## Stress at integration points or at nodes ?

## Stress at integration points or at nodes ?

(OP)

Think of that you have a 3-D model and you simulated some forces acting on it. And get a Stress contour ? it should have normally the most accurate results at integration points (or not)? but then you have 4 integration points so should we extrapolate the results to the nodes ? or for this element which value should we take ? how about the stress results at nodes arent they realistic ?

## RE: Stress at integration points or at nodes ?

The element shape function is used to extrapolate the integration point stresses out to the element nodes - these are in a useful location like a fillet radius free surface or a hole edge.

Adjacent elements combined with their shape functions will predict different stress values at their common nodes. The question then arises which stress do you believe? Most FE packages average the stresses for each element at the node.

If the unaveraged stresses are within a few percent of each other I go ahead and use averaged nodal stresses. If the unaveraged stresses are significantly different, I use peak unaveraged stresses, or refine the mesh to get a better result.

IN ALL CASES, UNAVERAGED NODAL STRESSES MUST BE CHECKED BEFORE USING AVERAGED NODAL STRESSES.

WHEN USING SHELL, BEAM, OR ANY OTHER ELEMENT FOR WHICH RESULTS ARE PRESENTED IN SOME FORM OF LOCAL SYSTEM, AVERAGED NODAL STRESSES SHOULD NOT BE USED UNLESS YOU REALLY KNOW WHAT YOU ARE DOING.

I can't count the number of expensive fatigue errors which I have seen as a consequence of averaged nodal stresses.

Amen

## RE: Stress at integration points or at nodes ?

It seems intuitive that this statement is true in a nonlinear analysis because you use the stresses at the integration points in the nonlinear iteration (say a Newton Raphson iteration is used to solve the nonlinear equations). However, the 'truth' of this 'stresses are most accurate at the integration points' statement doesn't seem so obvious in a linear analysis, since all you are doing is solving Ku=f, and there is no need to calculate the stresses at the integration points until you perform the post processing.

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

If "accuracy" means "correct for the linearized, discretized system of equations that was created to approximate a solid body deforming under load", then, yes, the stresses are correct only at the integration points, and are extrapolated to the edges of each element. A whole bunch of assumptions are built into the FEA element models, some of which imply that the stresses change only modestly from element to element. Thus, we conduct convergence studies, to show that stresses in a region of interest converge to a single value as the mesh is made finer and finer within the region.

If "accuracy" means "correctly reflecting the real-world solid mechanics which are being modelled", then we've got a whole 'nother can of worms to open...

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

Opethian, do it my way and all will be well with your world.

Do it any other way and learn to enjoy confusion, fear, late nights........................................

If you want to know why it is the way it is then read a book on FE theory, it's about masters level but not that bad if you are mathematically competent.

Alternatively if you have ABAQUS for example you can get it to print results at integration points, unaveraged at node, averaged at nodes etc. This is a good way to understand the differences. If you combine this with a good postprocessor like PATRAN you will also be able to look at the many different ways in which stresses can be dispayed as both contours and numerics and relate this to the numbers in your printout file.

## RE: Stress at integration points or at nodes ?

FEA is undergrad level course in my country and do not worry i passed several FEA courses . My question was about post processors mostly. If you have used hyperwork and abaqus post processors you will be confused soon for sure you will be confused about the programs. Best is to write your own post processor i guess.

## RE: Stress at integration points or at nodes ?

that's why (well, one reason why) there are canned programs. everybody knows (ok "knows") how they work and so it's "just" the application of the software to the particular problem that's the issue.

i agree you need to understand what the post processor is doing. run some simple patch tests. on a different level, i don't particularly care how the FEM extrapolates from the integration points to the nodes. i know that the stresses reported at the nodes are accurate for some cases and very inaccurate for others, depending on the program and the loading; this leads to increasing the mesh density to resolve the known issues (eg, put a single CQUAD4 in bending, it doesn't like it, so we don't do it !).

only MHO ...

