Which plate theories are used in Abaqus for composite plate modelling?

I am studying how delaminations in a composite plate affect its damping properties (by contact friction) and thereby to use for Structural health monitoring. Normally, in a metallic plate, no transverse normal stresses (S33) will be generated during pure bending. But, in a laminated composite plate, around discontinuities, S33 have been observed to be generated. But many plate theories neglect S33. But Reddys Layerwise theory and 3D plate theories consider S33 in the analysis.

I would like to know the theories used in Abaqus so that I can use it for my study.

Thanks

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Hi all,

I am using FRIC subroutine to calculate some parameters from CPRESS and NT11 (temperature) values at the contact interfaces. I want to visualize these calculated parameters in an ODB. Normally, outputs such as CPRESS or COPEN are defined as contact outputs and treated different than regular elemental outputs like stress or nodal outputs like displacement. Since FRIC subroutine loops only among the nodes at the contact pairs, it is not straightforward to write the calculated values to a state dependent variable (SDV) and visualize it (STATEV does not exist for subroutines that iterate on nodes).

So far, the closest I can get is to transfer the values to UFIELD subroutine through global variables and define a user-defined field in it. I can visualize these values by requesting the output 'FV' from the field output requests. However, my observation is that while visualizing, the values for the nodes at the interface are distributed to the whole element. Therefore, the visualized value is not the same as what is defined in the UFIELD subroutine for that node (it is exactly the half of it, since 4 nodes of the element has the value and for the other 4, the value is zero). Please check the picture "Visualization with UFIELD".

I want the results as they are for the CPRESS output (only at the contact interface). Please check the picture "Desired visualization form".

Is such a thing possible? How can I create a user-defined contact output?

Regards,

Ozan Celik

Dear All,

I hope all feel so fine with your simulations.

Recently, I am simulation a rolling process with a very high dynamic speed (around 5 m/s). Hence, the model is quite dynamic and not Quasi-static.

I use Abaqus Explicit for this simulation. The structure is composite material including several layers with different material properties. In this case, I am using ''Cohesive Surface'' definition as the surface interaction between different layers.

I have a big problem and there is an error during the simulation process so called:

''the ratio of deformation speed to wave speed exceeds 1.0000 in at least one element''.

I tried to find a solution for this problem such as:

Larger mesh ……. But the larger mesh decreases the solution accuracy.

Have you ever seen any solution for this case? In this respect, may I imagine to let me know please if you have already any hints or solution for this problem.

Thanks for your attention.

Kind regards,

Hassan

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Dear friends,

I am modeling contact problem with CEL method, which involves 2 kinds of materials (euler-to-euler contact) in euler domain and 3rd material in lagrangian (euler-to-lagrangian contact) domain. The modeling of contact between the euler and lagrangian is okay.

But the modeling of interaction of euler-to-euler contact seems to be lack of information in ABAQUS Documentation.

Does anyone know, how to set up the interaction property of euler-to-euler contact?

Is that possible to get contact information of euler-to-euler contact from ODB file, for example CPRESS, DSLIP or TEMP at the contact interface?

I am looking forward for your answers and help! Thank you very much!

Bingxiao Peng

from WZL RWTH AACHEN

How do i add a reinforcement beam to a door of a car at the front using the ls dyna contact card CONSTRAINED_RIGID_BODY ?

*DO,i,2,5,1

/SOLU

ANTYPE,,REST,,,0

w=i*(-8.01e-5)

DK,200, , w , ,0,UY, , , , , ,

DK,201, , w , ,0,UY, , , , , ,

SOLVE

FINISH

*ENDDO

This is a do loop in ansys apdl. I expect w to change every loop. However output window says i is an unknow parameter. Does anybody know what the problem is?

Dear everyone,

I am new to Abaqus and I have been struggling for the past 3 months about this issues I am facing on my modelling...

