Ansys Tips Notes

8/9/2019 Ansys Tips Notes

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ANSYS TIPS- Library Copy

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ANSYS® Tips ANSYS Tips and ANSYS Tricks

Peter Budgell

Burlington, Ontario, Canada

These tips and comments are intended for user education purposes only. They are to be used at your own risk. The contents are

based on my experience with ANSYS 5.3 -- more recent versions may change things. The contents do not attempt to discuss all theconcepts of the finite element method that are required to obtain successful solutions. It is your responsibility to determine if you havesufficient knowlege and understanding of finite element theory to apply the software appropriately. I have attempted to give accurateinformation, but cannot accept liability for any consequences or damages, which may result from errors in this discussion. Accordingly,I disclaim any liability for any damages including, but not limited to, injury to person or property, lost profit, data recovery charges,attorney's fees, or any other costs or expenses.

CONTENTS: Tip 1: Use AnnotationsTip 2: Making Room for AnnotationsTip 3: Using Parameters in AnnotationsTip 4: Use Small AnnotationsTip 5: Mathematical Functions AvailableTip 6: Start 16-Bit Applications before Starting ANSYS under Windows NTTip 7: Running ANSYS at Low Priority under Windows NT 4.0Tip 8: Operating on (Scaling) LoadsTip 9: Ramping Loads Down to ZeroTip 10: Starting ANSYS Graphs at t=0Tip 11: Pressure on LinesTip 12: Ramping Some Loads, Not OthersTip 13: Force and Pressure on Flat Plates or Flat ShellsTip 14: Linear and Nonlinear BucklingTip 15: Nonlinear Analysis and the Arc-Length MethodTip 16: Animating Results from a Nonlinear or Other AnalysisTip 17: Getting the Mass or Weight of a ModelTip 18: Using Fnc Calls from MacrosTip 19: Use ENSYM and ENORM to Turn Over Shell ElementsTip 20: Shell Types to TryTip 21: Moving a Model from ANSYS Mechanical to ANSYS Linear/PlusTip 22: Deleting Nodes with Nodal Coupling

Tip 23: Convergence with Shell Finite Element Models in Nonlinear Analysis under ANSYSTip 24: Working with Load Step Files in ANSYSTip 25: Plotting Shell Stress -- Surface, Mid-Plane Stress, Load Paths, ESYS and RSYSTip 26: Nodal Coupling (CP) versus Rigid Region (CERIG)Tip 27: Vibration Modes with Pre-stressTip 28: Creating New Elements by Copying or Reflecting Existing StructureTip 29: Adding to a Model Comprised of Elements and Nodes OnlyTip 30: Zero Mass Beam Elements Form Rigid RegionTip 31: Turn off Symbols When Changing a Model after SolutionTip 32: Are the "Free-Free" Vibration Modes Relevant?Tip 33: Selecting a CAD or FEA System -- Cover YourselfTip 34: Creating Lines Perpendicular to, or at Angle to Existing LinesTip 35: Use the /UI command in Your ANSYS Toolbar to Bring up GUI Dialog BoxesTip 36: Reaction Force, Nodal Force, and Load Paths

Tip 37: Inputting Temperatures with BF, BFE, and TUNIF in Structural AnalysisTip 38: ANSYS Toolbar UseTip 39: ANSYS Piping Element LengthsTip 40: Graphical Output from ANSYSTip 41: Check Nodal Loads at Bolts, Rivets, Spot Welds and LinksTip 42: Use QUERY to Check Results with PickingTip 43: Loads on Geometric Entities Overwrite Loads on Nodes and Elements -- Easy Error to MakeTip 44: Use Components for Load Input, and for Results ReviewTip 45: Simple Substructuring Examples -- Bottom Up and Top DownTip 46: Plot Applied Temperatures


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Tip 47: Skipping Over Statements in an Input FileTip 48: Static Analysis Followed by Transient AnalysisTip 49: File Compression for Model StorageTip 50: Organizing Large FEA ModelsTip 51: Selecting Nodes in a Stress or Strain RangeTip 52: Selecting Nodes that are Subjected to Nodal CouplingTip 53: /NOPR and /GOPR Speed Up Input Files and Macros

Tip 54: Using Commands IMMED and /UIS and /SHOW,OFFTip 55: What's the Bauschinger Effect? Comments on Material YieldTip 56: Thought ExperimentsTip 57: Control of MeshingTip 58: Four View PlotTip 59: Quick Review of Mode ShapesTip 60: Using ANSYS HelpTip 61: The FEA Job HuntTip 62: *VPUT and DESOLTip 63: How to Divide One Element Table Column by AnotherTip 64: Element Tables (ETABLE) and Arrays -- An ExampleTip 65: Error Estimation, PowerGraphics, and ERNORMTip 66: Concatenate and Mesh LastTip 67: ANSYS Output of Data to Files for Use by Other Programs

Tip 68: Writing Array Columns to Output or to FilesTip 69: Synthesizing Parameter Names and Manipulating Jobnames and Long Strings in APDLTip 70: Solid Elements 95 and 92 -- Efficiency and Interconnection

Tip 1: Use Annotations: Only a one-line title is possible on the ANSYS screen or plot. Considerably moreinformation can be included in annotations on the screen. The annotations are kept throughall plots until they are deleted with the command: /ANNOT,DELE or via picking with thegraphical user interface (GUI).At the top of the Annotation dialog box, there is a list box from which the user can chooseText, Lines, etc., on down to Controls. These selections bring up different menus. TheControls selection offers a SNAP setting that makes it much easier to get the text alignednicely. (Hint: ANSYS, Inc. should put this SNAP selection up front under Text, or even on

every menu.) Activate the Snap setting, then go back to Text to enter the annotations.Tip 2: Making Room for Annotations:

The /PLOPTS command controls what goes into the legend at the right (by default) side ofthe ANSYS screen and plot. If you turn off LEG2 (the relatively useless "view"information), you will get extra room at the bottom of the legend. This area can be used forannotations if the number of contour levels in stress plots is not too great (the default is fine).Tip 3: Using Parameters in Annotations:

Just as in a title created with the command /TITLE, ANSYS permits the use of a parameterin an annotation, as discussed in the Commands Manual description of the /TLABLEcommand. When typing the annotation using the GUI, include the parameter in percent signslike this: %pname% where pname is the parameter name. The parameter can contain eithernumbers or text. The value of the parameter will be plotted in the annotation string. TheANSYS function NINT can be used to round a number the nearest integer, sometimesimproving the appearance of the annotation for large numbers in which the fractional part isirrelevant (e.g. NINT(123.456789) = 123 ). For this, the parametric expression should beenclosed in percent signs. Annotations are usually created in the GUI, but can be enteredwith code like that shown below. Entering a single annotation line containing Result =%pname% generates log file contents such as:

! The following commands place an annotation on the screen.! For information only. Use at your own risk.


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! In this example, "pname" is a parameter with a numerical valuesuch as 123.456789/ANUM ,0, 1, 1.2303, -.74699/TSPEC, 15, .600, 1, 0, 0/TLAB, 1.010, -.747,Result = %pname%

The last line in the above example contains the string that the user types manually. The otherdata set up the string positioning on the screen, and the properties of the characters. To apply

the NINT function to the parameter, manually enter Result = %NINT(pname)% as theannotation:

! For information only. Use at your own risk.! Type the annotation in one line, so the log file contains:/ANUM ,0, 1, 1.2303, -.74699/TSPEC, 15, .600, 1, 0, 0/TLAB, 1.010, -.747,Result = %NINT(pname)%

The beauty of doing this is that if the value of the parameter pname should change, thenwhen the next plot command is executed, the annotation will automatically update toreflect the new value! Try it: after creating an annotation on the screen that includes a

parameter, change the parameter's value, then do a /REPLOT. Running a macro could getinformation that goes into the parameter that a /REPLOT will automatically put it on thescreen. This makes it possible to automatically include far more information than can go intothe title, and to do it for a series of automatically generated plots or graphs.Tip 4: Use Small Annotations:

The default character size setting for an annotation is 1. The size of an annotation can bedecreased using the GUI. A size of 0.6 is quite readable and permits far more information to

be packed into a plot. Note that there is a limit to the number of characters possible on anannotation line – this is character size independent.Tip 5: Mathematical Functions Available:

Under the Help listing for the *SET command there appears lists of mathematical functionsavailable in ANSYS. Another list is in the ANSYS User's Guide on APDL, Chapter 14 of theModeling and Meshing Guide. The commands are usable anywhere. They include:

ABS(X) Absolute value

ACOS(X) ArcCosine

ASIN(X) ArcSin

ATAN(X) ArcTangent

ATAN2(X,Y) ArcTangent of (Y/X) with the sign of each component considered (see aFORTRAN manual if you don't know what this means.)

