，其它离合器组件分离。此时输入轴1-2传递的发动机动力，经第一离合器C 1 1-5和第五离合器C 54-2驱动前行星齿轮齿圈4-3，由于前行星齿轮太阳轮4-4被制动器B 3-9制动，故动力由前行星齿轮行星架4-6输出至前行星齿轮机构动力输出轴4-7，由于第九离合器C 95-1接合，后行星齿轮机构5固连为一体，动力从输出组件6输出。
多盘式无级变速传动S模式如图4所示，接合第一离合器C 11-5、第四离合器C 4 3-4、第六离合器C 6 4-1、第七离合器C 74-5、第九离合器C 9 5-1，其它离合器和制动器分离。发动机动力经输入轴1-2传递到前行星齿轮机构太阳轮3-1，此时前行星齿轮机构3固连为一体，发动机动力传递到多盘式无级变速传动机构动力输入轴3-3驱动多盘式无级变速器3-5，多盘式无级变速器3-5通过调速控制机构使锥盘和T形盘的中心距a v改变以改变锥盘的工作半径，从而实现变速，多盘式无级变速器输出的动力经多盘式无级变速传动机构动力输出轴3-8传递到前行星齿轮机构太阳轮4-4，此时前行星齿轮机构4固为一体，再传递至后行星齿轮齿圈5-3，此时后行星齿轮机构5固连为一体，动力从输出组件6输出。
液压-齿轮串联复合传动HG1传动模式如图5所示，仅接合第二离合器C 2 2-2、第三离合器C 3 2-6、第五离合器C 5 4-2、第九离合器C 9 5-1、制动器B 3-9，其它离合器分离。此时输入轴1-2传递的发动机动力经第二齿轮副1-2传递到液压传动输入轴2-1驱动变量液压泵 2-3，进而带动定量液压马达2-4转动，定量液压马达2-4输出的动力经液压传动输出轴2-5、第三齿轮副2-7、第四齿轮副3-1，驱动前行星齿轮齿圈4-3，由于前行星齿轮太阳轮4-4被制动器B 3-9制动，故动力由前行星齿轮行星架4-6输出至前行星齿轮机构动力输出轴4-7，由于第九离合器C 95-1接合，后行星齿轮机构5固连为一体，动力从输出组件6输出。
液压-多盘串联HS1传动模式如图6所示，仅接合第二离合器C 2 2-2、第三离合器C 3 3-6、第四离合器C 4 3-4、第六离合器C 6 3-7、第七离合器C 7 4-5、第九离合器C 9 5-1，其它离合器和制动器分离，此时输入轴1-2传递的发动机动力经第一齿轮副1-4传递到液压传动输入轴2-1驱动变量液压泵2-3，进而带动定量液压马达2-4转动，定量液压马达2-4输出的动力经液压传动输出轴2-5、第三齿轮副2-7、第四齿轮副3-1、第五齿轮副3-2传递至多盘式无级变速传动机构动力输入轴3-3，驱动多盘式无级变速器3-5，多盘式无级变速器输出的动力经多盘式无级变速传动机构动力输出轴3-8传递到前行星齿轮机构太阳轮4-4，此时前行星齿轮机构4固为一体，再传递至后行星齿轮齿圈5-3，此时后行星齿轮机构5固连为一体，动力从输出组件6输出。
液压-齿轮-多盘复合传动HGS1模式如图7所示，仅接合第二离合器C 2 2-2、第三离合器C 3 2-6、第四离合器C 4 3-4、第六离合器C 6 3-7、第九离合器C 9 5-1，其它离合器和制动器分离，此时输入轴1-2传递的发动机动力经第一齿轮副1-4传递到液压传动输入轴2-1驱动变量液压泵2-3，进而带动定量液压马达2-4转动，定量液压马达2-4输出的动力经液压传动输出轴2-5、第三齿轮副2-7、第四齿轮副3-1，此时动力分两路传递：一路经第五离合器C 5 4-2驱动前行星齿轮齿圈4-3，另一路经第四离合器C 4 3-4驱动多盘式无级变速器，多盘式无级变速器输出的动力经多盘式无级变速传动机构动力输出轴3-8传递到前行星齿轮机构太阳轮4-4，两路动力在前行星齿轮行星架4-6处汇流，动力由前行星齿轮行星架4-6输出至前行星齿轮机构动力输出轴4-7，由于第九离合器C 95-1接合，后行星齿轮机构5固连为一体，动力从输出组件6输出。
