Log In. Are you new to this forum?Weld Plate Analysis of T-Joint using Abaqus/CAE
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Are you an Engineering professional? Join Eng-Tips Forums! Join Us! By joining you are opting in to receive e-mail. Promoting, selling, recruiting, coursework and thesis posting is forbidden. Students Click Here. Related Projects. Hi guys I'm a beginner student in Abaqus. I am modelling a quarter of a plate with a circular hole. I have to apply a load on the quarter of the plate, then study the stress and the mesh convergence. My supervisor asked me to apply a symmetry boundary conditions on this model i.
What does this mean, and how can apply these boundary conditions on the quarter of the plate. Please see the attachement. Thank you Frd. You should figure this out on your own. Read the documentation regarding boundary conditions.The specimen is subjected to pure Mode I loading. In some cases distributed pressure loads are applied to the cracked element surfaces as the crack initiates and propagates in the specimen. The results presented are compared to the available analytical solution.
In addition, the same model is analyzed using the XFEM -based low-cycle fatigue LCF criterion to assess the fatigue life when the model is subjected to sub-critical cyclic loading.
A plate with a circular hole is studied. The specimen, shown in Figure 1has a length of 0. Equal and opposite displacements are applied at both ends in the longitudinal direction. The maximum displacement value is set equal to 0. To examine the mesh sensitivity, three different mesh discretizations of the same geometry are studied. Symmetry conditions reduce the specimen to a half model. The original mesh, as depicted in Figure 2has plane strain elements.
The second mesh has four times as many elements as the original one, while the third mesh has sixteen times as many elements as the original one. In the low-cycle fatigue analysis, two steps are involved. A static step is used to nucleate a crack at the site of stress concentration prior to the low-cycle fatigue direct cyclic step, in which a cyclic distributed loading with a peak value of 1. Three different mesh discretizations of the same geometry are studied.
The second mesh has twice as many elements as the original mesh, while the third mesh has four times as many elements as the original mesh. The response of cohesive behavior in the enriched elements in the model is specified.
The maximum principal stress failure criterion is selected for damage initiation, and an energy-based damage evolution law based on a BK law criterion is selected for damage propagation. Figure 3 shows plots of the prescribed displacement versus the corresponding reaction force with different mesh discretizations when the XFEM -based cohesive segments method is used.
The figure clearly illustrates the convergence of the response to the same solution with mesh refinement. A plot of the applied stress versus the variation of crack length is presented in Figure 4 and compared with the results obtained by using the XFEM -based LEFM approach as well as the analytical solution of Tada et al.
However, as indicated in this figure, the crack initiates i. This value is in close agreement with the stress concentration factor of 2. In addition, the results in terms of crack length versus the cycle number obtained using the low-cycle fatigue criterion in Abaqus are compared with the theoretical results in Figure 5.
Reasonably good agreement is obtained.
Plate With a Hole
Subroutine for a user-defined damage initiation criterion with two different failure mechanisms. Problem description A plate with a circular hole is studied. Results and discussion Figure 3 shows plots of the prescribed displacement versus the corresponding reaction force with different mesh discretizations when the XFEM -based cohesive segments method is used.
Louis, Missouri, Figures Figure 1. Model geometry of the plate with a hole specimen. Figure 2. Original mesh of the half model for crack propagation in a plate with a hole. Figure 3. Reaction force versus prescribed displacement with different mesh discretizations XFEM -based cohesive segments method. Figure 4. Applied stress versus variation of crack length: XFEM and analytical solution.Author: Benjamin Mullen, Cornell University.
Problem Specification 1. Geometry 3. Mesh 4. Physics Setup 5. Numerical Solution 6. Numerical Results 7. To add deformation to the solution, first click Solution to add the solution sub menu to menu bar.
It should appear in the outline tree. In the details view window ensure that the Orientation is set to X Axis. To add the polar stresses, we need to first define a polar coordinate system. This will create a new Cartesian Coordinate System. To make the new coordinate system a polar one, look to the details view and change the Type Parameter from Cartesian to Cylindrical.
Now that the polar coordinates have been created, lets rename the coordinate system to make it more distinguishable. Right click on the coordinate system you just created, and go to Rename. For simplicity sake, let's just name it Polar Coordinates.