COS(X) Cosine

COSH(X) Hyperbolic cosine

EXP(X) Exponential

GDIS(X,Y) Random sample of Gaussian distributions where X is the mean, and Y is thestandard deviation. Might be used in a Monte Carlo Simulation to explore thedistribution of outputs based on randomized loadings and material properties.For an explanation, see a good modern engineering design textbook.


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LOG(X) Natural log (to base e)

LOG10(X) Log (to base 10)

MOD(X,Y) Modulus (X/Y), it returns the remainder of X/Y. If Y=0, returns zero (0)

NINT(X) Nearest integer (nice for outputs of stresses to /TITLE or annotations (see Tip3 above))

RAND(X,Y) Random number, where X is the lower bound, and Y is the upper bound.(Useful for Monte Carlo Simulation, etc.)

SIGN(X,Y) Absolute value of X with sign of Y. Y=0 results in positive sign.

SIN(X) Sine

SINH(X) Hyperbolic sine

SQRT(X) Square root.

TAN(X) Tangent

TANH(X) Hyperbolic Tangent

Note:The function form of the *GET commands can also be used to get information fromthe model -- see the APDL guide mentioned above for a listing of available functions.The APDL guide also gives functions to retrieve the values of parameters, bothnumerical and character. The *VFUN command has a list of functions that act on anarray entry. The Commands manual lists functions that act on Element Tables in the

section "POST1 Command for Element Table". Creatively used, the array andETABLE algebra commands can be surprisingly powerful.Tip 6: Start 16-Bit Applications before Starting ANSYS under Windows NT:

It has been my experience that some large commercial 16-bit applications will not start properly when ANSYS is already running. If you start them before launching ANSYS, therewill be no problem. If you intend to work with those 16 bit applications in the foregroundwhile the ANSYS SOLVE is running in the background, this will be a useful tip. I have seenother applications start up very slowly (e.g. Internet Explorer) or wait until ANSYS wasdone before proceeding (setup.exe for many Windows install programs).Tip 7: Running ANSYS at Low Priority under Windows NT 4.0: Under Windows NT 4.0 the priority level of individual processes can be user-adjusted. To dothis, bring up the Task Manager (right click on the Windows NT taskbar), and click the tab

for "Processes". Right click on the process titled "ANSYS.EXE", and "Set Priority >" comesup. Set the priority to "Low" to help make foreground applications run more smoothly whileANSYS is running SOLVE in the background. This may help more if you have a large RAMin the computer.When ANSYS has completed the SOLVE process, return the priority to "Normal" so thatANSYS is not slowed down when you start doing plots through the GUI.Tip 8: Operating on (Scaling) Loads:

You can operate on loads on nodes and elements in order to scale them up or down.Unfortunately, scaling loads on geometric entities (keypoints, lines, areas and volumes)


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seems not to be available. If any load on your structure has been applied to a geometricentity, rather than directly to elements or nodes, that load will be transferred to the elementsand nodes at solution time. The transfer will overwrite any scaling of loads that you haveapplied.So what can you do about this? Method 1: Transfer the loading from geometric entities to theelements and nodes, then write a load step file. This records loading on elements and nodes.

Delete the loading on geometric entities, then read the load step file that was just written. Now the loading can be scaled up or down freely. Method 2: For a faster method, see the"LSCLEAR,SOLID" command, which will not require writing a load step file. Method 3:Transfer the loading from geometric entities to the elements and nodes, then delete therelationship between geometry and the FEA mesh with the MODMSH,DETA command.Method 4: Transfer loading from geometric entities to the elements and nodes, then un-selectthe geometric entities, before executing SOLVE. The element and node loading can bescaled after it has been transferred from geometric entities. An un-selected geometric entitywill not transfer its loading to elements or nodes when SOLVE is executed. Warnings:Method 3 ruins the relationship between geometry and the mesh. Save the model under anappropriate file name before executing MODMSH,DETA. Method 4 is fine, as long as youdo not forget and re-select the geometric entities -- ALLSEL will do this.

Scaling displacements (nodal constraint values) is also possible. One thing that has notworked for me is an attempt to reduce applied displacements to zero by using 0.0 as thescaling factor. What did work for me was to use "_TINY" as a value, which multiplieddisplacements by a factor of roughly 10^(-31) and reduced loads to virtually zero. Attemptsto use 0.0 as the factor resulted in NO change to the applied displacements.Tip 9: Ramping Loads Down to Zero:

If you are ramping force, pressure, and acceleration loads up and down as part of an analysis,you may want to return loads to zero. I do this when I want to inspect permanent deformationthat results from plastic yielding. If you delete the applied load, the loading will dropimmediately to zero, even if you have load ramping turned on. The thing to do is to set theloading to virtually zero or the scaling factor to virtually zero, not delete the load. It isimportant to appreciate that to ANSYS, reducing a load to nearly zero is not the same thing

as deleting it (zeroing it), for the purposes of ramping loads. The time substep sizes to usewill depend on your model.Setting displacements to zero or near zero is, of course, very different from deletingconstraints.Tip 10: Starting ANSYS Graphs at t=0

Graphs start at the first data point, which means that if you do a time-history trace, you don'tget a t=0 data point. If you leave time as 0.0 on the TIME command, you get the default 1.0in your output. The only way to get a graph from zero that I have found is to do a first loadstep with "t" extremely small, in comparison to other times in the analysis, e.g. t=0.0000001.The load at this time must be appropriate so that the response ramps up correctly. (If yourintent was to ramp up from zero load, just leave the loads as zero.) The next load step

continues as usual.Tip 11: Pressure on Lines:Applying pressure on a line results in loads being applied to the nodes associated with thatline. The loads on the nodes that the FEA program applies will be appropriate given theformulation of the elements. If you want to apply a total force to the line, you can use a*GET command to find the length of the line, then divide the force by the length and use theresult as the pressure.

Note that pressure on a line acts in the plane of the area that is attached to the line. If twoareas are attached at 90 degrees or another angle, two loads are set up, acting in each of the


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area plane directions. You can use select logic on the areas to get some interesting effects asto the direction in which the applied forces act, but only if both areas are meshed, and theelements are selected. If you un-select one of the areas, pressure on the line will only beexerted in the direction of the area that is selected. The select logic must still be in placewhen you SOLVE, or else your carefully crafted load case can be overwritten. As above,transferring loads from geometric entities to nodes and elements, writing them as a load step,

deleting all the loads on geometric entities, and reading in the load step will protect your loadcase, and make scaling the loads possible. Alternatively, consider the "LSCLEAR,SOLID"command.

NOTE: Pressures on surfaces follow the deformed shape during a Large Displacement(geometrically nonlinear) analysis. Forces on nodes maintain their orientation in space, evenunder Large Displacement. This difference will govern how loads should be applied in somemodels.Tip 12: Ramping Some Loads, Not Others:

To hold some loads constant and ramp up or down others, run a first load step with all theloads at their starting values, ramping from zero only if appropriate.If you want, use an extremely small value on the TIME command, e.g. 0.0000001, and runthis as a first load step. Then set up a second load step, with ramping activated. Change those

loads to be ramped from their starting values to new values. Hold the other loads constant.The TIME command can be used with a new value, such as 1.0.An example is the application of a gravity load before other loads are to be ramped up fromzero. In some cases, this could give a more realistic assessment of nonlinear buckling caused

by applied forces other than gravity loading. (You will want to check the codes that regulateyour design work before deciding on this. Codes that I have seen were generally started

before FEA was widely available, and do not address this concept. Find out what isconsidered good practice in your industry.) Applying gravity first can give much betterconvergence when assessing the effect of thermal expansion moving structures acrossfriction contact elements, where the normal load on the contact elements is caused bygravity.I suspect that this is not possible with the Arc-Length method. I have not experimented withit, but do not see how controlled ramping of only some loads could be implemented underArc-Length control of applied loading -- any opinions?Tip 13: Force and Pressure on Flat Plates or Flat Shells

There is a rule of thumb, that if the out-of-plane deflection of a flat plate or shell is greaterthan half the thickness, then membrane forces start to become significant in resisting theapplied load. In ANSYS, this calls for activating a Large Displacement solution (a.k.a.geometric nonlinearity). Ignoring this can result in your design missing out on inherentstrength, OR in grossly inadequate underdesign. Know what you are doing.Tip 14: Linear and Nonlinear Buckling:

Linear eigenvalue (classical Euler) buckling is a "quick" check on a structure, but theANSYS manuals go to considerable pains to point out that in many situations, a Large

Displacement solution (geometric nonlinearity) needs to be run also as a check on the buckling adequacy of a design. As with linear buckling, nonlinear buckling may need to beassessed with respect to a number of load cases. In some structures, a diagonal tension fieldis developed in a web, and elastic buckling failure does not develop at the first eigenvalues

predicted. In other structures, buckling failure may occur before the first eigenvalue, andonly nonlinear analysis will predict this.Linear eigenvalue buckling has to assume that gap and contact elements are either closed andactive, or open and inactive. Nonlinear analysis will follow the effects of these elements asthey go in and out of contact, when the loading is applied.