Basically I am trying to simulate an indentation experiment on a biological tissue, with glass capillary as the tool for indentation. However, the tip of the capillary is very small compared to the actual size of the tissue, i.e. diameter of the capillary tip is ~0.7cm, the diameter of the tissue under indentation is ~50cm and the length of the tissue (modelling the tissue as a cylinder now) is ~500cm. The real experiment shows that the indentation can easily go to 16cm with NO penetration. However, building a model to simulate this indentation has been a headache for a past few months. Here are the approaches I have tried until now:

1. I use a global seed size of 3cm, and choose Tet linear (C3D4) to mesh the cylinder. And build the geometry as close to the real experiment, where there are 2 walls against the tissue under indentation. Please refer to the attached for the experimental geometry. The material properties for the tissue currently is 'Homogeneous, Isotropic' with a Young's modulus of 0.8MPa and Poisson Ratio of 0.48 (to simulate something like rubber). The simulation can go up to 16cm indentation. However, as shown from 'LD' and LD1' pictures, the distortion is very local and the element under indentation has a excessive dostortion.

2. I have tried all the other meshing scheme available from Abaqus 6.12, including the Hex (linear and Quad), Tet (Quad), Wedge (Quad). All these schemes cause the solution failed after along time of running. However, theory shows that Hex (Quad) and Tet (Quad) are the best schemes to simulate and estimate the stress?

3. I have also tried to assign local seed size (after partitioning the cylinder) so that the density of mesh close to the indentation area is very high. Please refer to 'Small Mesh' and 'Small Mesh 1' for the results. These methods usually takes a long time to run and always fail during the initial time increment, i.e. failed for small indentation depth.

With the approaches I have tried, my questions are:

1. Why only Tet Linear meshing scheme can simulate the experiment, but not the Hex/Tet Quad as I suppose Hex/Tet Quad can do a better job normally?

2. Why it looks like an element size of 3cm is the only choice, increasing and decreasing the global seed size can fail the solution and assigning different seed size on different parts of the cylinder can also fail the solution?

3. Is there a better way for modelling so that the deformation is more GLOBAL rather than the current excessive local deformation on the element under indentation?

I would like to take the chance to thank you all in advance in helping out! Really appreciated!

Thanks,

Dave

Dear everyone

My project is about the adhesional friction between rubber and a rigid surface. In this case I have modeled a block of rubber which is sliding on a real rough rigid surface in ABAQUS.

Cohesive bonding assumption via ABAQUS Standard was then implemented to model the resulting adhesional forces during the rubber sliding. In order to study the effects of the surface roughness, the surface was simulated by an analytical function in a sinusoidal shape. Hence, there are some small vertical gaps between the surface and some parts of rubber during sliding.

Then, two kinds of cohesive bonding are generated during the rubber block sliding:

1. Tangential adhesion bonding: on the zones that there is no vertical gap between rubber and surface which cause the shear stresses in rubber contact patch

2. Vertical adhesion bonding: on the zones that there is vertical gap between rubber and surface which cause the normal stresses in rubber contact patch

I want to know which of the above mentioned factors are more important to analyzing the rubber adhesional performance in reality? To put it in another way, which of them is generally more stronger or persistent during the sliding?

I need actually some probable experimental achievements or verified technical results for this topic. Could you give me a favor please?

Thank you very much in advance.

Best wishes,

Abolhasan

Hi everyone,

My project is about the trolological properities of piston ring and cylinder linder. I have a question to ask you. I have build a 2-D half ring in Abaqus/CAE and I hope to obtain the stiffness matrix of the half ring. The Yong's Modules of the half ring is 105GPa and the poisson's ratio is 0.3. The boundary conditions are as follows: the leftest node is totally fixed and the other nodes on the edge are limited in the U2 direction (seen in .cae file). I have write some commands in the end of the .inp file and the stiffness matrix can be output in .mix file(seen in attachment).

