Design of a Highly Dynamic Hydraulic Actuator?for Active Damping Systems in Machine Tools

Introduction

The demand on productivity and efficiency of machine tools has led to light weight machine designs in order to raise axes accelerations. On the contrary, the requirements of high quality work pieces lead to stiffer machine structures and therefore higher masses. Three different approaches to decrease the dynamic compliance of the machine tool’s structure without increasing its mass can be distinguished: design approaches, passive approaches and active approaches. While design approaches, such as topological optimization of structure parts, are to be considered in the early development phase of the machine tool, passive and active damping systems can be implemented within a machine tool already in production. Because of the fact that passive damping systems act only in a certain frequency bandwidth, active damping systems are developed. Active damping systems enable the user to act in a wider bandwidth, as those systems merge sensor, control and actuator into one system, and therefore adapt to the actual state of the machine tool. In order to work against the occurring vibrations, active damping systems load the structure with additional damping forces. The efficiency of an active damping system is therefore determined by the used actuator principle. Hence, some basic requirements have to be defined for an optimal dimensioning of the used actuator.

Due to economic requirements and the increasing demand for resource efficient machine tools, an additional active damping system should be used within the machine tool’s available energy sources (e.g., electric or hydraulic), and without the need of expensive peripheries. From the technical point of view, the used actuator requires a large frequency bandwidth and positioning velocity. At the same time, the actuator should be able to provide the necessary travel range and positioning accuracy. A high stiffness of the actuator is necessary if it is placed in the direct force flux. Moreover, nonlinearities are not desirable in order to get a high control quality. Regarding the integration into an active damping system, a compact actuator design is necessary.

Many active damping systems for machine tools use piezoelectric actuators to generate damping forces (Ast et al., 2008; Brecher et al., 2009; Munzinger et al., 2010; Simnofske et al., 2009). Isermann (2008) observed that piezoelectric actuators have a high power density (watt/kilogram), a compact design and an excellent dynamic behavior. But they are limited regarding their travel range. According to PI (2012), piezoelectric stack actuators are limited to approximately 2 ‰ of the stack length. In addition, piezoelectric actuators need expensive amplifiers, to generate the necessary power supply as a function of the calculated control signal. These drawbacks are behind the motivation to consider other actuator principles, for a use in active damping systems in machine tools.

In large scale machine tools, large levers and overhanging structural components lead to a need for long travel ranges of the actuator. For example portal milling machines may have a clearance height up to 10,000 mm and a travel of the RAM of up to 4,000 mm. Those structural components create a significant weak spot in the machine tool. Brecher, Manoharan, and Witt (2008) detected this weak spot to be a bending oscillation of the overhanging structure, which leads to a deviation between the work piece and the tool center point (TCP) of the machine. In addition to that, an unstable production process may occur due to pronounced negative real parts of the machine’s dynamic compliance.