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# Vibration Testing - Science topic

Explore the latest questions and answers in Vibration Testing, and find Vibration Testing experts.

Questions related to Vibration Testing

How to find the rotational mode shapes using an accelerometer that can measure the translational mode shape only in the modal analysis experiment for a cantilever beam?

Using Python, I would like to convert PSD (G

^{2}/Hz vs Frequency) diagrams to Acceleration vs Time diagrams. Would someone be able to provide some insight into this matter, because I would like to know first whether or not this can be accomplished and if so, how?How can I get the second and third natural frequency of a cantilever beam experimentally??

The problem that I am facing is that I got only the first natural frequency.. I can't get other frequencies..

Hello everyone,

Recently, I've been trying to conduct a vibration test to obtain the quality factor of a specimen. However, the support boundary inevitably affects the results.

I have tried elastic strips and EPE foams (supporting near the nodal points), but the quality factor varies (~10%) from test to test due to the slight difference in support. The quality factor of the specimen is about 10,000.

I wonder if there are practical treatments to minimize the variation of boundary effect so that the obtained quality factor can be more stable (~1%) during different tests?

Thanks a lot.

Hao

What is lattice mode in RAMAN vibration mode?
In general, can you help me about what the lattice mode is in the RAMAN analysis?

I'm working on Sb2S3 thin films. During the Raman analysis, I saw that there are lattice mode vibrational modes. How is it different from symmetric S–Sb–S stretching or symmetric S–Sb–S bending? In General, can you help me about what the lattice mode is in the RAMAN analysis?

I have data from shake table random vibration testing. the results are in terms of power spectral density vs frequency. how to convert psd to accelertion ?

How the vibration testing is used to detect damage in aerospace components using modal parameters?

What will be the accuracy, if only natural frequencies are used in damage detection?

Are natural frequencies upto 10th mode enough to detect damage and estimate severity?

I am doing testing of inertial sensors and I need what kind of processor is used for getting the data into my system and do some related calculation.

Hello

I am trying to perform a sine sweep test using OROS 38 DAQ, Modal analysis software and The modal shop shaker 2100E11. The DAQ software allows to specifying the peak to peak voltage from 0 to 10V and I am using natural air cooling for the shaker so the sine peak force is 220 N.

Only accelerometers are available to get the data and no force sensors are available. Is there any way to measure approximate values of force corresponding to the given peak voltage in the DAQ system?

Please help me out with this.

Thank you

I am running a real-time hybrid simulation by using an electric linear motor. I am trying to perform the free vibration test and in this process, I am measuring the inertia force of the specimen with a load cell. If I do not feed the force in the system it becomes a cyclic test. Under the cyclic test, the system delay is around 2ms and the measured frequency and the input frequency are the same and no damping is observed. However, When I am feeding the measured force into the system then observing a higher delay around 10/12ms and a higher frequency than the input frequency. Can anyone help me with the problem?

Good afternoon, I would like to ask you the method to get acceleration PSD profile from time-acceleration data which was obtained through vibration test.

First of all, I though that the time-acceleration data should be changed to frequency-acceleration data. Therefore, I used FFT.

And then, I though that transformed data to frequency-acceleration data should be squared and divided by their own frequency. Because as seeing Acceleration graph, the parameter of X-axis is frequency(Hz) and that of Y-axis is (g^2/Hz).

So, I wrote matlab code like below;

(Here, THM is the time-acceleration data)

time1=THM(:,1);

t_leng1=length(time1);

dt1=time1(2)-time1(1);

Freq1=(0:t_leng1-1)/dt1/t_leng1;

x=THM(:,2);

xft=fft(x);

xft=xft(1:t_leng1/2+1);

psdx = (abs(xft).^2/Freq1');

psdx = 2*psdx(2:end-1);

figure(1)

plot(time1, x);

hold

xlabel('Time(sec)');

ylabel('Acclearation(g)');

title('Time-domain Accelaration of X axis');

figure(2)

plot(Freq1(2:t_leng1),abs(psdx(2:end-1)))

hold

xlim([2 200])

xlabel('Frequency (Hz)');

ylabel('Accelaration (g)');

title('Frequency-domain Accelaration of X axis');

Freqq=Freq1(2:t_leng1)';

Xresopons=abs(xft(2:t_leng1))/t_leng1*2;

However, when running this code, the errors appear like below;

Error using plot

Vectors must be the same length.

