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对六连杆机械压力机碰撞噪声的调查报告罗彦鑫，杜如旭a中国，400044,重庆大学，国家重点机械连杆实验室 B香港，新界，沙田，中国香港大学的精密工程研究所摘要:传统机械压力机由曲柄和滑块构成，是一种最常用的冲压机。但它不能满足深拉操作，其中长时间停留在BDC是可取的，以避免裂纹或皱纹。这种动机的机械压力机是一六杆机构。然而,工作引入新机制产生噪声。本文提出一种研究机械压力机的噪音的分类学。调查噪声的根源采用噪声特征分析、刚体动力学分析和有限元法(FEM)。发现噪声是由齿轮的排序造成的。最后提出一种改进的设计，给出了一些建议,以减少噪音。 关键词：六杆机械压力机，特征分析，冲击噪声，有限元法（FEM）1.介绍常规机械压力机包括裂纹和滑块是一种最常用的冲压。其轨迹是不可控的下死点（BDC），因此它不能满足多种需求1。例如，在BDC长的停留时间是可取的，以避免裂纹或皱纹深冲操作2。这个机械压力机的动机被设计成五，六，九杆结构3，4。一般而言，在机械金属成型压机的设计，设计师大多关注的是运动学5。很少有人研究了动力学的动态，但它是非常重要的性能6当设计和建造一个商业印刷机时，那就有必要探讨其性能和持续改进。在这项研究中，一个六连杆（包括四连杆和曲柄滑块机构）已经建立一个通过300吨的最大负载能力的商业机械压力机。设计工程师们仔细检查了压力机的运动学和制造装配,均在设计规范内完成，因此，按工作正常。不过压力机产生噪音和不加载。声音强度大于90分贝，这会导致各种问题，如机器的可靠性和操作者的安全性。传动系统的机构示于图1（a），其CAD模型示于图1（b）。它包括七个部分：一马达（图中未示出），连接到飞轮通过高速轴，减速齿轮组（其中包括高速齿轮和所述低速齿轮），其控制所述接合的联接器的齿轮组和曲柄，曲柄滑块机构，和一个连接到滑块上的四杆机构。其中，四杆机构是关键，因为它压力机的动态性的理论。L1=45mm L2= 165mm L3=208 mm L4= 260mm L5=80010000mm e= 80mm图1（一）插图的六杆联动机制，该机制（二）物理模型本文旨在找出噪音的根源（次）。所提出的研究将调查噪音的各个方面，并提供完善的解决方案。本文的其余部分安排如下。在第2节中，噪声的特征分析。在第3节中，系统的动力学研究了刚体动力学分析和有限元分析。在第4节，一个改进的设计建议，以减少冲击噪音。最后，结论在第5节给出。2.噪声信号的特征分析2.1.实验装置正如前面提到的，噪声是设计的主要问题。第一步是噪声信号的分析。使用麦克风紧密放置在压力机的声音信号进行了测量。图2示出的实验装置，该装置主要包括一个麦克风（制造商：的BrelKJR，型号：型号4191），信号放大器（百灵达，型号：XENYX802），信号采集系统（声卡）和一台PC电脑。麦克风的频率范围为3.15赫兹40千赫。图2实验装置2.2.噪声信号和特征分析在实验过程中，压力机的运行速度设定为每分钟（SPM）（因此，其工作频率为1.67赫兹）100行程，无负荷施加和采样频率为48千赫。图3（a）示出了一个典型的噪声信号。图3（b）是放大在三个周期，从中可以看出，每个周期包括两个大峰，AB。 图3一个典型的噪声信号 图4噪声信号的频谱 图5噪声信号的包络谱 图6噪声信号的能量的时间 - 频率谱图4示出在图的信号的FFT频谱。 从图3中可以看出，该噪音信号具有三个主要组成部分，分别是157赫兹，781赫兹和1255赫兹。在157赫兹的组分是对应于齿轮啮合频率和它的能量是相当小的。在781赫兹的分量具有最大的振幅和负责的噪音。这将是研究的重点。图5显示了包络谱。从图中可以看出，主要的频率为23赫兹，这是峰值的发。这表明该噪音是由一系列的影响每个工作循环生频。图6示出了信号的时间 - 频率频谱。从图中，可以看出，沿781赫兹一系列峰的出现，以及它们的幅度变化时。然而，主要有两种高峰（图3对应的峰A和B）出现在每一个时期。此外，山顶A的振幅比B峰的大。基于以上的研究中，可以看出是（a）有对应于A和B分别在每个周期中两个大的影响，（b）该噪声的主要频率为780赫兹和（c）的冲击频率为23赫兹。有必要分析驱动系统的动力学找到噪声的根源。3.动态分析 3.1.影响力分析为了研究噪声的来源，系统的动态模型是必要的进一步机械调查，这可能反映真实的工作条件，应进行准确的负荷分析构造。在本研究中，压力机的动态分析使用商业软件RecurDyn的系统进行。该键之间的所有关节力由该模拟获得的，其中的齿轮之间的接触力引起了我们的兴趣。图7示出了齿轮之间的接触力（黑色），并且其分化（橙色）。检查力量的分化，可以看出，接触力迅速改变两倍方向，在A和B，当冲头向上移动。这很容易让人产生两个大的冲击。因此，一个响亮的噪音就会产生。图7齿轮之间的冲击力为了进一步调查噪声信号中的频率成分，但是，有限元分析（FEA）是必要。3.2.机械部件的自然频率振动通常是由结构的振动引起的这是众所周知的，因此，有必要找到该结构的固有频率。该分析方法可用于发现固有振动频率。然而，这是因为假设的不准确。在实用中，FEA是一个变量的方法来找到精确解微分方程用于验证结构的振动。有用于有限元分析等的Abaqus，ANSYS，Nastran软件，等一些商业软件7，8。在这个研究中，我们使用的Abaqus找到如表1中所示的机械部件的自然频。此外，该高速轴的第一模式形状是扭转，这是在转矩相同的方向。据认为，这种模式是负责对23赫兹的重复频率。和高速齿轮和低速齿轮的第二个自然第四固有频率接近的主要频率的噪音。此外，相应的模式形状是弯曲的齿轮齿。因此，齿轮的齿应该进一步调查。图8（a）示出低速齿轮的有限元分析模型。负载被施加到齿中的一个。所施加的负载量为1个单位的归一化力，在频率范围为1赫兹至1500赫兹。图8（b）示出了频率响应的结果。从图中可以看出，该齿轮具有在480Hz的，740H频率是负责巨响。有限元分析的频率（740赫兹）和实际噪声频率（780赫兹）之z分别三个主要的频率，和1350赫兹。最大频率分量是在740赫兹。据认为，这一间的差异可以归因于该有限元分析模型的简化。在结论中，我们相信，咋是由齿轮的碰撞产生的，而噪声可以通过消除齿轮间隙的减小。 表1模型的主要组成部分模态频率（HZ）模型 123456高速轴24.229.754.8285.3581.01035.4高速齿轮67.3360.5466.1753.0892.91189.0低速齿轮593.1619.01147.81205.31278.01477.0上链接67.074.780.5102.6185.2282.44.一种改进设计图8（a）有限元的齿轮小齿轮的模型（二）频率响应据较早提出的分析结果，联系的长度不应该改变，以保持设计的轨迹。另外，飞轮的惯性和高速的刚性轴等，也可以进行微调，以降低噪音。然而，这些解决方案是因为机械部件的不具有有效的小室用于改进设计，由于强度的限制。这是一个选项，通过消除齿轮之间的间隙，提高了设计。所提出的设计示于图9有两个齿轮安装在高速轴上，以消除齿轮间隙。在这个设计中，2组齿轮对的被利用和一个螺旋弹簧将安装在齿轮之间，以消除齿轮间隙。所述第一组齿轮对将扭矩传递到用于转矩的情况下，曲柄轴在顺时针方向转动。当扭矩方向改变时，第二组齿轮对的工作原理。因此，它认为，该设计将有效地降低噪音。然而，需要一种新的设计模型在未来的动力学分析。图9该齿轮对的一个改进的设计以消除齿轮间隙结论本文对一个六杆机械压力机的噪声提出研究。根据上面的讨论，以下结论可以得出：压力机的机械噪声包含了许多对应于压力机的各种部件的固有频率成分。当冲击发生时会产生噪声。噪声在操作过程中产生影响，变成变速四杆机构。变速产生变力，引起对方的齿轮冲击。