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

No, I've never used ABAQUS viewer. I gave PATRAN and ABAQUS as an example. ABAQUS results output type control is outstanding (meaning you can see exactly what is going on in the .dat file if you wish), and PATRAN can display combined numeric and contour plots at averaged, unaveraged, and integration points - just what you want if you be sure what's going on.

gwolf2

## RE: Stress at integration points or at nodes ?

Since it isn't often you know the exact solution, I would define 'accurate' relative to the numerical convergence of the stresses--any quantity is more accurate at a (first) location relative to another (second) location if its numerical convergence is quicker for the first location relative to the second. Now maybe this is dependent on your particular software, the way in which stresses are calculated--but the statement "...the stresses are correct only at the integration points, and are extrapolated to the edges of each element...." is incorrect in general (unless you can present proof otherwise). Stresses are calculated from the strains, which are just the spatial derivatives of the displacements (see, for instance, Szabo and Babuska, Finite Element Analysis--which is up to $190 on Amazon! When did they start gilding the pages?). Again, in a linear solution, there is no need to calculate stresses at the integration points, until you are finished solving the Ku=f equations. You can calculate stresses ANYWHERE in any element using just the solution vectors and the spatial gradients, and are not restricted to calculating stresses at just the integration points and extrapolating.

Unless of course that's just a restriction (stresses are computed directly only at integration points and are extrapolated to edges) your FEA software places on your analysis.

## RE: Stress at integration points or at nodes ?

"the statement "...the stresses are correct only at the integration points, and are extrapolated to the edges of each element...." is incorrect in general (unless you can present proof otherwise). Stresses are calculated from the strains, which are just the spatial derivatives of the displacements "

Yes, but no. Saying that the strains are calculated, and not the stresses, is pretty specious, since they differ by only the constitutive model (which for linear problems is a linear factor difference). The strains are no more calculated (again speaking of linear FEA) than the stresses are - both are buried in the FEM formulation, and thus are calculated implicitly whenever the K[x] = f matrices are solved. Backing them out and printing them or coloring the charts is more of an excersize in computation, not calculus. The calculus came in formulating the elements' stiffness matrices.

More importantly, in linear FEA, the displacements are known at the nodes, which are at the boundaries (generally the corners) of the elements. But the strains (and thus stresses) can only be calculated between the nodes, since the strain is derivative, i.e. the difference in displacements between two or more nodes. All element formulations I've ever seen and/or derived, have an implicit (assumed) stress distribution across the faces of the element (typically constant for linear elements), which allows the stresses and strains to be derived and the element stiffness matrix formulated. The result is that the stresses are known only by derivation from the displacements, which in linear models implies that the stresses or strains are "known" (in the sense of how they were modelled) in the center of the elements.

This basic idea, that the nodes (where the displacements happen) and the interior of the elements (where the strain and stress happen) is why you need to refine meshes to keep the difference in stress from one element to the next, is what we should all take away from this.

Textbook reference would be

The Finite Element Method in Engineering Science, Zienkiewicz, O. C., 1971, 1977. Might be a bit dated, but I doubt very much that they've changed how to compute derivatives in the last 40 years or so.## RE: Stress at integration points or at nodes ?

"This basic idea, that the nodes (where the displacements happen) and the interior of the elements (where the strain and stress happen) is why you need to refine meshes to keep the difference in stress from one element to the next, is what we should all take away from this. "

Should've been:

This basic idea, that the nodes (where the displacements happen) and the interior of the elements (where the strain and stress happen),

are different (one is a point in space, the other a volume, thus they physically cannot be the same)is why you need to refine meshes to keep the difference in stress from one element to the next, is what we should all take away from this.## RE: Stress at integration points or at nodes ?

This basic idea, that the nodes (where the displacements happen) and the interior of the elements (where the strain and stress happen), are different (one is a point in space, the other a volume, thus they physically cannot be the same) is why you need to refine meshes to keep the difference in stress from one element to the next

as small as practically possible to avoid errors, is what we should all take away from this....and if anyone knows how to avoid errors in posting, save iterations, please tell me.

## RE: Stress at integration points or at nodes ?

For example impose on an element a displacement field f whose degree is one order higher than the highest-order complete polynomial for the element interpolation. Then obtain the nodal dof by evaluating f at the nodal locations. Next if you seek locations in the element to evaluate the B matrix (used with the nodal displacements to compute the strains) such that the strain as calculated from f is the same as that calculated from B*d where d is the vector of nodal displacements, those locations will be at the Gauss integration points for lower order elements which are most commonly used and slightly different for higher order elements (which are seldom used).

