Hydro-mechanical Transmit Performance Analysis for a Continuously Variable Transmission

Transcription

1 Journal of Food Science and Engineering 6 (206) 2-3 doi: / / D DAVID UBLISHING Hydro-mechanical Transmit erformance Analysis for a Continuously Variable Transmission Liang He, 2, Wanlin Guo and Sihong Zhu 2. State Key Laboratory of Mechanics and Control of Mechanical Structures, Institute of Nanoscience, Nanjing University of Aeronautics and Astronautics, Nanjing 2006, China 2. Jiangsu Key Laboratory for Intelligent Agricultural Equipment, College of Engineering, Nanjing Agricultural University, Nanjing 2003, China Abstract: Based on an exact CAD model of hydro-mechanical continuously variable transmission (HMCVT) gearbox which can transmit 80 horsepower, virtual prototype of the HMCVT was built. Revolution speed of shafts, gears and clutches of the HMCVT were calibrated by using results obtained by theoretical calculation and test methods. The needed power and torques of both mechanical power input shaft and hydropower input shaft were calculated by simulation. Hydraulic power distributing ratio and power flow of the system was also studied. The analysis results showed that cycle power was produced inevitably when the output shaft speed of HMCVT change smoothly during mechanical and hydraulic working state HM to HM4, and the instantaneous maximum cycle power was 39.5%. Then the overall transmission efficiency of HMCVT was studied, and the maximum overall efficiency of the system was about 87%. The results of the studies gave references to select suited engine and variable displacement pump for the HMCVT, and to develop rational speed control strategies for the HMCVT by changing displacement ratio of variable displacement pump. Key words: Tractor continuously variable transmission, displacement ratio, rotational speed, torque power flow.. Introduction With development of urbanization and market economy of China which is a big agricultural country, agricultural mechanization is becoming much more important than ever. With the booming of tractor manufacturing industry, the number of large horsepower tractors in China increased a lot in recent years. When tractors do heavy load operations such as plough or rotary tillage, loads amplitude of the tractor fluctuate a lot. Transmission gearbox should shift to appropriate gears frequently to change rotational speed and torque of power output shaft so as to adapt external load changes []. When using traditional manual transmission gearbox, operator of tractor should shift gearbox frequently and appropriately by manual operation, which means a high working Corresponding author: Wanlin Guo, professor, research fields: nanomechancis and fracture mechanics. intensity to operator. In order to reduce operators working intensity and improve comfort, auto transmission gearbox is required, which means output power of engine will not be cut down when gearbox shift. Hydro-mechanical continuously variable transmission (HMCVT) is a kind of auto-transmission gearbox, whose planetary wheels structure is used to couple hydraulic power and mechanical power, and can shift with load output continuously. Key techniques such as mechanical transmission [-6] and control strategy [7-9] of HMCVT were studied by studiers widely, and lots of achievements were obtained. Several kinds of tractors with continuously variable transmission have been developed by some enterprises in western countries, and are sold in market now. Few tractor enterprises in China have developed continuously variable transmission, to say nothing of a HMCVT for large horse power tractors, and some key problems such as reliability of the

2 22 Hydro-mechanical Transmit erformance Analysis for a Continuously Variable Transmission transmission must be solved before the products put into market. Mechanical and hydraulic transmit processes in HMCVT used in big horsepower tractor are complicated. Various kinds of sensors must be installed in the gearbox to obtain information such as torque and impact loads in the transmission. Furthermore, a large amount of fuel will be consumed during the prototype reliable test stage, the cost is high, the period of the reliability test is long, and the data can be obtained during the test is limited. Therefore, it is urgent to study mechanical and hydraulic transmit characteristics in HMCVT for big horsepower tractors. In order to study the transmit performance of HMCVT carefully, and reduce the cost and shorten the period of the design, a kind of HMCVT for 80 horsepower tractor was analyzed by using multi-body dynamics method. Virtual prototype of the HMCVT was built in ADAMS. When five clutches engaged by sequences, and displacement ratio of variable pump changed according to the set rules, the rotational speed and torque of shafts, gears and clutches of the HMCVT were analyzed. The simulated results were consistent with results obtained by theoretical calculation and test results. Hydraulic power distributing ratio, power flow and overall efficiency of the gearbox were also studied. The analysis results show that cycle power was produced when the power output shaft speed of HMCVT change steadily from working state HM to HM4. The studies give references to develop speed control strategies for the HMCVT by changing displacement ratio of variable displacement pump, and also provide a basis for reliability and optimization analysis to the transmission parts in gearbox. 2. Virtual rototype Model of HMCVT 2. arameters of HMCVT The HMCVT studied in this paper uses double planetary wheels structure which is used to couple hydraulic and mechanical power. The gearbox mainly composed of a set of hydraulic system which includes a variable displacement pump and a fixed displacement motor, a set of double row planetary gear train and five wet friction clutches which were marked as C 0 to C 4. The double planetary wheels couple the power from hydraulic power input shaft and mechanical power input shaft, and output power to middle shaft or middle shaft 2. Then the power was output to power output shaft by engaging proper clutches. Fig. is transmission schematic diagram of the HMCVT. Mark is power input shaft from engine. Mark 2 is power input shaft from variable pump. Mark 3 is power output shaft from fixed displacement motor. Mark 4 is mechanical power input Fig. Transmission schematic diagram of HMCVT.

