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Recently I'm working on crack simulation in modal analysis using ANSYS APDL.
The process is essentially to solve a static problem in the first step, then solve a modal analysis problem based on the results in the first step (with prestress effect on).
In the first step I used XFEM to model the crack. The static problem is solved successfully. When I was trying to do the modal analysis, I got this error:
An unexpected error ( SIG$SEGV ) has occurred... ANSYS internal data
has been corrupted. ANSYS is unable to recover and will terminate.
Previously saved files are unaffected. Please send the data leading
to this operation to your technical support provider, as this will
allow ANSYS, Inc to improve the program.
Does this mean XFEM is not applicable to modal analysis in ANSYS?
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Just a follow-up. I reported the issue to ANSYS but didn't get any reply. But now I'm certain that the XFEM in ANSYS currently does not support modal analysis.
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I tried simulating the natural frequencies of a beam to compare with my analytical solution using Timoshenko beam theory. My simulation using COMSOL presents only the transverse vibration frequencies. Is it possible to simulate torsional vibration frequencies? Does anyone know the theory employed in the finite element analysis for eigenfrequencies using COMSOL?
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Thank you very much K. B. Mustapha I would search for that. Although, I have not noticed such within the eigenfrequency section of the app.
Thank you Jatin Poojary I would try that.
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Hello everyone,
I'm working on doing modal analysis of simple 1 dimension beam. I know that ANSYS APDL doesn't deal with specific dimensions. But when I do modal analysis to extract the natural frequencies of the beam, Ansys gives me different results from the results of the analytical solution ( manual ). Also, it gives me different results from results of a published paper.
so, my question is : what is the dimension of these results ? and how can I overcome this problem ?
Thanks in advance
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The dimension of the frequency information is 1/Time. The unit associated with Time is consistent with the other units you use in your model (Length, Force, and Mass).
Ex :
-> if you use m for lengths, N for forces, and kg for masses, you have seconds for time (and Hz for frequency). And Pa for pressure (Young Modulus).
-> mm / t / ton => seconds too. And MPa for pressure.
So, you have to check your length/force/mass units.
Be sure to read the frequency information (f), and not the pulsation information (w) (w=2.pi.f).
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I am currently working on my project of crack detection in cantilever beams using an Ansys, and I did modal analysis and harmonic response to find the natural frequencies and frequency response of the deformation.
I have found the maximum amplitude of all 6 modes and how does it really work in crack detection?
amplitude gets lower when frequencies get higher.?
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Nazanin Fallahi
Hi,
Thank you so much for your response and its so helpful.
I need to know about how it works with cracks detection?
absolutely the natural frequencies of the healthy and unhealthy varies.
also amplitude of the healthy and unhealthy varies as well. I need a clear idea about detection working?
it would be helpful if someone could clarifying this..
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Hello ,
I am performing Modal analysis for ultrasonic cylindrical horn at a frequency 20KHz. After performing the linear perturbation analysis to find the natural frequency of the horn the end result shows a distorted meshed figure. I am unable to understand where I am going wrong.
Any help related to this query would be appreciated.
Technical specifications of the horn.
Diameter : case 1: 2 cm , case 2: 6 cm
length :14.3cm
Natural frequency: 20KHz
material: Aluminium
Thank You so much.
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Its because of Hourglass energy effect.
Switch to element formulations with fully-integrated or selectively reduced integration.
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Hello World,
I am trying to perform the modal analysis and harmonic response analysis for forced vibration in rotating cantilever beam through ANSYS. Is there any way to incorporate the rotational velocity and add Coriolis and centrifugal effect for the analysis?
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Hello Amrit Tiwari ,
You have not the possibility to insert a rotational velocity in the Harmonic Response, because this kind of analysis uses mode superposition in RESPONSE to a load. The rotational velocity has its effect on the modes of the structure, then the loads from Harmonic Response are "projected" onto the structures modes and the response is than rebuilt up form each mode response.
In the Harmonic response only periodic loads are accepted (a constant acceleration is considered a periodic load with 0 frequency, therefore it is acceptable).
I suggest to perform a static structural analysis and use the pre-stressed state as input to the modal analysis (see project schematic in the annexed archive .wbpz).
Warning: the prestressed structure must have results different from zero everywhere, otherwise, there will be no difference with respect to fixed beam.
See the different response with and without rotational velocity when the prestressed state is different from zero.
Last remark: when I toggle "Coriolis effect" to on, the prestressed solution is zero everywhere, therfore the rotational velocity is completely ineffective. I don't know why, this is unexpected.
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I am trying to do modal analysis of tank using Ansys Mechanical but as the fluid is not having any elastic properties, the Ansys is throwing an error. I tried giving properties in the engineering property step but its not working.
Also I found that in some old research papers that they are modelling tank walls as SHELL181 and fluid as FLUID80 elements. Do I have to model the tank according to that way as I am not been able to find these elements in the ANSYS.
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this paper may help you :
Elkholy, S. A., Elsayed, A. A., El-Ariss, B., & Sadek, S. A. (2014). Optimal finite element modelling for modal analysis of liquid storage circular tanks. International Journal of Structural Engineering, 5(3), 207-241.‏
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I have a model of a building in which I have already defined the structural and non structural masses. Structural masses are defined through the density of the material while the non structural masses are defined as point masses. In this step I want to perform the modal analysis. In the subsequent steps I want to define a static analysis and later a dynamic analysis(earthquake). During the static analysis I want to exploit the gravity load for the structural mass while I want to use distributed pressure for the non structural. But if don't remove the point masses I would consider it twice. In the subsequent step I need these masses for the dynamic analysis. To resume, I have to apply the point masses in the first step, remove from the second and put in the third. Is it possible? When I try to do this from one step it's immediately eliminated from all the steps
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Hi Salvo Emiliano Emiliano Elmo,
You cannot activate and deactivate mass in any step. In other words, either you define mass or not for the whole analysis. But it’s different when it comes to the Gravity load. You can activate or deactivate the gravity load through the load manager dialog box in any step you want.
I hope this will help you.
Best wishes.
