Report on Vehicle Vibration and Noise

 Report on Vehicle Vibration and Noise

TABLE OF CONTENT

LIST OF FIGURES

Figure 1: Frequency spectra for gearbox vibration signal

Figure 2: Tooth meshing

Figure 3: Gear train Kinematic schematic

1. ABSTRACT

In various ways Engineers are trying to find the way to reduce the noise and vibration of engine by issuing the limit of maximum noise level of several of critical noise resource. Primarily motor operated vehicles are used to create too much noise and we cannot put noise limit all at once to control the same so from last 25 year we are trying to deal with it.

There are some international standard are fixed to calculate the intensity of noise emitted to the environment. In this repost we will discuss about various knocking and vibration details.

2. INTRODUCTION

Regular definition:

Tendency of one object force to anther object into vibration motion is known as forced vibration.

Definition in case of Engine:

The pulse signal is used to determine the RPM which starts and ends at the same time interval of the recording vibration signal for FFT calculation.

An example of vibration measurements of a gearbox is shown in below figure 1. The accelerometers attached on a bearing and it is close to the 27-tooth gear. This gear is rotating at1293 RPM. These data are used to calculate the gear meshing frequency (GMF) and it is equal to 528 Hz. The mating gear has 46 teeth and rotates at 759 RPM. The spectra with frequency in Hz and roader calculated from the recorded acceleration signal. The root mean squares (RMS) of the spectrum components associated with meshing of the 27-toothgear are nearly identical. The transmission units of vehicles do not operate at constant RPM. Their noise or vibration tests must simulate the operational conditions at any rotational speed from a wide working range. The tests of the truck gearbox noise are carried out in the range of the rotation speed from 1000 RPM to 2200 RPM. When testing noise of the gearboxes the rotational speed of the gearbox input shaft increases at a constant rate. The test takes about 40 seconds. During the run-up of the gearbox under the test, the sound pressure signal is recorded by measuring microphones located away from the gearbox housing at a distance of 1 m.

Figure 1: Frequency spectra for gearbox vibration signal

The multispectral is a three-dimensional chart. This chart contains rows of peaks that are similar to a mountain range, because of their continuity in the adjacent spectra. Some rows of peaks are parallel to the RPM axis and therefore do not depend on the rotational speed and the frequency of the peaks is constant. The origin of these peaks consists in the resonance of some parts of the gearbox structure. The other rows of peaks are diverging along radial straight lines that intersect at a virtual point corresponding to zero rotational speed. The frequency of these peaks depends proportionately on the rotational speed and therefore these vibrations are called forced. Only when the rotational speed changes can it is distinguished as resonant or forced vibrations.

3. LITERATURE REVIEW

Diagnostics is understood as identification of machine’s condition/faults on the basis of symptoms. Diagnosis requires a skill in identifying machine’s condition from symptoms. The term diagnosis misunderstood here similarly as in medicine. It is generally thought that vibration is a symptom of gearbox condition. Vibration generated by gearboxes is complicated in its structure but gives a lot of information. We may say that vibration is a signal of gearbox condition. To understand information carried by vibration one have to be conscious/ aware of a relation between factors having influence to vibration and a vibration signal. In order to detect (and diagnosis) an impending failure, a good understanding of the evidence relating to the failure mode and methods of collecting and quantifying the evidence is needed. Although many faults may be easily detectable by physical examination of component, using techniques such as microscopy, Xray, dye penetrates, magnetic rubber, etc., these methods usually cannot be performed without removal of, and in some cases physical damage to, the component. Whilst physical examination techniques still play a critical role during manufacture, assembly and overhaul, they are impractical in an operational large transmission system and other (non-intrusive) fault detection methods need to be employed for routine monitoring purposes. Most modern techniques for gear diagnostics are based on the analysis of vibration signals picked up from the gearbox casing. The common target is to detect the presence and the type of fault at an early stage of development and to monitor its evolution, in order to estimate the machine’s residual life and choose an adequate plan of maintenance. It is well known that the most important components in gear vibration spectraarethe gear meshing frequency (GMF) and its harmonics, together with sidebands duet modulation phenomena. The increment in the number and amplitude of such sidebands may indicate a fault condition. Moreover, the spacing of the sidebands is related to their source. source identification and fault detection from vibration signals associated with items which involve rotational motion such as gears, rotors and shafts, rolling element bearings, journal bearings, flexible couplings, and electrical machines depend upon several factors: (I) the rotational speed of the items, (ii) the background noise and/or vibration level, (iii) the location of the monitoring transducer,(iv) the load sharing characteristics of the item, and v) the dynamic interaction between the item and other items in contact with it.

