Any of the two basic methods described herein are used […]
Any of the two basic methods described herein are used to control the speed of the hydraulic motor. First, a variable displacement pump controls the flow to the motor. This configuration is often referred to as a hydrostatic transmission. Second, a proportional or servo valve powered by a constant pressure source (such as a pressure compensated pump) drives the motor.
The first method is energy efficient. However, the second method can exhibit higher responsiveness. The discussion will focus on the valve control method, not because pump control is not important, but because some problems with valve speed control may not be obvious. Problems with pump speed control will be discussed in subsequent releases.
The valve control of the motor speed uses a constant pressure source; a closed central four-way proportional/servo valve; the controlled motor drives the load torque at a certain speed; the amplifier provides a controllable current.
A typical valve control system is shown schematically. The design of the speed control system consists of two parts. The first is to optimize the component size to pass the maximum power to the load at the design point. The second is a feedback control method that is configured to provide a sufficiently wide range of speeds and speed control to maintain a constant speed of varying loads.
Consider an application, such as a conveyor drive, where the load changes at any given moment and the speed must be adjustable or synchronized with some other process action. The design point is a single worst-case operating point, in which case the load requires the most power. It consists of the load torque requirement TL, DP of the design point and the speed NM, DP required for the torque.
Because this is the worst case, if we design this point, all other operating conditions will be within the operating range of the hydraulic system - we will compete for a successful application.
The optimal size, also known as the optimal design, includes the evaluation and selection of three key hydraulic parameters: supply pressure PS, valve factor KV, T and motor displacement DM to deliver maximum power to the load design in the worst case. point.
Mathematically, when optimizing the design for a hydraulic motor application, we can create two equations, but evaluate three hydraulic parameters - supply pressure, motor displacement and valve factor. Therefore, you must know or specify one of the three before you can calculate the other two.
A complete development must be carried out to cover all three cases, starting with each of the three possible known quantities. For current development, we will only use one of three possibilities: the supply pressure is specified, so we need to calculate the displacement and valve factor.
The hole represents the power and return oil area of the servo or proportional valve. These orifices are commonly referred to as water inlets and outlets, respectively. However, in this article, more modern symbols will be used. The motor and constant pressure pump are represented in a conventional manner.
Of course, the supply pressure must be equal to the sum of the differential pressure losses of the valve ring and the motor. Figure 2 helps illustrate the various pressure drops. From the knowledge of hydraulic circuit analysis, it is further known that the valve pressure loss is related to the square of the flow through each valve coefficient.
The stall torque must be 11⁄2 times the design point operating torque. Of course, at the time of stall, only the leakage through the motor.
The motor displacement can be calculated. However, there is often a problem in the design of hydraulic circuits. Specifically, we need to understand the mechanical efficiency of the motor and its leakage flow even before we know anything about the motor! Fortunately, if we have some experience with hydraulic motors, we can guess reasonable motor mechanical efficiency and flow leakage values under stall conditions.
A reasonable approximation is that for most motors, the expected torque (mechanical) efficiency at stall will be 75% to 95%.
Determining traffic leaks is more complicated. This is the leakage flow through the motor when a full supply pressure is applied to the valve inlet and the load on the shaft is too large to rotate. Obviously, the only flow will be the flow that leaks through the internal gap of the motor. This value is small compared to the flow at full speed.
Suppose we apply a large load to the shaft, even if full pressure is applied and the valve is fully open, it is stopped. Motor flow only includes flow that will be squeezed by a small fixed internal clearance. Now consider the amount of pressure drop on a fully open valve with such a small flow. If the motor has a fairly high volumetric efficiency, it will be particularly small. When the maximum volumetric efficiency is between 80% and 95%, the valve pressure drop at stall will be about 3% to 10% of the applied supply pressure.
Using these approximations and some engineering implications, we can calculate a fairly good estimate of the required motor displacement. It is highly probable that accurately calculated displacements will not be available in the market. Therefore, we will have to choose the closest displacement, as well as the pressure, speed and torque requirements, and evaluate the suitability of the motor for this application.
The VCMM equation can be evaluated at the worst-case design point and then used to solve the required valve coefficients. Obviously, the speed and torque values must reflect the worst-case design points, but the volume and mechanical efficiency of the motor must also be the same.
Note that the efficiency of a full load operating point may be quite different from the efficiency at stall. Inexperienced designers may estimate that most motors have volumetric and mechanical efficiencies between 75% and 95%. However, experienced designers should use the values they feel appropriate.
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