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I did some measurements with Polytec laser vibrometer instruments to obtain the natural frequencies and mode shapes of the different beams I have in the laboratory.
After doing modal analysis, the mode shapes are not smooth although we did with a dense number of measurement points (we did the measurement in 2D and 1D and for both the mode shapes are not smooth). The material of the beam is concrete.
Who has the same experience?
I look forward to hearing your suggestion.
I can not repeat the tests and I may do smoothing. But all papers that used this instrument had the initial mode shapes, smooth.
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Dear Fahime, I am a colleague of Alexander from Polytec, I am working in the applications department. In order to support you, I would have some questions myself (sorry...):
- if you look at the measured deflection shapes in the PSV software itself: do they look smooth there and is the distorted shape only appearing when you do modal extraction (so curve fit) with an external software? Or is it already distorted within PSV?
- in case it is already distorted within PSV, could you try to animate just slightly next to resonance peaks to see if it looks ok there?
- easiest and most direct would be: could you upload the measurement data somewhere so that I can have a brief look at them?
Best regards,
Jochen
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Since Hinges don't get activated during FNA and I have to provide Links.
I need to know what should be the force deformation values and Effective Damping for Multi linear plastic links at the ends of beams and columns in Nonlinear modal analysis.
I can get the moment curvature curve from section designer but how to input it in a 2 joint link which asks for force deformation values.
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In your case, I would recommend to design and perform a test the emulates the behavior you expect from your link and get the data for its definition from there,
The easiest way to do it would be if you have experimental results for guidance. For example, a compression/tension link would be easy to define if you performed a compression/tension test on a sample of the material you want to represent (with the link element) but, here comes the difficult part, because you can either define the link with an "effective stiffness", hence providing only elastic behavior, or with a set of coordinates (force/deformation or moment/rotation in most cases), and since you are replacing a geometry, you also need to think about this.
About the proper selection of a link: multi-linear plastic or multi-linear elastic work in a similar way and are very versatile, you will only see the difference if you are performing a hysteresis analysis, where the elastic will... well, behave in an elastic way and the other not.
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So there is a simple beam element on which I was experimenting, trying to find answers to the questions: What is the difference between body load on a constrained assembly (bolted onto a baseplate) and direct nodal load on the constrains? Is there any of these setups that is close to Large Mass Method?
I got some very interesting findings that need explanation. I bolded out the questions that need an answer. Please find the simulation details in the attached images!
Findings:
- Free-free setups have many orders of magnitude lower response to the same unit acceleration load than fixed-free setups. Is this explainable physically, or is it only a simulation gimmick because of mode participation factors (99.99...% factor for rigid-body mode)?
- Also, when nodal acceleration is applied, if rigid-body mode is included, the response is the same as when omitted. Why? (.f06 was checked and it really uses the selected modes) And also if rigid-body mode is omitted from the body acceleration load type, the response is the same (whereas, when rigid-body mode is included, I got unit response because rigid-body mode dominates). Why?
- Enforced acceleration on one end of the free-free beam resulted in a fixed-free modal response. So I'm guessing enforced motion works like a fixed constraint in the modal analysis, then in the Frequency Response it works like a load. Is this statement true?
- In the fixed-free setups, nodal and body load types resulted in similar responses at the eigenfrequencies, but showed differences elsewhere. Why? Also, nodal load goes to unit response at 0Hz while body load goes to 0 response at 0Hz. Why?
Sidenote: I also made a reference run with SOL108 Direct Freq. Response with the fixed-free nodal load setup without damping, and got of course similar result to SOL111 same setup.
Thank you if anyone can answer my questions! The 'Help' of NX was not very helpful :(
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Hello,
I am studying the performance of 2 wing models (with the same size, constraints, loading and material properties but different material distribution) using Nastran. Wing A is lighter and 24% stiffer than wing B and that is demonstrated in the static analysis (under a distributed load on the top skin).
The two wings have about the same fundamental frequency (34.25Hz and 34Hz). Using the relationship f=sqrt(K/m), the difference between the two frequencies should be significant.
Also, the divergence analysis yields exactly the same displacement for the two wings.
Does any one know what could be causing such results? why the divergence displacement is not smaller for wing A (since it is stiffer) and its frequency is not much higher than wing B?
Thank you in advance for your input.
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Rossana Fernandes Can you share the two Nastran models (.bdf)?
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I am trying to perform a modal frequency analysis on valve by Abaqus, But the answer is zero in the first three modes.Anyone know about this?
In fact, I simulated simplified of this model and without any bolts and the simulation was done and ok. When I add the bolts to the model and apply the interaction between hole and screw , the simulation makes error and the first three value gets zero.
Figure 1 is a simplified model without bolts, which was successfully simulated.
Figure 2 is a model with bolts and nuts that I have trouble simulating.
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Nirmal Kushwaha Thank you so much.
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It is recommended to calibrate the measuring instruments before performing any experiment. Ideally, how often do we need to calibrate an accelerometer? Are there any simple and effective methods for calibrating accelerometers in the laboratory without seeking professional assistance?
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Stick the accelerometer to a weight of known mass. Excite the weight with a known force using instrumented hammer or shaker. The acceleration and force will be related by F=ma. If you know the mass m, you can work out an appropriate calibration factor quite easily.
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Hi,
I'm traying to perform a successive modal analysis using APDL, at the end of every analysis i need to save modal frequencies in a table by APDL commands. Please, is there anyone who can help?
kind regards
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Thanks all of u my colleagues, One way to do it is:
.........
.
.
.
.
/POST1
*do,j,1,nmodes,1
*GET,freq_tab(i,j),MODE,j,FREQ
kvector(i)=k*a
*enddo
......
.
.
*CFOPEN,'matable','txt',,'E:\PnC_test_2' !Set output path
*VWRITE,kvector(1),freq_tab(1,1),freq_tab(1,2),freq_tab(1,3),freq_tab(1,4),freq_tab(1,5),freq_tab(1,6),freq_tab(1,7),freq_tab(1,8),freq_tab(1,9),freq_tab(1,10),freq_tab(1,11),freq_tab(1,12)
(F8.4,' ',F12.5,' ',F12.5,' ',F12.5,' 'F12.5,' ',F12.5,' ',F12.5,' 'F12.5,' ',F12.5,F12.5,' ',F12.5,' 'F12.5,' ',F12.5)
*CFCLOS
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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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Hi everyone,
I want to extract the global mass and stiffness matrices of an Abaqus modal analysis, so as to post-processing it with Matlab.
