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For designing a component for vibration, the mass and materials are fixed for the problem. The component will experience a vibration in the 100-2500 Hz range. Now as the material and geometry are fixed, the mass and modules of elasticity is fixed. But we can increase the stiffness by increasing the fixed boundary conditions. In this way, the first natural frequency may lie above 2500Hz. Is this a good practice? We are bonding the component to the fixture to avoid dynamic coupling
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Dear Sunil, it is absolutely impossible to provide any relevant advice if you don't provide more details on your current design! (masses, interfaces, materials...)
Having a first frequency above 2.5kHz seems totally unachievable unless you are dealing with micro-technologies (objects <few cm overall size)
You might better use more flexible but properly damped structural elements (fibre composites better than metallic alloys or titanium)
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Compared to the old-fashioned and currently used emulsion type explosives, the explosive filling of the tunnel face with bulk charging provides better and higher quality vibration values. if you are drilling in the tunnel face with the Mwd (measurement while drilling) featured jumbo. Because with the mwd-capable machine, heterogeneous drilling is performed in the formation whose face surface is uneven and the drilling lengths are different. Therefore, a homogeneous charge in a heterogeneous face with an emulsion-type explosive of constant kilogram will be difficult. Therefore, I think that more stable vibration data will be obtained with bulk charging. What is your opinion?
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I obtained an empirical formula with 95% accuracy rate with emulsion type explosive. Thank you very much for your esteemed reply. I think I can get more accurate results with bulk charging. Thank you very much for your interest, Mr. Signh.
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Hello everyone,
I have a problem calculating the modal assurance criterion (MAC) of a experimental mode shape and a FEM mode shape. I can calculate the AutoMAC for each mode shape, for which the values are all correct. Both matrices show that the same mode shape gets a value of 1, while the rest is near 0.
However if I now apply the same formula to the normal MAC nothing seems right. The sensors for the experimental mode shape can measure displacement in one DOF. So at each node the displacement is a complex value in the direction of one of the local X, Y or Z-axis. The FEM mode shape contains real values at each node and the displacement can occur in all 3 DOFs.
I hope someone can help me resolve this problem.
Thanks in advance!
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you need to ensure that the mode shapes being compared are in the same format. Here's a suggested approach:
  1. Experimental Mode Shape: If the experimental mode shape has complex displacement values, representing motion along a single DOF, you can convert it to a real-valued format. For example, you can consider the magnitude of the complex displacement at each node as the mode shape value. This will result in a real-valued mode shape that represents motion in a specific DOF.
  2. FEM Mode Shape: Since the FEM mode shape already contains real values representing motion in all three DOFs, no additional conversion is required.
Once you have both mode shapes in the same format (real-valued), you can calculate the MAC using the standard formula. The MAC formula involves comparing corresponding displacement values at each node between the two mode shapes.
Remember to normalize the mode shapes before applying the MAC formula. Normalization helps in removing any scaling effects and ensures a fair comparison between the mode shapes.
I hope this explanation helps you resolve the issue and accurately calculate the MAC between your experimental and FEM mode shapes. If you have any further questions or need additional assistance, please feel free to ask.
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I applied 1000 N force and 900 hz frequency using the Steady State Dynamic Direct method but after 260 increments negative eigenvalue appeared in my analysis. Is there any way to avoid warnings and negative eigenvalue?
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Aynul Hossain Can you share your Abaqus model (.inp)?
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Hello everyone
I'm trying to do free and forced vibration analysis of a beam which rotate around of its ends with an angular velocity using ANSYS apdl 19.2.
How can I define the angular velocity and the radius of rotation in ansys in order to do the analysis ?
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Dear Mohamed,
To define the angular velocity and the radius of rotation for a rotating beam in ANSYS APDL, you can try to use the following steps:
1. Define the reference frame for the rotating beam. This reference frame should be centered at one of the ends of the beam and should have its z-axis aligned with the beam's axis of rotation.
2. Define the angular velocity of the beam. You can do this using the "D" command, which defines the angular velocity of a reference frame in terms of the angular velocity of a different reference frame. For example:
*D,REF,BEAM,WBEAM
This command defines the angular velocity of the reference frame "REF" in terms of the angular velocity of the reference frame "BEAM". The "WBEAM" parameter specifies the angular velocity of the beam in rad/s.
3. Define the radius of rotation of the beam. You can do this using the "C" command, which defines the position of a point in terms of the position of a different point. For example:
*C,P1,P2,X,Y,Z
This command defines the position of point "P1" in terms of the position of point "P2", with the "X", "Y", and "Z" parameters specifying the displacement of point "P1" from point "P2" in the x, y, and z directions, respectively.
To define the radius of rotation of the beam, you can use a "C" command to define the position of the end of the beam in terms of the position of the center of the beam. The radius of rotation will then be equal to the distance between the center of the beam and the end of the beam.
For example, if the center of the beam is at the origin and the end of the beam is located at (0, R, 0) in the reference frame of the beam, you can use the following "C" command to define the radius of rotation:
*C,END,CENTER,0,R,0
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In a linear forced vibration analysis, how could we determine the total time of the transient reponse ? If it is determined from the free vibration frequencies, i.e. total time= the period of the 10th mode. How many mode shapes do we need to account?
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Anders is perfectly right, knowing the free vibration frequency is not sufficient, what matters to set this transient time right is the associated structural damping. It depends primarily of the nature of this damping: while the tanDelta of a single piece of metal is only few 1/1000 (the material's loss factor), assemblies will increase it to several 1/100 for most industrial structures (friction between interfaces) - the more complex the structure the higher loss factor. It can be easily determined experimentally by measuring the acceleration decay averaged from several shock pulses here and there (rubber hammer for example)- this does not require complex vibration analysis equipment, using simply an oscilloscope with trigger mode is fair enough... It will remain an average/approximative determination compared to a full modal analysis as suggested by Anders, but still sufficient to get the right order of the number of periods you need to consider the vibration level as steadily established
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I would like to ask about the stability of the Newmark-Beta method used for the forced vibration analysis of laminated plate. In many books, stated that if the coefficients, alfa and beta are respectively taken as 1/2 and 1/4, the results are unconditionally stable which means the change of time step size does not effect the results. However, in my results, I have divided 1[s] into 40 and 100 time steps and it seems that the transverse mid-displacement time histories of the two cases do not match with each other.