它是一个选项，以避免对齿轮副之间的间隙，以减少噪音。签名分析，机械动力学分析和有限元法的组合是分析机器故障的根本原因的有效方法。除了上面给出的应用，它可用于涉及机械运动的其它许多应用。参考文献1Du R.Guo WZ.The Design of a new metal forming press with controllable mechanism.JMech Design 2003;125:582-592;2Yan HS, chen WR.A variable input speed approach for improving the output motion characteristics of watt-type presses.Int J Mach Tool Manuf 2000；42:675-690.3Tso PL,liang KC.A nine-bar linkage for mechanical forming presses.J Mach Tool Manuf 2002;42:139-145.4Meng CF,Zhang C,Lu YH,Shen ZG.Optimal design and control of a novel press with an extra motor. Mech Mach Theory 2004;39:11-818.5He K,Li WM,Du R.Dynamic modelling with kineto-static method and experiment validation of a novel controllable mechanical metal forming press.Int J Manuf Research 2006;1:354-378.6Su.S,Du R.Signature analysis of mechanical watch movements.Mech System Signal Process 2007;21:3189-3200.7Khelladi S,Kouidri S,Bakir F,Rey R.Predicting tonal noise from a high rotational speed centrifugal fan J Sound Vibration 2008;313:113-133.8Junhong Z,Jun H.CAE process to simulate and optimise engine noise and vibration.Signal Process 2006;20:1400-14097Procedia Engineering 29 (2012) 1486 14911877-7058 2011 Published by Elsevier Ltd.doi:10.1016/eng.2012.01.160Available online at 2012 International Workshop on Information and Electronics Engineering (IWIEE) An Investigation on the Impact Noise of a Six-Bar Linkage Mechanical Press Yanxin Luoa*, Ruxu Duba State Key Lab of Mechanical Transmission, Chongqing University, Chongqing, 400044, China. b The Institute of Precision Engineering, The Chinese University of Hong Kong, Shatin, N. T., Hong Kong Abstract Conventional mechanical press consisted of crack and slider is one of the most commonly used for stamping. But it cannot satisfy deep drawing operations, in which long dwelling time in the BDC is desirable to avoid crack or wrinkle. This motives the design of a six-bar linkage for the mechanical press. However, the working noise is introduced by the new mechanism. This paper presents a systematics study on the noise of the mechanical press. A combination of noise signature analysis, rigid body dynamics analysis and finite element method (FEM) are adopted to investigate the root cause of the noise. It is found that the noise is caused by the collation of the gears. Finally, an improved design is then proposed and some suggestions are given to reduce the noise. 2011 Published by Elsevier Ltd. Selection and/or peer-review under responsibility of Harbin University of Science and Technology Keywords: Six-bar mechanical press, Signature analysis, Impact noise, Finite element method (FEM) 1. Introduction Conventional mechanical press consisted of crack and slider is one of the most commonly used for stamping. Its trajectory is not controllable at the bottom dead centre (BDC), and hence it cannot satisfy the diverse needs 1. For example, long dwelling time in the BDC is desirable to avoid crack or wrinkle for deep drawing operations 2. This motives the design of the five-, six-, nine- bar linkage for the * Corresponding author. Tel.: +86-23-65106999; fax: +86-23-65105795 E-mail address: yxluo 1487Yanxin Luo and Ruxu Du / Procedia Engineering 29 (2012) 1486 1491mechanical press 3, 4. In general, in the design of a mechanical metal forming