4. SOURCE OF VIBRATION IN GEAR BOX

The main source of vibration in gear box is gear and bearings. Normally this sound is arrived when two gears are mesh with each other during the time of rotation while power transmitting. 

Firstly, we describe how vibrations arise. A gear train can be in several operating modes. One of these modes is the condition where the gear train is not transmitting the torque and the teeth move within their backlash range. This operation mode produces a gear rattle. This is not the main operating mode of the gearbox. For a description of the main sources of noise it will be assumed that the teeth are in mesh. Each pair of teeth first makes contact at a single point and then they slide over each other in the axial direction. The entrance into the mesh may cause small mechanical impact which excites structural vibration as it encounters defects on the raceway of the rolling bearings. These vibrations are transmitted by the gearbox structure. We can measure the frequency of these pulses, but the structural resonance is not usually identified. There are many resonance frequencies of the gearbox structure, but they are not associated with an individual part of the gearbox or a gear. The mechanical shock is not accompanied by damped oscillations as in the case of defects in rolling bearings. Vibrations of the gearboxes can sufficiently explain a phenomenon which is called a Para-metric excitation. If the individual teeth alternate in mesh then the stiffness of tooth meshing changes according to the number of tooth pairs in mesh. Moreover the point where the tooth is flanks touch moves in the radial direction. Without a doubt, a step change of stiffness occurs when the number of the pairs of teeth in meshes changes. For spur gear one or two pairs of teeth are alternately in mesh. An example which demonstrates the dependence of stiffness on the angle of rotation is in Figure 2. In this figure the dimensionless parameter𝜀𝛼which is called a profile contact ratio is introduced. This approximately indicates the average number of teeth in mesh during meshing cycles. The dependence of the tooth meshing stiffness on the angle of rotation for two values of the profile contact ratio is shown in Figure 2. The teeth of the gears with the profile contact ratio between 1 and 2 is called a low contact ratio (LCR) and teeth of the gears with the profile contact ratio equal to 2 is designated as a high contact ratio (HCR). Without calculation it is obvious that the gear meshing stiffness variations are smaller and therefore cause less parametric excitation of gearbox vibrations. There are various academic mathematical models describing dynamic behaviour of the gearing mesh. One of the simplest is designed for a kinematic scheme in

Figure 2. The equation of rotational motion of two gears which are connected by a spring-damper system is as follows:

Figure 2: Tooth meshing

J1̈𝜑1+b (t) (̇𝜑1−̇𝜑2) +c (t) (𝜑1−𝜑2) =M1

J2̈𝜑2+b (t) (̇𝜑2−̇𝜑1) +c (t) (𝜑2−𝜑1) =M2

Where M1 and M2 are torques (moments of force), 𝜑1and𝜑2are angles of rotation, J1andJ2are moment of inertia, band care damping and stiffness of the tooth contact, respectively. Equations of motion can be modified into the reduced form for modelling the deformation of the teeth.

x=x1+x2=r1𝜑1+r2𝜑

Wherer1andr2are radii of the pitch circle. The reduced form of the equation of motion is as follows

M * RED̈x+b (t) ̇exec (t) x=M1*(1∕r1) −M2*(1∕r)

For equilibrium state between the input and output torque

M2+M1r2/r1=0

The right side of the reduced equation of motion becomes zero. Because the right side of the differential equation is equal to zero the vibrations are not forced or free, but self excited. Coefficients of damping and stiffness are periodic functions of time that excite periodic solution of the no stationary differential equation. The coefficients and the equation solutions have a period TGMF = 1/Fame where Fame is the gear mesh or tooth meshing frequency. The periodical deformation of the spring causes angular vibrations of the shafts with the mounted gears. The angular vibrations and inertia of the gears cause the dynamic force acting between the teeth.