Has anyone ever faced with a problem like this and what could you advise me for this problem ?
Best Regards,
Ahmet
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I have solved the problem and want to share what the problem is. First of all, under a uniform constant load during time, there could not be such a transient response figure that looks like as if there would be a harmonic excitation. The problem is I have selected the duration of the analysis randomly such as 1[s]. I have taught that there would be no problem since Newmark-Beta constant average acceleration method (alpha=1/2-beta=1/4) is unconditionally stable. However, I have learnt that there is a limit calculating from the maximum angular free vibration frequency while selecting the time step size to make the calculations stable. I hope it works for all.
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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?
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I would say that you need more than one sensing point to describe your shape; then you may derive the shape spatially. This is what I achieved with full-field measurements, please have a look to:
If the spatial description is too coarse, other authors have tried with closely spaced accereometers to get the same spatial derivation I did with non contacting measurements.
Best regards,
Alessandro
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i newly started my phd in structural health monitoring and need to understand some basic terms.
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It would help if you show what you measure. I cant see any meaning to the y-axis on the right as a «system order». The height of the peaks is a measure of the modal mass and damping of your measured object.
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The aim is to study the effect of flow velocity of a certain fluid in a pipe, on the vibrational behaviour of the pipe.
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Thanks. That's exactly what I meant. Fluent is included in ANSYS.
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Using Python, I would like to convert PSD (G2/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?
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DearRavi Patel:
At the first, and as you know we can say the following information:
Vibration Research software uses Welch’s method for power spectral density (PSD) estimation. This method applies the fast Fourier transform (FFT) algorithm to the estimation of power spectra.
In regards to his method, Peter D. Welch said, “[the] principal advantages of this method are a reduction in the number of computations and in required core storage, and convenient application in non-stationarity tests.”
The process begins with Gaussian-distributed time-domain input data—i.e., a time history file.
Relying on the above، you can benefit from this valuable article about your topic:
"Calculating PSD from a Time History File"
I hope it will be helpful ...
Best regards ...
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Which is the most crack direction is studied in a metal cantilever beam? When the crack is vertical or horizontal? (I mean if the crack propagates vertically or horizontally?)
And how each type of excitation(Bending, Axial,..) is sensitive for each type of crack shape and orientation?
I am asking for test purposes .. So I can use the crack direction that is more sensitive for measurements in my experiment.
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I think it depend on the type of load applied.
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Is there any effect of damage shape on the damage severity? And which do you expect has more severity of damage the holes shape or longitudinal shape? And why?
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I think this article may be helpful
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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..
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To expand on the answer provided by Vyacheslav Ryaboy, when performing the increasing-frequency sine sweep test, you would pass through the first resonance frequency of the cantilever beam. This is what is referred to as "hitting" the first resonance frequency. If you continue to increase the frequency, you will eventually reach and pass through the second resonance frequency. If you continue to increase the frequency further, you will eventually reach and pass through the third resonance frequency.
Depending on the equipment at your disposal, you may wish to use an impact hammer to simultaneously excite a number of modes of vibration. The measurement system would likely have the ability to compute the transfer function between the response and the excitation force. The frequency response curve would generally show the first few natural frequencies of vibration of the cantilever beam, dependent somewhat on the bandwidth of the force impulse and also the point of application of the transient excitation force.
The links provided by Om Prakash Chhangani and Mohamed-Mourad Lafifi give the formulas for the first three natural frequencies of a cantilever beam, as well as a graphical representation of the associated mode shapes. There are some suggestions for how to go about experimental testing there too.
What test equipment do you have available to conduct the experiments with?
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How to represent natural frequencies and mode shapes in same matrix for many cases for damage detection purposes?
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As Christian Willberg already pointed out, a 2 DOF system can only have 2 eigenvalues and consequently 2 mode shapes. These mode shapes describe all possible movement/displacement fields of the system. That is to say, each system response is a linear combination of those 2 modes. Everything else is not possible. Thus, if you have 3 modes shapes you must have a 3 DOF system.
In the modal decomposition method, the modes can be used to decouple the equations of motion (provided that we fulfill certain assumptions with respect to the material damping).
Therefore, you must restate your question and provide more details on your system, if you expect a meaningful answer. At this point, the question you would like to be clarified is not physical. Please go back to your problem and think a bit about what you actually want/need to know.
Kind regards,
Sascha
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I have the following questions regarding vibration-based damage detection of a cantilever beam:
1-What is the purpose of discretizing the cantilever beam by finite element technique?
2-Do the number of discretized elements and their length affect the modal analysis( healthy and damaged natural frequency, mode shapes)?
3- Why do the biggest changes in natural frequency happened when the damage occurred near the fixed end and became smaller if the damage occurred far away from its fixed end?
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The reduction in stiffness will be high if the defect/crack is present in regions of high stress. In a cantilever beam, the maximum stress occurs at the fixed end. Therefore, if the defect/crack is near the free end you can expect a greater reduction in the natural frequency from that of an undamaged beam as compared to when the defect/crack is present elsewhere. You can find a discussion on the same in my research article. I have attached it below for your perusal.
Hope it helps.
Regards,
Jatin
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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?
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Thank you so much for your reply, and have a nice day. Best regard.
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I have calculated the gear meshing frequency of planetary gearbox to be 786 Hz. However, when a FFT is performed on the data acquired for the same planetary gearbox I could see peak around 645 Hz and not at 786 Hz.
The calculated mesh frequency was done based on the speed and number of teeth. But the signals acquired during operation was under loaded condition.
Does external load change the natural frequency and meshing frequency of gear?
Is there any reference to calculate the theoretical gear mesh frequency in relationship with load.