Now, we can define the radial stress using the new coordinate system. This will create "Normal Stress 2", and list its parameters in the details view.
We want to change the coordinate system to the polar one we just created; so in the details view window, change the Coordinate System parameter from "Global Coordinate System" to "Polar Coordinates". Ensure that the orientation is set to the x-axis, as defined by our polar coordinate system.
Abaqus plate with hole example - help
Now the stress is ready. Now let's add the theta stress. Now, change the coordinate system to Polar Coordinates, as you did for Sigma-r.Thanks for the video.
Basic knowledge, maybe, but a nice video as a reminder for not to take automatic meshing for granted or correct. A question: is it possible, using MidasNFX, to smoothen the mesh? The elements in the extreme corners of the patches now have quite a large slenderness? I missed that typing my previous reply. But the question still remains: how to smoothen a mesh in MidasNFX? Still interested in your reaction.
Unfortunately, there are no automated ways to do that in NFX… midas NFX is a bit limited compared to hypermesh for example. Hey I have a question regarding continuity…. We all know that the shape functions of order one i.
C1 are plate and shell elements where continuity in slope is also required…. If thats the case then 20 node brick elements… would be C1 continuous also… correct me if I am wrong. Hi Manisha, I hope I am not wrong, but I think higher order elements are better because they are C1 continuous which mean that their derivate is also continuous. To be verified though…. Higher order elements are not always better.
Do some reading on extrapolation of results at nodes. Thanks but what is the real impact on results? Work on shells meshing? Well Julien, better mesh means, of course, more accurate results. The mesh has to be good in areas where stresses are changing fast, otherwise, the mesh cannot capture the real behavior of the model you are trying to simulate. For all practical purposes, creating a mapped mesh for real geometries is quite unlikely.
What is the impact of the automatically created mesh on the results? Please run and share with us the results as a comparison. With all things considered other assumptions compared to realityI may say that the effect of the automated mesh is insignificant when compared to the mapped mesh.
The action is all happening around the holes, so in my opinion, the automated mesh had created nice uniform elements around the holes. Now perhaps in-between the holes there could be better formed elements shapes, but that could possibly be fixed with some quick meshing controls around the holes. In my opinion, the automated mesh is the better way to go; rather than be a master mesher, I would rather be considered a master engineer.
Meshing is just a tool as a means to those ends. I approve totally your view. FEA is a system in which all parts interact together. I think that the worst thing to do is to place some things into precise categories and to consider that the assumptions we made for those categories are always valid.
So, my purpose here is more to train about the best practices available because they may come handy sometimes.Log In.
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Already a member? Close this window and log in. Are you an Engineering professional? Join Eng-Tips Forums! Join Us! By joining you are opting in to receive e-mail. Promoting, selling, recruiting, coursework and thesis posting is forbidden. Students Click Here. Related Projects. I have a model of a deformable plate with a hole and a bolt that I have modeled as a rigid body.
The rigid body bolt is fixed to the 'ground' by fully restraining the reference point located at its center.
The plate is restrained on its far end in z, rx and ry. The corner node on this edge is also restrained in y. Clearly, there can be no rigid body motion other than x-translation of the plate.
However, I want the plate to move in x and ultimately come into contact with the bolt.The aim of this paper is the development of the two different numerical techniques for the preloading of bolts by the finite element method using the software Abaqus Standard.
Furthermore, this paper gave detailed guidelines for modelling contact, method for solving the numerical error problems such as numerical singularity error and negative eigenvalues due to rigid body motion or the problem of the extensive elongation of bolts after pretension which is occurring during the analysis. The behaviour of bolted joints depending on the two different approaches of pretension was shown on the example of an extended end-plate bolted beam-to-column connection under the monotonic loading.
The behaviour of beam-to-column connection was shown in the form and moment-rotation - curves and validated by experimental test. Advantages and disadvantages of pretension techniques, as well as the speed of numerical models, were also presented in this paper. The bolted joints remain the most common connection method in the construction and machine design.
As a structural component, it is often considered the critical part of an assembly. Due to the high cost of experimental test, development of the computer programs, especially the finite element modelling, gives accurate prediction of the real behaviour of such type of joints.
There are a lot of articles [ 1 — 7 ] which analyzed the behaviour of the assembly with bolted joints.