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After any Large Displacement nonlinear elastic buckling analysis (if it doesn't diverge), seewhether the elastic stress limits have been exceeded (this includes the surfaces of shellelements, and be careful that nodal averaging does not hide anything). If significantlyoverstressed, the structure may not be adequate.Combined bending and axial compression in a beam is a classic place where inadequacy instrength can be predicted in FEA only by Large Displacement nonlinear analysis (i.e. a linear

analysis says it is OK, but a nonlinear analysis shows it is NOT). For some structuresundergoing elastic Large Displacement analysis without contact and gap elements, the usermay want to consider a Southwell plot.If elastic stress limits are exceeded in the Large Displacement model, it may be desirable todo a combined Large Displacement and Plastic Deformation model. If the structure isoverloaded, it may begin to collapse (perhaps only locally), and the Arc-Length method may

be needed for convergence control. A need to strengthen the structure may be predicted oridentified. The material properties to use are application domain and industry specific -- start

by talking to your co-workers, supervisor, and suppliers.

Tip 15: Nonlinear Analysis and the Arc-Length Method:

The basic way to do nonlinear analysis in ANSYS is to use NR iteration and many default

settings. At times, convergence will become a problem; I've encountered this with shellstructures under compressive stresses. The arc-length method can sometimes cope betterwith nonlinear solutions, because of its ability to follow force-deflection curves that rise andfall. Be prepared for long run times if your model is large.My experience with the arc-length method is that in its default settings for step sizemultipliers, it does not give satisfactory results when compressing some shell-based models.What may work is to set a number of time substeps, such as 10, so that each substep is 1/10of the load step.Set the Arc-Length maximummultiplierMAXARC to 1.0so that nosubsteps largerthan 1/10 of theload step aretaken. Set theArc-LengthminimummultiplierMINARC to 0.1,so that thesmallest load

substep is 1/100of the full loadstep. I found thisto helpconsiderably. Youmay want to usera larger or smaller MINARC setting, but my experience to date suggests that one should notget greedy with MAXARC. Obviously, you may want to play with the number of timesubsteps.


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The solution may still diverge but it is likely that you will get more information than withoutarc-length analysis. You will want to set a termination condition for the analysis if bucklingis expected to result.I find it desirable to save the results at every time substep when doing this type of analysis (ithelps to have a large hard drive) in order to review the process. When you review the resultsof a single load case run under Arc-length control, the TIME value on the ANSYS plots

shows the decimal fraction of the full load being applied to the model. As you move forwardthrough the plots, if the load/displacement curve for the structure is falling, the decimalfraction will fall, even though some displacements are visibly getting larger.As mentionedabove,something I havenot tried is to getthe Arc-lengthsolution controlto ramp someloads and notothers, by having

run a preliminaryload step. Is thiseven possible? Ifnot, then the usermay face the

prospect ofgravity beingramped up anddown, in additionto other appliedloads, and the

physical realismof the model may

be affected.Tip 16:

Animating

Results from a

Nonlinear or Other Analysis:

It can be helpful to watch the increasing stress levels that result as a nonlinear analysisloading is ramped up. To create an animation, first run your analysis with loads ramped up,and a number of substeps. Have all substeps written to the results file. Do a stress plot ofinterest to set the type of stress plot to be animated by the macro that will be run. Make theANSYS Graphics window as small as you want the animation window to appear (most

screens will have lower resolution than a CAD workstation), keeping the aspect ratio correct.Smaller graphics windows result in smaller animation files, if size matters. Animation filesunder Windows NT (AVI files) from ANSYS often compress very well for storage purposes.Use the PlotCtrls menu selection on the Utility Menu, and choose Animate to get a sub-menuof choices. Choose "Dynamic Results" to create an animation of your saved load substepresults with the time shown in the legend. This seems to work only for the last load step(read the ANSYS macro). The resulting AVI file can be viewed with the media player,distributed, put on a web site, and so on. The media player can be stepped manually for slow


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viewing. It makes it easier to watch the changing stress pattern or deformation as nonlineareffects take over the model.In animating a changing stress or other contour plot, you may wish to specify the contourlevels before generating the animation file. View the load step or substep with the worstresults as part of deciding where to set the contour levels.I have not found that any of the ANSYS supplied animation macros do the one simplest

thing I want. Usually I want to animate every substep of every load step stored in the resultsfile. The following simple macro does this for me under Windows NT. There is virtually noerror checking in this macro. Note that this simple macro does not update element table dataat each frame. Consequently it will not work properly for plots of element table data. Ifstresses, strains, or other data with amplitude information are to be plotted, the user maywant to fix the contour map levels ahead of time. The user will want to set the displacementamplitude scaling with /DSCALE in advance--automatic scaling will not be satisfactory. Ingeneral, it may not be satisfactory to have /ZOOM,OFF active, since the view will change if

plots of significant deflection are included in the animation. Manually setting a view mayyield a better animation. Modify this macro as you wish.This macro must be called from within /POST1. The file that contains the results must havealready been selected, and a prototype plot command executed so that calling /REPLOT will

generate the type of plot the user wants:! --------------------------------------------------------------------! MY_ANIM.MAC A quick-and-dirty animation of all of the substeps! --------------------------------------------------------------------! For information only. Use at your own risk.! User must indicate how many frames are to be animated! This macro starts with the first substep in the results file! by using the SET,FIRST command internally! User implicitly indicates how many times to use the SET,NEXTcommand.! The number of frames needed must exist in the RST file, elseerrors.

! NOTE: This does NOT work for plots of data in an element table.! Plotting element table results would require a macro inwhich! the element table results are updated at each substep.!! Virtually NO Error Checking Is Performed ! ! ! ! !!! What will be plotted is based on /REPLOT therefore, on the lastuser plot executed! before this macro is called.! Scaling, etc. are all based on the last user plot. Only the SETvalue is updated.!! Call with:

!! my_anim, time_delay_for_frame, number_of_frames_including_first!ar11=arg1*if,arg1,eq,0,then

ar11=0.1*endif*if,arg2,ne,0, then

/NOPR/gsav,xxx,gsav,,temp/seg,delete


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/seg,multi,,ar11set,first/replot*do,_iii,1,arg2-1,1

set,next/replot

*enddo

/seg,offanim,1,1,ar11/gres,xxx,gsav/gopr

*endif

An alternative to this macro could step through all substeps on the RST file by using a *GETcommand of the type *GET,NTOTAL,ACTIVE,0,SOLU,NCMSS to check the number ofsubsteps as the SET,NEXT command is issued. The parameter NTOTAL will be re-set to 1when the animation is complete, and the *IF and *EXIT commands can check this and breakout of a do loop -- see Tip 59 below for the example of automatically plotting all modeshapes. The user would then not need to specify the number of substeps to plot, improvingthe automation, and letting the solver use variable substep sizes without the user having tocheck on the number of substeps that resulted.

Tip 17: Getting the Mass or Weight of a Model:A reader has been helpful by pointing point out that mass (or weight, depending on yourunits) of keypoints, lines, areas, or volumes in a model can be retrieved, when attributes have

been assigned to these entities, by using commands available in /PREP7. Using the graphicaluser interface, enter into "PreprocessorOperateCalc Geometric Items" to see the choices: "OfKeypoints, Of Lines, Of Areas, Of Volumes, Of Geometry". These items execute the "sum"commands: "KSUM, LSUM, ASUM, VSUM, GSUM" respectively. If no attributes have

been assigned to the geometric entities, unit densities are assumed in reporting mass andcenter of gravity information. After the execution of these commands, the *GET commandcan be used to assign to a variable the implied volume of an area (based on the thicknessassociated with its attributes) or the volume of a "volume". The volume of a series of areas or"volumes" can also be retrieved with the *GET command after a "sum" command is used.

