地下装载机电缆卷筒变速箱设计【CYE-1.5型电动地下装载机】【说明书+CAD+UG】
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齿轮噪声的结构传动装置和变速箱的作用Sneana iri KostiTeaching AssistantFaculty of Mechanical Engineering - KraljevoUniversity of KragujevacMilosav OgnjanoviProfessorFaculty of Mechanical Engineering University of Belgrade该齿轮组（变速箱）的噪音取决于（齿轮啮合，轴承运转等等）和箱体的绝热性能和配合形式。当然箱体的轻微振动可以预防或根据改进设计参数。噪音的传播可以通过开展对大功率传动机构结构中的能量释放过程和通过对外壳进行模态测试了解的机械装置这是令人兴奋。在对模态测试结果和噪声测试结果的对比中表明噪声的识别可能性结构来自变速箱的选择。对结果的比较和分析获得了导致齿轮传动单元全部光谱的创造性原因的精确测定。关键词：变速箱外壳，齿轮，振动，噪声，模态测试1、概述：在齿轮传动单元中，噪音是齿轮轮齿的啮合和滚动轴承产生的。冲击声，滑动，滚动等等。安装有弹性结构的机械零件可以通过它的整体结构吸收干扰的能量和传输。内表面这种能源的一部分以噪声的形式排放到周围环境中。另一部分经过衰减转化为热能。图1显示了扰动结构，阻尼和噪声排放到齿轮传动单元。这一进程的主要部分如下：初级声波直接造成齿轮网进入齿轮单元的内部。这些波穿透外壳壁的周围环境。波能量的一部分被减速机的壁所抑制。该传动装置的弹性结构零件（齿轮，轴，外壳等）吸收能量的主要干扰。这种能量以波的形式移动在这些装置中并且它很大的衰减。这部分的外表面喷出的次级声波。干扰能量在机体积振动可激发创造声波的自然能。该外壳的任务是双重的。它可以是初级和次级的绝缘体，同时也是高等声波的放大器。贝尔格莱德机械工程学院。所有版权都认为，噪音的产生机制在变速箱外壳方面的作用。这是文件2的一个延续，其中包含一个模式结构的数值分析和自然振动。该模态方式激发的简历通过调查激发模态振动。其他论文呈现数值预报的噪声（参考文献1和3）的可能性，以及通过几何形状的齿轮对经过处理10和9优化降低变速箱的噪音。变速箱声音强度的服务情况的影响已被Oswald, James, Zakrajsek等在参考文献 6证明了。由于工作速度的提升，影响声音强度的加剧，Oswald, Choy等人开发出了一种为使用变速箱动态仿真的全球性动态模型7，然后一个计算声强的程序8，他们所显示的变速箱噪声谱在一变形方壳一起振动。在那个例子中，然而，对一般噪音标准影响最大的是轮齿在啮合时产生的频率。此外，Sellgren和Akerblom的测试，为“沃尔沃”9和哈里斯等人的需要在11从事实出发观察问题，该外壳是不是有足够的刚性承载啮合齿的影响（传动时的增长误差）。这当然导致扰动和噪声分别地加剧。在参考文献12中，通过使用FEM和BEM，为减少震动，通过对现有外壳简单形状的优化，考虑到四舍五入对上部半径的影响。 本文的目的是确定噪声经过外壳壁所的发出结构，确立外壳壁和噪声发出之间的关系。图1一般结构中的干扰能量转化过程2、通过系统结构的能量传送干扰能量的传递从齿啮合地区到外壳壁是通过齿轮机构，轴和轴承（图2）实现的。能源是通过这些部分的波动被分散的。一个部分能源很大大的损失是通过连接被传送的（接触面）。一个值得注意的问题是能源在通过齿轮到轴上时损失了大部分。这种程度的减轻类似于从轴到轴承或是从轴承到外壳壁。球轴承的接触面这是相当于增加了能源的损失。所以，这指的是从啮合到外壳表面传输的方式有相当大的联系。齿啮合中声的功率Ws和扰动能量Wd之间的比例可以被定义作为扰动能源通过系统结构的传输因子。图2 通过齿轮传动结构的能源传输声音的功率（Ws1）一部分代表的一部分穿过墙壁来的内声能。这种能量的传输通过弹性波穿过墙壁的厚度实现的。声音的衰减是和声音的频率和墙的厚度成正比的。另一部分声辐射的传输从齿轮到的外壳壁的表面闭并以弹性波的形式释放到环境中（Ws2）。这两个部分的声功率造成辐射到周围环境的外墙的弹性结构中声音的被迫波。声音功率的第三部分是自然的自由振动的结果，也就是外壳壁（Ws3）的弹性波。通过使用测量模态阻尼，活跃能模态可以通过有限元方法计算得出。在这种情况下，总动能等于某些干扰的所有形式的总能Ekj。如果这种干扰是由齿轮啮合引起的，振动模态引起的能量是wn=j=1qwnj=j=1qEkjfnj (1) 其中q - 对模态形状的数量 。 fnj - 自然频率。可以将总功率传输因子（传递）分为两部分：st=wswd=wswdwvwv=wvwdwswv=sT1sT2 (2)Wv是总壁振动能量。传输因子的第一部分是振动功率成比例的，T1= Wv Wd，另一个因素是T2声音辐射和墙壁震动成正比。如果周围环境的物质密度2比壁密度小，传递因子T2变小。该外壳壁是用高密度1的铸铁或钢和高弹性波速cw1构成。声学空间提出了较小密度2和波速度为空气的cw2。干扰功率从外壳传输到周围环境的简化的公式是 参考文献14（图2）sT2=wswv=11+14p2cw2p1cw1+p1cw1p2cw22 (3) 通过使用这个公式和密度与钢铁的波动速度和空气的关系，我们可以得到声功率和壁振动功率之间的比值。这就意味着一个非常小的振动能量是以声音能量被传输的，即声功率传输。3、外壳壁振动的模态外壳壁被扰动产生自激，这种扰动通过轴承和轴，来自于齿啮合区，以自然频率振荡的2。自然振荡和弹性变形的波动是复杂的。主体部分激励振荡的过程也是复杂的，以及对振动能量的级别决定的，就是激发能量和释放能源之间所产生的比。为这种原因准确定义导致某些方式的震动，可能的震荡形式及自然振荡频率，通过FEM理论首先做出定义，然后通过在变速箱外壳测试模态完成。结果表明，只有一小部分模型在观察范围可达3000赫兹下产生自激的方式。结果的分析，得到了振荡模态和模态一致，则震荡模态会被激发：变形方向，如在激发行为的最大点发生变化并且，如果相应模态形状的激励频率等于自然频率。然而，振荡模态形状也可以被激励当激励频率与自然频率不一样时。某些方式的复杂机制激发，和数值模态分析和测试模态齿轮传动外壳的结果一样，在2，13有详细的对待方式。4、振动与噪音测量与分析在图1中提出的变速箱已被用作测试项目。振动和噪音的测量与分析已经通过PULSE-system的应用被执行，B&K。传输单元壁的模态测试已通过脉冲激励的手段被提出- 模态锤（图3），并已通过FFT的频谱分析仪分析振动测量措施。一些选择结果在图4中。振动已经在轴承领域被提出，通过使用压电加速度计（图3）冲击力（锤模态的冲击）已应用于外壳壁正交在壁上。图3变速箱住房模态测试在图4，列出相对振动反应在薄壁区域的冲击所造成的影响（图3）。这种反应在约2.4千赫的高密集自然频率时非常集中的。对于轴承领域的影响（薄壁区），该频率响应较低（反应 图 4b）。下一个响应图（c-图4）通过在齿轮齿上的响应获得的。冲击能（干扰能）必须传播经齿轮机构，通过轴和轴承各地，然后它激发外壳壁的自然振动。一个非常高的水平干扰耗能由于非常低的水平的自然振动引起的，但它得到高数值的自然频率的响应。图4外壳模态测试的频率响应1）影响了壳壁顶点，2）在轴承区域受到的影响，3）齿轮的影响.能量传送单元已被放入一镗孔里（图5），使声压可以用于分析。麦克风已放在变速箱的上面，0.5米的距离。图5 现场检测变速箱驱动器已通过电动变速器从次门相室到其相邻镗孔室的旋转速度的手段实现。目前已进行了测量使用激励从缺陷齿轮（图6A），并已达到目的。对于精确确定的外壳壁的影响，当上部（覆盖）的部分已被覆盖时实现测量，（图6B）及它已被关闭的情况（图6c中）。输入轴转速已被更改在每分钟140到1100的范围内。 180下的两个角缺陷对已由驱动齿轮弥补，摆脱哪一个是小，另一个是更大。一个强烈的冲击在这些齿的入口处实现了，并成为加强在齿更大的缺陷。在这些变速箱测试状态的照片。