Attached FFT plot.
Thanks in advance for sharing you knowledge.
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Gear meshing frequency is a kinematic (rotation speed-related) parameter. If you have another maximum in the spectrum under loading conditions, this effect can probably be related to another source (gear coupling or bearing). Of course, if you have reduced rotation speed under load (motor power drop), you will have shifted the frequency of gear meshing.
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I am analyzing the squeal of Disc Brake Assembly using Abaqus and I tried to find the unstable frequencies using complex frequency which caused the squealing noise: I am able to find the unstable modes but they are not forming a stable unstable pair. Also, the participation factor is coming for many modes. I am attaching the brake model for reference. Please help me to find unstable frequencies using Abaqus?
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Hi
For complex modal analysis, I fear we are all beginners.
Positive damping is conversion of vibration energy into heat, a nearly irreversible phenomenon, ie a power loss.
Negative damping implies power input, ie excitation.
Whether a mode get negative or positive damping depends on several factors. The cross product between transportation (rotation) and vibration matters, as does the mode inherent damping. Add to this effects from a spinning body where you get forward and backward rotating modes.
A well studied phenomenon that contain the same base physics is wind excitation of a circular elastically suspended body.
Anders advice wrt Jim Woodhouse is very good advice. In my mind, his work is sterling quality.
Hope this helps
Claes
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I have a simulation code for a Horizontal Washing Machine.
The code solves the equations of motions of the system by Matlab ode45 and plots the vibration response of the system at the transient state of performance.
In this code, the frequency (omega) is an exponential function of time, as it's stated below (and its diagram is attached to 'the question'):
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omega= (1-exp((-0.5)*t))*omega_0+(1-exp((-0.5)*heaviside(t-t1).*(t-t1)))*(omega_1-omega_0);
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ode command:
-----------------------------------------------------------------------------
[T,Y]=ode45(@snowa1,tspan1,initial_vector1);
plot(T,Y(:,1)-mean(Y(:,1)))
-----------------------------------------------------------------------------
The resulting displacement response is attached to the question.
It is desired to :
First, increase the frequency to omega_0 by exponential1
Then, increase it to omega_1 by exponential2
But 'the problem' is that:
the displacement response shows an unexpected increase in frequency at the beginning of the second exponential increase (it becomes 20 Hz, which is much larger than the maximum frequency in the simulation- 10 Hz).
Do you know what could be the reason for this response?
Any help would be gratefully appreciated.
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I understand that the 4DOF system for the washer can be quite lengthy. However, I'm unsure if your question is a control problem. If it is, and the 4DOF mathematical model can be expressed in such form:
x' = f(x) + g(xFc(ω)
where
  1. f(x) = [f1(x), f2(x), f3(x), f4(x)]T and g(x) are the nonlinear terms in column vector forms that you derived from the Lagrangian method,
  2. Fc(ω) is the control force that represents a function of ω in vector form, and
  3. ω (omega) is the control input,
then I think it is possible to design the spin speed profile for the control input, ω, so that the desired responses of x can be achieved. If you want to design the profile, you need to at least understand the mathematical equation for Fc(ω). Do you want to regulate the spin speed at 300 rpm, 600 rpm, or 1200 rpm? Because I see only the signal oscillates within the dimensionless amplitudes ± 4×10–3.
I have plotted the signal according to your suggestion, and compared it with Mahdi's original signal. Note that if t1 > 5/τ, then exp(–τ·θ(t – t1)·t) ≈ 0 after t1, because exp(–0.5·t) has decayed to almost zero.
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Can anyone suggest to me method to design fuzzy inference system (FIS) for MIMO structural damage detection (i.e. data distribution on the membership function, parameters of MF, Generate rules ... etc )
In my system there are 3 inputs and 2 outputs:
Inputs: Relative 1st Natural frequency , Relative 2nd Natural frequency , Relative 3rd Natural frequency
Outputs: Crack depth ratio , Crack Length
Note: I tried to use "genfis" By MATLAB it didn't give me reasonable results.
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Hi Murana Awad
Is it possible for you to share your data with me? I have some idea to solve your problem but I need your data.
I'm in touch by : ahmadi.v.1380@gmail.com
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I am doing vibrational analysis of mono-acetic acid after wavefunction and geometry optimisation. In the output file of vibrational analysis the following lines were present containing harmonic frequencies of mono-acetic acid.
****************************************************************
HARMONIC FREQUENCIES [cm**-1]:
-93.9446 -56.8345 -14.9036 79.1124
83.2432 123.7967 180.8622 415.9833
465.7580 564.8411 581.9038 795.9969
942.4123 1030.7040 1139.4888 1218.9753
1356.8784 1433.9457 1447.8727 1776.4078
2976.9823 3023.5483 3084.8354 3634.5117
PURIFICATION OF DYNAMICAL MATRIX
****************************************************************
HARMONIC FREQUENCIES [cm**-1]:
-55.0883 -0.0000 0.0000 0.0000
80.2766 101.0671 159.3476 415.9277
465.5977 564.8100 581.8316 795.9872
942.4062 1030.6968 1139.4801 1218.9731
1356.8769 1433.9456 1447.8725 1776.4022
2976.9823 3023.5483 3084.8354 3634.5117
ChkSum(FREQ) = 0.26286465E+05
Can some one help me understand what is meant by "Purification OF DYNAMICAL MATRIX" and interpret the above results CH3COOH should have 18 modes of vibrations (3N-6). but I am getting 20 modes of vibrations? Furthermore is -55.0883 cm-1 a false flag or is it because of incorrect optimisation of structure?
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Getting all zeros for the first 6 modes enforces the requirement of invariance of the PES with respect to translation/rotation. Harmonic frequencies obtained by finite differences may often lead, for lack of numerical precision, to non-zero values in these modes. The "purification" should consist (at least if it is the same as in GAMESS) in a procedure of transforming back and forth the matrix between cartesian and internal coordinates, which effectively helps zeroing the freqs and IR absorptions of the translational and rotational modes.