It takes into account the preload bolts in the joint. Maggi et al. Guidelines for modelling the bolt load are available in Abaqus [ 8 ], using the bolt load method and the adjust length method. The numerical problems such as numerical singularity error, negative eigenvalues, rigid body motion, and problem of the extensive elongation of bolts after pretension are very often during the numerical analysis.
All of these problems may stop the procedure or give the inaccurate results of the analysis. Contact surfaces with edges or corners also may create the convergence difficulties. Sharp corners may produce the element distortion and high stress level in the contact zones. Selamet and Garlock [ 9 ] give the solution for edges and corner numerical problems, as well as the comparison of explicit and implicit solution techniques.
Two different approaches to preloading of the bolts will be presented at the first part of the paper. The first technique was standardized in Abaqus software and uses bolt load for preloading. During the numerical analysis, using standardized Abaqus technique of preloading, all of the aforementioned numerical problems may occur. In order to avoid these numerical errors, the second technique of preloading will be presented which uses the initial stress.
The comparison between these preloading techniques will be shown on the example of an extended end-plate bolted beam-to-column connection.
Results obtained from numerical analysis will be compared with experimental test according to [ 10 ]. Numerical modelling of bolted joints is carried out by using the following parameters: geometric and material nonlinearities of the elementary parts of the joints, bolt pretension force, contact between connected plates, washers and plated elements, bolt-shank and hole, and friction. The nut and the bolt-head were considered as a single body with bolt-shank together with washers on both ends of the bolt.
Threaded part of the bolt-shank and the extended length of the bolt beyond each nut were ignored. Hexagonal shape of the bolt-head and nut was replaced with a cylinder. The typical bolted joint is presented in Figure 1. The plated elements in the joint were meshed with 8-node first-order linear hexagon brick elements with incompatible modes C3D8I. Each of these elements has 13 additional degrees of freedom DOF when compared to the fully integrated elements C3D8providing superior performance in bending dominant problems without shear locking behaviour or zero energy deformation modes; see Figure 2 a.
The 6-nodded linear triangular prism element C3D6 was used to model the bolts. The details of the finite element mesh are shown in Figure 2 b. Structure mesh technique is used for all parts of assembly. Previously, the bolts as well as bolt-hole region are quarter-partitioned. Then, they mesh with 16 elements around the circumference.
Numerical results are highly sensitive to the contact properties between components of the joints and the preloaded bolts. Small sliding surface-to-surface discretization method was considered for all the contacts.Abaqus Users. Search everywhere only in this topic.
Advanced Search. Classic List Threaded. Abaqus plate with hole example - help. This post was updated on. Re: Abaqus plate with hole example - help. This post has NOT been accepted by the mailing list yet. Could you attach your input file in this post I think I might have something for you.
In reply to this post by Mekanikal4. Hi there Thanks for the reply, I greatly appreciate it. The link obviously contains information directly linked to what I am concerned with. So the reason the graphs aren't exact is due to errors in the FEA? But what are those errors?
That's part of my confusion. Something to do with the way calculates? After further research I see that the whole free space node thing is to do with the fact that the node right at the hole and at the edge of the material are open to air in effect and so have nothing to push against.
Even reading what I have just written myself sounds so elementary but I don't know how better to explain it. So, judging by the graphs shown in the link, the bottom two match up with my CPS8R graphs. But the top two don't match up with my CPS4 graphs. Why is that?
What is it that causes such drastic differences in the graphs? I also tried to find info regarding the shear stresses in this problem S Am I correct in saying the shear stresses along the horizontal x-axis should be zero? So there shouldn't be any S12 values at all?
How does that work? First thing, the reason why the graphs shown in the link are better than what you're getting is because of the meshing. The mesh used in that model was sufficiently refined, and was following the physics of the problem i. The reason you wont get exactly analytical result is because of element displacement based formulation. For details read any good fea book; e. But the shear stresses? How is it that they should be zero? Is that to do with the fact that there is a condition which means the load is applied evenly across the top, no point load causing rotation?
So as the load is applied and the plate deforms, the entire plate deforms evenly? I understand that you're quite curious to understand what is happening See my friend there is no free lunch in this world.
With learning, there is no simple way. So I would expect you to go through if not thoroughly but at least simple glimpse. And I am disappointed to assume based on your question that you're looking for an easy answer.