图6变速箱测试：1）齿轮的缺陷，b）无盖室，三）封闭的室4.1模态反应和室振动之间的相关性该室模态响应（图7A）已被在变速箱下部轴承室的轴的两个大洞外侧测量，并已通过一个锤模式获得脉冲激励，这在方式意味着影响轴承的支持点的影响，即点在实际操作轴承引入的干扰点。通过此响应结果与模态分析比较，通过FEM手段执行，可以得出结论，通信时令人满意的，除了一定的偏差。数值和试验模态响应的对比并不是这篇文章的主题 - 在模态试验的室反应谱对在同一测点的振动光谱（图7b中），在与齿轮啮合的缺陷，在转数n=500 /分。振动频率和两个频谱基本相同。因此，它可以得出结论，变速箱室的震动是自然振荡的结果，而振动响应的强度（图7B条）与模态测试强度大致相同（图7A）。其中的差异目前仍然是其他主题，更广泛的考虑。图7齿轮箱室的振动光谱：1）在模态测试，b）在齿轮啮合与缺损第三个结论，可在振动频谱图7的基础上提出，是对指激励频率。在激励频率为为16Hz的旋转500 /分钟，及缺陷驱动齿轮。该振动频率的强度是主要的，但不是最重要。进入弹性结构取决于干扰能源的强度。4.2齿轮箱振动和噪音之间的相关性接下来的实验应确认的论点是：噪音通过变速箱排放到周围环境是室自然振荡的结果。为此，图8给出了两个噪声谱，1指出轴的瑕疵/损失，在500/min（图8A）及另一个指封闭齿轮变速箱具有相同齿轮和相同的旋转速度（图8B项）。通过比较这些频谱，及对比在图7中振动的频谱，可以得出结论如下：*开放式齿轮已取得复杂噪音的频谱。他们主要是初级声音波和已出现较高的谐波由齿的缺陷所影响*室壁以自然频率振荡并发出声波进入周围环境并进入自己室的内部。该声压强表示成m Pa（图8B号）已被测量在封闭变速箱0.5米的距离上。频率水平等于室自然频率强调在这个频谱上。这证实了上述论断，室的自然频率占声音频谱的支配地位。此外，其隔离室应造成声压水平降低，与无室（图8A）款齿轮相对比。然而，这并没有发生。噪音水平得到了显着提高。在室内，有一个相当大的开孔。外壳已充当谐振器框。该传动轴的角速度也速度显着影响的噪声排放水平。速度的改变导致激励频率，吸收干扰能量变化和噪音的水平对外壳的自然频率的改变。5、结论对发射机噪声结构的识别方法论已经研制成功。最初的假说已被证实，能量传输室的自然振荡模态结构和噪声发射频谱的全面关系。在输电系统（齿轮，轴振动，轴承）的齿啮合激励震荡中的干扰能源，以及外壳的自然振动。外壳壁作为基层声音主要绝缘体，为次级和第三级声波透射作为一个传动装置。这些声波可以详细的分离通过对该频谱结构的深入分析。图8 传输单元噪声频谱 1）开启变速箱 2）封闭变参考文献1Sabot, J.: Integrated Vibroacoustic Approach to Compute and Reduce Gear Transmissions Noise,- Proceedings of the International Conference on Power Transmission,Paris,pp.2039-2050,1999.2Ciric Kostic, S., Ognjanovic M.: Excitation of Modal Vibration in Gear Housing Walls, FME Transactions,Vol.34,No.1,pp.21-28,2006.3Houser, D.R., Sorenson, J.D., Harianto, J., ect.: Comparison of Analytical Predictions with Dynamic Noise and Vibration Measurements for a Simple Gearbox, Proc. Intl. Conf. on gears, Munich pp.995-1002,2002.4 Kartik, V., Houser, D.R.: An Investigation of Shaft Dynamic Effects on Gear Vibration andNoise Excitation, SAE Transactions,Vol.112, pp.1737-1746,2003.5Genechte, B.V., Pluymers,B., Vanmaele, C., ect.: On the coupling of Wave Based Models withModally Reduced Finite Element Models for Structural-Acoustic Analysis,Proceedings of ISMA Conf.,Leuven,pp.2383-2403,2006.6Oswald,F.B,Zakrajsek J.J.,Atherton,W.,Lin H.: H.Effect of Operating Conditions on Gearbox Noise, ASME Publication DE-Vol.43-2,pp.669-674,1992.7Choy,F.K.,Ruan,Y.F.,Zakrajsek J.J.,Oswald,F. B.:Modal Simulation of Gearbox Vibration with Experimental Correlation,Journal of Propulsion and Power,Vol.9,No.2,pp.301-306,1993.8Oswald,F.B.,Seybert,A.F.,Wu,T.W.,Atherton, W.:Comparison of Analysis and Experiment for Gearbox Noise,Proceedings of the International Power Transmission and Gearing Conference,Phoenix,Vol.2,pp.675-679,1992.9Sellgren,U.,Akerblom,M.:A Model-based Design Study of gearbox Induced Noise,Proceedings of the International Design Conference-Design2004, Dubrovnik,Croatia,pp.1337-1342,2004.10Chung,C.H.,Steyer,G.ect.:Gear Noise Reduction through Transmission Error ControlandGearBlankDynamicTuning,SAETransactions,Vol.108(2),No.6,pp.2829-2837,1999.11Harris,O.J.,James,B.M.,Woolley,A.M.: Predicting the Effects of Housing Flexibility and Bearing Stiffness on Gear Misalignment and Transmission Noise Using a Fully Coupled Non-Linear Hyperstatic Analysis,Romax Technology Preview Paper-Vehicle Noise&Vibration,88-NVH-AN-0299-1,2000.12Inoue,K.:Shape Optimization of Gearbox Housing for Low Vibration,Proceedings of the International Conference on Power Transmission,Paris,pp2053-2064,1999.13Ciric Kostic,S.,Ognjanovic M.:Gear Disturbance Energy Transmission through the Gear System and Frequency Spectrum,Proceedings of the International Conference Power Transmission, Novi Sad-Serbia,pp.167-172,2006.14Ognjanovic M.:The noise generation in machine systems,Faculty of Mechanical Engineering, University of Belgrade,Monography,Belgrade,1995.