The fact that you still get those -55. and +88. cm-1 is a possible sign of your structure not being close enough to a true minimum, but I can't be sure on this. In some cases, a solution is met by simply repeating the optimisation with increased numerical precision in the DFT integration scheme and/or SCF convergence.
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In this picture an experiment analysis by Shaker . What is the conclusion from picture?
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Hi All,
I have two vibration signals taken from the same source using two difference sensors. Therefore, the frequency resolution for both the vibration signals is different (One is 1.5 Hz, other is 7 Hz). What methods or tools could be applied to compare both the signals?
Attached are the two signals obtained from the two sensors.
Note - acceleration is Y-axis is considered to comparison.
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I guess, first you need to know what you would like to compare.
And then understand, what kind of spectral data you have.
In addition, if you want to compare the magnitude of the signal for different frequency resolutions, you need to be aware of the difference of a spectrum versus a spectral density.
E.g., a power spectrum gives you the total power within the frequency bin, so for a broad band signal, the power changes with the width of that bin (i.e. frequency resolution). On the other side, a power spectral density gives you the normalized power per unit frequency, so it stays constant for white noise independent of the frequency resolution, but it changes with analysis bandwith for a narrowband signal (e.g. sine wave) due to the normalisation by the frequency resolution.
I hope this somehow covers what you are looking for as your original question is quite broad.
By the way, the frequency resolution of the spectra you have is not defined by the sensor, but by the analysis process applied to the signals. It might be helpful to understand the complete measurement and analysis process in more detail. Depending on your task, it might even be possible to use the 2 sensors with the same analysis process, making it easier to compare (not: understand) the results.
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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
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Mr. Shaik:
I am not familiar with the specific equipment you use, but if your vibration exciter is a permanent magnet electromagnetic shaker, it might be possible that you can correlate the current input into the shaker’s coil with the output force. If you can set up the power amplifier so that it is in a controlled-current mode, then the current would be proportional to the voltage input to the power amplifier, and the force output of the shaker might be a predictable function of frequency and the current and voltage. I did something like this as reported in an article published in 1989, “Experimental active vibration damping of a plane truss using hybrid actuation,” which is posted on ResearchGate.
You might also find help in a detailed article (perhaps even a series of articles) about shakers written by George Fox Lang, published in the magazine “Sound and Vibration.” I don’t recall when Lang’s articles were published and I can’t find them in a quick look into my files, but I think they came out sometime between 1990 and 2010.
As you are probably aware, the conventional and most reliable method of measuring shaker force output is to position a dynamic force sensor between the shaker sting and the attachment point to the structure.
Yours is a well written question. Good luck finding a satisfactory solution.
William Hallauer
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i have a project to do the vibration analysis of underwater pipeline hence i dont know from whee to start and how to do please help
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You should be aware of the widely used document in pipeline industry titled: "DNVGL-RP-F105 Free spanning pipelines". Then ANSYS can be used for a simple beam modal analysis if needed to determine the governing vibration frequencies and mode shapes which are used therein, with due account for the effective axial equilibrium force in the pipeline.
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while performing i'm encountering error , i'm attaching the mess file and APDL commands
*** WARNING ***                          Entity 3 is undefined.  The MPDELE command is ignored.                 
 *** ERROR ***                            Element type 2 is not the same shape as FLUID30.  Switching to a         different shape is not allowed while elements of type 2 exist.  
please help
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I am interested in the cooperation.
Is there a possibility of doing the research here in Tunisia?
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I am an undergraduate working on modelling generator noise through an aperture and writing my literature review has been a little extra difficult.
I have been trying to understand the methods of noise analysis in acoustic engineering. There are four of them which are mostly used. And I need to understand each of them to be able to choose the method I would be using in my modelling.
If you have any material, book or journal that could assist me with my research work please do send in, would be very much appreciated.
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see
1-Fundamentals of noise and vibration analysis for engineers
Cambridge University Press
M P Norton, D G Karczub
Year:
2003
2-Noise and Vibration Analysis: Signal Analysis and Experimental Procedures
Wiley
Anders Brandt
Year:
2011
3-Fundamentals of Noise and Vibration Analysis for Engineers
Cambridge University Press
Norton M.P., Karczub D.G.
Year:
2003
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I would like to ask teachers: many experts have done a lot of research on the vibration analysis of cylinder panel, what is the research background of cylinder panel ? Such as submarines, aircraft, missiles and so on ? Do you have any pictures of this cylindrical panels applied to an actual object ? For example, the cylindrical panels on airplanes and submarines. It is better to have photos of actual objects or paper containing photos of actual objects.
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This chapter of the Marine Structural Design book might help you:
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This is the part of the input script. While running the simulation an error is showing
ERROR: Kspace style requires atom attribute q (src/KSPACE/pppm.cpp:208)
# vibration analysis simulation units  metal dimension 3 boundary p p f atom_style atomic neighbor 0.3 bin neigh_modify delay 2 every 1 newton on read_data vibration.data group Platinum type 1 group Argon type 2 group Carbon type 3 # force field pair_style hybrid/overlay lj/cut/coul/long 12.0 eam tersoff kspace_style pppm 1.0e-6 pair_coeff      1 1 eam Pt_u3.eam #Pt-Pt pair_coeff      * * tersoff BNC.tersoff NULL NULL C #C-C 3-3 pair_coeff      2 2 lj/cut/coul/long 0.0104 3.54 #Ar Ar Lj potential pair_coeff      1 1 lj/cut/coul/long 0.5650 1.066 pair_coeff      3 3 lj/cut/coul/long 0.0028 3.14 pair_coeff      1 2 lj/cut/coul/long 0.122 3.311 # Pt-Ar pair_coeff      1 3 lj/cut/coul/long 0.092 3.302 # Pt-C pair_coeff      2 3 lj/cut/coul/long 0.139 3.860 # Ar-C pair_modify       mix arithmetic #---------energy minimization------------------------------ minimize 1.0e-4 1.0e-6 100 1000 min_modify         dmax 0.4 min_style     cg
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Charges have to be set for each atom as the Kspace solver calculates long-range coulombic interactions and if you don't want to calculate these interactions, then remove the "Kspace" line.