齿轮-多盘复合传动GS传动模式如图8所示，仅接合第一离合器C 11-5、第四离合器C 4 3-4、第五离合器C 5 4-2、第六离合器C 6 3-7和第九离合器C 9 5-1，其它离合器和制动器分离。输入轴1-2传递的发动机动力经第一齿轮副1-4后分为两路：一路经第五离合器C 5 4-2传递至前行星齿轮齿圈4-3；另一路经第五齿轮副3-2、第四离合器C 4 3-4传递到多盘式无级变速传动机构动力输入轴3-3驱动多盘式无级变速器3-5，多盘式无级变速器输出的动力经多盘式无级变速传动机构动力输出轴3-8传递到前行星齿轮机构太阳轮4-4；两路动力在前行星齿轮行星架4-6处汇流，输出至前行星齿轮机构动力输出轴4-7，由于第九离合器C 9 5-1接合，后行星齿轮机构5固连为一体，动力从输出组件6输出。
液压-齿轮并联复合传动HG2传动模式如图9所示，仅接合第一离合器C 11-5、第二离合器C 2 2-2、第五离合器C 5 4-2、第八离合器C 8 2-8和制动器B 3-9，其它离合器分离。输入轴1-2传递的发动机动力经第一齿轮副1-4和第二齿轮副1-2后分为两路：一路经第一齿轮副1-4传递到液压传动输入轴2-1驱动变量液压泵2-3，进而带动定量液压马达2-4转动，定量液压马达2-4输出的动力经液压传动输出轴2-5、第八离合器C 8 2-8传递至后行星齿轮太阳轮5-4；另一路经第一离合器C 1 1-5和第五离合器C 5 4-2驱动前行星齿轮齿圈4-3，由于前行星齿轮太阳轮4-4被制动器B 3-9制动，故动力由前行星齿轮行星架4-6输出至前行星齿轮机构动力输出轴4-7，前行星齿轮机构动力输出轴4-7将动力传递至后行星齿轮齿圈5-3；两路动力在后行星齿轮行星架5-2处汇流，动力从输出组件6输出。
液压-多盘并联HS2传动模式如图10所示，仅接合第一离合器C 11-5、第二离合器C 2 2-2、第四离合器C 4 3-4、第六离合器C 6 3-7、第七离合器C 7 4-5和第八离合器C 8 2-8其它离合器和制动器分离，输入轴1-2传递的发动机动力经第一齿轮副1-4和第二齿轮副1-2后分为两路：一路经第一齿轮副1-4传递到液压传动输入轴2-1驱动变量液压泵2-3，进而带动定量液压马达2-4转动，定量液压马达2-4输出的动力经液压传动输出轴2-5、第八离合器C 8 2-8传递至后行星齿轮太阳轮5-4；另一路经第五齿轮副3-2、第四离合器C 4 3-4传递到多盘式无级变速传动机构动力输入轴3-3驱动多盘式无级变速器3-5，多盘式无级变速器输出的动力经多盘式无级变速传动机构动力输出轴3-8传递到后行星齿轮齿圈5-3，此时第七离合器C 7 4-5接合，前行星齿轮机构固连为一体；两路动力在后行星齿轮行星架5-2处汇流，动力从输出组件6输出。
液压-齿轮-多盘并联复合传动HGS2模式如图11所示，仅接合第一离合器C 11-5、第二离合器C 2 2-2、第四离合器C 4 3-4、第五离合器C 5 4-2、第六离合器C 6 3-7和第八离合器C 8 2-8，其它离合器和制动器分离，输入轴1-2传递的发动机动力经第一齿轮副1-4和第二齿轮副1-2后分为两路：一路经第一齿轮副1-4传递到液压传动输入轴2-1驱动变量液压泵2-3，进而带动定量液压马达2-4转动，定量液压马达2-4输出的动力经液压传动输出轴2-5、第八离合器C 8 2-8传递至后行星齿轮太阳轮5-4；另一路经第一离合器C 11-5再分为两路：一路经第五离合器C 5 4-2传递至前行星齿轮齿圈4-3；另一路经第五齿轮副3-2、第四离合器C 4 3-4传递到多盘式无级变速传动机构动力输入轴3-3驱动多盘式无级变速器3-5，多盘式无级变速器输出的动力经多盘式无级变速传动机构动力输出轴3-8传递到前行星齿轮机构太阳轮4-4；两路动力在前行星齿轮行星架4-6处汇流，输出至前行星齿轮机构动力输出轴4-7，前行星齿轮机构动力输出轴4-7将动力传递至后行星齿轮齿圈5-3；两路动力在后行星齿轮行星架5-2处汇流，动力从输出组件6输出。