(in Serbian)15Ognjanovic M.,Ciric-Kostic S.:Effects of Gear Housing Modal Behaviour at the Noise Emission,- Proceedings of the International Conference on Gears,Munich VDI-Berichte 1904-02,pp.1767-1772,2005. The Noise Structure of Gear Transmission Units and the Role of Gearbox WallsThe noise emission of gear units(gearboxes)depends both on the disturbances(gear meshing, bearing operation, etc)and on the insulating capabilities and modal behavior of the housing. Natural vibrations of the housing walls can be prevented or intensified depending on design parameters. The mechanism of exciting and emission of transmission noise is defined by carrying out the process of propagation of excitation energy through the structure of power transmitters and by modal testing of the housing. The results of vibration and noise testing in comparison to the results of modal testing give the possibility of identification of noise structure for the chosen gearbox. Comparison and analysis of the results obtained lead to precise determination of the causes of creation of the total spectrum of gear transmission units.Keywords: gearbox housing, gears, vibration, noise, modal testing1.INTRODUCTIONIn gear transmission units, sound is generated by excitation in gear teeth meshes and in rolling bearings. Impacts, sliding, rolling, etc. absorb disturbance energy in the elastic structure of machine parts and transmit it through the whole structure. Interior surfaces emit a part of this energy into the surroundings in the form of noise. Another part is converted into heat by damping. Figure 1 shows the structure of disturbance, damping and noise emission in gear transmission units. Some of the main parts of this process are as follows: Primary sound waves are caused directly in gear meshes and emitted into the interior of the gear unit. These waves penetrate through the housing walls into the surroundings. A part of the wave energy is damped in the gear unit walls. The elastic structure of the transmission unit parts (gears, shafts, housing, etc) absorb the dominant part of disturbance energy. This energy moves in the form of waves within these parts and a lot of it is damped. Exterior surfaces of the parts emit secondary sound waves inside the housing. Disturbance energy in the volume of the machine part may excite natural vibrations which can create tertial sound waves. The role of the housing is dual. It can simultaneously be the insulator of primary and secondary sound waves and the amplifier of the tertiary ones. The acoustic emission of gear transmission units has been treated in a number of papers but this oneFaculty of Mechanical Engineering, Belgrade . All rights reserved considers the mechanism of noise generation and the role of the gear box housing. This is a continuation of paper 2, which contains a numerical analysis of modal structure and natural vibrations. The manner of modal excitation is established by investigating excitation of modal vibration. Other papers present the possibility of numerical prediction of noise (papers 1and3), and the reduction of gearbox noise by optimization of geometry of gear pairs has been treated in10and9. The effects of service conditions on sound intensity of the gearbox have been shown by Oswald, James, Zakrajsek and others in paper6.Due to the