Hope this helps
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i have vibration data of bridge and process the data with different possible techniques such as FFT, Wavelet decomposition, Empirical mode decomposition, Hilbert Transform, Frequency domain decomposition, spectrograms, etc. but still i am not able to locate exact or pin point frequency of the bridge structure. in the following techniques the results are in a specific range of frequency such as 2.6 to 3.1 Hz. The data showing this specific range and not providing the exact frequency. because in this range any number can be the natural frequency of the bridge structure. so the actual frequency lies in this specific range but i want to go further in depth to figure out the exact value of the frequency in Hz. but the problem is i am not able to find any technique that will give me answer of my curiosity. i recorded data from accelerometers considering normal traffic conditions. there is not closure of bridge. vehicles are passing normal situation.if anyone have any clue about it or still my question is not clear so i can explain further
i attached some of my results
i will be very grateful to anyone who let me find my answer
thanks
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Dear Ali,
the maximum response recorded by your accelerometers is mathematically the CONVOLUTION of the structural response of the bridge (with a peak expected at each natural frequency) and the excitation by the trucks (with a peak at their own suspension frequency) meaning the frequency of the maximum response is a mix of both! Of course the truck's primary suspension frequencies vary from a truck to the next, still in the same range.
In addition the natural frequency of the bridge is load dependant (weight and position of the heavy trucks at the time of the record), and temperature-dependant (meaning day/night, clear day/cloudy and seasonal fluctuations...)
Back to your precise question, you have just NO WAY to determine the exact bridge natural frequencies by this "natural response" method. As suggested, you must close the bridge (however with the same added load than the usual traffic) and use for example a dropped weight to generate a broadband shock (preferably located in consideration of the anticipated mode shape, which allows you to trigger distinctively the first flexural/torsional modes). And still the exact frequency will vary with the previously mentioned load and temperature conditions... Sorry for you, that's just the complexity of structural vibration physics!
Back to Vahid's remark, we could only appreciate the likely natural frequency range of your bridge by knowing its precise construction and span, but the natural frequency of very long bridge spans can even be below 1Hz (remember the ill-fated Tacoma bridge, destroyed by wind gusts at approx. 0.5Hz resonance https://www.youtube.com/watch?v=j-zczJXSxnw). Obviously short bridges have higher natural frequencies but you can interpret it as revealing over-designed structures! The Tacoma bridge example is also a good evidence of the 3D complexity of the mode shapes (something also well evidenced by Eric's study, which provide a very accurate in-situ modal analysis). My guess is that in your case the 2.6 to 3.1 band correspond effectively to the interaction of the trucks suspensions with one of the main natural frequencies of this bridge... in this global 2.8 +/- 0.3Hz band! But don't ask for more precision...
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I have been given samples for 'FTIR for vibrational analysis of compounds'.
What short of information i can get through this and how to read the FTIR spetra to right results for nano materials. Is there any specific spectra table for inorganic compounds.
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I am not sure about the spectra table. However, there are many videos on youtube that show how to analyze the FTIR results. You can search and watch some videos for more details
Regards,
K
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The various parameters taken are:
1. Radius
2. Span
3. Diaphragm arrangement
4. Number of Elements.
5. Boundary condition
6. Cross-section.
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Does this include spring-dashpot elements?
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Dear Mathias D, according to Abaqus manual, element damping is not available for the random response analysis (see tables 1 and 2 https://abaqus-docs.mit.edu/2017/English/SIMACAEANLRefMap/simaanl-c-dynamicproc.htm#simaanl-c-dynamicproc-linear__simaanl-c-alineardynamics-simdamping)
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I need some help, I have run block lanczos modal analysis in ansys APDL.It shows that the 'solution is done ' . However, there is no any result listed in result summary. How to overcome this problem?
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Hi
Check your boundry conditions. because they are very crucial for performing an analysis on ANSYS. I recommend you instead of adpl just go into workbench and select modal analysis
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I've run a few simple random vibration analysis such as cylindrical tube and cantilever beam in LS-Dyna. Now, I try to understand the output results in terms of PSD and RMS values.  For instance, if I input a flat acceleration PSD of 0.02 g^2/Hz over a frequency of 0.1Hz to 1000Hz, how can I use this input to hand calculate the RMS von mises stress? I don't mind the math if someone could explain. Thanks
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You can make an approximation using Miles' Equation for a Single DOF system. The parameters that you need are : (a) input PSD level (b) Natural Frequency (Fn) (c) Q, Transmissibility at fn
Grms = √ ( π *fn*Q*PSDi)
fn = natural frequency
Q =Transmissibility
PSDi = Maximum input psd level in g^2/Hz from PSD v Frequency Curve
NB: This formula gives you the maximum Grms level based on a "White-Noise" input. i.e. maximum input psd level from your input curve. You can make an 'approximation' at the maximum G level that the component will be exposed to and using that G level you will know the exposed G level of the unit. From that level you can run another analysis and evaluate your stresses.
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During parameter study for a project (signal processing applied in mechanical vibration), when I decrease the frequency to a very low value (less than 200 mHz, for a interested range of frequency of a few hundred Herz), the frequency response I have gets "squishy" (instead of a single line going up or down, I have a waveform going up or down).
Checking on some forums online yield me the connection to "dirac delta function", but I do not understand it fully. So I'd like to ask if there is any connection between dirac delta function and frequency resolution, with regards to FFT in vibration analysis. Thanks for your help.
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Roberto Ferretti Don't worry about that. I have a... strange explanation for it. Both in terms of "why the grey line is heavily smeared" and "why do we see peaks at 50 150 Hz and so on".
The first is a combination of very low frequency resolution plus window function. The second... noise. Probably noise. Or some equivalent "freak accident". More recently, I have repeated the experiment under the same conditions, the change is only "hitting different point" - but considering the graph above is the FFT of input force signal, there should be no change. Yet, the recent experiments show practically no such peaks at all.