如图12所示，通过调节液压传动机构2的排量比，调节多盘式无级变速传动机构3的传动比和选择性控制所述离合器组件的接合，提供“液压传动H→多盘式无级变速传动S→液压-多盘复合传动HS”传动方式之间的无级调速切换。采用模式H起步，输出转速随液压传动机构排量比e的增加线性增大，当e＝1时，液压传动H模式达到正向最大值；当同时满足e·i S∈[n o(H)＝n o(S)]，且e∈[0，1]、i S在所确定传动比范围内时，液压传动H模式可同步切换到多盘式无级变速传动S模式，当多盘式无级变速器传动比i S从最大值变化到最小值时，n o(S)非线性地减小；当同时满足e·i S∈[n o(S)＝n o(HS1)]，且e∈[0，1]、i S在所确定传动比范围内时，多盘式无级变速传动S模式可同步切换到液压-多盘复合传动HS模式；当同时满足e·i S∈[n o(S)＝n o(HS2)]，且e∈[0，1]、i S在所确定传动比范围内时，多盘式无级变速传动S模式可同步切换到液压-多盘复合传动HS2模式，切换的位置不同，导致液压-多盘复合传动HS1模式输出值不同，但输出转速随液压传动机构排量比e的减小而线性减小。
如图13所示，通过调节液压传动机构2的排量比，调节多盘式无级变速传动机构3的传动比和选择性控制所述离合器组件与制动器组件的接合，提供“液压传动H→齿轮传动G→液压-齿轮复合传动HG2或液压传动H→液压-齿轮-多盘HGS2”传动方式之间的无级调速切换。采用液压模式H起步，输出转速随液压传动机构排量比e的增大线性增大，当e＝-1时，液压传动H模式达到负向最大值，当e＝1时，液压传动H模式达到正向最大值；当满足e∈[n 0(H)＝n 0(G)]时，此时可同步切换到齿轮传动G模式，齿轮传动G模式为定传动比传动；当满足e∈[n 0(G)＝n 0(HG2)]时，可以同步切换到液压-齿轮复合传动HG模式，当满足e∈[n 0(H)＝n 0(HGS2)]且e∈[0，1]、i S在所确定传动比范围内时，可以同步切换到液压-齿轮-多盘HGS2模式。
如图14所示，通过调节多盘式无级变速传动机构3的传动比，选择性控制所述离合器组件和制动器组件的接合，提供“液压传动H→齿轮-多盘复合传动GS→液压-齿轮-多盘串联复合传动HGS1”传动方式之间的无级调速切换。采用液压模式H起步，输出转速随液压传动机构排量比e的增大线性增大，当e＝-1时，液压传动H模式达到负向最大值，当e＝1时，液压传动H模式达到正向最大值当n o(H)＝n o(GS)且i S在所确定传动比范围内时，液压传动H模式可同步切换到齿轮-多盘复合传动GS模式；当多盘式无级变速器传动比满足i S∈[n o(GS)＝n o(HGS1)]且e＝1时齿轮-多盘复合传动GS模式可同步切换到多盘式无级变速传动HGS1模式，当多盘式无级变速器传动比i S从最大值变化到最小值时，n o(S)非线性地增加；液压传动机构排量比e＝1固定不变时，传动装置仅通过改变i S在所确定传动比范围内无级变速。
如图15所示，通过调节液压传动机构2的排量比，调节多盘式无级变速传动机构3的 传动比和选择性控制所述离合器组件的接合，提供“液压-齿轮串联复合传动HG1→多盘式无级变速传动S→液压-多盘串联复合传动HS1”传动方式之间的无级调速切换。采用液压-齿轮串联复合传动模式HG1起步，输出转速随液压传动机构排量比e的增加线性增大，当e＝1时，液压传动H模式达到正向最大值；当多盘式无级变速器传动比满足n o(HG)＝n o(S)且e＝1时齿轮-多盘复合传动GS模式可同步切换到多盘式无级变速传动HGS1模式，当多盘式无级变速器传动比i S从最大值变化到最小值时，n o(S)非线性地增加；液压传动机构排量比e＝1固定不变时，传动装置仅通过改变i S在所确定传动比范围内无级变速。实施例举例说明：