I have already read many comments on RG regarding this question and added the following lines at the end of my Input file to get the mass and stiffness matrix.
*STEP
*MATRIX GENERATE, STIFFNESS, MASS
*MATRIX OUTPUT, STIFFNESS,MASS,FORMAT=MATRIX INPUT
*END STEP
So far, everything is fine.
I have also written Matlab code that works perfectly for 3D solids or shell elements, but not when both solid and shell elements are together.
In my model, I had to use both solid and shell elements in Abaqus. Looking at the mass and stiffness matrices, I saw that there are 3DOF per node for solid elements and 6 DOF for shell elements, so I cannot use the attached Matlab code as the model has 3DOF for solid elements and 6DOF for shell elements.
My mass matrix looks like this:
1,1, 1,1, 5.115770202111807e-10
1,2, 1,1, 4.923414202110917e-12
1,3, 1,1, -5.733085750203833e-15
2,1, 1,1, 1.453330463639921e-11
2,2, 1,1, -2.461707101055460e-12
2,3, 1,1, 2.866542875101822e-15
1,4, 1,4, 4.295511112920469e-14
1,5, 1,4, -4.054846090561977e-14
1,6, 1,4, -8.889774326370160e-16
2,4, 1,4, 1.225209356335630e-14
2,5, 1,4, -1.103358818158063e-14
2,6, 1,4, -4.552490123962178e-16
...
131010,1, 131010,1, 6.793914400596321e-08
131010,2, 131010,2, 6.793914400596321e-08
131010,3, 131010,3, 6.793914400596321e-08
131011,1, 131011,1, 7.423556175182037e-08
131011,2, 131011,2, 7.423556175182037e-08
131011,3, 131011,3, 7.423556175182037e-08
As you see here, some parts have 3DOF and some parts have 6DOF. Some parts are like a lumped and some parts are like a consistent mass matrix.
In this case, how can one use Matlab in order to get a global stiffness and mass matrices from this?
I've been struggling with this problem for a long time and still haven't been able to solve it.
Here, I am attaching both the mass.mtx file and the Matlab code.
Any help would be appreciated.
function [matlab_matrix] = import_matrix(mtx_file)
%============== Import Stiffness Matrix ==============%
abaqus_stiffness_matrix = dlmread(mtx_file);
% merge node number info from column 1 and DOF info from column 2 and
% store in the 1st column of a new matrix
%If number of DOF are 2,multiply by 2.
matlab_nodes(:,1) = 2*(abaqus_stiffness_matrix(:,1)-1)+ ...
abaqus_stiffness_matrix(:,2);
% merge node number info from column 3 and DOF info from column 4 and
% store in the 2nd column of a new matrix
matlab_nodes(:,2) = 2*(abaqus_stiffness_matrix(:,3)-1)+ ...
abaqus_stiffness_matrix(:,4);
% extract the stiffness values from the .mtx file, and store in a double
% length vector
stiffness_values = [abaqus_stiffness_matrix(:,5); ...
abaqus_stiffness_matrix(:,5)];
% create a matrix of the new matlab node numbers, and a vector of indices
% of their position in the abaqus stiffness matrix
[matlab_matrix_indices, abaqus_stiffness_value_index] = unique( ...
[matlab_nodes; matlab_nodes(:,2) matlab_nodes(:,1)], 'rows');
% compile the stiffness matrix using the new node numbering convention
matlab_matrix = accumarray( matlab_matrix_indices, ...
stiffness_values(abaqus_stiffness_value_index), [], @max, [], true);
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I wish you all the best I apologize that this is not my specialty
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If there, can you recommend me a good software?
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I think this kind of photomicrographs is available in your university (Assiut University).
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Dear everyone,
I am trying to simulate the modal behaviour of an aircraft wing and its dependency on the temperature load. I have treated the aircraft wing as a catilever beam, with the airfoil cross-section.
However, when I am applying the temperature, it shows no change in the results. Can anyone tell me where I am going wrong?
Note: I am using the Beam189 element.
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SOLID226 and 227 are eligible for modal analysis of pre-stressed structures. These elements have structural, thermal and electrical degrees of freedom which enables performing a Multiphysics simulation.
Hope this answer helps!
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can you please tell how to solve the governing equation to obtain the frequencies to compare with the ansys result
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Hi
(I haven’t looked at your xml as I doing this from my iPad).
i guess the answer is Yes and No.
Yes, we can use analytical and numerical models, or tests, to find, as best possible natural frequencies. To what effect may be worth pondering.
There are some problem with pipelines.
1) they tend to be very long. Infinite, may be a reasonable approx. Unbounded systems don’t not have modes, they have wave propagation. They can be approximated using modes but you need a lot of them and the modal summation doesn’t always converge as one would want it to.
2) every frequency is a natural frequency for an infinite system, ie you get lots of closely spaced modes, which in turn provides a high modal overlap situation (search RG, it has been discussed before).
3) I imagine that numerical solvers may have a problem with closely spaced modal systems.
4) pipelines tend to lie on something, ie the couple to terrain, again a large system.
So, chasing this problem using modes may not be your best bet.
Wave propagation for analytical models and direct numeric solution for FE probably is the better way to go.
Hope this helps
claes
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I am analyzing the squeal of simple brake using Abaqus and I tried to find the unstable frequencies using complex frequency which caused the squealing noise; however, all of the modes have nearly zero dampings (lie on the imaginary axis) and the pairs of modes which become coupled and formed a stable/unstable pair couldn't be found. I attach the brake model in the comment. How can I found some unstable frequencies using Abaqus?
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Seyed Jamaleddin Mostafavi Yazdi sir, did you get what you were doing wrong?
I am also doing a similar analysis and stuck at the same problem.
Can you please help?