The *VGET command can also be used, where appropriate, in retrieving information madeavailable after one of the "sum" commands is executed.For some unstated reason, ANSYS will not directly give the total weight or mass of a model(retrieved from the mass matrices of the elements), except to print it to output during thesolution of a problem. The user can run a partial solve in order to get this weight or mass

printed reasonably quickly. In Imperial units, it may be desirable to convert between poundsmass and pounds weight. There is no *GET command that directly returns the weight ofselected elements. However, the volume of an element can be returned, and the volume of aset of elements can be put into an element table, and summed.You can get the weight of many models into a parameter by: step through all material types,selecting elements for each material type. Get the volume of those elements, and multiply bythe density of that material type. Sum the masses or weights of all the material types. Thiswill not include added mass and mass elements at nodes (check this carefully against theoutput mass in the solve module) or other things that I may not have thought of.Of course, you can get the weight (assuming you gave densities in the material definitions)

by removing all loads (don't let thermal expansion, nodal rigid region, nodal coupling,various gap and contact elements, or loads on constrained nodes trip you up -- use theminimal constraints needed to stabilize all bodies in 3-D), applying 1 g vertical, havingconstraints on vertical motion, running SOLVE in a linear analysis, and finding the verticalreaction force. In such a run, a combination of the FSUM (select vertically restrained nodes


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only, with all attached elements) and *GET commands in /POST1 might help you to get theweight into a variable. However, a partial solve will give the answer more quickly (but not

put it into a variable). Depending on your system of units, remember, you may want toconvert between weight and mass.I base my comment, about the inability of ANSYS to directly return the weight of the modelwith *GET, on comments in the manuals on Optimization. The optimization examples work

to reduce model volume, not weight.Tip 18: Using Fnc Calls from Macros:

Before using macros for the first time, read about the *USE command in the ANSYS Helpmanual, in addition to other relevant parts of the ANSYS manuals. The *USE command helpdiscusses the macro calling parameters and their local scope. Note a slight difference incalling parameters AR19 and AR20 when the *USE form of a macro call is used, versus the"unknown command" form.There are times when calls from macros directly to the Function form of an ANSYScommand will be the only way to get the function called with picking. It may be desirable tosent the user a message that explains why the picking has been requested. The function must

be called with the exact use of upper case and lower case characters. An example:Fnc_ENSYM will work, whereas fnc_ENSYM will not, because the capital F is missing.

Tip 19: Use ENSYM and ENORM to Turn Over Shell Elements:ANSYS has two commands, ENSYM and ENORM, for re-orienting shell elements so that aset of shell elements can all have their "top" surface face the same way. This makesapplication of pressure, contact elements, and review of results more feasible. Thisorientation should be done before running SOLVE; the results are not re-oriented in thedatabase when these commands are applied, nor in the results file, so if the elements are re-oriented after SOLVE, the stress results will no longer apply to the correct shell surfaces anda meaningless mess will result. These commands work with shell elements that are attachedto areas, as well as with independent shell elements. Note: If you clear the elements attachedto an area, then re-mesh, the new elements will have the same orientation as the area. (Hint:ANSYS ought to do this re-orientation for Areas, making it easier to pressurize the interiorsof containers defined with shell elements.)See HELP,ENSYM for information on what this command will do. ENSYM can be used to"flip over" a shell element so that the opposite side (Top or Bottom) is showing. To do thiswould require reversing the node order in the database so that Face 1 (Bottom) and Face2(Top) get switched.For more powerful capabilities in re-orienting shell elements, see HELP,ENORM. Thiscommand will search outward from a chosen element that the user considers "correct", re-orienting a connected set of shell elements so that they face the "same way" (this takes someinterpretation), even working around corners. It searches elements from the selected set ofelements, until it hits the edge of the model, or until two or more elements are attached toone element edge. The user should experiment with this command in order to understandexactly what it does, and inspect the model thoroughly after ENORM is applied, to verify

that the results are as desired. The correct use of ENORM can make the application of pressure or contact elements to a complex model substantially easier.It would be very helpful if ANSYS had a special command that would plot shell elementswith the sides colored according to whether they were FACE1 or FACE2 of the element.This command could be extended to color the (up to) six sides of solid elements, accordingto their face number. A similar command for plotting areas would help, too. It could even bedone for beams displayed with /ESHAPE showing the outer envelope. At present, withANSYS 5.3 running on Windows NT, I get different colors for Face 1 and Face 2 of shellelements when PowerGraphics is ON, and "No Numbering" plus "Colors" or "Colors and


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Numbers" has been chosen under PlotCtrls,Numbering. I have not seen this documented.This does not happen for areas, or for solid elements.Tip 20: Shell Types to Try:

I have used Shell 63 (for Elastic), Shell 43 (for Plastic), Shell 93 (8-Node, for Elastic &Plastic), Shell 143 (for Plastic), and Shell 181 (for Plastic). The Revision 5.4 for ANSYSwill include a bug fix for a Shell 181 problem. Shell 143 is no longer supported, but is still

embedded (hidden) in Revision 5.3 of ANSYS for compatibility reasons.I have recently found Shell 93 to be useful in modeling some curved structures, because ofits ability to follow curved surfaces. (Shell 63 elements are flat, and can make a mess of ageneral curved surface under free meshing.) Shell 93 gave me good convergence for bothelastic and plastic Large Displacement (nonlinear geometric) analysis. It does not like tofollow too large an angle of curvature with one element, so the number of elements on anarea fillet can be large. Set the angle subtended by Shell 93 elements during meshing to avalue that is small enough to avoid warning messages. Watch out for aspect ratio warnings.(Lack of warnings is not a complete guarantee of acceptable element shapes.) If the structurehas pressurized flat surfaces, Shell 93 often converges better when stress stiffening isactivated for Large Displacement analysis. Stress stiffening for Shell 93 is activated at thesolution phase of the analysis, whereas Shell 63 is (apparently) only stress-stiffened by

setting one of the KEYOPT values. (I have obtained different Large Displacementconvergences with Shell 63 with no stress-stiffening set, with the KEYOPT stress-stiffeningset, with stress-stiffening set in /SOLU, and with stress-stiffening set in both places.) LikeShell 63, Shell 93 also has the virtue of being supported by the Linear/Plus version ofANSYS for Large Displacement elastic analysis, so models can be moved back and forth.When forcing mapped meshing of curve-sided Shell 93 elements on a plane area byconcatenating perimeter lines, I have occasionally had mid-side nodes created, in the interiorof the area, such that there was too much element curvature distortion in the plane of theelement. One fix is to have the elements created with the sides straight, which is tolerable ifthe elements are flat, and if it does not cause trouble on the perimeter of the plane area beingmeshed. "Trouble" here means poor representation of curved boundaries--other elements onthese boundaries may need to curve to follow curved surfaces, or it may be desired to have acurved fit to an outside edge. If flat element sides cause trouble on the perimeter, then start

by meshing areas on the other side of the perimeter with elements that have curved sides--these elements could even be triangular. Next, mesh the area of interest with the elementssides set straight, then clear the surrounding areas, if the surrounding areas are not intendedto be meshed, or need better element shape control. This will leave the plane area of interestmeshed with elements that have straight edges in the interior, and curved edges on the

perimeter. This is illustrated by the following images of an intentionally extreme example. Inthe first image, a line plot of element edges shows extreme distortion in the plane. Anintended hole is meshed with triangles. All these elements are Shell 93, having mid-sidenodes.


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In the second image, meshing with mid-side nodes positioned on straight lines is beingchosen.

In the third image, the consequence of meshing the part with straight-sided elements isshown. The elements at the hole have a curved side, because the hole is already meshed withcurved-sided elements.


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Unsupported element types cannot be used in ANSYS Linear/Plus; neither can too large awavefront (can the PCG solver get around this?). The unsupported elements need to bedeleted or changed before moving the model (e.g. change SHELL181 to SHELL63). Then, ifthe number of entities does not exceed ANSYS Linear/Plus limitations, the database can bemoved to the other program.The next problem in moving models to ANSYS Linear/Plus, is that nonlinear material

models must be deleted in ANSYS Mechanical (Structural, etc...) before moving thedatabase to ANSYS Linear/Plus. This is because the ANSYS Linear/Plus program willcomplain that the material nonlinearity is included, but not accept the commands to delete it(Hint: ANSYS should add this delete function to Linear/Plus.) Of course, I found all this outthe hard way.On rare occasions, a model from a more recent version of ANSYS may be moved back to anearlier ANSYS version. If IGES is not satisfactory, a user could use CDWRITE to write outthe element and node model and other model data to a file (the DB option), then manually clean up the file so that the earlier version of ANSYS could accept it. This includesmodifying commands for element creation, after deducing what format is needed. Writingthe element data with EWRITE then cutting and pasting with the CDWRITE file may beeasier -- I haven't tried it. A user-written program can expedite cleanup for a large model.