主要参数为：i 1i 8＝1，i 2i 7＝1，i 2i 5i 6＝0.2，i 1i 3i 4i 7＝1，i 1i 3i 4i 5i 6＝0.2，k 1＝1.5，k 2＝2.5，i S∈[1.32，13.2]。
模式切换过程一：液压传动H→多盘式无级变速传动S→液压-多盘复合传动HS
液压传动H输出-输入转速关系为：
Figure PCTCN2022077551-appb-000002
多盘式无级变速传动S输出-输入转速关系为：
Figure PCTCN2022077551-appb-000003
液压-多盘串联复合传动HS1输出-输入转速关系为：
Figure PCTCN2022077551-appb-000004
液压-多盘并联复合传动HS2输出-输入转速关系为：
Figure PCTCN2022077551-appb-000005
Figure PCTCN2022077551-appb-000006
如图12所示，采用模式H1起步，输出转速随液压传动机构排量比e的增加线性增大，当e＝0.397时，H模式达到正向0.397n I；当同时满足e·i S＝5.0，且e∈[0，1]，且i S∈[1.32，13.2]三条件时，H模式既可同步切换到S模式，也可以同步切换到HS2模式。同步切换到S模式时，当i S从13.2变化到1.32，n o(S)非线性地从0.379n I增加到3.79n I；当满足e＝1时，S模式可同步切换到HS1模式，当满足e·i S＝5.0，且e∈[0，1]、i S∈[1.32，13.2]三条件时，S模式可同步切换到HS2模式，切换的位置不同，导致液压-多盘串联或并联复合传动模式输出值不同，但输出转速随液压传动机构排量比e的减小而线性减小。
模式切换过程二：液压传动H→齿轮传动G→液压-齿轮复合传动HG2或液压传动H→液压-齿轮-多盘HGS2
液压传动H输出-输入转速关系为：
Figure PCTCN2022077551-appb-000007
齿轮传动G输出-输入转速关系为：
Figure PCTCN2022077551-appb-000008
液压-齿轮并联复合传动HG2输出-输入转速关系为：
Figure PCTCN2022077551-appb-000009
Figure PCTCN2022077551-appb-000010
液压-齿轮-多盘并联复合传动HGS2输出-输入转速关系为：
Figure PCTCN2022077551-appb-000011
如图13所示，采用液压传动模式H起步，输出转速随液压传动机构排量比e的增大线性增大，当e＝1时，液压传动H模式达到正向最大值n I；当e＝0.6时，液压传动模式H可同步切换到齿轮传动G模式，此时齿轮传动为定传动比；齿轮传动G模式也可以同步切换到液压-齿轮并联复合传动HG2模式；当e＝0.8022时，i S＝10时液压传动模式H达到正向0.8022n I，此时液压传动H模式可同步切换到液压-齿轮-多盘并联HGS2模式，当i S从13.2变化到1.32时，n o(HGS)非线性地增加，传动装置通过改变e和i S在[0.82，1.79]n I范围内无级变速。
模式切换过程三：液压传动H→齿轮-多盘复合传动GS→液压-齿轮-多盘串联复合传动HGS1
液压传动H输出-输入转速关系为：
Figure PCTCN2022077551-appb-000012
齿轮-多盘式无级变速传动GS输出-输入转速关系为：
Figure PCTCN2022077551-appb-000013
Figure PCTCN2022077551-appb-000014
液压-齿轮-多盘串联复合传动HGS1输出-输入转速关系为：