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Greetings,
We know that locking in the finite element method (FEM) is a numerical artifact due to the choice of the approximation functions. A couple of implications of locking in a static analysis can be mentioned as follows
  1. A FEM model of a beam subjected to a point load at its tip can severely underpredict the tip displacement if the FEM model is prone to shear locking.
  2. A FEM model of a beam acted upon by pure bending moment would develop spurious membrane strains if the FEM model is prone to membrane locking.
As far as modal analysis is concerned, what is the effect of the locking phenomenon on the determination of the natural frequencies and mode shapes of the system?
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Hi all
The examples I provide do increase stiffness. The first example from internal maths, the second example from autoconstraining connecting dependent RBE2 dofs to ground and the last, connecting rotational dofs to ground.
If you look at hour glass modes where element formulation provides internal resonances, there I'd say that you lose stiffness .
Connecting beams to membranes, you will likely deal with all of the above.
To elaborate on the RBE2 - a rigid beam element - using it to connect a beam (6dof) to a solid (3dofs) - you get problems with the 3 rotational dofs. If the dependent dofs are on the solid, these get constrained to ground.
The practical workaround is to add a thin shell onto the solid element and to define it using a material with zero density to avoid spurious resonances in the thin shell.
Just my 2 cents
C
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If a model of beam without damping is considered we obtain the real natural frequencies. However, I would like to ask what shall be the nature of the natural frequencies if we consider the rotary inertia of the beam? Whether they will be real or complex?
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I agree with the previous answers. The natural frequencies are real. However, the answers refer to Timoshenko's beam theory, which also includes the effect of shearing. Shearing leads to the situation that the beam cross-section is no longer orthogonal to the beam axis. Additionally, Timoshenko also considers the rotary inertia of the beam. In Rayleigh's beam theory, only the inertia but not the shearing is included. The question refers only to this more simple beam theory. Bernoulli's most simple beam theory neglects both, rotary inertia and shearing.
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Performing modal analysis on freely hanged structure using impact hammer method. In this case roving acceleration method is followed and obtained different FRFs at different locations. so here my questions are
1. how to use these multiple FRFs to calculate single FRF of a structure ?
2. can we get identical FRFs from different locations ?
3. how many FRFs can a structure exhibit ?
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Nothing to add to the very clear guidance provided by Lubos, just proceed as suggested...
BTW no one answered yet your question 3, the answer is simply an infinity because your structure is continuous and every couple of points will provide a distinct FRF (a 2D infinity of possible excitation points x a 2D infinity of possible response points, divided by 2 for factoring reciprocity, possibly further reduced by symmetry, that makes infinity anyway)
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Hi, I am doing research about operational modal analysis using SSI. Generally, SSI only identifies natural frequencies, damping ratios, and mode shapes. But I am wondering how to estimate modal contribution of each identified mode using SSI. I did not find many papers related to this topic. I hope someone can recommend some work or give me some hints to me.
Thanks!
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hello
In order to do my thesis, I need to place the damper I developed at OpenSees on several asymmetric steel benchmarks. For this purpose, to develop asymmetric one- and nine-story structures, I decided to use the method given in the above article, but unfortunately in the article, the structures are parametric and I do not realize this.
Here are some basic questions for me.
One, how the mass and structural members of these structures are defined and modeled.
Second, how can a find period for these structures?
And finally, how is it possible to get the response of these structures under the earthquake record?
My strongest guess is that it is to define stiffness of each structural member as a coefficient of mass or vice versa.
Any help is greatly appreciated.
Take refuge in the right.
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In his name is the judge
thank all of you Juan Pablo Peñaloza & Kan-Jen Liu & Paul Mathew
Vizhalil my dear friends.
I think found my answer. for researching about kind of structure we use parametric structure to do complete research. for example to see affect of torsionally assymetry structures we develope parametric structures.
It means we Generate structure in 3 diffrent periods like 0.5,0.75,1 second. mass, dimension and stiffness of structure are not important in this case what is important is period only. in the second step we Generate 3 structures which have 3 diffrent torsional ratio (torsional ratio = (period of torsional mode)/(period of Transitional mode)) 0.5,1,1.5 for each single period. so totaly we build 27 diffrent structres which their elements, dimension or materials and etc are not important at all the only main things about them are periods and torsional ratio.
We call these structures parametric and we can use them for Exhaustive research about torsionally assymetry structure in our Purpose. i used parametric structures in case of see tlcgd affect on torsionally assymetry structure.
Take refuge in the right.
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Hi guys,
I am trying to perform a modal analysis of a pre-tension 2D array of 4-strut simplex tensegrity plate in Abaqus. Can anyone tell me how to apply pre-tension to the struts and cables for modal analysis in Abaqus?
For now, I was able to perform modal analysis without pretension and did get mode shapes. But only the first two were similar to plate mode shapes and the remaining modes have in-plane or only corner motion. I want to study the influence of pretension in mode shapes and for that, I need to apply pre-tension.
Here I am attaching the abaqus input file
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Run the simulation with two load steps, wherein initial step focus on pre tension and second load step focus on actual loads on the model...
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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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hi, i'm civil engineer phd
my qustion about how i can include a prestressed in my systeme (dam) for a modal analysis with Ansys apdl
think you
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That is a good question.
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Hi
I had a questiones about the operation of the Riley damping command in Opensees.
First, how does the Rayleigh Damping Command affect the damping in any structural mode?
Second, can the Rayleigh Damping Command be set to affect only certain modes of the structure?
Any help is greatly appreciated
I wish you all the best
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Rayleigh damping is defined as a linear combination, i.e.
D ='alpha M + beta K
also known as "Bequemlichkeitshypothese".
The mode shapes of the undamped system simultaneously decouples M, D and K. That means that all modes are damped by using this approach.
You might consider modal damping if only a subset of modes should be damped.
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I've extracted mass and stiffness matrices from a modal analysis in Ansys. I need to know how the elements of each matrix is related to node ID. Anyone can help me with an ansys command to find each node DOF location in mass and stiffness matrices?