press, designers concern mostly the kinematics 5. Few have studied the dynamics of the press, though it is very important its performance 6. When a commercial press is designed and built, its necessary to investigate its performance and make continuous improvements. In this research, a six-bar linkage (including a four-bar linkage and a crank-slider mechanism) has been adopted to build a commercial mechanical press with the capability of max loading of 300ton. Design engineers had carefully checked the kinematics of the press and the manufacturing and assemblies were done within the design specification and therefore the press works fine. Though, the press generates loud noise with and without loading. The sound intensity is higher than 90db, which causes various concerns, such as the reliability of the machine and the safety of the operator. The mechanism of the driveline is illustrated in Fig. 1 (a), and its CAD model is shown in Figure 1(b). It consists of seven parts: a motor (not shown in the figure) that connects to the flywheel through the high speed shaft, a reduction gear set (which includes the high speed gear and the low speed gear), a coupler that controls the engagement of the gear set and the crank, a crank-slider mechanism, and a four bar mechanism that connects to the slider. Among them, the four bar mechanism is the key as it doctrines the dynamic performance of the press. L1=45mm L2= 165mm L3=208mm L4= 260mm L5=80010000mm e=80mmFig. 1. (a) Illustration of the six-bar linkage mechanism; (b) the physical model of the mechanism This paper aims to find the root cause(s) of the noise. The presented research will investigate the noise from various aspects and provide improvement solutions. The rest of this paper is organized as follows. The signature analysis of noise is presented in Section 2. In Section 3, the dynamics of the system is studied by rigid-body dynamics analysis and FEM. In Section 4, an improved design is proposed to reduce the impact noise. Finally, conclusions are given in Section 5. 2. Signature analysis of noise signal 2.1. Experimental setup As mentioned earlier, the noise is the major concern of the design. The first step is to analyze the noise signal. The sound signal was measured using a microphone placed closely to the press. Figure 2 shows the experimental setup, the main apparatuses include a microphone (Manufacturer: Brel & Kjr, Model: Type 4191), a signal amplifier (Behringer, Model: XENYX802), a signal acquisition system (a sound card) and a PC computer. The frequency range of the microphone is 3.15 Hz 40 KHz.1488 Yanxin Luo and Ruxu Du / Procedia Engineering 29 (2012) 1486 1491Fig. 2. Experimental Setup 2.2. Noise signal and signature analysis During the experiment, the operating speed of the press is set at 100 stroke per minute (SPM) (thus, the operating frequency is 1.67 Hz), no loading was applied and the sampling frequency was 48 KHz. Fig. 3(a) shows a typical noise signal. Fig. 3(b) is a zoom-in of three cycles, from which it is seen that each period consists of two large peaks, A and B. 00.81.82-1-0.500.51Time /s0500100015002000250005001000150020002500Frequency /HzAmplitude781Hz1255Hz157HzFig.3. A typical noise signal Fig. 4. Spectrum of the noise signal 010203040506070809010002004006008001000Frequency /HzAmplitude23HzFig. 