As we know, the ring suffers radial contact pressure on the outside surface and the nodes' displacements along the outside surface are also radial. So I hope all the ring's analysis can be done under cylindrical coordinate system. I have looked up the manual, it seems that Abaqus just has one global Cartesian coordinate system and not be changed. Although we can build new cylindrical system or spherical system, they are all local system and the basic coordinate system would still be Cartesian.

My problem is: how can I obtain the stiffness matrix in form of cylindrical coordinate system，not rectangular coordinate system？If it can not be obtained from Abaqus, do you know how to transform it in form of cylindrical coordinate system?I have asked a lot people, but none of them know how to do that. Could you give me a favor?

Thank you very much in advance.

Best wishes,

Yuwei

I am working on the Finite Element Analysis of a metal forming process. I have created the rollers and the mandrel and the roller as analytically rigid surface and the blank as 3D deformable solid.

My issues are the follows :

1.I have specified the contact between the roller and the blank as surface to surface contact and that between the blank and the mandrel as surface to surface contact. The result turns out that the internal energy generated (ALLIE) is greater than the work done(ALLWK). Also, the total energy (ETOTAL) is not a constant- it is monotonously increasing.

2. I changed my contact definitions from surface to surface contact to general contact (All with Self) The total energy is found to decrease and the internal energy becomes less than the work done.

I have specified 10 as the frequency of ALE and 15 as the number of re-meshing sweeps per increment. I understand that the internal energy is 95% dependent on the material property and the rest 5% on the Artificial energies. I changed my material model from Isotropic hardening to Johnson Cook. The results show the same trend. I consulted a friend of mine who suggested that the increase in energy could be due to excessive mesh distortion and so I should try increasing the frequency of ALE.

I would like to have your comments on the following :

1. The change in trend of the total energy with change in contact definition.
2. Possible reasons for variation of Total energy
3. Generation of Viscous dissipation energy, though i have not specified any values for viscous dissipation.

I would be grateful if you could throw some light on these.

Hey all

As a part of my effort to learn how to use Abaqus, i was trying to simulate a familiar contact problem which i found in the internet.

fortunately, the document with the descripition of the problem included detailed guidance as to how to create the model.

after i have followed the precised phases needed, i completed the model and it ran OK. the results showed the same patterns of contact and general behavior but the stresses were 2 times greater.

i have no clue what the problem could be because i followed the steps precisely.

i am adding the PDF with the contact problem and the steps needed to define it.

THX a lot,

Iftah

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Hello,

I'm modelling an impact of a rigid body with a deformable plate (shell elements) in the dynamic explicit analysis. When defining contact parameters, I can add the damping coefficient. I'd like to know what value is typical for this parameter? After some computations, I've found that 0,01...0,05 was rather good for my model. Is that a reasonable value?

Thanks

Hi,

The problem I want to solve is: instabilities that arise due to
molecular potential (VDW interactions) when a rigid solid block
(contactor) approaches an elastic film bonded to a support (no
interactions or slip whatsoever between the film and the support). These
interactions come into play when the air gap between the elastic film
and contactor becomes less than some critical distance. I want to do
this in Abaqus Explicit.
a) I have tried using cohesive zone model traction-separation law (on
single element blocks) in VUMAT to model the air gap between the film
and the contactor (giving tractions calculated as the derivatives of VDW
potential wrt the displacements). But the film is getting compressed
instead of getting attracted to the contactor (because van der waals
forces increase as the distance or teh air gap decreases). Is this
expected or is it because of the stiffness of cohesive zone (since
traction separation law need stiffnesses to be specified to calculate
stresses and these stiffnesses are derivatives of tractions and are
finite. So the cohesive zone model has some finite stiffness to it), and
is producing compressive stresses in the film when the contactor is
brought closer to the film. Is this behaviour anticipated? I can give
you the .inp and .f files if you want. please help... am still using the
3 element model to check this and stuck in this problem since months.
thanks!!

regards,
praveena

Hi..