Figure PCTCN2022077551-appb-000015
如图14所示，采用液压模式H起步，输出转速随液压传动机构排量比e的增加线性增大，当e＝0.8时，H模式达到0.8n I；当同时满足i S＝10两个条件时，H模式可同步切换到GS模式。同步切换到GS模式时，当i S从13.2变化到1.32，n o(GS)非线性地从0.75n I增加到2.12n I；当满足e＝1时，GS模式可同步切换到HGS1模式，GS模式与HGS1模式呈e倍正相关，切换到HGS1模式时，n o(HGS1)非线性地从0增加到2.12n I；切换的位置不同，导致齿轮-多盘复合传动或液压-齿轮-多盘串联复合传动模式输出值不同，但输出转速随液压传动机构排量比e的减小而线性减小。
模式切换过程四：液压-齿轮串联复合传动HG1→多盘式无级变速传动S→液压-多盘串联复合传动HS1
液压-齿轮串联传动HG1输出-输入转速关系为：
Figure PCTCN2022077551-appb-000016
多盘式无级变速传动S输出-输入转速关系为：
Figure PCTCN2022077551-appb-000017
液压-多盘串联复合传动HS1输出-输入转速关系为：
Figure PCTCN2022077551-appb-000018
如图15所示，采用液压-齿轮串联传动HG1起步，输出转速随液压传动机构排量比e的增加线性增大，当e＝1时，H模式达到正向最大值0.6n I；当同时满足i S＝8.3两个条件时，HG1模式可同步切换到S模式。同步切换到S模式时，当i S从13.2变化到1.32，n o(S)非线性地从0.379n I增加到3.79n I；当满足e＝1时，S模式可同步切换到HS1模式，GS模式与HGS1模式呈e倍正相关，切换到HS1模式时，n o(HS1)非线性地从0增加到3.79n I；切换的位置不同，导致液压-齿轮串联复合传动或液压-多盘串联复合传动模式输出值不同，但输出转速随液压传动机构排量比e的减小而线性减小。
多盘式无级变速器的锥盘和T形盘具有大的锥顶半角,故接触区的当量曲率很小，接触应力低，且在接触表面形成一层坚韧的牵引油膜来传递动力，因而承载能力和寿命高，在正常使用情况下可连线使用多年，并可承受一定的冲击载荷；速度稳定性好，传动效率可高达90％，液压元件的比功率大，但传动效率偏低，通常在80％-90％之间，若取多盘式无级变速器传动效率为90％，液压元件的总效率为80％，当采用多盘-液压串联HS1模式时，系统传动效率＝90％×80％＝72％，当采用多盘-液压并联HS2模式时，设两路的输入功率相同，则系统的传动效率＝0.5×90％+0.5×80％＝85％，与液压-多盘串联传动HS1相比，其效率提高了13％，设9/10的输入功率从多盘路通过，1/10的输入功率从液压路通过，则系统传动效率＝0.9×90％+0.1×80％＝89％，与液压-多盘串联传动HS1相比，其效率提高了17％，故此模式能够有效提升系统的传动效率，但其调速范围有限，输出转速no(HS2)在正向调速区间[0，2.98]n I内变化，适用于区域内高效无级调速，而液压-多盘串联传动HS1虽然传动效率偏低，但其调速范围宽，输出转速no(HS1)在[0，3.9]n I范围内变化，适用于大范围非线性无级调速。
应当理解，虽然本说明书是按照各个实施例描述的，但并非每个实施例仅包含一个独立的技术方案，说明书的这种叙述方式仅仅是为清楚起见，本领域技术人员应当将说明书作为一个整体，各实施例中的技术方案也可以经适当组合，形成本领域技术人员可以理解的其他实施方式。
上文所列出的一系列的详细说明仅仅是针对本发明的可行性实施例的具体说明，它们并非用以限制本发明的保护范围，凡未脱离本发明技艺精神所作的等效实施例或变更均应包含在本发明的保护范围之内。