Error in FFT_PSd (line 59)

plot(Freq1(2:t_leng1),abs(psdx(2:end-1)))

Here are my questions.

1. Is the correct method to obtain the acceleration PSD graph?

2. if it is correct, how can I solve this matlab error.

3. If it is incorrect, please let me know the method.

Thank you.

Hi,

Can we use FDD algorithm for forced vibration response (sinusoidal response produced from eccentric mass vibrator)?

Thanks

I am dealing with vibration signals which were acquired from different systems. They are mostly non-stationary and in some cases cyclostationary. What are the less expensive methods for removing noise from the signals? It can be parametric or non-parametric.

During the correlation of vibration measurement and FE frequency response simulation data, the MAC value can be calculated in 2 ways, with and without considering the effect of phase data. These are called real and complex MAC values respectively. The real MAC value calculation is relatively simple and takes into account only the directional vector magnitudes. How can the phase data be included so as to get a more accurate correlation?

Good morning

I would like to perform a ground vibration test (GVT) on a wing model. Can a someone suggest a good material on that (including how to perform it)?

Thank you in advance for your input and time.

I am currently working on my Masterthesis and i am about to make some tests about vibration. I have searched many biological and psycho-physical Articles. But i didn't find any data when it comes to vibration thresholds. I know they are hard to measure, but is there any data, in order to have a guideline?

I am working on vibration test.

my question is how much 'g' level can we accept for a automotive components when it is subjected to forced vibration in terms of 10-40 'g' at various frequencies range of -0-2000 Hz.

eg: input is 10g at 500 Hz, and output vibration level is 200 g at 500 Hz , Is it accepted or on what basis can we have any criteria?

Please let me know the standard procedure or acceptable procedure for validating the vibration test fixtures.

1. My argument is that the only fixtures are to be validated in the test frequency range for any resonance condition and resonance should not occur with in the test frequency range. It is not meant for fixtures along with the units.

2. My QA agency says that when validating fixture means along with the unit. And says that when accelerometer monitoring is on the unit, resonance condition should not occur. If at all occurs, the transmissibility at resonance frequency should not be more than

3.But I am totally not agreeing with my QA.Who is correct?Please explain. Thanks

We showed the explicit analytical expression of the complex wave number of a longitudinal wave in a viscoelastic rod and a flexural wave in a viscoelastic beam. (1)

Frequency response functions (FRFs), such as the mobility, are evaluated using the complex wave number calculated by the proposed formulas. This means the direct problems. (Estimation of the frequency response functions from the given viscoelastic data)

On the other hand, there might be inverse problems. (Estimation of the viscoelastic data from a given frequency response function)

We tried to estimate the viscoelasticity from the mobility data of a beam specimen numerically. We searched for the viscoelasticity (E, eta) to minimize the root mean squared residual error of the absolute mobility. But the derived viscoelasticity had the different frequency dependency from the original one. Then we calculated the mobility by using the derived viscoelasticity. The results agreed with the originally measured one.

It seems that the solution of the inverse problem is not unique. Do you have any idea to solve this kind of inverse problem?

If this type of inverse problem can be solved properly, I think it can be applied to the measurement techniques. We are looking forward to your answer and advice.