My model consisits of two steel plates. Top plate has curved bottom and bottom plate is rectangular. Top plate is resting over the bottom plate. There are 3 steel pintels between the bottom face of top plate and top face of bottom plate to prevent running of top plate. Sliding is allowed between the contact surface of top and bottom plate. Pintels are fixed to the grooves made in the bottom face of top plate and frictionless hard contact is defined there.

Loading - Vertical pressure of 5MPa on top face

Lateral displacement of 1.5mm on the side face of top plate

BC- Bottom surface is fixed.

Steps- 1. Vertical load

Step-2. Lateral displacement

When I am doing the analysis results for very low value of vertical load are coming fine for both steps. But as I increase the vertical load to the above mentioned value the RF2 value in step 2 starts decreasing after reaching its peak value in 7th increment. According to me it should attain a constant value of RF2 as soon as sliding starts between the two surfaces.

Any help regarding this is highly appreciated

Thanks

Neha

Two recent publications in International Journal of Solids and Structures regarding the Analysis of Knoop Indentation.

Maybe they are of some interest to contact mechanicians.

Knoop indentation tests have long been a standard method for material
characterization due to the fact that they provide an easy, inexpensive
non destructive and objective method of evaluating basic properties from
small volumes of materials. In spite of the broad use, Knoop
indentation has never been analysed and its methodology is basically
empirical. the two publications present an extensive finite element study
on the adhesionless contact of elasto-plastic flat surfaces by Knoop indenter.

1. Analysis of Knoop indentation

http://www.sciencedirect.com/science/article/pii/S0020768310003343

2. Analysis of Knoop indentation-strain hardening effects

http://www.sciencedirect.com/science/article/pii/S0020768311002678

hi all,

I have some problems with contact elements in ANSYS.

My model is a cube of matrx containing two fibres with periodic boundary conditions. The interface between Matrix and Fibre is modeled with contact elements and cohesive zone law.

In 2-D everything works out just fine when bringing up displacement in x-direction. Here is a movie I made of the Contact Status over the loadsteps:

http://dl.dropbox.com/u/4920002/2D_contactmovie_ContaStat_0.4.avi

Doing the same in 3D brings up some problems. The Contact status is not behaving as expected.

http://dl.dropbox.com/u/4920002/3D_contactmovie_contstat.avi

Any ideas what is going wrong in the 3D case?

Thanks,

Andy

Hello, I am modeling the spherical indentation of a poroelastic layer. I am modeling the indenter as a rigid analytical surface. As you know, the total pressure in a porous medium at each point is the sum of the elastic stresses from the matrix, and the pore pressure caused by fluid pressurization. Abaqus provides both the pore pressure and of course the rest of the stress quantities.

As I'm analyzing an indentation problem, my final result needs to be a force quantity (spherical indenter thus non-constant area), or I need to be able to convert the stresses to force by having some knowledge of the area. There are a few outputs that look interesting to me for this purpose, but all correspond with elastic stresses it seems, and not the pore pressure. So if I were only dealing with elastic stresses, I could use CFORCE (CNF2) and sum over nodes for which contact is closed, or alternatively use CFN which is a whole surface output and is the total force due to contact pressure.

These values don't show any stress relaxation however, so it makes me believe that they are not the actual "total" but just the elastic part of the stress tensor. I can't use the pore pressure values and get the force resulting from that part because the problem then becomes converting the pressure to force. I see two output quantities for areas but they don't seem to be useful: CAREA is the "total" area in contact so that won't help because pore pressure varies across the surface. The other is CNAREA which is the "contact nodal area" which is only available for the master surface. This still looks like the most useful to me, but I am not finding any description in the documentation more in-depth than that. Is this the effective area at each node, so that if I multiply the pressure by this and sum over all nodes, I will get what I need (sort of)? I say "sort of" because although the horizontal components cancel out overall (due to symmetry), since I'm looking at each node individually then I will also need to be able to project this onto the z axis for each node.

I'm basically looking for a way to convert the simulation output to force so that it can be compared with the experimental results. Does anyone have an idea/experience with this?

Thanks!