increasing working speeds, which affects intensification of sound intensity, Oswald, Choy and others have developed a global dynamic model which they use for simulation of gearbox dynamics 7, then a programme for calculation of sound intensity8 and they have shown, on a housing with one deformable side, that its vibrations participate in the spectrum of gearbox noise. In that example, however, the biggest influence on the general level of noise is made by the frequency of gear teeth meshing. Further, Sellgren and Akerblom, in testing for the needs of Volvo9and Harris and others in11 observe the problem by starting from the fact that the housing which is not rigid enough can have significant influence on teeth meshing(increase of transmission error).This certainly leads to intensification of disturbances in meshes and intensification of noise separately. In 12, Inoue has, by using FEM and BEM, and for the purpose of reducing vibrations, worked on optimization of simple shapes of the housing by considering the influence of the radius of rounding of the upper part. This paper aims at determining the structure of noise emitted by the housing walls and establishing the correlation between the housing vibrations and the noise emitted. The noise structure in comparison with disturbance conditions (gear meshing) defines the main phenomenon in the mechanism of noise generationFigure 1.General structure of the process of disturbance energy transformation2.TRANSMISSION OF ENERGY THROUGH THE SYSTEM STRUCTURE Transmission of disturbance energy from the zone of teeth meshing to the housing walls is realized through the gear bodies, shafts and bearings(Fig.2). Energy is distributed through these parts by wave motion. A part of the energy is transmitted through the joints (contiguous surfaces) with considerable losses. A significant amount is lost at the passage of the energy from the gear to the shaft. The degree of reduction is similar at the passage from the shaft to the bearing and from the bearing to the housing. It is considerably increased at the passage of disturbance energy over the contiguous surfaces of the bearing balls. So, this refers to a relatively high number of contacts on the way of transmission from the teeth mesh to the housing surface. The ratio between the sound power Ws and the disturbance power in the teeth meshes Wd can be defined as a factor of transmission (transmissibility) of disturbance energy through the system structure.Figure 2.Transmission of energy through the gear transmission structureOne part of the sound power (Ws1) represents a part of the inside sound energy which comes through the walls. This energy transmission is realized by elastic waves through the wall thickness. The sound reduction is proportional to the sound frequency and the wall thickness.Another part of sound radiation is transmitted from the gear contacts to the surfaces of the housing walls and emitted in the form of elastic waves into the surroundings (Ws2).The first two parts of the sound power create forced waves in the elastic structure of the walls which radiate sound into the surroundings.The third part of the sound power is a result of natural free vibrations, i.e. elastic waves of the housing walls(Ws3).By using the measured modal damping, the modal kinetic energy can be calculated by means of FEM method. In this case, the total kinetic energy is equal to the sum of the kinetic energy Ekj of all modal shapes for a