S represents the equivalent stiffness of the test piece, primarily the bolt. m₀ is the equivalent mass of the test piece, Fₚ is the pre-tension force applied to the bolt-nut assembly, and f₉ refers to the rolling friction force associated with lateral vibration of the test piece. Unlike conventional hydraulic servo systems, the force behavior of the test piece is unique. Before the test piece becomes loose, if the loading force from the power mechanism is less than the friction force generated by Fₚ on m₀, the test piece only undergoes elastic deformation without any movement. Once the loading force balances with the friction force, the test piece begins to move, and no further elastic deformation occurs. When an alternating load is applied, the test piece alternates between these two stress states.
The design of the hydraulic power mechanism focuses mainly on determining the rated flow of the servo valve, specifically the no-load flow Qom, and the effective working area of the hydraulic cylinder. Although optimized design of hydraulic servo systems is well-established, this system has its own distinct characteristics. First, the load is more complex—depending on the stiffness value Ks, the specimen may behave like an elastic load when Ks is small or resemble a friction load when Ks is large. Second, the power source used in the test bench is limited, providing only about 75 L/min at 21 MPa. These factors significantly influence the output characteristics of the power mechanism and the load trajectory.
In conclusion, the guiding principle for designing the hydraulic power mechanism of the lateral vibration test bench under limited oil source power and complex system dynamics is to ensure sufficient dynamic capacity while minimizing the maximum load flow of the valve. This reduces the demand on the oil source power. Using two valves in parallel helps address the issue of high no-load flow affecting system bandwidth. This approach represents a meaningful attempt in designing such systems.
To ensure stable operation, the minimum outlet flow rate of the pump is set slightly higher than the total system leakage, which meets basic process requirements. However, this flow is not enough to dissipate the heat generated by the mechanical and volumetric inefficiencies of the pump. To solve this, two measures were implemented: forced cooling of the pump casing using low-pressure oil, with the forced flow rate set at half the maximum flow of the constant pressure variable pump. However, due to the high flow rate, the pump shaft seal was damaged from excessive backpressure, leading to unnecessary flow consumption. A low-resistance suction line design was introduced instead, limiting the oil suction velocity to 0.2 m/s, using reducer joints and rubber hoses with parabolic shapes at the pump inlet. This minimized resistance and controlled negative pressure, preventing air separation in the oil. Air in the oil reduces the bulk modulus, which affects both damping ratio and natural frequency of the system.
Since the oil source unit was commissioned in October 1999, it has allowed simultaneous coarse and fine adjustments, with each movement speed meeting process requirements. The working pressure point set by the hydraulic pump has remained unchanged, and the system oil temperature has not exceeded 55°C even under original cooling conditions. The annual average equivalent electric power consumption is also lower than that of the previous system.
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