CN101169159A - Large damping magnetic levitation high-speed rotation system device - Google Patents
Large damping magnetic levitation high-speed rotation system device Download PDFInfo
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Abstract
一种大阻尼磁悬浮高速旋转系统装置，属于磁悬浮技术领域。它包括转子组件、两个径向磁悬浮轴承组件、一个推力轴承组件和驱动电机，其特征在于：还包括只对转子组件具有支承阻尼而无支承刚度的辅助磁悬浮支承组件。本发明的辅助磁悬浮支承组件，不提供支承刚度，仅提供可控附加阻尼，可以避免对低转速下转子的刚体模态振动产生超静定约束，同时通过调整可调电位器灵活改变支承阻尼以满足系统动态性能的要求。这一技术方案简单易行，效果好。旋转机械转子高速、重载、细长发展的要求，使机械的动力学问题日益突出，本发明的研究结果能为解决相关实际问题提供思路与借鉴。
The utility model relates to a large-damping magnetic levitation high-speed rotating system device, which belongs to the technical field of magnetic levitation. It includes a rotor assembly, two radial magnetic suspension bearing assemblies, a thrust bearing assembly and a drive motor, and is characterized in that it also includes an auxiliary magnetic suspension support assembly that only has support damping for the rotor assembly but has no support stiffness. The auxiliary magnetic suspension support assembly of the present invention does not provide support stiffness, but only provides controllable additional damping, which can avoid super-static constraints on the rigid body modal vibration of the rotor at low speeds, and at the same time flexibly change the support damping by adjusting the adjustable potentiometer to achieve Meet the requirements of system dynamic performance. This technical solution is simple and easy to implement, and has good effect. The requirements of high speed, heavy load and slender development of rotating machinery rotors make the dynamic problems of machinery increasingly prominent. The research results of the present invention can provide ideas and references for solving related practical problems.
Description
技术领域technical field
本发明的大阻尼磁悬浮高速旋转系统装置，属于磁悬浮技术领域。The large-damping magnetic levitation high-speed rotation system device of the present invention belongs to the technical field of magnetic levitation.
背景技术Background technique
和传统轴承相比，磁悬浮轴承与转子无接触，支承功耗小，使用寿命长；不需要润滑和密封，可长期用于高低温等特殊环境中；维护费用低、便于主动控制等等，因而被认为是支承技术的一次革命，是目前唯一投入实用的主动支承装置。磁悬浮轴承主要用于刚性转子系统，由于旋转机械转子高速、重载、细长发展的要求，现正逐步扩展到柔性转子系统。Compared with traditional bearings, magnetic suspension bearings have no contact with the rotor, support low power consumption, and have a long service life; do not need lubrication and sealing, and can be used in special environments such as high and low temperature for a long time; low maintenance costs, easy to actively control, etc., so It is considered to be a revolution in support technology, and it is the only active support device put into practical use at present. Magnetic suspension bearings are mainly used in rigid rotor systems. Due to the requirements of high-speed, heavy-load, and slender development of rotating machinery rotors, they are gradually being extended to flexible rotor systems.
但是，国内外相关研究结果表明，将磁悬浮轴承运用于柔性转子系统存在着较大的困难。主要原因是磁悬浮轴承的等效刚度及等效阻尼受控制参数稳定区域的限制，一般比动压滑动轴承小2～3个数量级。在系统接近或越过弯曲临界转速时，因阻尼过小，转子振幅过大，容易导致系统破坏。因此，针对柔性转子系统的研究一直是该技术领域的热点和难点。为了减小磁悬浮轴承转子系统的振动，国内外许多文献从两方面进行了研究。一是采取同步振动抑制技术，二是运用现代控制理论或鲁棒控制理论设计控制方案以提高支承阻尼。However, the relevant research results at home and abroad show that there are great difficulties in applying magnetic suspension bearings to flexible rotor systems. The main reason is that the equivalent stiffness and equivalent damping of the magnetic suspension bearing are limited by the stable area of the control parameters, which are generally 2 to 3 orders of magnitude smaller than the dynamic pressure sliding bearing. When the system approaches or exceeds the bending critical speed, the damping is too small and the rotor amplitude is too large, which easily leads to system damage. Therefore, research on flexible rotor systems has always been a hot and difficult point in this technical field. In order to reduce the vibration of the magnetic suspension bearing rotor system, many literatures at home and abroad have carried out research from two aspects. One is to adopt synchronous vibration suppression technology, and the other is to use modern control theory or robust control theory to design a control scheme to improve support damping.