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Lumping of the Mass Matrix
The bandwidth of the consistent mass matrix computed from Eq. (4.18d) depends on the node-numbering scheme. To reduce computational effort, we generally approximate it by a lumped (or diagonal) matrix; the same is also done for the consistent heat capacity matrix given by Eq. (4.18e). The use of node points as quadrature points to numerically evaluate Eq. (4.18d) will give a diagonal mass matrix but will result in some zero diagonal entries for axisymmetric problems. Zero or negative diagonal entries in mass matrix can have disastrous consequences. The following two techniques are commonly used to obtain a lumped mass matrix:
Row-sum technique: Elements in each row of the consistent mass matrix are summed, and the result is placed on the diagonal. It can sometimes produce negative masses.
Special-lumping technique: Entries of the lumped mass matrix are set proportional to corresponding diagonal elements of the consistent mass matrix with the constant of proportionality selected to conserve the total mass. The positive definiteness of the consistent mass matrix requires that its diagonal entries be positive.
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Dear researchers
What is your opinion about the criterion recommended in seismic codes for determining scaling period, which are used to scale ground motion records?
As you know, the mentioned criterion is the period of the structure’s dominant mode, which has the largest modal participating mass ratio (usually the first vibration mode). Hence, the period of the mode with the second largest modal participating mass ratio is not considered in the scaling process. Consequently, although this criterion usually results in the largest value of scaling period, it is not logical ones.
This is especially important when Tuned Mass damper (TMD) or Base-Isolation system is utilized, which cause the modal properties of the structures to change.
I used a new criterion based on the weighted mean value of the periods for the structures equipped with TMD.
Have you used any criteria other than the criterion mentioned in the seismic codes?
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Dear Mikayel Gregor Melkumyan , it is my pleasure if u can read my latest article
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I am running modal analysis using Ansys software. i need to extract the effective mass of each mode using APDL and Workbench
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Hi Mostafa Sleem. It can be easily done by switching output to file. Hope this helps.
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hello everyone, i'm trying to model a 3D one story buiding with mass IIrregularity, the model has 4 elastic columns and rigid diaphragm. the model has no beam, so i used rigid diaphragm to equal dof of top nodes of columns. i use a model from a reasercher's article, he note that the modal frequences of first three mode is 1.16 , 1.96 , 2.37 HZ respectively. i builded the model in etabs and can verify my model with 0.33% error, but i couldn't verify model in opensees. i trying almost everythings to verify it but Unfortunately all my work was unsuccsesfull. in my reaserch i found that mass moment of inertia is principle parameter in model with mass IIrregularity, so i try figure out how modal analysis run in etabs and opensees and what difference between them. i found that etabs centeralized all masses in center of mass and then make mass matrix with 3 parameters like that [mass in x direction 0 0 0 mass in y direction 0 0 0 mass moment of inertia about hight axis ] but i wondering opensees generate mass matix for each node(also we can use center of mass for assigne mass) and also mass matrix for 6 dof model with 3 limited dofs (dofs limited in z direction and rotation about x ,y axis) is 6*6, so i'm confused to understad how opensees derrive mass matix and calculate modal frequences. i try both assigning mass to nodes and center of mass. in first condition i assigned mass moment of inertia like that mass n "mass in x direction" "mass in y direction" 0. 0. 0. "mass moment of inertia about height axis" etabs calculate mass moment of inertia with multiple each mass of nodes by distance of node from center of mass then sum these valus and attain mass moment of inertia abuot height axis. also we can use another formula for it. i try all patterns i found but i can't verify my model because i can't understand how opensees run modal analysis and make eigen values. for best understanding i upload my model in opensees and etabs. I'm so thankful if you can help me. with best regard
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I hope my research will be of use to you
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I am trying to simulate response of structure with TLD using ANSYS APDL. I mesh the fluid volume and tanker so that they can have coincident nodes, merge the nodes, flagged on FSI for the interface areas plus I assign SURFACE-To_SURFACE contact between tanker and structure slab (which am not sure if it's the right way to assign connection between tanker and top slab). when i run modal analysis for the fluid filled tanker (with out attaching it to the structure) i find frequency value which is almost equal with Housner's equation, but when i do the modal analysis for the structure+TLD and check the deformed shape the fluid elements penetrate the structure boundary and don't know how to correct it. I would appreciate if any one can give me some recommendation how to solve the problem. For your reference i have attached the snip of the deformed shape. Thank you in advance
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Dear Dr Tuong
Thank you for taking the time to respond and helping me get the materials i needed.
And thank you for your compliment, yeah we have beautiful country full of hospitable people. I would love to if you can visit Ethiopia some day, am sure you will enjoy your stay...
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Hello.
I modeled a complex ring in abaqus and with adding below command in keyword, extracted Stiffness matrix with Coordinate format. But output is a .mtx file that listed stiffness for each node (or element). What to do i to assemble this file to reaching to Global stiffness matrix (stiffness matrix for whole part, i mean m(d2x/dt2)+Kx=F)?
*STEP, name=exportmatrix
*MATRIX GENERATE, STIFFNESS
*MATRIX OUTPUT, STIFFNESS, FORMAT=MATRIX INPUT
*END STEP
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Nils Wagner so in your opinion for this work, it's better to extract this kind value using from K=F/dx? So that force is applied on part and then deformation is measured and stiffness is extracted used from K=F/dx.
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I am trying to model the combined effect of a static and dynamic load on a cantilever beam. I have calculated the dynamic response as follows
1) Natural frequency and mode shape extraction (modes 1-5), see image
2) Steady-state modal analysis gives the steady-state response across a range of frequencies. I am looking for deflection with direct model damping of 2%. If I am not mistaken, this form of analysis require a load independent of time and essential applies harmonic loading to each of the natural frequencies and calculates the response at each frequency.
I am unable to apply a static load at this stage (steady-state modal dynamics) to observe the combined effect. See image for the deflection as a function of frequency. This is precisely the plot I want if I was just considering dynamic response. But surely, this would change significantly with a static load.
I tried setting up a static load in a step prior to the frequency and steady-state modal dynamics step but that was of no use.
Any tips on how to do this.