Mechanical power is transmitted from a driving gear to the driven gear by means of the force F Tasting along the line of action. This force is compensated by the same and anti parallel force F Sating at the shaft support point. The simultaneous acting of these forcesFTandFSto the driven gear results in torque, see Figure 3. The teeth contact stiffness is not equal toe constant value, but it is oscillating in synchronism with tooth meshing frequency due to the oscillation of the number of the tooth pairs in contact and moving the teeth contact point along the tooth flank. The oscillation of the tooth contact stiffness causes the self excited angular vibration of the driven gear, which results in the time varying forcesFTandFS. The force acting at the shaft support is dynamic as well and excites vibration of the gearbox housing and consequently emitting noise.

The contact of engaged teeth is not a point but a straight line, see figure 2. The distribution of forces along this line may be non-uniform. Analysis of the solution of the reduced differential equation is an academic issue, the test engineer or designer prefers an experiment. There could be an entire book devoted to experiments with measuring vibrations of the gearbox housing, noise radiated by the gearbox and angular vibrations of the gearbox shafts. The relationship between the angular vibrations and the linear vibrations of the gear train will be illustrated by the example in Figure 3.4. Both signals which are acceleration have the same frequency range up to one hundred times the speed of rotation. The linear acceleration is measured by an accelerometer which is placed on the bearing of the shaft where is the gear is mounted and is orientated in perpendicular direction to the shaft axis. Angular acceleration is measured by an increment rotary encoder (IRC). A description of the method of calculating the angular acceleration and how to assign records to the selected gear is included in this book.

Figure 3: Gear train Kinematic schematic

5. VIBRATION ANALYSIS OF VEHICLE

Gears are important element in a variety of industrial applications such as machine tool and gearboxes. An unexpected failure of the gear may cause significant economic losses. For that reason, fault diagnosis in gears has been the subject of intensive research. Vibration signal analysis has been widely used in the fault detection of rotation machinery. The vibration signal of a gearbox carries the signature of the fault in the gears, and early fault detection of the gearbox is possible by analyzing the vibration signal using different signal processing techniques. In this paper, a review is made of some current vibration analysis techniques used for condition monitoring in gear fault.

Minimum Quantity Lubrication(M.Q.L)

This was the live project which we perform in Lucent Energy Pvt. Ltd. Nagpur, MQL system reduce the coolant quantity which reduce the cost of the process as well as reduce surface roughness.

6. CONCLUSION

Gearbox vibration signals are usually periodic and noisy. Time-frequency domain average technique successfully removes the noise from the signal and captures the dynamics of one period of the signals. The study of vibrational analysis have been done in this report.

7. REFERENCES

[1] Munro, R.G. and Yildrim, N. (1994) Some measurements of static and dynamic transmission errors of spur gears. Proceedings of the International Gearing Conference 1994, University of Newcastle upon Tyne.

[2] Chung, Ch-H., Steyer, G., Abe, T.et al.(1999) Gear Noise Reduction Through Transmission Error Control andGear Blank Dynamic Tuning, SAE Paper 1999-01-1766.

[3] T ̊uma, J. (2006) Simple gear set transmission error measurements. Proceedings of Thirteenth InternationalCongress on Sound and Vibration (ICSV13), July 2–6, 2006, 8 p.

[4] Henriksson, M. and P ̈arssinen, M. (2003) Comparison of Gear Noise and Dynamic Transmission ErrorMeasurements. Proceedings of Tenth International Congress on Sound and Vibration (ICSV10), Stockholm,pp. 4005–4012. (Paper 443