Tip 22: Deleting Nodes that have Nodal Coupling:When deleting a set of nodes for which some were members of coupled node sets, delete thecoupling equations BEFORE deleting the nodes. Otherwise, unwanted coupling equationsmay be active if you create more nodes. The coupling equations are not automaticallydeleted when the nodes are deleted--is this a bug? Select the nodes to be deleted, then deletenode coupling equations for which any nodes are selected, then delete the nodes. (You willhave had to first delete the elements.) Clearing solid model entities is the same as deletingthe elements and nodes simultaneously.I find it very helpful to turn on the symbols for nodal coupling when checking for proper useof these details.Tip 23: Convergence with Shell FEA Models in Nonlinear Analysis under ANSYS:

First, remember that there are three basic kinds of nonlinearity: (1) Large Displacement(geometrically nonlinear) analysis, and (2) Plastic Material properties are the obvious types.In addition, nonlinear solutions occur (3) when nonlinear elements such as gap elements,hook elements, and surface contact elements are used. Because of (3) it is clearer to refer to a"linear" analysis as "small displacement elastic", since "linear" may be perceived as meaningthat there are no nonlinear elements present. A nonlinear analysis will take longer, usuallyconsiderably longer, than a linear analysis. For a large finite element model, it helps to havea computer with an extremely fast CPU, large RAM, large hard drive, and fast hard drivedata transfer (high-speed SCSI may help on PC's) for nonlinear analysis.In ANSYS, the Shell 63 element will do Large Displacement, but is NOT capable of materialnonlinearity (plasticity). Shell 43, Shell 143, and Shell 181 are capable of both LargeDisplacement and material nonlinearity. These four elements are 4-node quad elements.

ANSYS also has an 8-node shell element, Shell 93. The Shell 93 element is capable of bothLarge Displacement and material nonlinearity. Shell 93 has the advantage that it can follow acurved surface. There are also shell elements for composite materials and for P-elementsolutions. I will restrict my comments to the basic shell elements: 63, 43, 143, 181, and 93.The elements should have acceptable aspect ratios, not be ridiculously large or small, not be

pathologically deformed, and not generate warnings about being warped. If warped quadelements are unavoidable during meshing, it may be desirable to use either small triangles, orthe Shell 93 element. Note that within the ANSYS manuals, high order elements are notconsidered to be ideal for nonlinear work. However, I seem to have had some success with


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the Shell 93 element (can't say if the results were ideally accurate). You can evaluate themodel quickly by doing a partial solve (Partial Solu in the GUI), only generating the elementmatrices, and getting warnings (if any) and other information in the ANSYS Output window.If a Large Displacement solution is chosen, some solutions are improved by setting StressStiffening before running the solve process. Stress stiffening for elements 63, 43, 143, and181 can (apparently) only be set with one of the KEYOPT values (Keyopt(2)) for the

element (see Options when using Add/Edit/Delete to add element types with the GUI). Some beam elements are like this, too. It apparently (I find the manuals difficult to interpret onthis) can NOT be set within the Solve module, even though the GUI has a selection box forStress Stiffening. However, I seem to have had convergence differences with Shell 63 withstress-stiffening set and not set in the solve module. For Shell 93, stress stiffening IS setwithin the Solve module, by choosing it under Analysis Options in the GUI (SSTIF). The useof stress stiffening for convergence improvement is contraindicated by some conditions suchas the substantial use of nodal coupling or nodal constraint equations... see the ANSYSmanuals on this. Note that SSTIF is NOT the same thing as the command PSTRES.A second thing that helps many nonlinear solutions (both Large Displacement and plastic) toconverge when substeps are being used is to activate the Predictor (PRED) in the Solvemodule. (This may be more of a hindrance than a help when gap and other nonlinear

elements will be changing status frequently.)There are other settings that can be tried when attempts at convergence are not working. Iusually stick to letting the program decide how to use Newton-Raphson iteration andadaptive descent in the Solve module. Under the Nonlinear settings of the GUI, the user canmodify the Convergence Criteria. I often use only convergence on forces (not moments)when analyzing shells if I am not inputting any moments directly. I usually reduce thenumber of Equilibrium Iterations to 15 when doing shell models, preferring to use smallersubsteps instead. However, in a model with gap or contact elements it may be desirable tohave a much larger number of Equilibrium Iterations. I rarely try Line Search.Making a good choice of time substep sizes is critically important in getting models toconverge. If shell models of flat plates subjected to pressure or perpendicular forces areincluded in the analysis, the shell will at first act as a flat plate in bending. Once the shell hascurved, by movement as small as half its thickness, the shell will start to carry the appliedload with membrane forces. In a model of this type, starting with very small substeps (e.g.1/100 of the full load) may be needed to achieve convergence. I would start with a very smallfirst substep, but allow the largest substep to be as large a fraction as 1/4 of the applied load.If there are no perpendicular loads but the loading is causing Large Displacement, or if

buckling is to be considered, it is likely that small timesteps will be needed toward the end ofthe force application ramp. Where there is no pressure or perpendicular force on flat shells, Iwould start with a substep such as 1/10 or 1/4 of the applied load, but allow a minimumsubstep as small as 1/100 of the full load. If these approaches will not work, it is likely thatconvergence control commands in addition to time substep size will need consideration.If the structure is buckling or undergoing plastic failure, or "simply will not converge" it may

help to use the Arc-Length method. As I have noted elsewhere, I don't use the default Arc-Length settings. I usually start with a number of substeps (NSUBST), and don't let the Arc-Length solver increase the size of a step beyond my maximum substep size. I let the Arc-Length solver use a minimum step size that is 1/10 or 1/100 of my substep size. I let the Arc-Length solver use a maximum step size multiplier of one. The Arc-Length method canfollow a rising and falling force-displacement relationship. I findPlotCtrls/Animate/Dynamic_Results to be useful in reviewing the behavior during an Arc-Length analysis, and other nonlinear analyses. I prefer to save the results at every substepwhen doing this (Output Ctrls). When using Arc-Length analysis, it is usually desirable to set


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a criterion to stop an analysis (NCNV). I usually use maximum displacement as the criterionfor shell work.Remember to ramp up your loads, permit automatic time stepping, and in the NSUBSTcommand, allow the program to bisection by setting the maximum number of substepsgreater than the minimum number of substeps.If you are having trouble with convergence, save the results at intermediate substeps so you

can review the stress and displacements. If you are doing combined Large Displacement and plastic deformation, and having trouble with convergence, consider a study in which you do(1) an elastic small displacement analysis as a check on element shape, loading, andconstraints, (2) a Large Displacement elastic solution, and possibly (3) a plastic smalldisplacement solution. If these work without significant warning messages, you should bemaking some progress. If gap or contact elements are being used, consider (4) softening thenormal and tangential stiffness values in a preliminary analysis (KN and KS). You can also(5) try relaxing the convergence criteria on force and/or moment error. If desperate, acoarsely meshed model may improve speed enough for you to study what helps get ananswer. These preliminary studies may help you to find what settings help you to getconvergence or discover modeling problems before you do more time-consuming accurateanalysis. If you are trying a new technique, consider testing it on a toy-sized problem, before

applying it to a large industrial-sized problem that runs for hours or days, in order to learnthe peculiarities and pitfalls of a particular time-consuming method.If gap or contact elements are the only nonlinearities in a model, consider substructuring thelinear regions of the model. This can result in a tremendous increase in solution speed. Ifonly a sub-region of a model will behave in a nonlinear manner, it may reduce solution timeto substructure the region that can be regarded as acting in a linear manner. This speedupeffect or may not occur with large displacement modeling, when the substructure itself will

be undergoing large displacement -- I have done only limited testing of this technique. See below for a brief discussion and for simple examples of substructuring.Tip 24: Working with Load Step Files in ANSYS:

Load step files can be used to automate the application of a number of different load cases ona structure. A load step file contains loads on elements and nodes. It does NOT contain loadson geometric entities. Consequently, a load step file can be generated after all loads fromgeometric entities have been transferred to a model. After all loading on geometric entitieshas been deleted, the load step file can be read back in, recovering all applied loads.Alternatively, consider the "LSCLEAR,SOLID" command. These loads can then be scaled.The user needs to be careful when manipulating load step files. The load step files maycontain the KUSE instruction telling ANSYS to re-use the TRI file if the constraints have notchanged. If the user deletes a load step file, changes the order of their execution, or manuallymodifies their contents, invalid analysis might result.If the model is re-numbered after load step files are generated, the node and element numbersin the load step file will no longer be synchronized with the model, and will be invalid. Away around this is mentioned elsewhere in these notes (See Tip 21).