My FE model consist of Metal plasticity and contacts (node to surface and finite sliding,friction co eff-0.1).With elastic material model,it worked fine but as soon as I incorporate plasticity (hardening=isitropic) some excessive deformation issues are coming up and the run gets diverged.element types I used are both C3D8 and C3D8I.material is aluminium.excessive distortion is coming at contact zone.What should I do to get over of that ?

Please suggest something.

Hello all,

This is my first post here. This seems be an informative and helpful forum.

I am trying to model the rotation of a "hard" shaft as a result of motion of 2 strips of soft materials (please see attached figure).

The strips are deformed due to external forces and as a result, contact pressures are applied on the shaft. The forces are then varied on the strips leading to net rotational forces on the shaft.

I am interested in the torque applied on the surface of the shaft and the angle of rotation (from some abituary changes in forces). I am not interested in the kinematics of the problem.

I have set up the model as follow in ABAQUS/Standard:

• The material of the strips are described by a custom user routine and have been tested to be working correctly.
• There are BC's constraining the end of the strips.
• The shaft is fixed in all degrees of freedom except with rotation with axis out of plane (UR3).
• The actual model in Abaqus is a 3D model. The strips are represented with membrane elements, and the shaft by 3D deformable elements.

I have tested the model with frictionless contact. The contact surfaces are the edge of strips and the surface of the shaft. The model worked as expected, the strips conform to the surface of the shaft and slides across the surface of the shaft (due to absent of friction).

However, when I add in friction, the solution fails to converge. This was most likely due to opening and closing of the contact surface when the shaft is turning (i.e. the strips exert some turning force, turning the shaft. But near ends of contact, they come off contact and possibly lead to "chattering" as described in ABAQUS manual). Would anyone be able to give me pointers into preventing this from happening or some other set up to get a converged solution?

Hy everyone,i searched everywhere and i can't seem to find the equation for the deflection surface in the case of a circular plate simply supported loaded by a eccentric point force.....i found the solution for the clamped case....can anyone please help me?Thank you,Marius

I am pursuing my Ph.D.in Soft Contact Manipulation. I want to understand Hertz Contact Model very clearly.Can any one help me to understand it or Give me some helping material. It is very difficult to understand it from Contact Mechanics (K.L. Johnson).

Dear sirs,

I would like to seek your assistanceon using the Abaqus software .

I am trying to study the stresses inpavement layers , by applying movingloads caused by vehicles. However I have not been able to get a visualrepresentation of the moving load.

I would like to know how to get acycle of movement , to and fro ...over a period of years .

I shall be grateful if you alsocould send me a similar case on the lines of Abaqus/CAE truss tutorials .

with regards

Teeha Mufti

i m doing a project on ball bearing. there is a point contact between the ball and the races. when the force is applied on the inner race, the point contact turnes into elliptical surface to surface. please guide me how to model this problem in ansys.

I am fairly new to ANSYS and have been trying to model the interaction between glass micro-spheres and silicone polymer micro-structures for my dissertation research. I've been working through the online manual, but have been getting very odd results when running the simulation. Any help at all would be appreciated. The summary of my problem is as follows:

Modelling:

It may help to first refernce the first two images at : http://bdml.stanford.edu/twiki/bin/view/Main/AnsysModelling

I have a single, thin wedge shape modelled with SOLID95 elements. The body of the wedge (everything but the round tip) is comprised of 20-node bricks. The rounded top of the wedge is modelled with 10-node tetrahedron, with pyramidal interface eelements between the tets and the bricks. The material model for the wedge is a linear isotropic one with ballpark values of: E=1e6 and PR=0.49. I also gave it a friction coefficient of 1.

Contact Elements:

The contact surface is showin the third picture in the above link.

I set the type to CONTA174 and, using the meshing tool, meshed the shown areas with triangular elements on the round top of the wedge, and quad elements on the front face. I then set the R1 value in the real constant set to the desired radius of my sphere, which is the rigid target. Using the same real constant set, I set the type to TARGE170 and created a pilot node element where I wanted the center of the sphere and then created a sphere element on top of the pilot node. The sphere was the appropriate size, so I believe the shared real constant set is correctly set up.