## RE: Stress at integration points or at nodes ?

As far as stresses being superconvergent, I doubt that's even numerically possible, since stresses are derivatives of the displacements (which can be superconvergent, depending on the formulation), and anyone who has ever tried to compute derivatives of anything knows how crazy derivatives of apparently smooth data can look. I would be most appreciative if you would supply an actual reference, say a published paper, and not a statement from a User's manual for the FE software. I have accepted anything's possible, but it would be nice to have proof found in the open literature.

## RE: Stress at integration points or at nodes ?

How is the stiffness matrix derived for a given element?

To derive the algebraic equations for a given element's stiffness matrix, one must first solve implicitly for the stress and strain fields internal to the element. This is done by making certain assumptions about the boundary conditions (small displacements, linear constitutive model, constant tractions at the boundary faces, etc.) and then applying complex transforms to solve the stress/strain fields directly via numerical integration.

Thus, after actual load and displacement vectors are solved for, the strain and stress can be backed out from the stiffness matrix formulation as algebraic sums. The stress and strain were already solved for, by the person who computed the equations for the stiffness matrix. All FEA models report element stress tensors based upon the element stiffness model, and these stresses are calculated at, and are most "accurate" at, the integration points used to derive the element stiffness model. If you try and apply your own "stress calculation" to derive the stress field from the solved displacements, you will have a result that is less accurate than using the already solved equations to find the stresses at the integration points that were used to derive the element stiffness.

So, fine, Mr. Prost, you are technically correct in the statement "stresses are calculated AFTER you compute the displacement vector "u",". Yes, generally, the stresses are computed after first inverting and solving the Kx=f matrices.

But, the rest of your statement "...by computing, as you know, the spatial derivatives." is not correct, at least not for linear solutions; the stresses in a linear FEA again, are found from implicit equations, not by recomputing the difference equations that you've already solved by integration. So, I and others, are also correct. And the reference I gave is a textbook, one of the original ones for FEA, that describes the above in excruciating detail. I should know, as I used it for many years to write and solve my own FEA codes.

## RE: Stress at integration points or at nodes ?

The superconvergent-like behavior of stresses at Gauss points for Q4 or Q8 elements is shown to be true in many textbooks..for example look on page 231 of Cook fourth edition. My example gave the reason for this and was actually taken directly from Cook.

## RE: Stress at integration points or at nodes ?

Which Cook book are you referring to, is it :-

"Finite Element Modeling for Stress Analysis"

or

"Concepts and Applications of Finite Element Analysis" ?

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

To illustrate, I think I don't lose any generality by describing how the finite element matrix is constructed in 1D, for rods. We are solving Ku=f; 'u' is the solution vector, the displacements, we are trying to compute. 'K' is the stiffness matrix. 'f' is the force vector, which represents how we have loaded our body. In 2D, those are point loads, and distributed stresses (which are specified, so they are known and don't have to be computed).

Assuming linear elasticity, stiffness matrix elements k(i,j) look like (AE/l), where A is the area of the beam, E is the Young's modulus, 'l' is the length of the rod.

Stresses and strains are then computed AFTER you compute the displacement vector 'u', and are therefore never needed for the finite element solution process until the post processing (you might need to compute stresses to compute Stress Intensity Factors, for instance).

Certainly these equations are derived from the equilibrium of stresses, and you are computing spatial derivatives to derive the stiffness matrix elements k(i,j), so you 'sort of' are computing quantities that look 'like' stresses. However, just because you are computing spatial derivatives, doesn't mean IMO that you are computing strains and stresses. Therefore, I still cannot see reasons why and no one has provided a reference as yet that shows that stresses at integration points are more accurate (as measured by their convergence rates) than stresses at the nodes. I'd love to hear the explanation and see a demonstration, though. There's a lot to be said for personal experience, that's for certain, but its hard to use that in say, an internal or external report, so that observation is of little use to many FE analysts such as myself, since it represents something akin to hearsay or anecdotal evidence.

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

To try and answer your basic question..."why are stresses more accurate at integration points than at nodes" consider the following....

Assuming that we are talking about isoparametric element formulations then most element stiffnesses are computed using Gaussian integration to compute the stiffness (K=sum(Bt . C . B * wt * vol))where B is strain/displacement matrix, C is stress/strain matrix, wt is Gaussian weighting factors (dependent on number of integration points used), and vol is the volume fraction (also associated with the number of integration points used). If you investigate the Gaussian integration method you find that the results of the integration are most accurate at certain points (the Gauss points) for a given number of integration points (in either 1-D, 2-D, or 3-D).