3 Hydro-mechanical Transmit erformance Analysis for a Continuously Variable Transmission 23 shaft. Mark 5 is hydraulic power input shaft. Mark 6 is middle shaft. Mark 7 is middle shaft 2. Mark 8 is power output shaft. When the tractor starts to move forward or backward, clutch C 0 installed on hydraulic power input shaft marked 5 engages, and the gearbox works in pure hydraulic transmit stage HM0 during this time. When the five clutches from C to C 4 engaged by sequences, and displacement ratio of variable pump changes according to the set rules, gearbox works in hydro-mechanical working state from HM to HM4 respectively. By shifting reverse fork to make double gear engage with idle gear or not can drive the tractor forward or backward. When the five clutches from C 0 to C 4 engaged by sequences, and displacement ratio of variable pump changed between - and continuously, gearbox works in hydro-mechanical working state from HM0 to HM4 respectively. Rotational speed of power output shaft marked 8 increases steadily begin with 0. Rotational speed of transmission shafts can be defined by equations from () to (5). n ne 8HM 0, iii 4 5 0i2 (e [0,-0.6] ) () n ii n k e, 45 8HM ( 2 ) ( k2) i4i5i8i2 i (e [-0.4,]) (2) -n ii n k e 45 8HM 2 [( ) )] kiiii 457 i (e [0.8,-] ) (3) n ii n k e, 45 8HM 3 ( 2 ) ( k2) i4i5i9i2 i (e [-0.8,] ) (4) -n ii n k e 45 8HM 4 [( ) )] kiiii 456 i (e [0.8,-] ) (5),, Where: n 8HM0, n 8HM, n 8HM2, n 8HM3, n 8HM4 is rotational speed of power output shaft when gearbox works in working state from HM0 to HM4 respectively. n is output rotational speed of engine; e is displacement ratio of variable pump; k, k 2 are annular gear and sun gear ratio of planetary gear trains K and K2 respectively, and the value of k, k 2 is 2.56 and 3.56 respectively. Transmission ratios of gears are as follows, i =.44, i 2 = 0.97, i 3 =.48, i 4 = 0.67, i 5 =.96, i 6 =.02, i 7 = 3.52, i 8 = 2.77, i 9 = 0.79, i 0 = 2.77, i = 0.82, i 2 = Modeling of HMCVT Virtual rototype Multi-body dynamics model of HMCVT was built in ADAMS based on its accurate three-dimensional model. All parts material were set proper properties. Joints among all moving parts were established as follows. arts such as gears, bearings and keys mounted on shaft rotate with the shaft synchronously were merged with the shaft as one component. arts in clutches C 0 to C 4 rotate with a gear synchronously were merged with the gear as one. All the shafts supported by bearings in the gearbox were simulated by setting revolute joints between the two parts. rocess of clutch engaging was simulated by using a coupler joint to couple the rotational speed of the gear and the shaft. Rotational motion drive was set to power input shaft, and the corresponding torque was set to the power output shaft. Virtual prototype model of HMCVT was shown in Fig Speed Calibration of HMCVT To verify the virtual prototype of HMCVT, rotational speed of all shafts were calibrated. Simulation results were compared with the results obtained by prototype testing, and the virtual model was modified to make sure all the rotational speed consisted with test results. Fig. 4 is test bench of the HMCVT.