(References)

**(1)Propagation of stress waves in viscoelastic rods and plates**

Ryuzo Horiguchi, Yoshiro Oda and Takao Yamaguchi

**, Vol.2, No.1, pp.24-39, 2018.**

*Journal of Technology and Social Science*(2) Stress Waves in Viscoelastic rods beams and plates_ICMEMIS2017.pptx

(Presentation slides)

From my experience, I know that amplitude vs frequency curves are generated as output from measurement points in a shaker test. (Amplitude can be displacement, velocity or acceleration. Acceleration is obtained by default from the accelerometer, which can be internally converted into displacement or velocity by the data acquisition system).

I have seen certain test results where curves are generated for different engine orders and these have been plotted on the Campbell diagram against the engine rpm. I have also observed that for the same frequency, at the same measurement point, the amplitude of acceleration/displacement is different for different engine orders. How is this data extracted from the default amplitude vs frequency data ?

What can be some new and innovative ways to develop a condition monitoring set-up using Vibration Signature of any system (say a coupling, an engine)

Because I could only use the compression type accelerometer which is more sensitive to the zero shift in the impact test, I cannot obtain the accurate signal, and thus cannot obtain the correct velocity, displacement using common integration approach as well. The velocity and displacement shows unrealistic trend to increase continuously without approaching zero.

So my problem is how to process the acceleration correctly with matlab (preferred) or any other software, the detrend function in matlab seems to be unable to handle this problem.

After reading several relevant literatures I found that the discrete wavelet transform (DWT) method and the empirical mode decomposition (EMD) method may be appropriate to be applied. Have you ever met similar problem during the impact tests? Looking forward to hearing your experiences and thanks in advance.

The brief introduction is enclosed as attachment.

The guitar tuning is the initial state on guitar by which all possible states are known. The guitar tuning is a point in tuning space represented by a 6-digit number such as 0 7 5 3 4 5, commonly known as Open D Minor when intonated at D2 pitch level as D A D F A D.

In guitar tuning space every tuning is one step away from every other tuning because for any two tuning vectors there is a difference vector which changes one tuning into another using a single operator.

Given unknown tablature for the song “Happy Birthday” written in an unknown tuning there is one and only tuning which makes sense of the unknown tab, which otherwise seems to be nonsense for every other tuning. Clearly if every possible tuning is used to interrogate the unknown tab, the authentic tuning will be found. Also, if the tuning used to write the tab is D A D F A D, and the tab is interrogated using D A D F# A D, the song will be recognizable but every note falling on the F# string will be off-key and the error can be used to deduce the string should be tuned to F, not F#.

Most common tunings are within 3 strings of each other, but clearly there cannot be more than five steps to go from any tuning to any other.

In affine theory and music copyright law every possible guitar tuning is a simple mathematic derivative of music that any trained musician can make, but clearly there is one and only one authentic guitar tuning used to compose and record guitar. Finding the tuning used to record guitar is possible by interrogating the music, but an additional step is required to determine both tuning and intonation, which surprisingly is only slightly more difficult than find the tuning for unknown tabs.

I see this question as similar to the 3-color map problem in topology because the guitar is finite so there must be a way to learn the tuning from a record. If there is a way to learn the guitar tuning, then there must be a deductive system of logic and implication that can find the best-possible tuning.

Tablature makes more sense as the number of tunings considered increases but clearly considering every possible tuning is not required. In practice, there is certainly a maximum number of tunings that are used in popular music, on a order of magnitude between 10 and 100. The number of useful tunings is clearly highly determined, but how? Why isn't every tuning just as good as any tuning?

If it is possible to learn guitar by auditor surveillance of a record, then can anyone calculate the minimum number of guitar tunings that must be considered? How many times does tablature need to be re-written until the sequence of tabs converges on the best-possible? I realize that there is no way to prove the absolute best but in any specific collection the best is always clear. Can tabs just get better and better in definitely?