Another answer (and a more technical one) is the influence of powerline (AC 50 Hz) and some internal harmonic of the measuring system under certain conditions...
Yeah, it is still quite fuzzy.
But still, thanks for your help.
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After taking the frequency spectrum of an impulse force (from a shaker, sampled from a vibration analysis test), there are noticeable peaks at 50 and 150 Hz (presumably due to some electrical/electronic artefacts or some mismatch impedance between the shaker's coil and its structure). This specific experiment aims to study the effect of frequency resolution on the spectrum.
When the frequency resolution is smaller (same bandwidth, or F_max - F_min, is used, and more spectral/FFT lines are used), these sudden peaks start to appear. It should be noted that the original input signal is the same. The question is why this happens?
Thanks for your help
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The graph shows the conversion of a "knocking force" (man-made impulse) from time to the frequency domain. Thus, the peaks should not be there in the first place. Time duration of the signal or the measurement time (defined by the software) is 1/delta f (inverse of the frequency resolution).
About the different heights, it can be contributed to the different coefficients in Fourier transform (this is made by having different frequency resolution)
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Dear all
I am working on dynamic vibration analysis of FGM plate with piezoelectric patch
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Hello folks,
i am doing a vibration experiments to analyze the transfer function. But, before that I want to see the effect of FFT-Lines (Band width kept constant) on my experiments. Higher the FFT-Lines will be, finer will be the frequency resolution, which is actually good to analyze the result (if measurement time is not a big factor). Now, I wanna know is there any limitation for frequency resolution? Does my frequency spectrum will not show any change after a certain frequency resolution factor?
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If you can get access to them, read the Keysight application notes on spectrum analysis. Many of the same principles apply when using an FFT, which is spectrum analysis. The old application notes are as good or better than the modern ones. They used to be Hewlett Packard and then Agilent. Their technical publications are excellent.
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Our research is turing to nonlinear vibrations of plates and solids of acosutic wave device strcutures. It is high frequency vibration, and associated phononmena can only be considered with nonlinear analysis.
Please share your experiences and publications on nonlinear Rayleigh-Ritz method with us.
Joint work on the problem is possible for co-authored publications.
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Dear professor Ji Wang
RR method can be used vastly when it is difficult to obtain analytical solution. One of the key benefits of this approach is its capability to be used for variety of boundary conditions.
For the nonlinear vibration it is used to transfer the space-time problem to the time problem. I mean, after applying this method the vibration problem converts to the a 1dof system, which is just time-dependent (mD[X,2]+cD[X,1]+kX=f). now you can implement a variety of methods for this current problem, namely, multiple scale, straight forward expansion, harmonic balance, etc.
Hope it helps
All the best
Mahdi
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I use a piezoelectric actuator for active control of a flexible structure, and measure the vibration amplitude at the same place using an accelerometer. When actuating the structure with a shaker, I see less noise, but when actuating with the piezoelectric actuator, there is too much periodic noise in the response, and it is not usable.
Any idea how I can reduce or remove it?
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I know there is an electric circuit especially for removing and attenuating the piezoelectric noise. although the software are able to attenuate some noise, but I have tried and I've seen that the hardware filters are much more effective.
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Dear researches
In general, the value of modal participating mass ratio (MPMR) for each vibration mode represents the participation of each mode in the structural responses.
when a structure equipped with TMD, a vibration mode is added to the others. MPMR value of the added vibration mode can be significant even when TMD mass is very small and as a result, TMD has no effect on the structural responses.
How can this contradiction be justified?
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I second Einar Strømmen & Pranoy Nair responses, also I would like to add a point that when we add a TMD we are introducing damping into the system. I am not really sure how you would define MPMR in such a system as modal analysis works only for undamped systems.
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I am conducting a project where underground blasting effects are analysed on soil and structures above the soil. I've currently only have a very basic Abaqus model of a soil block with an implicit dynamic step. A time history amplitude was induced on this basic block model of soil. It shows the stress wave propagations. It has parameters for the Mohr-Coloumb plasticity. I am looking for books or any literature to further understand how the vibration of the blast will effect the soil layer, and possibly different layers of soil (rocks, clay, silty, soil etc). Any help or advice is appreciated!
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ERT
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Dear friends
I'm working on the validation of an FIV case, which for making phase difference between force and displacement, I used an initial velocity in the structural, but after several coupling step, the drag force becomes constant, therefore the structure doesn't fluctuate, the drag force is more than the amount mentioned in the original case which shows vibration in the structure.
any suggestion will be appreciated
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dear Dr. Thomas Frank
thanks for spending your time and answering my question.i will check points you mentioned.
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In ABAQUS help I read the applications of direct solution in steady state dynamic analysis as follows , what are the examples of applications and where I can NOT use this method ? I would appreciate a comprehensive answer , thank youhttp://ivt-abaqusdoc.ivt.ntnu.no:2080/v6.11/books/usb/default.htm?startat=pt03ch06s03at09.htmlThe direct-solution steady-state analysis procedure can be used in the following cases for which the eigenvalues cannot be extracted (and, thus, the mode-based steady-state dynamics procedures are not applicable):
  • for nonsymmetric stiffness;
  • when any form of damping other than modal damping must be included; and
  • when viscoelastic material properties must be taken into account.
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Aha, I see.
Well, even with zero damping, you will get finite response as long as you do not hit exactly the resonance frequency.
To better understand, take a look here - the sdof system is key to all linear dynamics
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Damping of oscillation by interstitial atom movement through lattice shows a damping peak. Why the damping is frequency dependent, while interstitials transit from site to site, and restoring forces are not strictly hookean? An article claims, such fitting of damping data with mechanical impedance models have no theoretical justification, ( , , )why?
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Dear Sumit,
I think snoek effect is related with hydrostatic relaxation and orientation.
relaxation.