4 24 Hydro-mechanical Transmit erformance Analysis for a Continuously Variable Transmission Fig. 2 Virtual prototype of HMCVT. Fig. 3 Rotational speed of power output shaft and overall efficiency of HMCVT. To ensure the fuel efficiency engine is required to running in the fuel economy point and output a fixed rotational speed. But rotational speed of power output shaft of gearbox should change correspondingly in a certain range to adapt to tractor different working conditions, which conflicted with fixed output speed of engine. Requirements of different working conditions were met by changing output speed of engine, but the requirements of energy saving and emission reduction were not satisfied. HMCVT can achieve the fuel economy goal by keeping engine output speed constant, but only changing displacement ratio of variable pump and engage different clutch to adapt to different working conditions. Rotational speed of power output shaft of the HMCVT was simulated when working in different working states. The simulation results were both consistent with results obtained by theoretical calculation and test methods, which was shown in Fig. 3. When gearbox working in pure hydraulic transmit stage and displacement ratio of pump was -0.2, the maximum error between simulation and test results was 5 percent. Minus displacement ratio means pump output reverse hydraulic oil.

5 Hydro-mechanical Transmit erformance Analysis for a Continuously Variable Transmission 25 Fig. 4 Test bench of the HMCVT. : Engine; 2: Rotational speed and torque sensor; 3: Electromagnetic valves; 4: HMCVT; 5: Rotational speed and torque sensor; 6: dynamometer. Fig. 5 Rotational speed of shafts and clutches of HMCVT. When displacement ratio of pump changes from - to, rotational speeds of power input shaft, mechanical power input shaft, annular gear K and five clutches C 0 to C 4 were shown in Fig. 5. The simulation results of the transmission shafts and the clutches were also consistent with test methods. The maximum error was 3 percent. 4. Torque Characteristics Analysis Torque and power transmitted in hydraulic and mechanical transmission lines are important characteristics of the gearbox which should be studied carefully. Due to the high cost and complicated structure of the HMCVT, sensors of torque and rotational speed were not installed to all shafts. Torque and rotational speed of the shaft could not be measured directly in gearbox test bench except for the input and output shafts. Therefore, it is a time and cost saving method to study characteristics of all transmission parts in gearbox by using virtual prototype model. Output rotational speed of engine is set,500 rpm which is the engine fuel economy working point. Engine maximum output power is about 20 kw at this time, which is about 9% of the full power of engine. The engine worked in heavy load operation mode, the remaining power of engine was considered as engine s power reserve. When displacement ratio of pump changed from - to, the motor speed changed from -2,239 rpm to 2,239 rpm, rotational speed of power output shaft increased steadily from 0 to 2,264 rpm. The rotational speed and torque simulation results were shown in Fig. 6.