I am doing a 1D FE code for a simple expansion chamber, but at the expansion and contraction area jumps, i cannot use my regular shape functions, beacuse they just take length into account, in order to consider area jump i was trying to use an impedance element, since the impedance will change depending on the area, but i am confused on the point that how to give it correctly into code and what all variables i should use while defining the impedance element. Also, i am doing just 1 degree of freedom model i.e, pressure.

thank you

Hi everybody,

I'm going to do dynamic analysis of a vibrating string in ABAQUS. The string is vibrating at a specific frequency and at time t1 boundary conditions are altered in such a way that tension is released while the string is still vibrating. How to model it in ABAQUS?

Thanks for your time

Hi all,

I was simulating Eigenfrequencies of an sign post 3D object. I obtained many frequency values including pure imaginary,for the first natural frequency.

It shows the pure imaginary frequency for rotation (attached image), but I don't understand the meaning of pure imaginary frequencies.

Would anybody help me to understand this obtained result? Or is this wrong result?

I'm using the ERA/OKID algorithm for the identification of a real structure from noisy experimental data (full scale free vibration tests).

The algorithm provides at least two real-valued modes for which damping=100%. How can I interpret this result?

Thank you ,

Anastasia

Hello, I am trying to calculate the element modal strain energy for a simple mass spring system as shown in the attachment. I have calculated the first 3 modal frequencies as [4.4504,12.4698, 18.0194] rad/s. The values i calculated for mode shapes of the first mode are [2.3305, 1.8689, 1.0372], while the second and third modes have a mode shape of [-1.8689,1.0372, 2.3305] and [1.0372, -2.3305,1.8689] respectively. Currently I am unable to match the modal strain energy ratio for each member at the first three modal frequencies, and i need some help understanding the process of calculating the modal strain energy.

Thank You.

Method1: I have the time domain acceleration data at a specific measurement point obtained from hammer testing. I also have the excitation data in time domain from the hammer used (method is roving accelerometer method). I do all the necessary filtering to remove noise and convert both data into frequency domain (FFT) and then take the ratio (output/input), which would be the FRF.

Method 2: After this, I use the shaker to directly compute FRF (using data acquisiton systems) at the same measurement point.

Will the 2 FRF curves using the above methods be similar (not necessarily in amplitude level, but in terms of behavior and peaks)? As per my understanding, it should be the same. Just wanted to make sure.

I have results for repeated vibration test (excitation and response) in time domain. I need to calculate average frf (using cpsd 'Sxy/Sxx') of these test results.

- when i use [abs{(sum(S
_{xy}) / number of test) / (sum(S_{xx}) / number of test)}] i got something different than what i have for the single test, also it looks much noise. - but when i use [{(sum(abs(S
_{xy})) / number of test) / (sum(abs(S_{xx})) / number of test)}] i got better smooth results.

which one of these two methods is true and is there a nice reference to read in deep ?

thank you

We are developing a new evaluation of acoustics properties of wood and wood panels with small dimensions and in the free-free scheme for vibration testing is necessary a microphone to pick up the signal.

- For study of vibration of cylindrical shell i need to find Governing equation by using energy function with Ritz method.
- Use of Hooks law ,strain vectors ,resultant force and moment relation and strain energy and kinetic energy find the equation for frequency.

I know that commonly random vibration analysis is performed as implicit analysis and uses modal superposition method, but can it be performed as explicit analysis? How can we do that? and what are the pros and cons relative to implicit analysis? Any cited sources would be greatly appreaciated

Daniel Bernoulli established the principle that the number of modes of vibration in a system is equal to the degrees of freedom. In his method of hanging weights on a string, I assume without any weights the string has one mode. Clearly, a system can have at most one fundamental. But the problem is how can the string be defined by a fundamental frequency while radiating a complex sinusoidal wave?

I am confused, therefore, about how to count the degrees of freedom of movement that exist in a string detained between two point with constant tension, length, mass, and composition. We assume the parameters allow the string to be subject to harmonic motion when deformed. The question seems a simple geometric problem.