certain disturbance. If this disturbance is caused by gear meshing, the power of modal vibration iswhere q the number of modal shapes and fnj - the natural frequency. It is possible to divide the total power transmission factor (transmissibility) into two parts:Where Wv is the total wall vibration power. The first part of the transmission factor is proportional to the vibration power ratio T1= Wv Wd, and the second one T2 is proportional to the sound radiation in comparison with the wall vibration. The transmission factor T2 is smaller if the material density 2 of the surroundings is smaller in comparison with the wall density. The housing walls are made of cast iron or steel with the high level of density 1 and with the high elastic wave speed cw1. Acoustic space is presented with the much smaller density 2 and the wave speed cw2 for the air.The simplified formula 14 of disturbance power transmission from the walls into the surroundings (Fig.2) isBy using this formula and the values of density and wave speed for steel and for the air, we obtain the ratio between the sound power and the wall vibration power.This means that an extremely small part of vibration energy is transmitted into sound energy, i.e. sound power.3. MODAL VIBRATION OF THE HOUSINGThe housing walls excited by disturbances which, through the bearings and shafts, come from the zone of teeth meshing, oscillate with natural frequencies 2. Elastic deformations at wave motion and natural oscillation are complex. The process of excitation of main shapes of oscillation is also complex, as well as determination of levels of oscillation energy, that is the ratio between the excitation energy and the emitted energy. For the purpose of accurate definition of the causes leading to excitation of certain modes, possible shapes of oscillation and natural frequencies have been firstly defined by applying the FEM method, and then modal testing of the gearbox housing has been performed. The results show that only a small number of modes have been excited in the observed range of up to 3000 Hz. The analysis of the results obtained shows that modal oscillation will be excited if it is coincided by modal:directions of deformations,if the excitation acts at the point of the biggest deformations andif the excitation frequency is equal to the natural frequency of the corresponding modal shape.However, the modal shape of oscillation can also be excited when the excitation frequency is not identical with the natural frequency. The complex mechanism of excitation of certain modes as well as the results of numerical modal analysis and the results of modal testing of the gear transmission housing are treated in a detailed manner in 2,13.4. MEASURING AND ANALYSIS OF VIBRATION AND NOISE The gearbox presented in Fig.1 has been used as the subject of testing. Measuring and analysis of vibrations and noise have been performed by the application of the PULSE-system, B&K. Modal testing of the transmission unit housing has been carried out by means of impulse excitationthe modal hammer(Fig.3),and by measuring of vibrations which has been analyzed by using a FFT frequency analyzer. Some of the chosen results are presented in Figure 4.Vibration has been carried out in the area for bearings, by using a piezoelectric accelerometer(Fig.3). Impact force (impact of the modal hammer) has been applied around the housing walls orthogonal on the wall.Figure 3.Gearbox housing modal testing In