6 26 Hydro-mechanical Transmit erformance Analysis for a Continuously Variable Transmission Fig. 6 Simulation of rotational speed and torque on mechanical and hydraulic power input shafts. When gearbox works in pure hydraulic driving stage and hydro-mechanical driving stages, due to limitations of ground adhesion the maximum output torque allowed is about 2,887 Nm [4]. ower input rotational speed of engine was set to a fixed value, and output rotational speed of fixed displacement motor was set different values according to different displacement ratio of pump. Corresponding load of torque was set at the end of the power output shaft. Driving torque in hydraulic and mechanical power input shafts could be obtained by simulation respectively. When rotational speed of power out shaft increased from 0 to 2,264 rpm, torque and rotational speed of hydraulic and mechanical power input shafts were shown in Fig. 6. When power output shaft speed was lower than 397 rpm, the maximum output torque is -2,887 Nm due to limitations of ground adhesion. The engine out power was less than 20 kw during this period. When output rotational speed was higher than 397 rpm, the maximum output torque of gearbox was no longer limited by ground adhesion. The maximum out power of engine was 20 kw, and with the speed keep increasing, torque output decreased gradually. When rotational speed of power output shaft got the maximum value 2,264 rpm, and the output torque was -506 Nm. Minus torque means torque with the opposite direction of power input shaft. When gearbox worked in pure hydraulic stage HM0, rotational speed of mechanical power input shaft was -,042 rpm due to constant output rotational speed of engine. Displacement ratio of pump changed from 0 to -0.6, and rotational speed of hydraulic power input shaft increased from 0 to 685 rpm. The unique shaft output power was the hydraulic power input shaft during this period, the driving torque was 957 Nm and kept a constant value. When gearbox worked in hydro-mechanical stage HM, rotational speed of mechanical power input shaft was -,042 rpm, and output torque was 747 Nm. Displacement ratio of pump changed from -0.6 to, rotational speed of hydraulic power input shaft decreased to 0 from 685 rpm, and then output reverse speed from 0 to 42 rpm. The output torque of hydraulic power input shaft was 22 Nm and kept constantly due to limitations of ground adhesion. When gearbox worked in hydro-mechanical stage HM2, rotational speed of mechanical power input shaft kept constantly, and output torque decreased from,390 Nm to 84 Nm gradually. Displacement ratio of pump changed form to -, rotational speed of hydraulic power input shaft changed from -,42 rpm to 0 rpm, and then output a reverse speed from 0 to,42 rpm. The output torque of hydraulic power input shaft changed from -390 Nm

7 Hydro-mechanical Transmit erformance Analysis for a Continuously Variable Transmission 27 to -235 Nm approximate linearly. When gearbox worked in hydro-mechanical stage HM3, rotational speed of mechanical power input shaft kept constantly, and output torque decreased from,462 Nm to 834 Nm gradually. Displacement ratio of pump changed from - to, rotational speed of hydraulic power input shaft changed from,42 rpm to 0 rpm, and then output a reverse speed from 0 to -,42 rpm. The output torque of hydraulic power input shaft changed from 42 Nm to 237 Nm approximate linearly. When gearbox worked in hydro-mechanical stage HM4, rotational speed of mechanical power input shaft kept constantly, and output torque decreased from,459 Nm to 842 Nm gradually. Displacement ratio of pump changed form to -, rotational speed of hydraulic power input shaft changed from -,42 rpm to 0 rpm, and then output a reverse speed from 0 to -,42 rpm. The output torque of hydraulic power input shaft changed from -40 Nm to -235 Nm approximate linearly. 5. Hydraulic ower Distributing Ratio Analysis Hydraulic power distribution ratio is output power ratio of hydraulic driving component and total output power of gearbox without cycle energy loss [3]. M n M n (6) where: ρ is hydraulic power distributing ratio. 5 and 8 are transmission power of hydraulic power input shaft and power output shaft respectively. M 5 and M 8 are transmission torque of hydraulic power input shaft and power output shaft respectively. n 5 and n 8 are rotational speed of hydraulic power input shaft and power output shaft respectively. Fig. 7 is hydraulic power distributing ratio in different transmission stage. When gearbox worked in pure hydraulic stage HM0, mechanical power input shaft did not output any power. The only power output into power output shaft was the power from hydraulic power input shaft, and they should be equal when friction loss was neglected, which can be expressed as ρ =. When gearbox worked in hydro-mechanical stage HM, and rotational speed of power output shaft increased gradually, hydraulic power distributing ratio ρ increased from -0.4 to 0.24 gradually. When gearbox worked in the other hydro-mechanical stages HM2 to HM4, when rotational speed of power output shaft increased gradually, hydraulic power distributing ratio ρ all increased from to 0.24 gradually. All transmission stages except pure hydraulic stage had a Fig. 7 Hydraulic power distributing ratio versus different gear ratio.