FYI, I am actually trying to model a chassis experiencing cyclic and static loading from an engine and am using this to fully understand the problem.
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Gabriel Zenkner, I had the same issue and I think I have found the solution. Create 2 steps, the static and the dynamic, and submit the analysis. Then, once you have the solutions to both problems, go to Visualization Module -> Tools -> Create Field Output from Fields. There, you can select Fields from each Step and Frame Increment. Once you pick one, it automatically adds the Field Variable to the big white upper box where you can write whatever expression you want. Also, to the right, you have a list of available operators. In the following example, I wanted to obtain the combined stress state due to static and harmonic loads, therefore: s3f1_S + complexReal(s2f2_S) (stress tensor due to static loads on frame 1 + real part of stress tensor due to harmonic loads on frame 2). The complexReal() operator is necessary since it appears that the harmonic stress tensor in stored in complex form.
Even though I selected to compute the real response only in the steady-state dynamic step, some errors show up when I tried to do the sum without the complexReal() operator.
Hope you find it helpfull.
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Is there any software or excel sheet to calculate the mineral proportions (modal % of minerals) based on bulk-rock analysis. Your kind help is highly appreciated.
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You can use the GCDkit software for normative calculations (no modal % of minerals). You can download it free. It works under the R program and you can use an excel file. It is very easy to use.
Good luck
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Hi
I know the procedure of modeling modal analysis to obtain natural frequencies(I mean using Linear perturbation step>>frequency and without defining any other loading....) but the problem is all examples I have seen is just for model made of just one material(most examples for cantilever beam) but my model is consist of two or more material( 2layer soil and layer of polymer between these soil layers), for obtaining natural frequencies in this case, my exact question is: I have to run modal analysis for each material separately or all together?!
I should mention that the polymeric membrane part does not have any type of support and its just in interaction with soil( normal and tangential behaviour), and if I run the modal analysis for this part alone natural frequencies for polymeric membrane is zero. while the soil is fixed at bottom and sides and by running modal analysis once for soil alone and another time soil with polymer,the natural frequencies are the same for both conditions. so i'm willing to know any advise and special point to obtain natural frequencies procedure for model consist of two or more materials in abaqus.
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Interesting
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Hello,
I tried to perform a response spectrum in my structure. However I can't proceed because of the error (ERROR: Data file file.db does not exist for RESUME). My modal analysis run successfully but I am getting this error when performing response spectrum. How can I fix this error?
I am currently using ANSYS 2020 R2 student version.
Thanks
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If I have a rotor on two bearings, under what conditions should I use pre-stressed modal analysis and when should I run a rotordynamic analysis i.e. using the rotordynamics options in a modal analysis?
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Normally a rotor on bearings does not require pre-stresses modal analysis, since neither the rotor or the bearing stiffness does not change with loading (in this case transverse loading). In order to use pre-stressed modal analysis, you need to have loading which has an effect of changing the overall stiffness of the system. For instance, if you are interested on the modal analysis of a blade on a rotating hub, the rotational speed causes a radial force on the blade which extends it. This extension increases the overall stiffness of the blade. Therefore, results of standart modal analysis corresponds to zero rotational speed. At any other rotational speed, you need do pre-stressed modal analysis to determine the actual mode shapes and natural frequencies corresponding to that speed of rotation.
In rotordynamic analysis, gyroscopic effects are included in the analysis as a result, the modal analysis is done depending on the speed of rotation, since gyroscopic effects depend on the rotational speed. Moreover, rotordynamic analysis may include the stress stiffening (as in the blade example, pre-stressed modal analysis) or spin softening effects depending on the options selected.
If you think that, stress stiffening and/or spin softening effects are important you can turn on these option the rotordynamic analysis. This depends on the rotordyanmic system. If you are not sure what to do, you can always perform two different analysis with and without these effects included and compare your results to see if these effects make a significant change.
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When I heat the structure and run the analysis, the deformation by temperature occurs, and we want to re-run the modal analysis on the deformation results.
It takes a long time to enter boundary conditions at once and run simulations.
Currently, ET is using 227.
If there's a way to solve it by using commands, could I know how to use?
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Taeho Kim Can you share the corresponding .cdb file?
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Hello there! I'm new in using Ansys APDL. I'm trying to do the modal analysis of a simple unimorph piezoelectric cantilever beam, but the result is not what i expect to achive.
I'm wondering that there are some inputs that i miss because of my poor capacity in using APDL, so i'm asking for some suggestions: are there any specific commands that i have to use? I follow a scheme like:
1. Add materials in element type (solid45 for brass, solid5 for pzt, circu94 for the electric part that i don't need for this first analysis)
2. Choose the material props thar i've found on datasheets
3. Modeling the cantilever (2 volumes one on the other): 4.85x1.0x0.4 mm brass and 4.85x1.0x0.5 mm for pzt
4. Glue the volumes
5. Attribute the materials and mesh the model with mapped mesh (0.5mm per edge)
6. Put the boundary conditions on the blocked area (pzt+brass) UX=0, UY=0 and UZ=0
7. Solve modal analysis
This is the result of the first mode:
I'm so upset! I hope someone could help me!
Thanks!
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Dear all
I am trying to perform a modal analysis of a beam yet I am keeping recieving the following warning:
''The model dimensions in the solver unit system were determined to be very large. This may lead to numerical accuracy issues. Check results carefully.''
How to possibly fixe that?!!
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Have you tried changing the solver unit system to, for example, mm?
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I've done the Modal Analysis followed by Structural analysis in Ansys Mechanical & now I want to export my model in State Space Form. So that i can use it in Matlab.
Everything works well except loading of input vectors, I've followed guidelines from the link below:
example:
my query is that the given link provides accessibility to define DOFs in ux, uy & uz. However, I want to Include variable forces in x,y & z direction that will be controlled by my Control Algorithim in Matlab. how can I add forces as input?
The link also says if i don't define any input array it will define all the Loading vectors in the system as Inputs(by Default). But it's not working as it says.
Note: I've defined "Nodal forces" using "Named Selection"
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The page you requested is unavailable and will not be restored. For documentation please contact your software provider.
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Hey there!