5. Envelope spectrum of the noise signal Fig. 6. Energy-time-frequency spectrum of the noise signal Fig. 4 shows the FFT spectrum of the signal in Fig. 3. From the figure, it is seen that the noise signal has three main components at 157 Hz, 781 Hz and 1,255 Hz respectively. The component at 157 Hz iscorrespondent to the gear meshing frequency and its energy is rather small. The component at the 781 Hz has the largest amplitude and is responsible for the noise. It will be the focus of the study. Fig. 5 shows the envelope spectrum. From the figure, it is seen that the main frequency is at 23 Hz,which is the occurring frequency of the peaks. This indicates that the noise is caused by a series of ABABAB1489Yanxin Luo and Ruxu Du / Procedia Engineering 29 (2012) 1486 1491impacts in each working cycle. Fig. 6 shows the time-frequency spectrum of the signal. From the figure, it is seen that along the 781 Hz a series of peaks appear, and their amplitudes changes from time to time. However, there are mainly two high peak (corresponding to Peaks A and B in Fig. 3) appeared in each period. Moreover, the amplitude of Peak A is larger that of Peak B. Based on the study above, it can be seen that (a) there are two large impacts in each cycle corresponding to A and B respectively, (b) the main frequency of the noise is 780 Hz; and (c) the impact frequency is 23 Hz. It is necessary to analyze the dynamics of the drive system to find the root cause of the noise. 3. Dynamics analysis 3.1. Impact force analysis To investigate the sources of the noise, the dynamic model of the system is necessary for further investigation of mechanical, which may reflect real working conditions, should be constructed for an accurate loading analysis. In this research, the dynamic analysis of the press is carried using a commercial software system RecurDyn. All joint force between the linkages are obtained by this simulation, among which the contact force between the gears is attracted our interesting. Fig. 7 shows the contact force (in black) between the gears, and its differentiation (in orange). Examining the differentiation of the force, it is seen that the contact force quickly changes its direction twice, in A and B, when the punch is moving up. This is very likely to create two big shocks. Consequently, a loud noise will be generated. Differetiation (N/s)Punch position (mm)Gear Force(N)Fig.7. Impact force between the gears In order to further investigate the frequency components in the noise signal, though, Finite Element Analysis (FEA) is necessary. 3.2. Natural frequencies of the mechanical components Its well known that the vibration is usually caused by the vibration of structure, and therefore, its necessary to find the natural frequency of the structure. The analytical method can be used to find the natural frequency. However, it is not precise because of the assumptions. In practical, FEA is a variable method to find the precisely solution of differential equations for verifying the vibration of a structure. There is some commercial software for FEA such as Abaqus, Ansys, Nastran, and etc7, 8. In this research, we used Abaqus to find the natural frequencies of the mechanical components as shown in Table 1. Also, the first