8 28 Hydro-mechanical Transmit erformance Analysis for a Continuously Variable Transmission pure mechanical shift where the hydraulic transmission power was 0, which meant hydraulic power distributing ratio ρ = Cycle ower Analysis Cycle power was a kind of useless power produced in closed loop transmission system. This part of energy was transited into heat [4, 5]. The cycle power has a big influence on the overall transmission efficiency. The power transmitted in mechanical power input shaft, hydraulic power input shaft and power output shaft in gearbox were shown respectively in Fig. 8. When gearbox worked in pure transmission stage HM0, and rotational speed of power output shaft increased gradually, mechanical power input shaft did not output any power, output power of hydraulic power input shaft increased from 0 to 69 kw. No cycle power was produced during this period. When gearbox worked in hydro-mechanical stage HM, and displacement ratio of pump changed from -0.6 to 0, rotational speed of power output shaft increased gradually, due to limitations of ground adhesion, output power of mechanical power input shaft kept a constant value which was 82 kw. Output power of hydraulic power input shaft decreased from 0 kw to 0. The product of torque and rotational speed of hydraulic and mechanical power input shaft was negative, which meant ρ < 0. art of the power from mechanical power input shaft was offset by output power from hydraulic power input shaft, which meant cycle power was produced. Output power of gearbox was less than the total power transmitted by mechanical and hydraulic power input shaft. The energy loss of the system decreased from 20 kw to 0 gradually, and cycle power was always twice of the power transmitted by hydraulic power input shaft. Output power from power output shaft increased from 72 kw to 82 kw gradually. Energy loss percentage decreased from 20% to 0. Thereafter displacement ratio of pump changed from 0 to. Fixed displacement motor and hydraulic power input shaft both output a reverse rotational speed. The product of torque and rotational speed of hydraulic and mechanical power input shaft was positive, which meant ρ > 0 and no cycle power was produced during this period. Output power of gearbox was sum of the total power transmitted by hydraulic and mechanical power input shaft. Output power of mechanical power input shaft kept 82 kw and output power of hydraulic Fig. 8 ower transmitting on hydraulic and mechanical power input and output shafts.

9 Hydro-mechanical Transmit erformance Analysis for a Continuously Variable Transmission 29 power input shaft increased from 0 to 25 kw. Therefore total transmitted power of gearbox was increased from 82 kw to 07 kw. When gearbox worked in hydro-mechanical stage HM2, and displacement ratio of pump changed form to 0, rotational speed of power output shaft increased gradually, output power of mechanical power input shaft decreased from 52 kw to 20 kw, and output power of hydraulic power input shaft decreased from 37 kw to 0. Cycle power was produced was during this period. Total transmission power increased from 5 kw to 20 kw. Total energy loss of gearbox decreased from 74 kw to 0 gradually and energy loss percentage decreased from 39.5% to 0. Displacement ratio of pump changed form 0 to, and no cycle power was produced during this period. Output power of mechanical power input shaft decreased from 20 kw to 92 kw, and output power of hydraulic power input shaft increased from 0 to 28 kw. Total transmission power was 20 kw and kept constant thereafter. When gearbox worked in hydro-mechanical stages HM3 and HM4, the situation of power flow was almost the same as work condition of hydro-mechanical stage HM2. The maximum energy loss percentage was 39.5%, which was shown in Fig. 8. When gearbox worked in hydro-mechanical stage HM displacement ratio of pump was from -0.6 to 0, stage HM2 displacement ratio from to 0, stage HM3 displacement ratio from - to 0, stage HM4 displacement ratio from to 0, cycle power was produced. Therefore, when tractor works for a long time under heavy load conditions such as plough or transportation, the gearbox should keep away from the periods when cycle power was produced in stages HM to HM4. ure mechanical transmission shift was highly recommended to maximize efficiency of gearbox, which meant displacement ratio of variable pump 0 was required. The analysis results provide a fundamental basis for speed control strategy of variable pump. During the process of transmission, the period when cycle power was produced was transient, and the processes only last for a short time. Therefore, the tested results on test bench showed that the gearbox could obtain relatively high overall transmission efficiency when proper control strategy was adapted. 7. Overall Transmission Efficiency Analysis When engine worked in fuel economy point and mechanical efficiency of parts of the gearbox were considered, overall transmission efficiency η overall can be calculated according to power transmitted in hydraulic and mechanical power lines and mechanical efficiency of different parts. When gearbox worked in pure hydraulic stage HM0, overall transmission efficiency of gearbox could be described by Eq. (7). overall 5 4 in g rb sb M (7) in 0. abs( e) 0.68 (8) M where: e is displacement of variable displacement pump; in is overall input power of gearbox. η g, η rb, η sb, η c, η -M are the transmission efficiency of gears, rolling bearing, sliding bearings, clutches, variable displacement pump and fixed displacement motor system respectively, and η -M could be described by Eq. (8). The efficiency value of η g, η rb, η sb, η c was 0.98, 0.99, 0.95 and 0.98 respectively. As operating time of clutch engages was approximately to 2 seconds, and had little influence on the overall efficiency of gearbox. Therefore, the efficiency of clutches was neglected in the study. When gearbox worked in pure hydro-mechanical stages HM and HM3, overall transmission efficiency of gearbox could be described by Eqs. (9) and (0) respectively.