Claims (10)

  1. 一种集齿轮-液压-多盘为一体的多模式复合传动装置，其特征在于，包括输入组件(1)、液压传动机构(2)、多盘式无级变速传动机构(3)、前行星齿轮机构(4)、后行星齿轮机构(5)、输出组件(6)、离合器组件和制动器组件；所述离合器组件将输入组件(1)的输出端分别与液压传动机构(2)的输入端、多盘式无级变速传动机构(3)的输入端和前行星齿轮机构(4)的输入端连接；所述离合器组件将液压传动机构(2)的输出端分别与多盘式无级变速传动机构(3)、前行星齿轮机构(4)和后行星齿轮机构(5)连接；所述离合器组件将多盘式无级变速传动机构(3)与前行星齿轮机构(4)连接、将前行星齿轮机构(4)与后行星齿轮机构(5)连接、将后行星齿轮机构(5)与输出组件(6)连接；
    所述离合器组件包括第一离合器C 1(1-5)、第二离合器C 2(2-2)、第三离合器C 3(2-6)、第四离合器C 4(3-4)、第五离合器C 5(4-2)、第六离合器C 6(3-7)、第七离合器C 7(4-5)、第八离合器C 8(2-8)、第九离合器C 9(5-1)、制动器B；所述第一离合器C 1(1-5)用于选择性地将输入轴(1-3)连接到多盘式无级变速传动机构(3)；所述第二离合器C 2(2-2)用于选择性地将输入组件(1)与液压传动机构(2)输入端连接；所述第三离合器C 3(2-6)用于选择性地将液压传动输出轴(2-5)通过第三齿轮副(2-7)连接到多盘式无级变速传动机构(3)；所述第四离合器C 4(3-4)用于选择性的将发动机输出轴(1-6)或液压传动输出轴(2-5)连接到多盘式无级变速传动机构动力输入轴(3-3)以共同旋转；所述第五离合器C 5(4-5)用于选择性地将发动机输出轴(1-6)连接到前行星齿轮齿圈(4-3)以共同旋转；所述第六离合器C 6(3-7)用于选择性的将多盘式无级变速传动机构动力输出轴(3-8)连接到前行星齿轮太阳轮(4-4)以共同旋转；所述第七离合器C 7(4-5)将前行星齿轮太阳轮(4-4)连接到前行星齿轮行星架(4-6)以共同旋转；所述第八离合器C 8(2-8)用于选择性地将液压传动输出轴(2-5)通过第八齿轮副(5-5)连接到后行星齿轮太阳轮(5-4)以共同旋转；所述第九离合器C 9(5-1)用于选择性地将后行星齿轮行星架(5-2)连接到后行星齿轮齿圈(5-3)以共同旋转；所述制动器B(3-9)用于固定前行星齿轮机构太阳轮(4-4)；
    通过调节液压传动机构(2)的排量比、调节多盘式无级变速传动机构(3)的传动比和选择性控制所述离合器组件与制动器组件的接合，提供输入组件(1)与输出组件(6)之间的液压传动H、齿轮传动G和多盘式无级变速传动S中的任意一种或任意两种组合或三种组合的传动方式。
  2. 根据权利要求1所述的集齿轮-液压-多盘为一体的多模式复合传动装置，其特征在于通过调节液压传动机构(2)的排量比和选择性控制通过控制所述第二离合器C 2(2-2)、第八离合器C 8(2-8)和第九离合器C 9(5-1)的接合，提供输入组件(1)和输出组件(6) 之间的液压传动H。
  3. 根据权利要求2所述的集齿轮-液压-多盘为一体的多模式复合传动装置，其特征在于，通过控制所述控制第一离合器C 1(1-5)、第五离合器C 5(4-2)、第九离合器C 9(5-1)和制动器B(3-9)的接合，提供输入组件(1)和输出组件(6)之间的齿轮传动G。
  4. 根据权利要求3所述的集齿轮-液压-多盘为一体的多模式复合传动装置，其特征在于，通过调节多盘式无级变速传动机构(4)的传动比和选择性控制所述第一离合器C 1(1-5)、第四离合器C 4(3-4)、第六离合器C 6(4-1)、第七离合器C 7(4-5)、第九离合器C 9(5-1)的接合，提供输入组件(1)和输出组件(6)之间的多盘式无级变速传动S。