I guess it should be straightforward, but i can´t find out how to extract the eigenvectors from a simple free-free (no bondary conditions) modal analysis in ANSYS workbench . Does anyone know how can i extract these eigenvectors from ANSYS workbench? Maybe using the Command window... I've searched in the folder with the output files and i can't find the eigenvectors anywhere. But i'm sure the program creates these at some point in the solution to obtain the mode shapes and natural frequencies.
Thank you all!
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Hello, this research is to execute a modal analysis of the functionally graded piezoelectric material rectangular plate and the results shown are from the ANSYS workbench. May be helpful for you.
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I created a finite element model of steel truss bridge and tried to do a modal analysis to generate the natural frequencies and mode shapes. Unfortunately, i don't get deformed shapes, but rather the SAP 2000 software will add more constraints at the middle spans of the bridge. Furthermore, negative Eigen values are also generated. Please i need help. Find the screen shot of the bridge before the modal analysis and after the modal analysis.
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The negative eigenvalues represent buckling modes and refer to the reversed direction of loads. have you considered changing the load's directions?
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I did two simulations with the same geometry one is Quad dropped (4 nodes/elements) and the other is kept (8 nodes/elements). Maintaining all bc and settings the same , i get different frequencies in modal analysis in Ansys. For example 100 Khz vs 125 Khz , the shapes are correct but the frequencies is totally different .
Any ideas why is that ?
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ANSYS Workbench in the modal analysis of a structural frame. The structure is imported as solid body NOT surface model.
The contact is detected automatically by the program. Also you can assign the same contact region with different contact types by selecting the contact bodies manually and assign them the contact type as you wish(Bonded, Frictional). There is drawback in workbench is that you can't select the element type directly "Only through command snippet" or you can play with the meshing options "For example you can convert the "20 Nodes 3D- SOLID186" to the "8 Nodes 3D- SOLID185" by setting the midside nodes to "Dropped" or Element order to "linear"in newer ansys versions.
So for contact you can do the same by using with the mesh options.
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I am performing modal analysis of a Free-Free 2D beam (1 x 0.02 m) and I am trying to compute the mode frequencies and mode shapes up to the lowest 32 frequencies. Here I observed that the mode shapes started to repeat, such as
1,2,3 are rigid modes
4,5,6,7,8,9,10,12,13,14,16,17,19,20,21,23,24,26,27,29,30,32 are elastic modes
11,15,18,22,25,28,31 modes are similar to 2,4,5,6,7,8,9 modes
I used FENICS and verified in ABAQUS and got the same results.
Is this a phenomenon or a simulation error?
I will be using this mode data to perform Fluid-structure interaction using the modal expansion method.
Can anyone suggest what is criteria for choosing the number of mode shapes for this method?
Should I include the repeated modes in the modal expansion method?
Also, can anyone refer me to some books or articles related to the modal expansion method and modal analysis?
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Dear Chaitanya,
The load may be arbitrary, but you need to know the frequency content of the excitation. You need to estimate the frequency content from a preliminary CFD analysis. You may consider your structure rigid first to have an idea about the frequency content of the excitation or you may specify a shape for your elastic structure to do this. After this step you can identify which modes of the elastic structure will contribute to the result. This is a must step to identify the modes to be used.
Best,
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Dear all,
In ansys mechanical, I performed static structural analysis with adaptive convergence and I want to retrieve the same convergent mesh automatically generated tu use it in modal analysis. Is there any way to do it? Thanks.
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Nice question! You can try linking the static structural analysis (adaptive converged) to the modal analysis. My additional suggestion is that; apply real tests and adjust the mesh layout based on test results. If you can't do this, validate with articles that apply real tests.
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Hi All
I want to simulate an ambient vibration to be used as an input to a Numerical Model. I want this to represent the vibrations occurred to the building through the wind, vehicle/human movement, etc.
The sole purpose of simulating this ambient vibration is to input it to the numerical model in order to find out the natural frequencies, mode shapes and damping ratios of the building through Operational Modal Analysis
Thank you
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Hello Navindra,
If your only goal with this simulation is to find the eigen modes/frequencies and damping, a gaussian white noise bandpass filtered to the range of your suspected natural frequencies of interest (e.g. number of floors/10 estimates f0 for buildings) will do the job without fail. The magnitude of the white noise must be sufficient to excite your structure. (usually ~0.2g suffices in my experience for buildings) I suggest using Automated FDD for OMA to ease your calculations.
I hope this helps! Good luck with your research :)
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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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Hello,
For the last weeks I have been comparing modal analysis results for ANSYS, Inventor and analytically. For solid square and rectangular beams the results were quite similar, however, now that I use a C channel beam the results from ANSYS are totally wrong. Does someone know why this is? I have sent the command file, which I used in ANSYS mechanical APDL.
The formula used is f = (K/(2*pi*L^2))*sqrt(E*I/m)
From the formula the first four eigenfrequencies should be: 14.8, 93.08, 261.03 and 511.91. Inventor showed similar results, but ANSYS fourth eigenfrequency is only 90.11. If someone knows why this is the case, please tell me. I have already checked the weight, Moment of Inertia, length etc.
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Fedde Engelen If you want to treat the cantilever as a plane problem, the boundary conditions for UX,RX,RY are missing at nodes 2-784.
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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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"I have mode shapes and eigen frequencies from FE simulation. I have modal matrix and eigen frequencies corresponding to each mode. The mode shape is mass normalized. I am required to correct each mode for damping, i.e to add 1% modal damping to each mode. do you know how to add 1% damping to each mode. what should I do
Structural dynamics, Vibration, modal analysis
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The mode shape is not altered by hysteretic damping distributed homogeneously as a material property, it is simply the amplitude of the response to a dynamic force input which is reduced. This corresponds to a slight out-phasing between the applied force and the response acceleration/displacement, that requires, as rightly said by Alessandro, to use complex (vectorial) formulations.
If the damping model is not hysteretic (i.e. an intrinsic property of the material's young modulus), the physics turn far more complex and the mathematisation non-linear (e.g. viscous or frictional damping). In such cases the mode shape is altered and reveals generally amplitude-dependant...