Thus the strains (and stresses) that are computed as eps=B . U and sig= C . eps at the Gauss points are more accurate (optimal) than those computed at other non-optimal points as the B matricies used in the stiffness formulation are then consistent with those used to compute strains...Indeed the B matricies at some points (not Gauss points) can give very bad results for some formulations while the matricies at the Gauss points give good results..In other words being consistent when computing strains at the same points as those used when forming the stiffness gives the best results.....

For elements using other formulations and/or integration methods the above might change but most numerical integration methods have optimal points at which the integrations should be performed.....

Hope this helped....

Ed.R.

## RE: Stress at integration points or at nodes ?

Chapter: Isoparametric Elements

Section: Stress calculation

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

Yes, I said so, have the degree and thesis in Aeronautical Engineering to prove it, have done so multiple times in the course of my career for multiple different problems, and cited the reference used, although I also extended said reference in several instances, to include thick-wall shell elements, elements undergoing rapid dynamic loading, and elements undergoing rotation and associated body forces...amongst other things. The reference given discusses the formulation of stiffness matrices using the techniques I described earlier, the use of Gaussian integration and errors associated with it (which EdR also describes), and how the use of Gaussian integration allows you to make statements regarding the solution's error bounds (whereas other techniques, such as Galerkin method don't allow for such treatment of error bounds beyond the technique of refinement and iteration). And your assertion that I don't know what I'm talking about irks.

Your next statement:

"To illustrate, I think I don't lose any generality by describing how the finite element matrix is constructed in 1D, for rods. "

Is specious. Yes, you can use "strength of materials" to compute 1-D and perhaps even 2-D (beam bending) element problems, then extend those to 3-D for frames and trusses. You can certainly ignore the rest of the FEA textbook, if you've ever read one, which I am beginning to doubt.

To extend the FEA method to even a 2-D general plane stress element (having both shear and normal tractions applied at the boundaries, having any general shape within certain limits) and then prove that the element is robust (capable of solving the problem under all possible combinations of shape and load) requires that you solve for the stress and strain field within it, in order to arrive at the stiffness matrix formulation. AFAIK, there are no shortcuts to that process, and the extension from 2D plane strain elements to 3D general-purpose elements is made using the same procedure.

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

BTW, I am intimately familiar with about 75% of the Bathe FE book, since I needed it for explanations of the nonlinear deformation coding in ADINA. I don't recall any such superconvergence property, but I admit I could just have forgotten.

I have also taken 2 classes from Szabo himself (he was one of my thesis advisors, so he supervised the research), and have a PhD in mechanical engineering (which means dissertation of course--"Large indentations of nonlinear viscoelastic materials, with an application to cell poking"), research in analyzing nonlinear deformations of viscoelastic materials. I wrote two research FE codes, in addition to writing material constitutive subroutines that interfaced with ADINA. Does that mean I think I know everything? Of course not. This makes me the opposite of that--I am confident in what I know, but I am acutely aware of my ignorance in other areas. I will check out the one reference cited here (can't get the other one at the library), and maybe even work out an example or two to check if the same superconvergence property holds with higher order (p>2) elements.

It's possible I have read this before, with a proof, about the superconvergence property, and perhaps I forgot what the explanation was. I am sorry if I offended you; the explanation for the superconvergence property isn't so obvious, and I'm like most engineers, I think--I don't automatically assume when someone makes a statement that seems so obvious to him/her, that the statement is still true.

One last thing--because you are using the principle of virtual work (which is expressed in integral form) to create you FE matrices, you are not strictly enforcing the equations of equilibrium (which are most often written with stresses), point by point. This is why the FE equations are called the 'weak form' as compared to the 'strong form' of the equations of equilibrium. Therefore just because are using the equations of equilibrium to formulate the FE matrices, doesn't mean that equilibrium occurs at every point (or for that matter, any point in particular) in your domain. This also leads me to the conclusion that my stresses aren't more accurate say at the Gauss points in particular compared to anywhere else just because I am computing spatial derivatives just at those Gauss points. Point wise equilibrium just isn't being enforced point by point, so why should anyone assume engineering computations any particular point are more accurate than at any other point?