Figure 4,diagram(a)presents relative vibration response caused by impact in the area of the thin wall (Fig.3).The response is very intensive for the high natural frequency of about 2.4 kHz. For impact in the area of bearings (area of thick wall),the response for that frequency is less(response diagram4b).The next response diagram(c-Fig.4)is obtained by using impact in the gear tooth. Impact energy (disturbance energy) has to be transmitted through the gear body, through the shafts and across bearings, and then it excites natural vibrations of the housing walls. A very high level of disturbance energy dissipation causes very low level of responded natural vibrations, but it is obtained response of high number of natural frequencies.Figure 4.Frequency response of housing modal testing: a)impact on the top wall of the housing, b)impact on the housing in the area of bearings, c)impact on the gearsThe power transmission unit has been placed into an anehoic chamber (Fig.5) so that acoustic pressure could be used for the analysis. The microphone has been placed above the gearbox , at the distance of 0.5m. The gearbox drive has been realized by means of an electric variator with the rotation speed from the next door room in relation to the anehoic chamber. Measuring has been performed by using the excitation from the gears with defects (Fig.6a) which have been caused on purpose. For the purpose of determining the effects of the housing walls, measuring has been performed when the upper part (cover) of the housing has been removed(Fig.6b)and in the situation when the housing has been closed(Fig.6c).The rotation speed of the input shaft has been varied within the range of 140-1100 min-1.Two defects under the angle of 1800 have been made on the teeth of the driving gear, out of which one is smaller and the other one is bigger. An intense impact is realized at the entrance of these teeth, and it becomes intensified at coming of the tooth with a bigger defect. The photographs of the gearbox in the state of testing are presented in Figure 5Figure 5.Gearbox testing: a) gears with defects, b)housing without cover, c)closed housing 4.1 Correlation between modal response and vibrations of the housingThe modal response of the housing (Fig.7a) has been measured at the lateral side between two bigger holes for the bearings on the lower part of the gearbox housing, and it has been obtained by impulse excitation by means of a modal hammer, which means by the impact at the point where the bearing is supported, i.e. at the point where real disturbance at bearing operation is introduced. By comparing this response with the results of the modal analysis performed by means ofFEM, it can be concluded that correspondence is satisfactory but that there are certain deviations. The comparison between numerical and experimental modal responses is not the subject of this paperit compares the spectrum of housing responses at modal testing with the spectrum of vibrations measured at the same point (Fig.7b) at meshing of gears with defects, at the number of revolutions n=500 o/min. Frequencies of vibrations are mostly the same in both spectrums. Therefore, it can be concluded that the gearbox housing vibrations are the consequence of natural oscillation and that the intensities of vibration responses (Fig.7b) approximately correspond to the intensities of response at modal testing (Fig.7a).The differences which are still present can be the subject