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Hi, I am currently simulating pipe using modal analysis.
How can I export a text file that includes node location for Ansys 2020? Currenlty if I export the total deformation text file it only shows node number and deformation value. Kindly help!
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macros?
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Dear researchers
I modeled a specified structure with the same properties (mass and stiffness) (apply mass through defining nodal masses instead of calculating from gravity loads) in both Etabs and Sap2000 software and observed differences between the modal results of the structure (the value of natural frequencies). After various investigation (all factors affecting the stiffness of the structure), I have found that the mentioned differences are arisen from difference between the mass matrices created in Sap2000 and Etabs software.
Since, unlike SAP2000, it is not possible to output the mass matrix and nodal masses of the structure in Etabs, the reason for the difference in mass matrices cannot be identified.
In your opinion, what is the reason for the difference in the mass matrices created in Sap2000 and Etabs? And which one is more correct?
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Hi,
OAPI is useful to extract the stiffness and mass matrices from SAP2000.
Let try to do it in Etabs to recognize where are the differences.
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Hi,
For a rectangular plate of a given dimension and material, we have the expression for modal density as n = Area/(4*pi) *((rho*h)/D)^(0.25) , where D is transverse stiffness of the plate, rho is the density of the material, h is the plate thickness.
This gives a constant number.
1. What value of modal density is high?
2. If I look at the FRF for a rectangular plate, I can see that as the frequency increases, the resonance peaks are closer to each other. Is this related to the modal density? If yes then why is modal density a constant number?
regards,
Vachan
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I think the modal density starts to become high when modes tend to overlap at their -3 dB skirts and it becomes difficult to read out half max values from each mode.
In reality the answer depends on the use. For having enough modes within a 1/3rd octave band you may have one answer, lets say > 5-6 modes. For octave bands it might be less dense and it depends on the frequency as the bands tends to get wider at higher frequencies.
The density grows faster in 3d systems like room modes than 2D, like plates, or 1D like wires or pipes.
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For IUGS classification(QAPF) modal(volume% ) analysis of the plutonic rock is necessary.
In thin section how can it be done?
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You can use an image analysis software, like ImageJ, on a whole thin section scan/photo stitching. You need to carefully separate each phase based on its color. This is simple with femic/sialic minerals (adjusting contrast/brightness may saturate all sialic to white and all femic to black), but can be tricky with minerals of similar color (e.g. feldspars). In the latter case, you may need to manually contour grain by grain, before estimating area% with an image analysis software.
There are also techniques that can map an area of a thin section to estimate composition and area% of phases (EBSD, X-Ray maps, EDS-SEM), but I don't think they can be performed with high accuracy on entire thin sections.
I strongly recommend ImageJ. There are also many tutorials on youtube.
If you need a quick estimate and don't require high accuracy, another option is to use visual estimation diagrams like the one attached. This is an example for sedimentary rocks. I am sure that you may find something better for plutonic rocks.
Best,
Samuele
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Hello,
I am trying to export Mass and Stiffness matrices using Ansys Workbench via Modal analysis modular, with APDL commands inserted. However, there is no .matrix file generated. Could anyone please tell me what I'm doing wrong? The figure illustrating the procedure as well as the commands is attached.
Thank you.
Rui Wang
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Dear Wang,
You can use this command to Export the stifness, damping and mass matrices of the structure as txt file. You should add a command under solution part of your Analysis. Afterward, you will Kdemse.matrix, Mdense.matrix and Ddense.matrix files in the path of your Analysis such as ......_files\dp0\SYS\MECH
! Stiffness
*DMAT,MatKD,D,IMPORT,FULL,file.full,STIFF
*PRINT,MatKD,Kdense.matrix
! Mass
*DMAT,MatMD,D,IMPORT,FULL,file.full,MASS
*PRINT,MatMD,Mdense.matrix
! Damping
*DMAT,MatCD,D,IMPORT,FULL,file.full,DAMP
*PRINT,MatCD,Ddense.matrix
Best regards,
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The displacement mode shapes are determined by running the modal analysis on AnsysWorkbench. How we can have the strain mode shape and the modal strain energy using ansysWorkbench. Example: cantilevered beam.
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for example, 750 mmx12x12. E = 2E9 Mpa, poisson ratio 0.3. My question is: what are the steps to follow on Ansys Workbench in order to have strain mode shape and strain modal energy?
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I am working on a brake squeal problem and for the unstable modes I am getting a pair of stability readings in Hz (where one is positive and the other is of the same magnitude but opposite sign) which I dont know how to interpret. I understant that the real part of the eigenvalue obtained is the represents the stability but I don't understand how you can show this as a reading in Hz.
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Hi Carlos,
Are you using ANSYS?
FEA codes have different ways to handle damping in modal, harmonic and transient analysis, and i'm sure it's possible to switch from one form to another.
The form you're mentioning is the complex one, and it's by far the hardest to make sense of in physical terms. I usually try to express the damping as a modal damping ratio (or damping factor), or as a exponential decay or as a logarithmic decrement. You should be able to find the transformation equations in vibration textbooks.
Once you get the complex eigenfrequency, you have a real part (stability, i.e. related to damping) and the imaginary part (the oscillation frequency). The mode is said to be dynamically stable if the real part is negative.
Hope this helps.
Best,
Maurizio
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Dear researchers
The modeling of P-Delta for the structural elements (specially columns) reduce the stiffness of the structure and consequently increase the period of vibration. Now, these changes are negligible when the structure and P-Delta effects are modeled in OpenSees software. Instead, the mentioned changes are more noticeable when modeling is perform in the softwares such as Etabs and Sap2000.
How to apply P-Delta effects is the reason for difference changes due to P-Delta modeling in different softwares.
Now with all these interpretations, what is the most accurate method for modeling P-Delta effects? In other words, In which software are stiffness changes calculated more correctly due to modeling of P-Delta effects?
As you know, In Etabs and Sap2000 software, P-delta effects are applied to the structure by defining a gravity load combination, while this procedure is performed in OpenSees using geometric transformation (second-order P-Delta effects) and its effects on stiffness matrix.