但是，同步振动抑制技术难以解决实际柔性转子系统的不平衡振动问题。这是由于转子在亚临界及超临界状态将产生弯曲变形，不平衡质量所引起的振动与转子弯曲变形状态有关，远比刚性转子系统复杂，因而目前同步振动抑制技术主要针对刚性转子系统。However, the synchronous vibration suppression technology is difficult to solve the unbalanced vibration problem of the actual flexible rotor system. This is because the rotor will produce bending deformation in the subcritical and supercritical state, and the vibration caused by the unbalanced mass is related to the bending deformation state of the rotor, which is far more complicated than the rigid rotor system. Therefore, the current synchronous vibration suppression technology is mainly aimed at the rigid rotor system.
国内外一些研究结果表明，采用现代控制理论或鲁棒控制理论设计合适的控制方案可以改善系统的动态性能，但目前尚不能做到大幅度提高系统在亚临界及超临界运行时的支承阻尼，而且研究对象多为实验系统，实际应用不多。2000年轴承制造著名企业日本“光洋株式会社”研发中心研究人员Hirochika Ueyama在瑞士举办的第7届磁悬浮轴承国际会议上发文认为，“尽管一些公开文献表明，采用现代控制理论(LQG、H∞、μ理论等)能够解决这个问题，但这仍然是一个具有挑战性问题。为了避免这个问题，相关实际应用项目采用刚性转子结构”(Hirochika Ueyama，Helium Cold Compressor with Active Magnetic Bearings，Proc.of the 7th Int.Symp.on Magnetic Bearings，Zurich，Switzerland，August 2000，1～6)。Some research results at home and abroad show that using modern control theory or robust control theory to design a suitable control scheme can improve the dynamic performance of the system, but it is not yet possible to greatly improve the support damping of the system in subcritical and supercritical operation. Moreover, most of the research objects are experimental systems, and there are not many practical applications. In 2000, Hirochika Ueyama, a researcher at the R&D Center of Koyo Co. , Ltd., a well-known bearing manufacturing company in Japan, published a paper at the 7th International Conference on Magnetic Suspension Bearings held in Switzerland. μ theory, etc.) can solve this problem, but it is still a challenging problem. To avoid this problem, related practical application projects adopt rigid rotor structure” (Hirochika Ueyama, Helium Cold Compressor with Active Magnetic Bearings, Proc. of the 7 th Int. Symp. on Magnetic Bearings, Zurich, Switzerland, August 2000, 1-6).
发明内容Contents of the invention
本发明的目的，在于将磁悬浮轴承运用于柔性转子系统，即提供一种振动幅度小、稳定性强的大阻尼磁悬浮高速旋转系统装置。The object of the present invention is to apply the magnetic suspension bearing to the flexible rotor system, that is, to provide a large damping magnetic suspension high-speed rotating system device with small vibration amplitude and strong stability.
这种大阻尼磁悬浮高速旋转系统装置，包括转子组件、两个径向磁悬浮轴承组件、一个推力轴承组件和驱动电机，其特征在于：还包括只对转子组件具有支承阻尼而无支承刚度的辅助磁悬浮支承组件。This large-damping magnetic levitation high-speed rotating system device includes a rotor assembly, two radial magnetic levitation bearing assemblies, a thrust bearing assembly and a drive motor, and is characterized in that it also includes an auxiliary magnetic levitation that only has support damping but no support stiffness for the rotor assembly Support components.
一般磁悬浮轴承转子系统主要由转子组件、两个径向磁悬浮轴承组件、一个推力轴承组件和驱动电机组成。在合适控制参数作用下，磁轴承对转子产生支承刚度与支承阻尼，决定了整个系统的动态性能。由于控制参数稳定区域有限，支承阻尼的选择受到限制。A general magnetic suspension bearing rotor system is mainly composed of a rotor assembly, two radial magnetic suspension bearing assemblies, a thrust bearing assembly and a drive motor. Under the action of appropriate control parameters, the magnetic bearing produces support stiffness and support damping for the rotor, which determines the dynamic performance of the entire system. Due to the limited stable region of the control parameters, the choice of support damping is limited.