of other, broader considerations.Figure 6.Spectrums of gearbox housing vibrations:Figure 6.Spectrums of gearbox housing vibrations: a) at modal testing, b) at meshing of gears with defect The third conclusion which can be made on the basis of the spectrum of vibrations in Figure 7b refers to the frequency of excitation. The frequency of excitation is 16Hz for the rotation speed of 500 o/min and for two defects at the driving gear. The intensity of vibrations for this frequency is important but not the highest one. The intensity of disturbance energy which is entered into the elastic structure depends upon it.4.2 Correlation between gearbox housing vibrations and noiseThe next experiment should confirm the thesis that the noise emitted into the surroundings by the gearbox is the consequence of natural oscillation of the housing. For that purpose, Figure 8 presents two spectrums of noise, one referring to gears with defects/damages, at 500o/min(Fig.8a)and the other one referring to the closed gearbox with the same gears and the same speed of rotation(Fig.8b).By comparing these spectrums with each other and with the spectrums of vibrations in Figure 7,the following can be concluded: A very complex spectrum of noise of open gears has been obtained. They are mainly primary sound waves and their higher harmonics that have arisen by impact of teeth with defectsThe housing walls oscillate with natural frequencies and emit sound waves both into the surroundings and into the interior of the housing itself. The spectrum of acoustic pressure expressed in m Pa (Fig.8b) has been measured at the distance of 0.5 m above the closed gearbox. The levels for the frequencies equal to the natural frequencies of the housing are emphasized in this spectrum. This confirms the above mentioned thesis that natural frequencies of the housing are dominant in the sound spectrum.Besides, the housing with its insulation should lead to reduction of the level of acoustic pressure in comparison with the noise of the gear without the housing (Fig.8a). However, this has not happened. Significantly increased noise levels have been obtained. At the housing, there is a considerably big open hole. The housing has acted as a resonator Box.The angle speed of the driving shaft also significantly influences the level of the noise emitted. The change of speed results in the change of excitation frequency, the change of disturbance energy absorbed and the change of the level of noise for corresponding natural frequencies of the housing.5. CONCLUSIONThe methodology of identification of the power transmitter noise structure has been developed.The initial hypothesis has been confirmed that the modal structure of natural oscillation of the power transmitter housing and the spectrum of the noise emitted are in full correlation.Disturbance energy in teeth meshes excites vibrations of the transmission system (gears, shafts, bearings)as well as natural vibrations of the housing. The housing walls act as an insulator of primary sound waves, as transmission of secondary sound waves and as a generator of tertiary (structural)sound waves. Detailed separation of these sound waves can be performed by a deeper analysis of the structure of the spectrums presented.Figure 7.Spectrums of transmission unit noise: a)open gearbox, b)closed gearboxREFERENCES1Sabot, J.: Integrated Vibroacoustic Approach to Compute and Reduce
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