Interstitial atoms in ternary b.c.c. alloys exhibit hydrostatic relaxation, if an applied stress has a dilatational stress component. The effect is absent in binary alloys. Hydrostatic relaxation, in contrast to orientation relaxation effect, involves long-range diffusion; in order to understand various aspects of the process. The Snoek relaxation is associated with the redistribution of interstitial atoms in the bcc lattice under the application of the oscillatory stress. Addition of substitutional solutes introduces new peaks or broadening of the normal snoek peak. The relaxation due to interstitial solutes in bcc metals is often affected markedly by the presence of a small amount of substitutional solute atoms. This effect may be originated from the interaction between the substitutional and interstitial solute atoms, or the s-i interaction.
Example, damping of steel tuning forks goes through a maximum as a function of temperature and that the location of this maximum is dependent upon the frequency of the fork.
Ashish
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Hello,
I want to analyze a wheel-tire system. In order to do that I need Young's Modulus, Shear Modulus, Loss Factor and Poisson's Ratio. I have already checked some articles but I need a specific TPE or rubber. Because I will produce this system after my analyses. How can I find these properties?
Thank you.
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Dear sir,
Rubber properties vary from each and every blend based on each constituents added per 100 parts of rubber, their process, shelf life, etc. Better to obtain properties from RMA or original manufacturer.
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I am trying to regenerate the Natural Frequencies of a clamped-clamped viscoelastic core via. ACM (Approached Complex Eigenmodes) method as is in the article : "Linear and nonlinear vibrations analysis of viscoelastic sandwich beams" (DOI: 10.1016/j.jsv.2010.06.012).I am trying to use semi analytical Galerkin method in order to do so. the problem with the use of ACM method for modeling the core my answers are close but when I increase the number of modes to solve the problem in more precise way my answer get worse! and also there is no convergence happening with increasing the number of the modes. what the problem could be? and would it be a better approach to solve the frequency dependent problems which wouldn't be highly costed in solving?
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Hi. There are two ways. the first one is to construct a mechanical model to describe the viscoelastic behaviour (in frequency domain) and include it in the stiffness matrix of the model (In this case, the Kglobal matrix will depend of the frequency and you will need use a recursive algorithm to compute the modal parameters). the second way is to express your model in a time domain and use the FFT.
I describe the fist way in my reseach
Uncertainty propagation analysis in laminated structures with viscoelastic core
WP Hernández, DA Castello, TG Ritto - Computers & Structures, 2016
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I have been doing research on extracting information like acceleration, distance, etc. from an analogue signal by performing analogue signal processing on any type of signal. I haven't found any certain techniques which would help me in providing a formula or any sort of information which would lead to extracting information by analogue signal processing. If anyone is aware of analogue signal processing, could you please enlighten me if it is even possible to do this?
After spending hours, I am at a stage where I feel like it's not even possible to do this.
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I think so. The most acceptable method is to convert the measured non-electric quantity (acceleration, distance, etc.) into an electrical quantity (voltage, current, resistance). After receiving the analog electrical signal corresponding to the measured value, you can further work on the processing of this signal. First of all, it will be necessary to choose the appropriate type of sensor (accelerometer, piezo sensor, displacement sensor, etc.). The analog signal received from the sensor will most likely have to be amplified, and then it can be processed using an analog-to-digital converter. But the easiest way is to convert the signal from the sensor into electrical voltage with a value proportional to this value and then using the capabilities of a conventional multimeter, you can control this value. If the received DC signal, then using a multimeter you can easily control the voltage in the range of 1 mV - 1000V. If the AC voltage is less than 200mV, then it must be amplified so that it can be measured with a multimeter.
I wish you success
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Hi All
How can I define frequency dependent damping in Ansys ?
it seems to be Damping frequency, but I am told that the value I input here, is used only to calculate the stiffness matrix coefficient, so If I input multiple values for damping , they will get overwritten
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You might find something useful here, although Im not entirely sure of what you are asking for.
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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.
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I am looking for ways to do vibrational analysis for the molecule I studied. But I found very little about how to self-define coordinates. And I don't know the meaning of 's' or 'k' in the first column of automatically generated .dd2 file.
Does anyone know the logical way to do it?
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I am using FFT(Matlab) to convert time domain curve in frequency domain but it is showing me the number of samples in X-axis instead of frequency. How to obtain the frequency in X axis in frequency domain curve obtained from time domain using FFT?
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For example, if you have N data in a range of T seconds, then the steplength in frequency domain is 1/T Hz. Keep in mind, FFT will also return N data in frequency domain.
Thus, what you get is exactly [0:(N-1)] * 1/T Hz in frequency domain.
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I have frequency domain result for pressure. After converting it to time domain by applying inverse Fourier transform it's amplitude is changing. It is not matching with the respective (1/frequency) values in time domain curve. Is there any normalization applied while converting by inverse Fourier transform?
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Yes, there is often a constant scaling factor, such as a 1/(2*Pi), or square root of that, as I vaguely recall. Each FFT package seems to use its own favorite scaling.
You must look up the detailed description of the inverse FFT package that you are using. It should place the equations for the forward and the inverse FFT side by side in the description, from which you can see exactly how the scaling is being handled. You generally have to check for every different FFT routine you use (MatLab, MathCad, R, Numerical Recipes, etc...)
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Hi all
in an HCF analysis for a Shell body, how do I use probability in Stress level in wohler curve ?
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Hi Farzad, could we consider bootstap fir this problem? Let s try
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We know that IRF for an SDoF is:
h(t-ta) = ( 1/m*wd ) * e- kesi * wn * (t-ta) * sin( wd * (t-ta))
what is the form of an MDoF function, with this in mind that we do not want to use modal analysis to decouple the equations?
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Hello Mohammad Maldar
The Impulse response function for a MDoF will be given by the same expression as for a SDOF system with the exception that the terms in the formula will now be matrices and vectors.
That is h(t-ta) = (M*wd )-1 * ekesi * wn * (t-ta) * sin( wd* (t-ta))
where M is the mass matrix ; K is the stiffness matrix
wn = sqrt(KM-1) wd = wn*sqrt(I -D2)
I = unit matrix ; D = damping matrix
The user should know how to compute a function of matrix A: f(A). This involves the diagonalization of A by determining its eigen values and eigen vectors.