本发明装置在一般磁悬浮轴承转子系统的基础上增加了辅助磁悬浮支承组件。通过增加该组件，提高系统的支承阻尼，降低转子的振幅，提高系统的稳定性。和一般径向磁悬浮轴承组件不同的是，辅助磁悬浮支承组件对转子不提供支承刚度，仅提供可控附加阻尼，以避免对低转速下转子的刚体模态振动产生超静定约束，为此设计了相应的控制器电路。The device of the invention adds an auxiliary magnetic suspension support assembly on the basis of a general magnetic suspension bearing rotor system. By adding this component, the support damping of the system is improved, the vibration amplitude of the rotor is reduced, and the stability of the system is improved. Different from the general radial magnetic suspension bearing assembly, the auxiliary magnetic suspension bearing assembly does not provide support stiffness for the rotor, but only provides controllable additional damping to avoid super-static constraints on the rigid body modal vibration of the rotor at low speeds. corresponding controller circuit.
在由传统轴承(滚动轴承、滑动轴承)支承的转子系统中，有一些实际应用采用了多点支承方式，但各点支承均对转子产生支承刚度与支承阻尼。为了减小支承对转子的超静定约束，通常采用提高加工精度或调整轴承中心高来实现。传统轴承与转子有接触，摩擦阻力大；另外这种避免多点支承产生超静定约束的方法使得机械结构复杂，调整效果不理想，影响轴承的使用寿命和系统性能。In the rotor system supported by traditional bearings (rolling bearings, sliding bearings), some practical applications use multi-point support, but each point support produces support stiffness and support damping for the rotor. In order to reduce the statically indeterminate constraint of the support on the rotor, it is usually achieved by improving the machining accuracy or adjusting the center height of the bearing. Traditional bearings are in contact with the rotor, and the frictional resistance is large; in addition, this method of avoiding multi-point support to produce super-static constraints makes the mechanical structure complex, and the adjustment effect is not ideal, which affects the service life of the bearing and system performance.
在前述中提到，国内外对于磁悬浮柔性转子系统，一是采取同步振动抑制技术，二是运用现代控制理论或鲁棒控制理论设计控制方案以提高支承阻尼。由于转子在亚临界及超临界状态将产生弯曲变形，不平衡质量所引起的振动与转子弯曲变形状态有关，远比刚性转子系统复杂，因而目前同步振动抑制技术主要针对刚性转子系统。运用现代控制理论或鲁棒控制理论设计合适的控制方案可以改善系统的动态性能，但由于控制参数稳定区域仍然有限，目前尚不能做到大幅度提高系统在亚临界及超临界运行时的支承阻尼。As mentioned above, for the maglev flexible rotor system at home and abroad, one is to adopt synchronous vibration suppression technology, and the other is to use modern control theory or robust control theory to design a control scheme to improve support damping. Since the rotor will produce bending deformation in the subcritical and supercritical state, the vibration caused by the unbalanced mass is related to the bending deformation state of the rotor, which is far more complicated than the rigid rotor system. Therefore, the current synchronous vibration suppression technology is mainly aimed at the rigid rotor system. Using modern control theory or robust control theory to design a suitable control scheme can improve the dynamic performance of the system, but because the control parameter stability area is still limited, it is still not possible to greatly improve the support damping of the system during subcritical and supercritical operation. .
本发明装置中采用的辅助磁悬浮支承组件，不提供支承刚度，仅提供可控附加阻尼，可以避免对低转速下转子的刚体模态振动产生超静定约束，同时通过调整可调电位器灵活改变支承阻尼以满足系统动态性能的要求。这一技术方案简单易行，效果好。旋转机械转子高速、重载、细长发展的要求，使机械的动力学问题日益突出，本发明的研究结果能为解决相关实际问题提供思路与借鉴。The auxiliary magnetic suspension support assembly used in the device of the present invention does not provide support stiffness, but only provides controllable additional damping, which can avoid the ultra-static constraints on the rigid body mode vibration of the rotor at low speeds, and at the same time, it can be flexibly changed by adjusting the adjustable potentiometer Support damping to meet the requirements of system dynamic performance. This technical solution is simple and easy to implement, and has good effect. The requirements of high speed, heavy load and slender development of rotating machinery rotors make the dynamic problems of machinery increasingly prominent. The research results of the present invention can provide ideas and references for solving related practical problems.