The reader should take note of the ANSYS user guides comments on the LSCLEARcommand. This deletes all loads and resets all load step options to their defaults. This can"clean up" the load step data before using LSREAD to read a load step file for modification.What this implies is that the load step execution process does NOT execute an LSCLEARcommand when a load step file is read in. If it did, then ANSYS would have to implementsubstantial checking to see whether a TRI file was safe to re-use, under the frontal solver(TRI file re-use saves considerable time). Load step implementation can cause havoc whenthe user employs load step files in a manner for which the method was not designed. It mayhelp to read the contents of the LSSOLVE.MAC macro in predicting what will happen, and


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to see what LSSOLVE does to avoid trouble. The LSSOLVE.MAC macro at ANSYS 5.3includes some undocumented commands including DMARK, FMARD, SMARK, BMARK,and a *GET command that retrieves the error number in the /SOLU process. It also uses an"LSCLEAR,SOLID" command that removes loads on geometric entities before reading inload step files. It selects all DOF labels, sets xCUM labels to "replace", and does a few otherthings. I do not consider the manuals to pursue this topic adequately -- a user ought to read

the macro.The ANSYS manual comments on the LSREAD command. The command does NOT clearALL current loads on the model when it reads in a new load step file (it does clear some...read the manual).When using load step files: If loads on nodes and elements are set with BF and BFEcommands (for example applying temperatures for a thermal deformation stress analysis),then if you set up a subsequent load step, if these temperatures are to be returned to ambientit may be necessary to use the BF and BFE command to set the nodes and elements to thereference temperature (by default 0) rather than just deleting the loads using BFDELE andBFEDELE and using BFUNIF to input the uniform temperature. It may help to usecommands such as "nsel,s,bf,temp,-999,99999" and "esel,s,bfe,temp,-999,99999" to select allof the nodes or elements to which temperatures have been applied, if you are going to change

them. Be very careful with the BFE command. If you set the value of the temperature at, forexample, four locations on an element with BFE, and in a later load step set the value at onlytwo locations within an element, the temperature at the other two locations will still be"hanging around" at the previous value. It is very easy to make this mistake when running aseries of load step files. (Another thing I found out the hard way, in a model where both

piping creation commands and beam elements were used.)If the user is deleting displacement constraints using DDELE, and then writing an additionalload step file, the old constraint may still be present when the series of load step files is readin under LSREAD; check for this in your results. Be careful with this. It may compromisethe use of load step files, or require some intervention like writing an input file that calls loadstep files in using LSREAD, implementing fix-up commands as needed -- be careful that aTRI file is not re-used because a load step file contains "KUSE,1" when your changes toconstraints mean that a new TRI file should be generated. Statements in the LSSOLVE.MACmacro can provide guidance on using LSREAD effectively. You may need to look inside theload step files with a text editor. Be warned that changing the contents of load step files witha text editor can be tricky because of unintended side-effects.In general the user will have to be careful that the "residue" from the loads anddisplacements from one load step do not appear inappropriately in later load steps. This istrue when generating the load step files in the first place, and may apply when reading inload step files with LSREAD. As noted, LSSOLVE.MAC uses cleanup statements.The user will have to be careful to change loads between load steps in a manner consistentwith getting smooth ramping of loads and displacements, for those cases when this isdesired, either for transient analysis, or for good nonlinear analysis convergence, or when

intermediate results are desired at in-between loads.Before reading in load step files to solve with LSREAD, ensure that loads on geometricentities and elements and nodes have been deleted, unless you are keeping them intentionally(as noted, loads on geometric entities overwrite loads on elements and nodes). As noted,LSSOLVE.MAC in ANSYS 5.3 contains the command "LSCLEAR,SOLID" to remove thesolid model loads on the model before proceeding.If Large Displacement analysis is going to be used in analyses run by load step files, the

NLGEOM flag must be set in the first load step file. There will be no NLGEOM commandgenerated in subsequent load step files. Because ANSYS does not permit the kind of analysis


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to be changedwhen applyinga series of loadsteps, errormessages will

be result if the

user changesthe value of

NLGEOM inthe middle of aset of load step files.Tip 25: Plotting Shell Stresses -- Surface, Mid-Plane Stress, Load Paths, ESYS and

RSYS:

In the ANSYS database, shell stresses (and strains) for the basic shell elements (63, 43, 143,181, and 93) are reported at the top and bottom surfaces of the shell element. The user canhas four options in ANSYS 5.3 for plotting shell stresses (and strains). Three of them areselected with the commands: "SHELL,TOP" , "SHELL,MID" or "SHELL,BOT". These willcause plotting of shell stresses (and strains) to be based on the values at the top surface, mid-

plane, or bottom surface of each shell element. This is a bit misleading. The mid-plane stressis based on the average of the stresses at the top and bottom (this may not be correct, at least

for some elements, considering Section 2.3.4 of the Theory manual, which refers to stress on

the mid-plane of a shell element separately from the top and bottom, and forms the force per

linear unit from a weighted average of top surface, mid-plane, and bottom surface stress --

what's going on here?). What constitutes the top and bottom of a shell element depends onthe element's orientation when it was defined (see elsewhere in these pages). It is possible tohave adjacent elements, one with a "top" surface pointing upward, and its neighbour with the"top" surface pointing downward. In complex structures it happens all the time. If nodalaveraged plots are done, for example with "PLNSOL,S,EQV", when either top surface or

bottom surface plotting is chosen, then with such adjacent elements, the plotted top surfaceand bottom surface results will get blended, causing a misleading mess to be displayed. (SeeTip 19 for commands that can re-orient shell elements.)More insight into the flow of stress in a model can be gained by plotting the stress vectors,using the "PLVECT,S" command. With shells, these vectors will be plotted for the mid-

plane principle stress components. At times you will want to use vector graphics with nohidden surface removal, to give the best view of these vectors. If there is local compression,the vectors point inward. These vectors can give insight into load paths in a structure.Where there are intersections of planes of shell elements, e.g. corners or "Tee" intersections,or where elements of differing thickness meet, the averaging of node stresses can renderlocal stress plot information meaningless at the intersection. This is true of both surface andmid-plane stress plots. This is one way in which excessive stresses will be unintentionallymissed.

Any time that nodal averaged plotting is done, it is possible for the averaging to "wash out"local stresses that may be important, yet it is common to do nodal averaged plots because oftheir much cleaner appearance (I do them myself). The fourth option in plotting shell stressesis to switch on the ANSYS Powergraphics feature. This causes shell results to be displayed,even averaged, for the visible surface. Options activated with the AVRES and /EFACETcommands can refine the way the results are plotted under Powergraphics (look them up).Powergraphics has the options to discontinue the averaging of stress contours where thereare certain discontinuities in the material or geometry in the model. I'm going into this detail,


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because a high stress that is washed out by nodal averaging could be a stress that causesserious fatigue or other damage, such as cracking, or a weld being torn apart.The only shortcoming is that Powergraphics will not work with mid-plane stress. The userhas few options here. Sometimes it is important to select only regions of a model when doingnodal averaged mid-plane stress plots (using "SHELL,MID", without Powergraphics) so thatthe averaging does not wash anything out. A mid-plane stress plot without Powergraphics

can be done for element stresses, using a command like "PLESOL,S,EQV". This will lookmessy, but at least it doesn't hide an extreme stress. An alternative I used is discussedelsewhere in these pages: I wrote a macro to get the mid-plane averaged stress (allcomponents) at every node of every element (a given node has different results withreference to each of the elements to which it is attached, so a given node will be looked up asmany times as the number of elements to which it is attached), and transfer it to the top and

bottom surfaces, so that Powergraphics would plot mid-plane averaged stress neatly, withdiscontinuities. CAUTION: This ruins the results database. The macro is extremely slowto run. The method (under Powergraphics) does, however, give far better looking plots thanusing the "PLESOL,S,EQV" command to plot mid-plane element stresses without nodalaveraging (without Powergraphics).

LOAD PATHS: The macro I mention above could be modified to multiply the mid-planeaveraged stress components by the local shell element thickness at each node. The resultingvalues would yield a contour plot of force per linear inch (or other dimensional unit)"averaged" at the mid-plane of the shell -- this could help to make load paths visible in acomplex shell structure. "PLVECT,S" plots that would now show arrows corresponding tothe load-per-unit-length on the mid-plane and show the principal directions in which it

points, helping to illustrate the load paths. This macro would also ruin the database for anyother use. Before plotting "load-per-unit-length" data, the user needs to decide how to orientthe results data coordinate systems with RSYS for information such as Sx or Sy that containsdirection information (stress and strain with EQV does not contain direction information).