## RE: Stress at integration points or at nodes ?

I tried to point out in an earlier post, that the idea that stresses are accurate only at integration points is strictly true only for the FEA model that was constructed, not for the "real world", and you can find different stresses for the same real world points by refining or modifying the FEA mesh. Which implies what I think you are stating. Or perhaps not.

The problem is worsened (and I'm not going to address it any further as you sound like the expert here) by the idea of nonlinear, viscoelastic materials, where you use (correct me if I'm mistaken, it's been 25 years) separation of variables to solve for the instantaneous elastic stress, then apply the Duhamel integral to solve for the decaying time history of the viscous part of the constitutive relations. The "stresses are correct at the integration points" may be applicable to the instantaneous/elastic FEA model solution (depends on whether you are using Gaussian integration or another method), but the time history part I think induces a whole 'nother set of error analysis ideas more applicable to those used in finite-difference (i.e. CFD) codes... and the combination of the two techniques in the same solution would make me throw up my hands.

## RE: Stress at integration points or at nodes ?

It turns out there is a strong theoretical basis for this claim, which appears to have been proved for a few element types, but not all. Starting with Barlow, "Optimal Stress Locations in Finite Element Models," Int. J. for Num. Meth. in Engrg, Vol 10, 1976 pp. 243-251; Barlow demonstrated with a few examples that show that evaluation of stresses is 'optimal' at the Gauss integration points:

1) for beam with cubic shape functions if the exact solution is a polynomial of order quartic 2) for 8 node isoparametric plane elements (what many of you call QUAD8, right?) if the exact solution is a cubic, and 3) for 20 node isoparametric solid elements if the exact solution is a cubic. Barlow did not specifically use the term "superconvergence." However, this paper is a very nice introduction to the way an engineer might try to show 'optimality' of a particular computation.

While those Barlow examples might seem to be too specific, in that we cannot assume a priori that the 'exact solutions' are polynomials, there are at least 3 theoretical treatments that demonstrate the validity of the "stresses are more accurate..." statement: Strang and Fix, An Analysis of the Finite Element Method, Zienkiewicz, The Finite Element Method (I have access to the late 80s edition) and Wahlbin, Superconvergence in Galerkin Finite Element Methods. The last one is interesting, but readable only if you have some idea what a L2 space is! Wahlbin uses a common technique--first guess that the convergence rates for a certain calculation are of a certain form, say the 'norm' raised to a power 'q', then prove that the power 'q' is less than 1, and therefore the computation is 'superconvergent'. (Proving that 'q' is some value is IMO the most significant contribution the mathematicians have made to the theoretical development of the finite element method.) In a review by DN Arnold, a couple of results are relevant to this discussion were made clear: in one dimensional problems, and continuous piecewise quadratic elements, the finite element solution (that is, the displacements) are superconvergent at the mesh points and element midpoints, while the derivatives (which you use to compute stresses) are superconvergent at the 2 Gauss integration points. The derivation of the superconvergent results for 2D and 3D is more problematic, and restricted to meshes with high degree of symmetry. Nevertheless, this is an excellent theoretical treatment of the subject.

The reason I think I 'missed' this result, that 'stresses are more accurate at the integration points,' is that my background is in p-version, which uses higher order shape functions, and generally does NOT use the isoparametric elements. It seems that this result "stresses are more accurate..." is confined mainly (but not exclusively) to isoparametric elements of the type most FE analysts use (which doesn't make the isoparametric elements inferior of course, it's just a property specific to those element types). I still think it is incorrect to state, though, that 'stresses are superconvergent at the Gauss points' in general for all element types; still it might be true, perhaps it just hasn't been proved for all element types as far as I can tell.

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

particularly if you consider that "most" of the time we average the stress at the node (using the different nodal stresses from the different elements) ? ... most of the time i use element centroidal stress.

## RE: Stress at integration points or at nodes ?

The 1D arguement is specious (meaning fallacious), because it can be argued either way. I.e. I can just as readily argue that you must know the stress before you can start computing the displacements, it's just a multiplication factors, so who is right - depends on which way you started. You choose to start from displacements, I could just as easily derive the element model starting from stresses.

But whatever. You found my reference, read it, and it said what I said it said.

BTW, you're welcome.