10 30 Hydro-mechanical Transmit erformance Analysis for a Continuously Variable Transmission overall overall ( me ) g rb hy g rb M sb in (e [0,] ) (9) ( me ) 2* g rb hy g rb M sb hy (e [-,0] ) (0) where: me is the mechanical power input to power input shaft, hy is the hydraulic power input to 2 power input shaft. When gearbox worked in pure hydro-mechanical stages HM2 and HM4, overall transmission efficiency of gearbox could be described by Eqs. () and (2) respectively. overall overall ( me ) 2* g rb hy g rb M sb hy in in (e [0,] ) () ( me ) g rb hy g rb M sb (e [-,0] ) (2) Overall efficiency of the gearbox can be calculated by equations from (8) to (0), the results were shown in Fig. 3. When gearbox worked in pure hydraulic stage HM0, and displacement ratio of pump e changed from 0 to -0.6, overall transmission efficiency of gearbox increased from 54% to 60%. When gearbox worked in hydro-mechanical stage HM, and displacement ratio of pump e changed from -0.6 to 0 and then 0 to, the overall transmission efficiency of gearbox increased from 62% to 87% at first and subsequently decreased from 87% to 8%. When gearbox worked in hydro-mechanical stage HM2, and displacement ratio of pump e changed from to 0 and then 0 to -, the overall transmission efficiency of gearbox increased from 38% to 8% at first and subsequently decreased from 8% to 77%. When gearbox worked in hydro-mechanical stage HM3, and in displacement ratio of pump e changed from - to 0 and then 0 to, the overall transmission efficiency of gearbox increased from 42% to 87% at first and subsequently decreased from 87% to 8%. When gearbox worked in hydro-mechanical stage HM4, and displacement ratio of pump e changed from to 0 and then 0 to -, the overall transmission efficiency of gearbox increased from 38% to 8% at first and subsequently decreased from 8% to 77%. Overall transmission efficiency of gearbox was low during the working stage HM0 due to the relative low efficiency of variable displacement pump and fixed displacement motor system. When gearbox worked in hydro-mechanical stages from HM to HM4 and displacement of pump was 0, the overall efficiency was no less than 80%. The maximum overall efficiency can be reached which was 87% when gearbox worked in hydro-mechanical stages HM and HM3 and displacement of pump was 0. The analysis results were according with the test results, where the maximum overall efficiency was about 86% [6, 7] which was a relatively high efficiency for variable transmission of large horsepower tractor and almost reached the efficiency of mechanical transmission. 8. Conclusions Virtual prototype of a kind of HMCVT for 80 horsepower tractor was built by using multi-body dynamics method in ADAMS. The model was calibrated by the test results. When gearbox worked in five different working states from HM0 to HM4, rotational speed of shafts, gears and clutches of the HMCVT was simulated. The simulation results were consistent with test results. ower distribution in hydraulic and mechanical transmission lines of the gearbox was studied. Rotational speed and torque of different shafts and hydraulic power distribution ratio of gearbox was obtained by simulation. The results showed the maximum hydraulic power distribution ratio was ower flow and the overall efficiency of the