Note that K = [ k11 k12.......k1,n ; k21 k22..............k2,n; ..................; Kn,1 kn,2 ........................kn,n]
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Hi all!!!
I am trying to obtain frequency parameters of one dimensional nano-structure which is placed on an elastic foundation by using MATLAB code. But I am getting complex numbers (purely imaginary) and complex conjugates for frequency parameters . What do those imaginary values mean and why am I getting this?
Thank you in advance.
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These are the roots of the characteristics equation or Eigenvalues of system matrix. Imaginary parts represents oscillation or damped natural frequencies
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I want to do a free vibration analysis for a plate. In this research, I want to use the Ritz method. For Ritz-shape functions, which are the best polynomial functions that give the most accurate answer?
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you can find one of them in the link below
fenix.tecnico.ulisboa.pt/.../Shape_functions.pdf
and the other one in my derivation for static deflection of simple (beam/plate)under uniform load
Load = w
Shear = w (x - a/2)
Moment = w/2 (x^2 - ax), Moment = 0 at x=0, so integration constant =0
slope = w/2 (x^3/3 -ax^2/2 + a^3/12) slope = 0 at x = a/2
deflection = w/2 (x^4/12 - ax^3/6 + a^3 x/12 ) , def. = 0 at x=0, constant =0
deflection = w/24 ( x^4 -2 a x^3 + a^3 x)
shape function = ( x^4 -2 a x^3 + a^3 x) x ^m
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Hello everyone!!!
I am very curious to know about the real life applications of various elastic foundations such as Winkler elastic foundation, Pasternak elastic foundation etc.
Thank you in advance
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Dear Subrat,
The Winkler elastic foundation concept is valid for both the shallow (strip footing, individual footing & raft foundation etc and deep foundation.) The applied load on these foundations are always lesser than the ultimate loqad by considering higher factor of safety so Winkler model concept is valid. However, these models have its own limitation, the main limitation is the predicted settlement will be considered only at the point, it does not consider the continuity effect..The application of these model for predicting the response for sheet pile wall or the pile subjected to only lateral load or combination of vertical and lateral load is questionable. Also, Winkler model theory is not very much useful in case of ground movement predictions due to earthquake. But, still Winkler model based analysis is valid for shallow foundation subjected to pure vertical load..Ok all the best in your research..
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Hello everyone,
I want to simulate SKF bearing 6308 in ANSYS Workbench explicit dynamic. I created the model as below, but for simplifying the model, I used 3 balls for the bearing. 
The inner race of the bearing has a local fault. As it clear on the outline, I used frictional contact (coef=0.005), an angular velocity of 1000 RPM for the inner ring, fixed support of the outer surface of the outer ring, and end time 0.018 (for one complete revolution). I did not consider any radial load. After solving the problem, I realized that only the inner ring rotates and the cage and balls slip a little (3-5 mm). For this problem, I changed the frictional coef from 0.005 to 0.1 and even no separation. But, still, I did not see any rotation in balls. They just slip a bit with the inner ring. One more time, I assigned clearance between balls, races, and the cage. Again, it did not work (clearance 0.01 mm).
It was weird why balls do not follow the rotation of the inner ring. As a result, I created a simple model similar to bearing the cross section with a ball between them. As shown in the picture below. 
By moving the upper part to the right, again the ball remains constant without any rotation. It just has a bit displacement (slipping). 
I really appreciate if anyone can help me step by step to solve this problem. 
Thank you. 
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Hi
i thin you must attention to connection and interaction type between bodies
suggest test different connection type and select better of that
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I'm trying to solve motion equations of one degree of system with tuned mass damper. the principal mass is subject to withe noise excitation. how can i have obtained the frequency response and the temporal response??
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Hello Aymen, if your system is linear, you can use Frequency Response Functions (FRF) analytically derived from differential equations to calculate system response by the simple formula: Sij(jω)=|Wij(jω)|^2 x S0(jω), Sij - response spectrum, S0 - input spectrum, Wij - FRF. White noise excites all natural modes.
Good source here:
If you solve equations numerically, just apply at input a white noise signal and see response.
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Hello,
A structure is tested for a Random Vibration (say Z axis, 0.1PSD, 20-3000Hz), where I obtained response acceleration from measurement system. Now I wanted to obtain same output(amplitude of modes) by giving Harmonic(sine) sweep as input(in Frequency domain) to Shaker machine. I read some on-line papers explaining conversion/equivalence from PSD to harmonic but none of them are actually working. Can someone please refer some good source?
Thank you,
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The basic assumption of the random vibration is that you can't predict the actual time realisation of the process, only the statistical measures, which assuming a Gaussian process is the standard deviation. So you can only compare a harmonic and random it in an average sense, which is to have the same input power. And that is what the PSD gives you. PSD is usually approximated as the square of the signal's Fourier Transform. So if you know what PSD value your random process has in a fequency bin, then you can work out what amplitude of sine wave would give similar power. But the only way you can relate the two are through the PSD, comparing actual time histories is irrelevant
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Hi All
For an ndof vibration system of mass and dampers, we know the equation of motion for imposed force.
but in finite element codes, it is also possible to apply a constant displacement and calculate the reaction force in the fixed end of the system as the response function .
Anyone could help with how to write the equation of motion for such a vibrating system ?
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What are the advantages and disadvantages of performing numerical integration from acceleration to displacement in the time domain and frequency domain, respectively?
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When you deal with high frequencies, time domain integration such as trapezoidal method may give incorrect results. One thing to remember in frequency domain integration is that waveform needs to be demeaned and padded for DFT to avoid aliasing (caused by cyclic convolution property of the inverse Fourier transform)
I suggest the following paper, which nicely covers the topic:
Brandt, A. and Brincker, R. (2014). “Integrating time signals in frequency domain – Comparison with time domain integration,” Measurement, 58: 511-519.
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