I seems to me that the string has only one mode of vibration which is d

^{2}y/dt^{2}= 0 but then there is a second mode, which you can see with the naked eye, where the shape of the string concatenation is not a circle in cross-section so the axis of the string elliptic rotates.This makes me ask if d

^{2}z/dt^{2}= 0 is also a boundary condition?If we imagine the string is plucked so the deformation vector is only in the direction of the y-axis, then clearly the concatenary does not remain in the x-y plane. But the string does not necessarily assume a circular cross section either. In fact, I cannot see any reason why the cross-section would be circular at all.

Some say the string has n modes or even an infinite number of modes

**, which I think more closely applies to air and electromagnetic fields that have a broad frequency response but not to the detained string defined on the interval 0 to 1 where x is fixed and there must be a fixed point.***at the same time*I have an intuitive answer for the degrees of freedom of string movement that I hope someone can state formally. There must be two modes of string vibration, a dominant and subdominant string mode. Then we have the string tone and overtones as two sets.

I think the string has only 2 possible modes of vibration, perhaps dy/dt and dz/dt, or it may be a radial and rotational mode, but such that the fundamental has only one mode possible at a time, but there is a secondary mode induced by the fundamental on any mode in the frequency expression that is not isochronus because d

^{2}y/dt^{2}is not zero Perhaps there is one degree of freedom that is perpendicular to the string axis and then there is a rotational degree of freedom perpendicular to dy/dt?I don't see the equation for the string concatenation form is any where stated in the literature. Am I missing it?

Is there any way that dy/dt and dz/dt could have different frequencies so that the string radiates a complex sinusoidal function that is the sum of two frequencies? I know that doesn't make sense, but there must be a simple explanation for tones and overtones that is easy to understand and that has been so far overlooked.

I am assuming that the deformation and concatenary forms are states of system and not modes.

The S1 circle is a boundary that classifies pitch in the Z/12Z system as the chromatic circle, which can also be a spiral.

The string concatenary is the shape the string returns to when plucked, which is a standing wave with the boundary condition for a standing wave d

^{2}y/dt^{2}= 0.The string is closed by the octave into a chromatic circle and also open to union with any pitch by adjusting the frequency of the fundamental.

The fundamental of the string has only one degree of freedom and therefore it cannot have more than one mode of vibration.

The string in classic theory is defined on the interval (0, 1) which is an unordered pair but the musical string is on an ordered pair since the frets define the lower segment.

There are 3 sets: observed pitch, fret positions value, string pitch values. The intersection of the pitch values and fret values on the string is a graph that is a restriction R x R onto R, which I think is the basis for both the S1 and d

^{2}y/dt^{2}= 0 boundary condition.So I am wondering if the standing wave boundary is d

^{2}y/dt^{2}= (0, 1).There must be a theorem that say the string can only have one boundary condition if it has only one degree of freedom, right?

The string is the smallest set that contains an image of every set in music so long as the string is at least 12 in size. There is the graph of pitch and position and then there are the tuning function f and the intonation function g.

The tuning function f maps the observed pitch to the string position and the intonation function g maps the string position to the observed pitch. f and g are composable functions that make an identity. This makes the string a homomorphism that is an arrow with a 1 or just a 1.

Now my question concerns how the integration of the boundary condition can be determined. It must be a simple sum if there is a smallest possible Lebesgue measure of 1 step that is countably additive so I think I can conclude the volume of the concatenary is 1. Since everything has to add up to 1, or at least a whole number. Then the boundary conditions are two ways to say the same thing. "S

_{12}= 1".Please help me to formalize the string finite state model. Its a 300 year old puzzle not yet solved.

Two time series data (1. Engine RPM wrt time and 2. Seat Acceleration wrt time) were analyzed using LMS Testlab and LMS AMESim post processing tools. It was observed that the order plot of acceleration wrt rpm obtained from both softwares using the same window type and trend removal option do not match. Further investigations reveals that using the same sampling frequency in AMESim and Testlab produced a result that have similar trends with Testlab result. However, the accelerations amplitudes have significant deviation from Testlab results.