mode shape of the high speed shaft is torsion, which is in the same direction of the torque. It is believed that this mode is responsible for repetitive frequency of 23 Hz. And the fourth ABBA1490 Yanxin Luo and Ruxu Du / Procedia Engineering 29 (2012) 1486 1491natural frequency of high speed gear and the second natural of low speed gear are close to the mainly frequency of the noise. Moreover, the corresponding mode shape is bending of the gear tooth. Therefore, the gear tooth should be further investigated. Fig. 8(a) shows the FEA model of the low speed gear. The loading is applied to one of the teeth. The applied loading is a normalized force of 1 unit, in the frequency range from 1 Hz to 1,500 Hz. Fig. 8(b) shows results of the frequency response. From the figure, it is seen that the gear has three main frequencies at 480Hz, 740Hz, and 1,350 Hz respectively. The largest frequency component is at 740 Hz. It is believe that this frequency is responsible for the loud noise. The discrepancies between the FEA frequency (740 Hz) and actual noise frequency (780 Hz) may be attributed to the simplification of the FEA model. In conclusions, we believe the loud noise is generated by the collision of the gears, and the noise can be reduced by eliminating of the gear clearance. Table 1. Modal frequencies of the main components of the press (Hz) Fig. 8 (a) FEM model of the gear pinion (b) frequency response 4. An improved design Based on the analysis result presented earlier, the length of the linkages should not be change as to keep the trajectory of the design. Also, the inertia of the flywheel and the stiffness of the high speed shaft etc., can also be fine-tuned to reduce the noise. However, these solution is not effective because of the mechanical components has small room for improving the design due to the strength constraints. Its an option to improve the design by eliminating the clearance between the gears. The proposed design is shown in Fig. 9. There are two gears mounted on the high speed shaft to eliminate the gear clearance. In this design, two set of gear pair are utilized and one spiral spring will mounted between the gears to eliminate the gear clearance. The first set of gear pair will transmit the torque to the crank shaft for the case of the torque is in clockwise direction. When the torque changes its direction, the second set of gear pair works. Therefore, its believed that the design will reduce the noise effectively. However, the dynamics analysis is needed for the new design model in the future. Mode 1st 2nd 3rd 4th 5th 6th High speed shaft 24.2 29.7 54.8 285.3 581.0 1035.4 High speed gear 67.3 360.5 466.1 753.0 892.9 1189.0 Low speed gear 593.1 619.0 1147.8 1205.3 1278.0 1477.0 Upper linkage 67.0 74.7 80.5 102.6 185.2 282.4 1491Yanxin Luo and Ruxu Du / Procedia Engineering 29 (2012) 1486 1491Fig. 9. An improved design of the gear pair for eliminating the gear clearance 5. Conclusions This paper presents a study on the noise of a six-bar mechanical press. Based on the discussions above, following conclusions can be drawn: The mechanical noise of the press contains a number of components corresponding to the natural frequencies of various components of the press. The
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