Simulation:

I constrained the bottom of the wedge in all DOF, and put a simple displacement load on to the pilot node of the sphere. I want the sphere to move straight down a certain distance, and observe how the wedge deforms. I didn't change much in the way of thesolution controls. I set it up as a large displacement static problem with a 250 set as the number of substeps and 150 as the minimum number of substeps. The simulation fails each time, however, due to an error in element formulation. It's easy to see why there is an error when I plot the deformed shape of the last solution set. The elements become highly distorted almost as soon as the target comes in to contact with the contact surface. This can be seen in the last image in the link above.

As I said before, I'm a bit new to contact simulation in ANSYS. Maybe I've made a silly, rookie error. Any guidance at all would be greatly appreciated.

Paul

Hi,

Can anybody throw some light on the wear coupling characteristics between SS347 and 17-4 PH, as well as between Phosphor Bronze and 17-4 PH Steel?

It will be helpful if you could refer to some literature where I can learn about Wear Coupling Characteristics...

Thank You,

Navanee

Hello,

I am using Abaqus for nonlinear FEA of contact between material1 (soft foam like) and material2 (hard plastic type) in vertebrae for my thesis. I have encountered many problems during this exercise, because I have to include the results of various contact conditions in this situation with friction. I am trying to get results for

• Hard contact (using penalty constraint enforcement)
• Soft contact (using linear and exponential Pressure-Overclosure relations)

The major problem is convergence during both analyses.

In hard contact the problem was the contact formulation - surface restrictions that limits some connectivity characteristics (e.g 3D solids/faces joined at nodes, or, T-intersections) , it was because of some solid elements (of material2- hard) which were connected via only one node (instead or being connected via edge or face). Mesh refinements were not so much helpful, because they only divide elements, refinement has no thing to do with smoothness of connection of already existing large-sized elements. I was able to get results only with selecting Finite Sliding option, although no sliding occurs in physical model because or interdigitation). This matter put a question mark on my results.

In soft contact, the problem was pressure-overclosure relation. It is really difficult to decide what should be C_o (clearance at zero pressure) and P_o (pressure at zero clearance), because the nodes and elements faces of material1 and material2 are nearly at zero distance (<10E -4) from each other. By selecting a linear pressure-overclosure relation and setting the value of contact surface stiffness "'K'" roughly equal to maximum value available in the siffness matrix of an element at the contact surface, I started my analysis. Soon I realized that the stresses at contact interface are higher even than stresses of Hard contact. so I reduced the value of "'K'" and re-run the analysis, repeating the analysis and reducing "'K'" till I got interface stresses equal in both cases finally. This value of "'K'" is now K_o for my analysis.

I have K_o , but the problem is still C_o and P_o, I can calculate P_o from P_o = (K_o) x (C_o), because C_o is very very small (less than 10E -4) , and if multiplied with (K_o = 350) gives some thing really funny ( i.e Pressure required to clost the clearance will be 0.0017). I have tried to run the analysis with some higher values of P_o but the analysis is not converging.

Anyone who has some idea on how to tackle such problems, please help.

Hello,

I would like to know how contact elements are different from " general "ie non contact finite elements.

Also I would like to know how Lagrange multiplier method is used in contcat analysis.

My understanding of the Lagrange multiplier is like this: when we have to impose a constraint equation

[C][D]=[Q] as constraint between DOFs, we incude a term Y*{[[C][D]-[Q]} where Y is the Lagrange multiplier ,in the potential energy functional and minimize the functional to get the algebraic equations.

My understadning is: during contact analysis , we take an assumed state of contact and hence a constraint equation between the contacting and contacted nodes and the Lagrange method is imposed in the same way as above where this time [D] stands for the DOF which are "assumed" to contact.

Kindly advise if my understanding is correct.