  5. 根据权利要求4所述的集齿轮-液压-多盘为一体的多模式复合传动装置，其特征在于，通过调节液压传动机构(2)的排量比和选择性的控制第二离合器C 2(2-2)、第三离合器C 3(2-6)、第五离合器C 5(4-2)、第九离合器C 9(5-1)、制动器B(3-9)的接合，通过选择性的控制第一离合器C 1(1-5)、第二离合器C 2(2-2)、第五离合器C 5(4-2)、第八离合器C 8(2-8)和制动器B(3-9)的接合，提供输入组件(1)和输出组件(6)之间的各自相异的液压-齿轮HG；
    通过调节多盘式无级变速传动机构(4)的传动比和选择性的控制第一离合器C 1(1-5)、第四离合器C 4(3-4)、第五离合器C 5(4-2)、第六离合器C 6(3-7)和第九离合器C 9(5-1)的接合，提供输入组件(1)和输出组件(6)之间的齿轮-多盘复合传动GS；
    通过调节液压传动机构(2)的排量比、调节多盘式无级变速传动机构(3)的传动比和选择性的控制第二离合器C 2(2-2)、第三离合器C 3(3-6)、第四离合器C 4(3-4)、第六离合器C 6(3-7)、第七离合器C 7(4-5)、第九离合器C 9(5-1)的接合，通过选择性的控制第一离合器C 1(1-5)、第二离合器C 2(2-2)、第四离合器C 4(3-4)、第六离合器C 6(3-7)、第七离合器C 7(4-5)和第八离合器C 8(2-8)的接合，提供输入组件(1)和输出组件(6)之间的各自相异的液压-多盘HS。
  6. 根据权利要求5所述的集齿轮-液压-多盘为一体的多模式复合传动装置，其特征在于，通过调节液压传动机构(2)的排量比、调节多盘式无级变速传动机构(4)的传动比和选择性控制第二离合器C 2(2-2)、第三离合器C 3(2-6)、第四离合器C 4(3-4)、第六离合器C 6(3-7)、第九离合器C 9(5-1)的接合，通过选择性的控制第一离合器C 1(1-5)、第二离合器C 2(2-2)、第四离合器C 4(3-4)、第五离合器C 5(4-2)、第六离合器C 6(3-7)和第八离合器C 8(2-8)的接合，提供输入组件(1)和输出组件(6)之间的各自相异的液压-齿轮-多盘复合传动HGS。
  7. 根据权利要求5所述的集齿轮-液压-多盘为一体的多模式复合传动装置，其特征在于，通过调节液压传动机构(2)的排量比、调节多盘式无级变速传动机构(3)的传动比和选择 性控制离合器与制动器组件的接合，提供“液压传动H→多盘式无级变速传动S→液压-多盘复合传动HS”传动方式之间的无级调速切换。
  8. 根据权利要求6所述的集齿轮-液压-多盘为一体的多模式复合传动装置，其特征在于，通过调节液压传动机构(2)的排量比、多盘式无级变速传动机构(3)的传动比和选择性控制离合器与制动器组件的接合，提供“液压传动H→齿轮传动G→液压-齿轮复合传动HG2或液压传动H→液压-齿轮-多盘HGS2”传动方式之间的无级调速切换。
  9. 根据权利要求6所述的集齿轮-液压-多盘为一体的多模式复合传动装置，其特征在于，通过调节液压传动机构(2)的排量比、调节多盘式无级变速传动机构(3)的传动比和控制离合器组件的接合，提供“液压传动H→齿轮-多盘复合传动GS→液压-齿轮-多盘串联复合传动HGS1”传动方式之间的无级调速切换。
  10. 根据权利要求5所述的集齿轮-液压-多盘为一体的多模式复合传动装置，其特征在于，通过调节液压传动机构(2)的排量比、调节多盘式无级变速传动机构(3)的传动比和控制离合器组件的接合，提供“液压-齿轮串联复合传动HG1→多盘式无级变速传动S→液压-多盘串联复合传动HS1”传动方式之间的无级调速切换。
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