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

Following is a journal paper which discusses about the accuracy of stresses in FEA. You will see much more references in this paper.

The superconvergent patch recovery and a posteriori error estimates. Part 1: The recovery technique

O. C. Zienkiewicz, J. Z. Zhu

International Journal for Numerical Methods in Engineering, Volume 33 Issue 7, Pages 1331 - 1364

Following link also provides some more reference papers:

h

A.A.Y.

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

Is there a guarantee that because the stresses are superconvergent somewhere (for the sake of example, say they are superconvergent at the Gauss points) that equilibrium of stresses is obtained at those Gauss points?

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

"I still don't think you need to know anything about the stresses to compute the displacements."

i think you need to read what you type, or maybe think alittle more about what you're typing ...

you can't have displacement without stress (nor stress without displacement, generally). Basic FEA assumes a stress distribution in the elements, and determines displacements so that the resulting internal stresses balance the applied loads.

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

elements typically assume a displacement field (there were some early on that assumed a stress distribution). and after a bunch a math it solves for those displacements, and then derives strains and so stresses.

but i could contend that if you assume a constant displacment field then you are also assuming a constant stress field (assuming a constant thickness element).

## RE: Stress at integration points or at nodes ?

Think a bit more....For a linear displacement field (two nodes) the strains are constant (eps = du/dx, etc.); thus the stresses are constant......thus a constant displacement field would give zero strains and stresses......

Ed.R.

## RE: Stress at integration points or at nodes ?

stress based elements (actually the assumed field is a force distribution) were initially developed for the aircraft industry and they work very well in that environment because the important parameter across element boundaries is equilibrium of forces (i.e. shear flow). There is no enforcement of strain compatibility across element boundaries and it would be like having cracks along adjacent edges between nodes. Strain compatibility (enforced by displacement based elements) was considered more important by others in the field who were not necessarily dealing with semi-monocoque structure. Displacement based elements give better representation of stiffness, if used appropriately.

However, displacement based elements have their problems too, i.e. excessive stiffness of full integration elements, shear strain locking, hour glass modes, etc. Thankfully a number of commercial programs have developed "softening" features over time to minimize these issues.

## RE: Stress at integration points or at nodes ?

I also add my two-cents... for what they are worth of...

In FEA packages (and by consequence in FEA literature) there are A LOT of different formulations for the same "kind" of element. Some hypothesize a stress field, other hypothesize a displacement field, some have auxiliary nodes at the edges' midsides, other have auxiliary nodes "inside" the element, and so on and so on.

However, for every 3D element I am aware of, there is always an hypothesis made on the (internal)force/displacement fields. The undirect demonstration of this is, for example, when you launch the solution with ANSYS: the first solver message before starting the equilibrium iterations is the output of a "force norm", determined on the basis of the [F] matrix AND the force-field coming from the elements' formulation. If stress/force didn't come into play in the computation of the equiibrium, there would be no way to decide the solution has come to convergence (i.e. there are simultaneously an equilibrated force field and a compatible displacement field).

Even if you use a direct solver, the internal [u] matrix depends upon the element formulation, for the exact same reason: you MUST give an equilibrated force field as a response of a compatible displacement field, and vice-versa.

So, IMHO, to return to the original question: the Gaussian results as regards stresses are <generally> considered "the most accurate" from a numerical point of view simply because the element formulations are <generally> based upon these points and not upon the corner nodes. OK, there is also to say that the positions of the Gaussian nodes DO ARE calculated AS A FUNCTION OF the nodes' positions, so there is IMHO no point in saying that they are "more realistic" or not.

Just to make another example, in the case of the components/assemblies I usually analyze, the UNAVERAGED nodal results, by someone considered as a "must-be", are complete rubbish. I know of other situations where, instead, averaging like I do may lead to dangerous under-estimations.

Literature makes a good work in telling which kind of results, and in which applications, should be sought from the integration points, which instead from the nodes, which from the overall element, and so on.

AFAIK, there is no univoque answer to the original question.

"Gaussian better than Nodes? It depends..."

Regards

## RE: Stress at integration points or at nodes ?

In my experience of displacement based elements I find that nodal extrapolation of element stresses give higher stress values for many structural applications. Although the mathematical correctness of stresses and strains at gauss points is true ... it really stops there when it comes to the interpretation of a continuous structural domain. Higher stress levels are usually associated with outer and inner structural boundaries (however, it is load type dependent). For 2D elements it is imperative that nodal averaging is only carried out for similar materials, and similar thicknesses, etc....