What other parameters should be checked to obtain same results from Testlab and AMESim?

Best Regards,

David

I'm carrying out an experimental modal analysis on the fixture of rotating beams. The rotary beams are attached to a D/C motor and I measured the vibrations at the fixture of the motor. I got peaks at very low frequencies before and after turning on the motor.

Thanks in advance

Bell plates are hand bells made from sheet material having a particular shape. When the plate is struck the resulting vibration and sound continues for a long period of time, even when held by hand.

I am doing a project to investigate and understand the vibration characteristics of bell plates including the development of a computer model to predict the low level of damping achieved. Well anyone can help me out on a better understanding of the bell plate vibrations? Since the shape of my bell plate model is triangular kinda shape with a tang at the edge, any references on developing the FEA computer model by using ANSYS or COMSOL ?

Hello All,

I am simulating vibration of a cantilever in Abaqus using dynamics implicit in two cases; automatic time increment and fixed time increment. As far as I know, results obtained from automatic time increment should always be reliable/correct but surprisingly, in some cases, I can see that by reducing the time increment from the automatic one, I get different results. I am wondering if anybody has dealt with this problem. If this is the case, then we can not simply rely on the results of automatic time increment.

I appreciate your answers.

One of the tests is using FRF(frequency response function),but besides FRF which else tests can be done to get the damping value?I want to perform structural vibration analysis of a washer and so i first have to find out the damping by testing and then use that value for simulation in ansys workbench.

i have to do vibrational analysis of a washer from washing machine.i need damping values for it which i will get from frf(frequency response function) testing.but how do i find damping from frf?

There are many studies for predicting the ground PPV (peak particle velocity) due to blast and its effect on structures. I wonder if there are empirical formulas for frequency distribution of the resulting vibration. I need the frequency distribution to build the time history for the vibration that will be used for more advanced structure analysis.

I have interest in structural vibration especially stationary and moving loads on structures.

The pulse of the vibration within the ring itself is what I'm interested.

My guess is that at least four wave formations occur, these being

(1) a sine curve which originates at the place where the hammer strikes (and its mirror image),

(2) A damping cosine curve of one half the period of the sine curve,

(3) four sub-sine curves at the nodes of the first sine curve space at 45 degrees off the place of the hammer striking in a square of equilibrium in the ring, and

(4) a combination of 1 and 2.

Is there any material on this that presents the answer?

Is there any experimental method to validate the numerical results that obtained from the simulation of rotating blade (flapwise and chordwise vibration) ?

I think that it's related to the geometry of the specimen under test and the frequency of the impact hammer?

I have a cantilever beam with cracks, I found out natural frequencies experimentally.

The natural frequencies of intact cantilever are less than that of same cantilever with cracks.

I made literature and I found some papers show that cracks caused increase in the natural frequencies and other papers show reduction in them ...

I'm confused ,because logically the cracks should reduce the stiffness of the beam.

What is the physical meaning of that increase in the natural frequencies?

Thank you in advance

vibration analysis enveloping and demodulation

I want establish FE model for the disc brake to calculate the vibration responses...

When a coarse set of SIMO FRF is available, the driving point is just one. But when the FRF are estimated very close to each other in spatial domain, and might be very similar and all with PI-constrained phase or one sided imaginary part (apart from little measurement errors), how can we discriminate the best/true driving point FRF?

Can anyone answer me the simplest way to do it.....

My recent perception experiments (The Tension Theory of Pitch Production and Perception) show that the physical dimension (length of a string or diameter of a membrane) is an inherent source of force for the vibrating body. I am now trying to see how tension meters measure the tension of a string since they do not seem to take all the physical dimensions of the string (or other bodies) into account.

Say I hit the free end of cantilever beam hinged at other end with a hammer and find the frf of middle point on the beam.

In the second case I put the shaker at the free end of the same hinged cantilever beam and calculate the frf at the middle point on the beam.

