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O+P Fluidtechnik 5/2017

O+P Fluidtechnik 5/2017


CONTROLS AND REGULATIONS FORSCHUNG UND ENTWICKLUNG PEER REVIEWED the number of switching valves and used pressure rails are determined by using measurement data of three working cycles. To analyze the energy consumptions of the proposed system, simulations are conducted. For the simulation, an 18 ton hydraulic excavator was modelled and a conventional LS-system is also simulated as a reference. As the result of the simulation, the newly designed system consumed approximately 20 % less energy than the LS-system in a 90˚ dig and dump test cycle. 2 SYSTEM CONFIGURATION The newly designed hybrid hydraulic system is shown in Figure 02-1. In this section, some main subsystems which make up the hybrid system are explained. 02-2 Efficiency of the ICE and efficiency map of the variable displacement pump 02-3 The supplied pressures of the linear actuators in real working cycles 2.1 ICE AND PUMP The firstly introduced subsystem is the supply consisting of the ICE, a variable displacement pump and accumulators. The efficiency map of the ICE with a rated power of 125 kW based on measurements is shown in Figure 02-2 (left). The efficiency of the ICE depends on the rotation speed and output torque. At lower rotation speeds the efficiency is higher than at higher rotation speed. However, the engine is driven at a high rotation speed in the conventional machines, because it has to supply the peak power or peak flow rapidly. The new hybrid system, on the other hand, is operated at 1100 min -1 . The maximum available torque is 700 Nm and maximum output power at this speed is 81 kW. Even if the same ICE is used, the maximum output power is changed very much according to its rotation speed. For example, when the measured ICE is 82 O+P Fluidtechnik 5/2017

CONTROLS AND REGULATIONS driven at 1800 min -1 , the maximum output power is 120 kW which is approximately 50 % higher than the maximum power of the lower rotation speed. To avoid this negative effect, the accumulators are used to compensate the power deficit occurring during the high power demand. In this study, to avoid the engine stall the braking torque produced by the pumps and other auxiliary components, which are two small pumps for cooling fan and pilot pressure, an electric generator and an air compressor, is limited to 650 Nm, which is equal to 75 kW at 1100 min -1 . And it is assumed that almost 13 kW is always needed to drive the auxiliary components. Although the pump displacement volume is normally determined by the maximum required flow rates to the actuators, the displacement volume of the hybrid system is determined based on the maximum available engine torque and the expected pressure level, because the pump is only used to charge the accumulators without any relations to the loads, thus the pump displacement D pump is described as D pump 2πη = p pump.HM T charge.low pump.max () 1 where, T pump.max is the available torque for the main pump, p charge.low is the lowest charging pressure and η pump.HM is hydro-mechanical pump efficiency. The efficiency map of the main pump is shown in Figure 02-2 (right). The iso-lines show the efficiency and the data was obtained by measurement of a variable axial piston pump. To avoid overloading the ICE, the input power of the main pump must be lower than the available power of 62 kW. On the other hand, the input power of the main pump in a slightly lower range than the available input power limit produces the efficient braking torque of the ICE. In this manner, the accumulator which needs to be charged is selected by the solenoid valve SV1 and the pump displacement is controlled according to the pressure along the approximated pump output power limit of 55 kW by parallel-delayed springs. After both accumulators are fully charged the pump displacement is returned to zero by SV3. Although the solenoid valves of SV1, SV2, SV3 and SV5 are electrically controlled, they are distinguished from electric control used in recent high performance machines, because they can be controlled by five pressure switches of PS1, PS2, PS3, PS4 and PS5 and very simple PLC or relay systems. All pressure switches give ON signals once the pressure becomes lower than each set lower pressure limit and it is returned to OFF after the pressure increases higher than each upper pressure limit. Each pressure limit is summarized in Table 2-1. The solenoid valve SV1 is switched when PS3, PS4 and PS5 are OFF and PS2 is ON. SV3 is switched when both PS1 and PS2 are OFF. Pressure Switches Lower Pressure Limit Upper Pressure Limit PS1 335 bar 350 bar PS2 195 bar 210 bar PS3, PS4 and PS5 50 bar 330 bar Table 2-1 The set pressures of each Pressure Switch 2.2 VALVE SYSTEM FOR LINEAR ACTUATORS In this section, the selection of the pressure levels and a newly designed valve system are discussed. The aim of the valve system is to eliminate the pressure sensors and complex controller as well as to reduce the number of switching valves. At first the pressure supply side is designed. The absolute requirement is that the highest pressure p hp must correspond to the required maximum pressure of the conventional machines to keep the same performance. Firstly two pressure rails (HP: high pressure rail, MP: mid pressure rail) were chosen in this study because a small number of pressure rails makes the system simple and the costs and the frequencies of switching low. To determine the pressure level of MP p mp , the measurements of three working duty cycles, shown in Figure 02-3, are used. These figures show the required pressures to drive the boom, arm and bucket cylinders. Cycle (A) and (C) are 90˚ dig-dump load cycles and (B) is a trenching cycle. The common point of the different cycles is that the machine is firstly exposed to heavy loads during the digging and boom up motion and then the loads become considerably lighter while the machine returns to the initial position. The p mp is set to 60 % of the p hp , i.e., 210 bar, because the pressure differences between p mp and required pressures are relatively small during the heavy loads and the switching frequency can be kept low. To reduce the number of valves, the fluid is basically supplied by MP. When load acting on the linear actuator increases, the pressure of the return side decreases. If at least one pressure of the three return sides becomes lower than the lower pressure limit of PS3, PS4 and PS5, the valve SV2 is switched and HP begins to supply the fluid to all linear actuators. Once the valve is switched, it is kept until the pressure becomes higher than the upper pressure limit again. To design the return flow side, load quadrant diagrams are used, as shown in Figure 02-4. The upper left diagram shows each quadrant. The x-axis illustrates the velocity of the cylinder and y-axis displays the force acting on the cylinder. The cylinder is extended against a resistive load in Quadrant I, is retracted with an assistive load in Quadrant II, is retracted against a resistive load in Quadrant III and is extended with an assistive load in Quadrant IV. Energy recovery is possible in Quadrants II and IV. To analyze the operating points of cylinders during three cycles, the measured forces and operating speeds are plotted in the diagrams. The circle diameters indicate the frequency of each operating point. The horizontal lines in the diagrams show the combinations of supply and return side. The colors of the lines represent the supplying pressure rail. The pressures HP and MP are already defined by the supply side discussion, so that the combinations of HP→HP, HP→MP, HP→TP, MP→HP, MP→MP and MP→TP are available. Here TP is the tank rail. The direction of the arrow indicates the drive direction of the cylinder, the pressure rail written on the left side of the arrow is connected to the piston side and the pressure rail written on the right side is connected to the rod side. In the range of Quadrant I and IV, the combinations allowing higher cylinder forces than each operating point can drive the cylinder and the combinations producing smaller forces can drive it in Quadrant II and III. The combination closest to the operating point is able to reduce throttling losses the most. The boom cylinder is operated only in Quadrant I and II. The combinations of HP→MP, MP→HP and MP→MP in Quadrant I and HP←HP, HP←MP and MP←MP in Quadrant II are ineffective, because they are used very seldomly, so that the HP and MP are not connected to the return side. Therefore, the combinations are limited to MP→TP, HP→TP, TP←MP and TP←HP, which produce large throttling losses at the most frequent load forces located between 200 to 400 kN in Quadrant I and can no longer recover the large potential energy of the front end attachment. To avoid these, the O+P Fluidtechnik 5/2017 83


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