附图说明Description of drawings
图1为装置的机械结构总装图。图1中标号名称：1、6为径向磁悬浮轴承组件，2.转子组件，3.推力轴承组件，4.高频电机组件，5.辅助磁悬浮支承组件，7.基座，8.推力保护轴承，9.径向保护轴承，10.径向传感器，11.轴向传感器。Figure 1 is the general assembly drawing of the mechanical structure of the device. Designation of labels in Fig. 1: 1 and 6 are radial magnetic suspension bearing assembly, 2. rotor assembly, 3. thrust bearing assembly, 4. high frequency motor assembly, 5. auxiliary magnetic suspension support assembly, 7. base, 8. thrust protection Bearing, 9. radial protection bearing, 10. radial sensor, 11. axial sensor.
图2为辅助磁悬浮支承组件的控制器电路图。Fig. 2 is a controller circuit diagram of the auxiliary magnetic suspension support assembly.
图3为试验模态分析激振位置示意图。在转子上布置11个激振点与1个拾振点，其中第5点同时为激振点与拾振点。Figure 3 is a schematic diagram of the excitation location for the test modal analysis. Arrange 11 vibration excitation points and 1 vibration pickup point on the rotor, and the fifth point is both the vibration excitation point and the vibration pickup point.
图4为无辅助磁悬浮支承条件下，第9拾振点与激振点间的跨点频响函数。Figure 4 is the cross-point frequency response function between the ninth vibration pickup point and the excitation point under the condition of no auxiliary magnetic suspension support.
图5为无辅助磁悬浮支承条件下，各阶固有频率对应的复振型及阻尼值的模态软件分析结果。Figure 5 shows the modal software analysis results of the complex vibration modes and damping values corresponding to the natural frequencies of each order under the condition of no auxiliary magnetic suspension support.
图6为有辅助磁悬浮支承条件下，第9拾振点与激振点间的跨点频响函数。Fig. 6 is the cross-point frequency response function between the ninth vibration pickup point and the excitation point under the condition of auxiliary magnetic suspension support.
图7为有辅助磁悬浮支承条件下，各阶固有频率对应的复振型及阻尼值的模态软件分析结果。Figure 7 shows the modal software analysis results of the complex vibration modes and damping values corresponding to the natural frequencies of each order under the condition of auxiliary magnetic suspension support.
图8为第1组控制参数、无辅助磁悬浮支承条件下，第9拾振点与激振点间的跨点频响函数。Figure 8 is the cross-point frequency response function between the ninth vibration pickup point and the excitation point under the condition of the first set of control parameters and no auxiliary magnetic suspension support.
图9为第1组控制参数、有辅助磁悬浮支承条件下，第9拾振点与激振点间的跨点频响函数。Figure 9 is the cross-point frequency response function between the ninth vibration pickup point and the excitation point under the condition of the first group of control parameters and the auxiliary magnetic suspension support.
图10为第2组控制参数、无辅助磁悬浮支承条件下，第9拾振点与激振点间的跨点频响函数。Fig. 10 is the cross-point frequency response function between the ninth vibration pickup point and the excitation point under the condition of the second set of control parameters and no auxiliary magnetic suspension support.
图11为第2组控制参数、有辅助磁悬浮支承条件下，第9拾振点与激振点间的跨点频响函数。Figure 11 shows the cross-point frequency response function between the ninth vibration pickup point and the excitation point under the second set of control parameters and the auxiliary magnetic suspension support.
图12为第3组控制参数、无辅助磁悬浮支承条件下，第9拾振点与激振点间的跨点频响函数。Figure 12 is the cross-point frequency response function between the ninth vibration pickup point and the excitation point under the third group of control parameters and without auxiliary magnetic suspension support.
图13为第3组控制参数、有辅助磁悬浮支承条件下，第9拾振点与激振点间的跨点频响函数。Figure 13 is the cross-point frequency response function between the ninth vibration pickup point and the excitation point under the third group of control parameters and the auxiliary magnetic suspension support condition.