So what differences I may find between the two frfs?

I suppose that both will be same only difference I feel is the impluse excitation by hammer will consists of all the frequencies while the shaker test will contain only the frequencies that the shaker is configured for.

Please correct me If I am wrong.

How do you calculate vibrational frequencies at different temperatures? This can not be done on Gaussian.

I have some problems in understanding criteria of computing MAC.

This my second question on the same topic.

I have an accelerometer on the tip of a wind turbine blade and after exciting it with white noise I got the mode shapes.

I made modal analysis on the ansys apdl and got the mode shapes.

I want to check the pair of modes (experimental and analytic)

What is the simple way to get MAC for both?

How to extract the data?!

As every one knew that mufflers are place in a automobile to reduce the noise of the engine and car body currently I'm working on exhaust mufflers and specially on exhaust absolute mufflers but I couldn't find required theory for it.

I performed modal analysis using Ansys and experimental modal analysis on a wind turbine blade.

The deviation between FE and experimental measurement for higher modes (third mode shape) is so big

274 Hz from ansys and 306 Hz from experiment.

I am familiar with the theory, it's well documented in the manual. What I am looking for is some sort of pdf/links etc. that has some tutorial examples.

I am hoping to analyse vibrations occurring in a MEMS device, and calculations indicate I will need to measure frequency and amplitude between 1 - 150 kHz range. The sensors I have looked at (e.g., accelerometer) top out around 20 kHz. What might one suggest?

-Ryan

I want to use Buckinghum Pi Theorem to relate a similitude of the blade to the prototype blade.

I began as stating the natural frequency is function of ( Young's Modulus, Length of the blade, Cross Section area,Moment of inertia,Density)

Fn (fn,E,L,A,I,Rho) ,then I got the dimensionless pi groups that relate natural frequency of similitude to the prototype !!!

Is this correct procedure or should I derive the natural frequencies equations first ?

I would like to build a shaker running at variable frequencies between 50Hz to 1kHz at 20 micron displacement with a load of 20 gram using PZT piezo stack actuator. I am trying to decide whether to go for a close loop voltage driven or an open loop charge driven system. Which system is preferable in such a situation.

In exhaust pipe lines at the clamps, because of vibrations, the clamps welds fails. Can I replace the clamps by rubber dampers near the chase portion? What kind of rubber material can I use?

I have data which is in .hpf format that I am trying to convert .xlsx file. Can any one help me out?

Basically, I am trying to detect damage on beam structures for a cantilever support. Suppose, I am measuring first natural frequencies from the experiments at several times of undamaged state of structure. I will get different first natural frequencies value in each trial with slight variation. Now how do I take single value of natural frequency ? Which method should I use to get mean value?

If anyone can help me by telling me the source of vibrations (may be low or high frequency) except engine vibrations in aerospace structures?

I recorded the raw response signal and repeated the set of experiments to reduce uncertainty in the measurement by averaging process.

In modal analysis , most of the researchers are following the average of 'n' FRF measurements to reduce the noise. Is this correct?

Modal parameters that need to be extracted from experimental modal analysis. So I want to use RFP method because it deals with multi FRF signal which suited for SIMO and MIMO method. If any of you having good material, kindly share with me.

The signals acquired from both accelerometer and impact hammer through DT9837 DAQ Card using MATLAB. Will it affect the frequency of the structure if data starts to capture after nth sample?

Can anyone tell me that which method of balancing is being used in overhung fans? For rotors, single plane and double plane balancing can be selected by seeing their D/W ratio where "D" is diameter of the rotor and "W" is the width of the rotor. What will be the width of the of Fan or Blower if the same is the method used to select type of balancing?

Is there anybody performing random vibration analysis? Random vibration analysis results do not include phase data, so stress invariants such as von mises equivalent stress can not be directly calculated. Is there any method or tool which you use to estimate Von Mises Equivalent Stress in random vibration analysis?