11 Hydro-mechanical Transmit erformance Analysis for a Continuously Variable Transmission 3 gearbox were studied. The simulation results showed that cycle power was produced during hydro-mechanical working stages from HM to HM4. The instantaneous maximum cycle power was 39.5%. Therefore the gearbox should keep away from the periods when cycle power was produced in stages HM to HM4. ure mechanical transmission shift was highly recommended to maximize efficiency of gear box, and the maximum overall efficiency can be reached which was 87%. The studies provide a fundamental basis for speed control strategy of variable pump, and also provide a basis for reliability and optimization analysis to the transmission parts in gearbox. Acknowledgement The authors acknowledge the support of roject supported by recommend international advanced agricultural science and technology plan of Ministry of Agriculture of China (Grant No. 200-Z8), and the National Natural Science Foundation of China (Grant No ). References [] Xu, L. Y., Zhou, Z. L., Zhang, M. Z. et al Design of Hydro-Mechanical Continuously Variable Transmission of Tractor. Transactions of the Chinese Society for Agricultural Machinery 37 (7): 5-8. [2] Zhang, M. Z., Zhou, Z. L., Xu, L. Y. et al Design of a Multi Range Hydrostatic Mechanical Transmission for Farm Tractors. Journal of Xi an University of Techno logy 9 (6): 8-2. [3] Liu, X. Q., Zhou, Z. L., Xu, L. Y. et al Application of Virtual Assembly Technology in Hydro-mechanical Continuously Variable Transmission. Journal of Henan University of Science and Technology: Natural Science 28 (): [4] Ni, X. D., Zhu, S. H., Ouyang, D. Y. et al Design and Experiment of Hydro-mechanical CVT Speed Ratio for Tractor. Transactions of the Chinese Society for Agricultural Machinery 44 (4): [5] Zhu, S. H., Ni, X. D., Zhang, H. J et al. Doubleplanetary Gear Train Busbar Bodies of Automatic CVT. China: [6] Yuan, S. H., Wei, C., and Zhang, Y. C Influencing Factors for Dynamic Characteristics of Hydro-mechanical Continuously Variable Transmission. Transactions of the Chinese Society of Agricultural Engineering 24 (2): [7] Yuan, S. H., Huo, G. Y., and Zhang, B. B Variable arameter ID Control on the Hydro-mechanical Stepless Transmission. Chinese Journal of Mechanical Engineering 40 (7): 8-4. [8] Wei, C., Yuan, S. H., Hu, J. B. et al. 20. Theoretical and Experimental Investigation of Speed Ratio Follow-up Control System on Geometric Type Hydro-Mechanical Transmission. Chinese Journal of Mechanical Engineering 47 (6): 0-5. [9] Wang, G. M., Zhu, S. H., Wang, S. H. et al Speed Ratio Control of Tractor Hydraulic Mechanical CVT. Transactions of the Chinese Society of Agricultural Engineering 29 (7): [0] Hiroyuki, M., Keiji, O., Tsutomu, I. et al Development of Hydro-Mechanical Transmission (HMT) for Bulldozers. SAE 94772, Journal of assenger Cars 03): [] Editor New Transmission to Challenge the ower Shift. Frican Farming 5/6: [2] Takahisa, N., and Koichi, M Development of Super All Terrain Vehicle and Hydro Mechanical Transmission System. SAE 02448, Journal of assenger Cars : [3] Xu, L. Y., Zhou, Z. L., Zhang, M. Z. et al Characteristics Analysis of Hydro-Mechanical Continuously Variable Transmission of Tractor. Journal of China Agricultural University (5): [4] Cao, F. Y., Wang, J., Zhou, Z. L. et al ower Analysis of Hydro-Mechanical Differential Turning Mechanism of Dongfanghong 302R Tractor. Transactions of the CSAE 2 (3): [5] Zhang, M. Q., Hu, Q. C., Zhu, X. J. et al ower Flow Analysis and Efficiency Calculation for Closed lanetary Gear Systems. Journal of Mechanical Transmission 3 (5): [6] Wang, G. M., Zhu, S. H., Shi, L. X. et al Simulation and Experiment on Efficiency Characteristics of Hydraulic Mechanical Continuously Variable Transmission for Tractor. Transactions of the Chinese Society of Agricultural Engineering 29 (5): [7] Wang, G. M., Zhu, S. H., Shi, L. X. et al Control and Interaction System for Tractor Hydro-mechanical CVT. Transactions of the CSAE 46 (6): -7.