具体实施方式Detailed ways
结合图1所示，本发明装置包括径向磁悬浮轴承组件1、6，转子组件2，推力轴承组件3，高频驱动电机组件4，基座7，推力保护轴承8，径向保护轴承9，径向传感器10，轴向传感器11。磁轴承在5个自由度上对转子提供支承作用，对于每一个自由度，由传感器、控制器、功率放大器、磁轴承、转子等组成闭环系统。在合适控制参数作用下，磁轴承对转子产生支承刚度与支承阻尼，决定了整个系统的动态性能。由于控制参数稳定区域有限，支承阻尼的选择受到限制。As shown in FIG. 1, the device of the present invention includes radial magnetic
本发明装置在一般磁悬浮轴承转子系统的基础上增加了辅助磁悬浮支承组件5，该组件在机械结构上与一般径向磁悬浮轴承组件相同。和一般径向磁悬浮轴承组件不同的是，该组件不提供支承刚度，仅提供可控附加阻尼，可以避免对低转速下转子的刚体模态振动产生超静定约束。为了达到这一目的，控制器的结构和参数选择与一般控制器不同。控制器电路图如图2，电路中各器件类型、编号及标称值见表1说明。The device of the present invention adds an auxiliary magnetic
表1.电路图中各器件类型、编号及标号Table 1. The type, number and label of each device in the circuit diagram
该电路结构简单，仅由电阻、电容、二极管、可调电位器和1片集成电路组成。其中由U1A与周围电阻电容组成比例运算放大器；由U1B与周围电阻电容组成微分电路；由U1C与周围电阻组成比例系数为1的运算放大器；由U1D与周围电阻组成跟随电路。The circuit is simple in structure and only consists of resistors, capacitors, diodes, adjustable potentiometers and an integrated circuit. Among them, a proportional operational amplifier is composed of U1A and surrounding resistors and capacitors; a differential circuit is composed of U1B and surrounding resistors and capacitors; an operational amplifier with a proportional coefficient of 1 is composed of U1C and surrounding resistors; a follower circuit is composed of U1D and surrounding resistors.
该电路输出一定大小的比例信号和微分信号。采用该电路作为控制器时，输出比例信号使得辅助支承产生合适的控制电流，形成对转子的正刚度。该正刚度和辅助支承本身所固有的位移负刚度相抵消，这样辅助磁悬浮支承对转子支承刚度为零。而通过调整可调电位器W1，可以改变控制器输出微分信号的大小，从而改变辅助磁悬浮支承对转子的支承阻尼。The circuit outputs a proportional signal and a differential signal of a certain size. When the circuit is used as a controller, the proportional signal is output to make the auxiliary support generate an appropriate control current and form a positive stiffness to the rotor. The positive stiffness cancels out the inherent displacement negative stiffness of the auxiliary support itself, so that the stiffness of the auxiliary magnetic suspension support to the rotor is zero. By adjusting the adjustable potentiometer W1, the magnitude of the differential signal output by the controller can be changed, thereby changing the support damping of the auxiliary magnetic suspension support to the rotor.
下面通过系统试验模态分析和系统高速旋转实验进一步说明本发明的效果。The effects of the present invention will be further illustrated below through system test modal analysis and system high-speed rotation experiments.
系统试验模态分析System Test Modal Analysis
在实际系统运行前，通过对稳定悬浮转子系统进行试验模态分析，可获得系统的固有频率、阻尼及振型等系统的动态性能参数，预知系统的运行状况。Before the actual system operation, through the test modal analysis of the stable suspended rotor system, the dynamic performance parameters of the system such as the natural frequency, damping and mode shape of the system can be obtained, and the operating status of the system can be predicted.
如图3所示的位置分布，在转子上布置11个激振点与1个拾振点，其中第5点同时为激振点与拾振点。采用脉冲锤对激振点实施敲击，采用加速度传感器测得拾振点的响应，采用“北京东方振动和噪声技术研究所”研制的分析软件“智能数据采集和信号分析系统”(版本号：DAST2006)进行模态分析。As shown in the position distribution in Figure 3, 11 vibration excitation points and 1 vibration pickup point are arranged on the rotor, and the fifth point is both the vibration excitation point and the vibration pickup point. The pulse hammer is used to strike the excitation point, the acceleration sensor is used to measure the response of the vibration point, and the analysis software "Intelligent Data Acquisition and Signal Analysis System" (version number: DAST2006) for modal analysis.