Nodal extrapolation for linear material analysis is appropriate and more prudent. However, non-linear material analysis is where nodal averaging becomes more complicated. Again, since the stress/strain levels are correct at the gauss points it could mean that nodal extrapolation will generate stresses in excess of the non-linear material definition described in the data file (over and above those for "true" stress adjustment). I feel that in non-linear material problems it is more appropriate to use gauss point values for stress interpretation but nodal extrapolation for interpretation of strain. Strain is a better result to interpret in non-linear material problems.

For 3D problems, again boundaries give the highest stress levels (in general). Extrapolation to nodes is again more appropriate (conservative) but it may prove more beneficial to "skin" the solid boundary with 2D elements and only plot the results for the "skin" elements.

## RE: Stress at integration points or at nodes ?

outpost11 - could you please expand on or explain your reasons for this statement?

## RE: Stress at integration points or at nodes ?

A useful approach for interpreting stress levels on the boundaries of 3D solids is to "skin" the boundary with 2D face elements. However, you do need a pre-processor (like PATRAN) to achieve this if it is a complex solid. The idea is to create 2D face elements on the exposed faces of the solid elements i.e. QUAD on the exposed faces of a HEX model, or TRI on the exposed faces of a TET element model. The nodes on the boundaries must be compatible i.e. use TRI6 with TET10 etc. One should never use a TET4 model other than some preliminary guideline analysis. Assign some very small thicknesses and/or material props to these "skin" elements without ill-conditioning the problem. Remember, you must have strain compatibility between the 2D and 3D elements so the stress levels in the 2D elements will be what you would expect to find on the surface boundaries of the 3D solid elements.

## RE: Stress at integration points or at nodes ?

I have always used the same material property for the 2D elements as the solid elements but small thicknesses for the geometric properties. If you use a different material property you need to make adjustments to the thicknesses to arrive at the same material constitutive equation (stress/strain relationship) which is messy and not recommended. It is the strains that match to the solid so the same material with small thicknesses would give the best result.

## RE: Stress at integration points or at nodes ?

But the main point of my original question was, why is this beneficial?

Assuming you are comparing surface stresses of the 3D solid where there are no applied pressures, no supports and no contact and thus only a 2D stress field exists, would you really expect the shell element stresses to yield more accurate stresses than those given by the solid elements?

## RE: Stress at integration points or at nodes ?

Having similar nodal boundaries addresses the compatibility issue. It is the solid that dictates the strain ... the 2D element goes along for the ride (small thicknesses) but it is a convenient way of outputting results.

You may find that this approach is useful for addressing stress concentrations (for fatigue problems) in complex solid geometry. If a 3D automated TET10 mesh is used there can be issues associated with nodal extrapolation and nodal averaging. Irregular shaped TET elements are common and the locations of some gauss integration points for adjoining elements may be inconsistent in relation to the boundary position. This can lead to poor nodal averaging or even un-averaged peak values.

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

My point is merely an approach to assist in getting stresses calculated on the boundary itself. That's all.

If can also be very useful in keeping output result files (and post-processed) much smaller in size. If one has a number of load cases that can surely add up.

Just a suggestion ... employed by many analysts out there!

## RE: Stress at integration points or at nodes ?

## RE: Stress at integration points or at nodes ?

However, I have also had difficulty at times trying to get this to work in non-linear problems (i.e. surface contact and large displacement) mainly due to the 2D elements away from the area of interest producing undesirable performance etc....

For complex CAD solids the local mesh refinement can get very challenging, especially with PATRAN.

PATRAN can give "unknown failure" for solid meshing. The most dreaded response in the program......

## RE: Stress at integration points or at nodes ?

I find that with the appropriate tools (mentioning no names)there is often little difference in speed between hand crafting the right hex mesh the first time and iterating a couple of TET meshes. We are of course talking about the 1-2% accuracy range here for fatigue. - and I don't hand craft everything these days, often using hybrid hex/tet meshes.

I remember when this was all fields.....................................

gwolf

## RE: Stress at integration points or at nodes ?

Were the fields arrayed as far as the eye could see?

## RE: Stress at integration points or at nodes ?