系统仅由磁悬浮轴承支承，选取控制器比例系数kpr＝4.16，积分系数Kir＝22.21，微分系数kdr＝7.2×10-3，微分时间系数Tdr＝1.12×10-5s。通过试验模态分析，可得各原点及跨点频响函数。以第9拾振点为例，其跨点频响函数如图4，各阶固有频率对应的复振型及阻尼如图5。系统各阶固有频率如下：The system is only supported by magnetic suspension bearings. The proportional coefficient k pr =4.16, the integral coefficient K ir =22.21, the differential coefficient k dr =7.2×10 -3 , and the differential time coefficient T dr =1.12×10 -5 s are selected. Through the test modal analysis, the frequency response functions of each origin and cross-point can be obtained. Taking the 9th vibration pickup point as an example, the cross-point frequency response function is shown in Figure 4, and the complex mode and damping corresponding to the natural frequencies of each order are shown in Figure 5. The natural frequencies of each order of the system are as follows:
N1＝1169rpm，N2＝3549rpm，N3＝12817rpm，N4＝28682rpmN 1 =1169 rpm, N 2 =3549 rpm, N 3 =12817 rpm, N 4 =28682 rpm
图5中显示相应阻尼比为：Figure 5 shows that the corresponding damping ratio is:
ζ1＝0.02，ζ2＝0.05，ζ3＝0.07，ζ4＝0.03ζ 1 =0.02, ζ 2 =0.05, ζ 3 =0.07, ζ 4 =0.03
在上述基础上，增加辅助磁悬浮支承，其等效阻尼deqr＝5.5 5×104Ns/m。通过试验模态分析，可得各原点及跨点频响函数。仍以第9拾振点为例，跨点频响函数如图6，各阶固有频率对应的复振型及阻尼如图7。系统各阶固有频率如下：On the basis of the above, the auxiliary magnetic suspension support is added, and its equivalent damping d eqr =5.5 5×10 4 Ns/m. Through the test modal analysis, the frequency response functions of each origin and cross-point can be obtained. Still taking the 9th pickup point as an example, the cross-point frequency response function is shown in Figure 6, and the complex mode and damping corresponding to the natural frequencies of each order are shown in Figure 7. The natural frequencies of each order of the system are as follows:
N1＝1128rpm，N2＝5912rpm，N3＝13072rpm，N4＝28765rpmN 1 =1128 rpm, N 2 =5912 rpm, N 3 =13072 rpm, N 4 =28765 rpm
图7中显示相应阻尼比为：Figure 7 shows that the corresponding damping ratio is:
ζ1＝0.02，ζ2＝0.32，ζ3＝0.18，ζ4＝0.04ζ 1 =0.02, ζ 2 =0.32, ζ 3 =0.18, ζ 4 =0.04
由图4、图6以及阻尼比的激振实验测试结果可以看出，增加辅助磁悬浮支承后，400Hz以下的各阶固有模态(包括第1、2阶弯曲固有模态)对应阻尼比有了较大提高，系统频响幅值则有较大程度降低。这说明，增加辅助磁悬浮支承能有效抑制转子的第1、2阶弯曲固有模态振动，有利于转子越过弯曲临界转速。It can be seen from Fig. 4, Fig. 6 and the excitation test results of the damping ratio that after adding the auxiliary magnetic levitation support, the corresponding damping ratio of each order of natural modes below 400Hz (including the first and second bending natural modes) has increased. Greater improvement, the system frequency response amplitude has a greater degree of reduction. This shows that the addition of auxiliary magnetic suspension supports can effectively suppress the first and second order bending natural mode vibrations of the rotor, which is beneficial for the rotor to cross the bending critical speed.
与上述类似，下面给出另外三组不同磁悬浮轴承控制参数下的系统试验模态分析结果。三组试验对应的磁轴承控制参数如表2。Similar to the above, the modal analysis results of the system test under three other groups of different magnetic suspension bearing control parameters are given below. The magnetic bearing control parameters corresponding to the three groups of tests are shown in Table 2.
表2.3组试验模态分析对应的磁轴承控制参数Table 2.3 Magnetic bearing control parameters corresponding to group test modal analysis
在各组试验时，系统先仅由磁轴承支承，可得各原点及跨点频响函数；然后增加辅助磁悬浮支承，其等效阻尼deqr＝5.55×104Ns/m，可得组合支撑下各原点及跨点频响函数。In each group of tests, the system is first supported only by magnetic bearings, and the frequency response functions of each origin and span point can be obtained; then the auxiliary magnetic suspension support is added, and its equivalent damping d eqr = 5.55×10 4 Ns/m, the combined support can be obtained The origin and cross-point frequency response functions are shown below.
仍以第9拾振点为例，图8、图10与图12分别表示无辅助磁悬浮支承的各试验组跨点频响函数，图9、图11与图13分别表示增加辅助磁悬浮支承的各试验组跨点频响函数。由各组试验的频响函数图可以看出，不同磁悬浮轴承控制参数下，增加辅助磁悬浮支承后，400Hz以下的各阶固有模态(包括第1、2阶弯曲固有模态)对应阻尼比均有了较大提高，系统频响幅值则有较大程度降低。这说明，不同磁悬浮轴承控制参数下，增加辅助磁悬浮支承均能有效抑制转子的第1、2阶弯曲固有模态振动，有利于转子越过弯曲临界转速。Still taking the 9th vibration pickup point as an example, Fig. 8, Fig. 10 and Fig. 12 respectively show the cross-point frequency response functions of each test group without auxiliary maglev support, and Fig. 9, Fig. 11 and Fig. Frequency response function across points of the test group. From the frequency response function diagrams of each group of tests, it can be seen that under different control parameters of the magnetic suspension bearing, after adding the auxiliary magnetic suspension support, the corresponding damping ratios of the natural modes below 400Hz (including the first and second bending natural modes) are uniform. With a greater improvement, the system frequency response amplitude has a greater degree of reduction. This shows that under different magnetic suspension bearing control parameters, adding auxiliary magnetic suspension bearings can effectively suppress the first and second order bending natural mode vibrations of the rotor, which is beneficial for the rotor to cross the bending critical speed.
系统高速旋转试验System high speed rotation test
高速试验中，选取磁悬浮轴承控制器比例系数kpr＝4.16，积分系数kir＝22.21，微分系数kdr＝7.2×10-3，微分时间系数Tdr＝1.12×10-5s，辅助磁悬浮支承等效阻尼deqr＝5.55×104Ns/m，将转子稳定悬浮，并通过内置高频电机带动转子由0rpm稳定运行至16000rpm。运转过程中，转子实际转速由光电传感器测量，采用“北京东方振动和噪声技术研究所”研制的分析软件“智能数据采集和信号分析系统”(版本号：DAST2006)对电涡流位移传感器输出进行分析，可得转子振动的三维谱振图如图14。In the high-speed test, the proportional coefficient k pr =4.16 of the magnetic suspension bearing controller, the integral coefficient k ir =22.21, the differential coefficient k dr =7.2×10 -3 , the differential time coefficient T dr =1.12×10 -5 s, and the auxiliary magnetic suspension bearing Equivalent damping d eqr = 5.55×10 4 Ns/m, stably suspends the rotor, and drives the rotor to run stably from 0 rpm to 16000 rpm through the built-in high-frequency motor. During operation, the actual speed of the rotor is measured by the photoelectric sensor, and the analysis software "Intelligent Data Acquisition and Signal Analysis System" (version number: DAST2006) developed by "Beijing Dongfang Vibration and Noise Technology Research Institute" is used to analyze the output of the eddy current displacement sensor , the three-dimensional spectrogram of the rotor vibration can be obtained as shown in Figure 14.
由图14可以看出，系统已稳定越过5900rpm左右的一阶弯曲临界转速和13000rpm左右的二阶弯曲临界转速。It can be seen from Fig. 14 that the system has stably passed the first-order bending critical speed of about 5900 rpm and the second-order bending critical speed of about 13000 rpm.
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