Note: The Output Data section on Shell63, Shell43, and Shell93 includes In-plane element X,Y, and XY forces called TX, TY, and TXY. Consequently, shell "force per unit length" datacan be obtained directly in an Element Table very quickly, though with a resolution of onevalue per element. (For Shell63, 43, and 93, use SMISC setting 1, 2, or 3 when generatingthe element table data.) The Theory Manual uses the term In-plane forces per unit length


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while the elements manual refers to just forces as above -- a simple test I ran shows the datato be force per unit length. The elements manual ought to clarify this. The Element Tabledata can be contour plotted, but there are no principal stress style vector plots of table data.(Clarification: PLVECT can plot vector arrows based on 3 ETABLE columns, but not thedouble-headed arrows for an ETABLE as in a principal stress vector plot.) The Elementsmanual shows the TX, TY, and TXY values not being available under "Miscellaneous

Element Output" at every node, only at the centroid. The Elements manual does notexplicitly show that S,EQV or S,INT stress information can be extracted at the mid-plane.Their value is extracted with the component name method. Brief experimentation shows thatif the command "SHELL,MID" is followed by "ETABLE,SEQVMID,S,EQV" that thecolumn called SEQVMID will contain an average SEQV value for the mid-plane. If"SHELL,TOP" or "SHELL,BOT" is called, the ETABLE value of SEQV will change if theupdate command "ETABLE,REFL" is executed. Warning: When plotting ETABLE shellelement element table data with PLETAB the plot information legend will read TOP, MID,or BOT according to the current setting of the SHELL command. This bit of informationDOES NOT reflect the SHELL surface setting conditions in effect when the ETABLE datawas stored, and could be misleading. For this reason, the label used for the column shouldindicate the shell layer setting in use when the element table data was loaded, as with

"SEQVMID" above. Doing an element table update with ETABLE,REFL will re-fillcolumns with results data. A change of the SHELL layer setting can change stress results thatare loaded in an update. Consequently, loading shell element data must be handled verycarefully in order that the layer choice is controlled. Element table data from the CALCmodule (adding columns etc.) is NOT updated and has to be explicitly re-calculated.

NOTE Also: The direction of the element table load-per-unit-length TX, TY, and TXY is astaken from the element in Element Coordinates. Unlike SX or SY, the values of TX, TY, andTXY appear to be insensitive to the RSYS setting. The Element Coordinate System will varyorientation from element to element, particularly under free meshing, and affects theusefulness of TX, TY, and TXY data. The element table data can be processed by the user toyield a new table column containing the "load-per-unit-length intensity" in the sense of aMohr's circle, giving rapid if somewhat coarse plots of load path information along the shellmid-plane. The plots will usually be more informative without nodal averaging. Section 2.3.4of the ANSYS Theory manual discusses Forces and Moments per unit length on shellelements -- the suggestion is that internally, at least for some shell elements, the mid-planestress is NOT simply the average of the top and bottom stresses. The way around the

problem of element coordinate systems being arbitrarily oriented is to define local coordinatesystems before meshing areas (or otherwise generating shell elements) and use ESYS to getall shell elements oriented with the local coordinate systems. ESYS assigned to elements can

be modified after the fact but before SOLVE, by using the EMODIF command in /PREP7. Itmay be desirable to have a local coordinate system aligned with each flat area to be meshedwith shell elements so that all shell element coordinate systems can be aligned in the plane ofthe area -- a time consuming process unless a macro is used. Curved surfaces would be

difficult.The problem of orienting coordinate systems in the plotting of results is illustrated by theimages below. The first shows 3 elements that were created during free meshing. Theelements are plotted using vector graphics, with the element coordinate systems shown. Eachelement has its coordinate system oriented differently. The image below it lists the elementsand their node numbers. Look at the sequence of node numbers for the three elements to seewhy the element coordinate systems point in such different directions.


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The next two images show a plot of TX done from an element table. The element table wasfilled by the TX values for the elements (this is the load-per-unit-length in the elementcoordinate system X direction). The values differ so much from element to element becauseof the difference in the element coordinate systems. The plot consequently tells us too little.The following element plot of Sx shows the stress in the X direction. The results are shownin the global coordinate system.


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The final images in this section show a group of Shell63 elements that have had theirelement coordinate systems aligned with local coordinate systems at the time of the creationof the elements, by the use of the ESYS command. This will permit element table results TX,TY, and TXY to be aligned in a known manner. This also permits Sx, Sy, and Sxy to bealigned in the plane of the elements creation if RSYS,SOLU is active when plotting stressresults. Knowlege of the alignment of the loads and stresses can make plots more useful in


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understanding load paths, reduce the total number of plots required in model assessment, andhelp facilitate an evaluation of loading on welds. The first plot with vector graphics showsthe elements with their element coordinate systems. Note that they are aligned. There are twolocal coordinate systems at work in this example -- they are numbered 11 and 12 and theirsymbols are plotted. Elements have been created aligned with number 11 in one plane, andaligned with number 12 in the other plane of elements. A line pressure has been applied in

the global -Y direction. The second plot with raster graphics is of Sx at the shell mid-plane.Because RSYS,SOLU was active when the Sx plot was generated, there are Sx values shownin all elements. If RSYS,0 were active when the Sx plot was done, the plane of elements thatis perpendicular to the global X axis would show zero stress in the X direction in thisexample.


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There is an alternative to using ESYS and RSYS,SOLU to align element coordinate systemsfor the purposes of stress plots like Sx, Sy, and Sxy. During postprocessing in /POST1, alocal coordinate system can be aligned with the plane of shell elements of interest, and RSYSset to that local coordinate system, before plotting Sx, Sy, or Sxy. However, this would donothing for TX, TY, and TXY which depend on the element coordinate system and aregenerated in an Element Table.I leave the topic of whether to plot surface or mid-plane shell stresses to the reader todetermine. Too much is industry or application domain specific. Hint: Check mid-plane plus

both shell surface stresses. Surface stresses and strains can cause local bending, cracking, breaking of protective coatings, fatigue, and imply possible overload or prying of welds andfasteners, and can highlight other troubles.Tip 26: Nodal Coupling (CP) versus Rigid Region (CERIG):

I have seen analysts mistakenly use nodal coupling where rigid region constraint equationsshould have been employed. (The nodes concerned were not at the same location in space.)Rigid region constraint locks together a selected set of nodes so that they translate ANDrotate in space as if they were locked together by an infinitely stiff structure. Nodal couplinglocks together selected degrees of freedom (translation and/or rotation) individually, so thatthe same degree of freedom value will result for the nodes in the coupled set. Nodal couplingwill not combine the rotations and translations that are necessary to imply rotation as a rigid

body in space.


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Note that rigid region constraint may not be appropriate for Large Displacement, when thedisplacement rotations are significant (sin(theta) differing from theta, etc.). This is becauseANSYS uses a linear approximation to the rigid body rotation matrix. A rigid regiongrouping can be implied by tying nodes together with extremely stiff beam elements (zero-mass beam elements a few orders of magnitude stiffer than the structure to which they areattached.) The beam elements should have the advantage that they work under Large

Displacement. The beam elements should not be too stiff, or ill-conditioned matrices couldresult. If the beams are of very widely varying lengths, then some may be too stiff, others tooflexible -- remember that flexibility is proportional to length cubed.I ran a model in which about one thousand beam elements were used to position gapelements. These beam elements would ideally have been infinitely stiff. I needed elements,instead of nodal coupling or constraint equations, because of thermal expansionconsiderations. The beam elements were widely varying in length. This created solvertrouble, until I wrote a macro that assigned each beam element a unique REAL value, whichset values for each BEAM4's Ixx, Iyy, Izz, and Area as a function of the element's length. Ifound it sufficient to set their stiffness a couple of orders of magnitude stiffer that contactstiffness for the gap elements.Turning on the symbols for nodal coupling and for nodal constraint equations is very helpful

in reviewing the correctness of a model.Tip 27: Vibration Modes with Pre-stress:

Calculation of natural frequencies and modes of vibration CAN be done with pre-stressing ofthe structure under ANSYS. There is a "PRESTRESS" flag to set under modal analysis. Thisis available in the dialog box for Modal Analysis Options. First, do a static analysis with the

prestress flag set. Exit Solution (click Finish or enter "/fini"). Re-enter Solution, and do amodal analysis with the prestress flag set again. This does not seem to work when the stressrun is done with Large Displacement activated.I leave the question of how a performer plays music with a hand saw and a violin bow as an"exercise for the reader" :-)Tip 28: Creating New Elements by Copying or Reflecting Existing Structure:

In order to create new elements by reflecting or copying existing elements, there are a fewthings to do. First, select the elements to be copied and get their nodes with NSLE. Copy orreflect the nodes, noting the nodal number offset that will be used -- write it down. Copy orreflect the elements, using the nodal offset number that you wrote down. ANSYS shoulddefault to a nodal offset number equal to your highest numbered node. If you make itsmaller, you run the risk of changing the location of nodes that already exist, resulting in alovely mess. If you are running something like ANSYS/ED you may want to compress yournode numbers first, for if a node number results that exceeds the ANSYS/ED limit, the

program will terminate immediately (the more recent ANSYS revisions may give a non-fatalwarning message and quit some time later if you don't clean up). You could compress thenode numbers, and then make the offset number equal one plus the difference between themaximum node number of the whole model and the lowest node number of those nodes to be

copied or reflected. You can find these node numbers with *GET commands. (Rememberthat compressing node or element numbers will destroy synchronization with Load Stepfiles.)The same nodal offset number will need to be used if nodal coupling is to be copied as well.In order to copy nodal coupling, use "Generate Coupled DOF Sets with same DOF" forwhich you will need the same nodal offset number. Do a replot to see the newly creatednodal coupling. Caution: Be sure that if nodes were deleted earlier, that nodal couplingequations that in the past included those deleted nodes were also deleted. If you forget, youmay get a pretty mess.


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Remember that if there are nodes on the plane of reflection, new nodes will overlay them.Merge commands may be wanted for the nodes on the reflection plane. Now the tricky part:elements lying in the reflection plane (shell elements will do this) get generated with thenode order reversed, because of the mirror imaging. They Will Not Merge with the elementfrom which they were reflected. They may have to be deleted, depending on what you aretrying to accomplish. Alternatively, do not select elements that lie in the plane of reflection

when reflecting the structure. You still need to reflect the nodes on the plane of reflection, inorder to reflect the elements that will join them to the remainder of the reflected structure, sothe nodal merge will still be needed.Tip 29: Adding to a Model Comprised of Elements and Nodes Only:

It may happen that a model that consists of nodes and elements only has to have a sectionreplaced, or requires the addition of more structure. The way to attach new geometry ontoexisting nodes and elements is to: (1) Place keypoints on the nodes onto which newgeometry is to be built (i.e. grafted). (2) Join these keypoints with lines. (3) Set mesh densityalong these lines to only one element. (4) Build new geometry outward from these keypointsand lines. This gets messy if you are building solids. (5) Mesh the new geometry. (6) Selectthe nodes (new and old) along the interface between the old nodes and the nodes of the newgeometry. (7) Merge ONLY these nodes along the interface using the NUMMRG,NODE

command. Alternatively (much more work unless a macro is written or the CPINTFcommand is used correctly), fully couple the PAIRS of nodes with the CP command. In theevent of elements with mid-side nodes, lines will have to be created curved so that a singleline spans three keypoints placed on the three nodes along the edge of an element. It is

probably advisable to connect elements with mid-side nodes to other elements with mid-sidenodes.This attaches the new geometry and mesh to the old elements and nodes. Be sure to doublecheck that the merging has been done correctly and according to your intentions -- I havefound this to be a surprisingly error-prone operation.Tip 30: Zero Mass Beam Elements Form Rigid Region:

An analyst could use very stiff beam elements (a few orders of magnitude stiffer than thesurrounding structure) in order imply a rigid region grouping of nodes, which works underLarge Displacement (a CERIG group does not work with large displacement). This is an oldFEA trick -- it is not perfect. A separate material should be created for these beams, and begiven zero mass (set the material density to zero) so that no gravitational or other inertialload acts on the material. A thermal expansion coefficient should be input if appropriate -- itwould usually be identical to the coefficient value for the structure that it approximates.I wrote a macro to create a rigid region using beam elements. It is called after the set ofnodes to be connected is selected. The lowest numbered of the set of nodes is attached toeach of the other nodes in the set by a beam element. The beam element to use has to be setup in advance, and the appropriate MAT, REAL, and TYPE set by the user. A macro likethis is very fast to run. Caution: Such a macro would become complex if it checked forduplicate nodes at the first node location (ANSYS can't use zero length beams), and checked

for widely varying beam lengths. This is not a guaranteed method.Tip 31: Turn off Symbols When Changing a Model after Solution:If you have run SOLVE, the results database will be full of data. If you then change a model,and create anything that plots a symbol, all symbols become active, and plots becomeextremely slow. Turn off symbols with /PBC,ALL,,0 to speed things up. I put this commandin the Toolbox for convenience. I have found that plotting can become slow with very largemodels when loads have been applied, and even when applied and deleted. PresumablyANSYS is checking to see if any symbols should be shown. The plotting speeded up


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considerably when symbols were turned off with "/PBC,ALL,,0" even though there were, infact, no symbols to be plotted.Tip 32: Are the "Free-Free" Vibration Modes Relevant?:

Simple supports on a structure may be appropriate for static analysis and gravity loading,since the structure will "sink" until the simple support reacts enough to withstand the appliedload. If a modal vibration is excited, small amplitude vibrations

may result in very little response from the support, and vibrationsimilar to a structure that is free in space may result (this isobviously very problem dependent). If so, it may be desirable torun a modal vibration analysis with no constraints. More than sixmodes must be requested, since the first 6 represent the freetranslation and rotation, and give Zero eigenvalues. A betterapproach would be to characterize the flexibility of the constraint

points. With some structures, you may get a few surprises, astorsional and other vibration modes appear.Tip 33: Selecting a CAD or FEA System -- Cover Yourself

It is common to evaluate a few CAD or FEA packages when trying to make the right choicefor a purchase. Watch out for this stunt: (I've seen it done, and been threatened with it once

(I laughed at her).) A losing vendor writes a letter to your boss, or even to the head of yourcompany, claiming that the engineers are incompetent (stupid, uninformed, can't spell, and soon) and making a huge mistake. If the boss is not an engineer and cannot understand theissues, this could get awkward. (Certain Dilbert cartoons come to mind.) Warn your boss(es)in advance that a few vendors pull this move and that you and your group will evaluate the

products in a thorough manner. Write down some criteria and your assessments. Also, becareful that you cannot be accused of leaking information unfairly from one vendor toanother -- date your correspondence carefully, and work through your purchasing departmentif that is appropriate at your firm. Some sales-types are very greedy for their commissions,and petulant when they lose. (Names will not be mentioned, to protect the guilty. If you've

been around the block a couple of times, perhaps you can make a few guesses.)(The ANSYS vendor I've dealt with has been very professional.)

Tip 34: Creating Lines Perpendicular to, or at Angle to Existing Lines

When creating structures in the /PREP7 portion of ANSYS, I find that the commands thatcreate lines that are perpendicular to existing lines, or at an angle to existing lines, areextremely useful. Look at the commands LANG, LTAN, L2ANG, as well as the others.These commands break lines where new lines intersect, even though the original lines arealready attached to areas. Since I often model shells that are to act as if they are weldedtogether, I need the lines to be shared where areas contact each other. These commands givethe connectivity I need.The command that meets another line at an angle may do better if it is entered manually,with the first guess of the contact point set at 0.0, 0.5, or 1.0, depending on your intention.

This often succeeds when the interface command fails.Tip 35: Use the /UI command in Your ANSYS Toolbar to Bring up GUI Dialog BoxesTake a look at the /UI command in ANSYS. You can use it in your Toolbar to activatecertain GUI dialog boxes with one-click simplicity, instead of finding your way though themenu system. I sometimes get odd results from the Hard Copy command when I do this -- Ihave no idea why.Tip 36: Reaction Force, Nodal Force, and Load Paths

I worked on a model subject to aerodynamic pressure and gravity load. We needed to knowthe load that the structure would apply to its foundations. Printing the Reaction Force would


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give this value, however the +/- sign is in the direction of the force that the constrained node(or nodes) applies TO the structure. If nodes are selected with the three commands

NSEL,S,D,UX $ NSEL,A,D,UY $ NSEL,A,D,UZ the Nodal Force at the constrained nodescan be printed. This is the force with which the nodes press on the supports. (NOTE: Youmay need to include nodes where there are constraints on rotation, depending on what youare modeling.)

WARNING: A number of things can go wrong with this approach.1. If you ask for nodal forces without limiting the node selection to nodes where there

are constraints, you will get nodal forces wherever forces and pressures have beenapplied to your structure. (For the curious, printing nodal forces when only pressurehas been applied to shell and high order elements will illustrate that FEA softwareinputs a complex set of forces and moments because of how the elements are derivedfrom first principles. What is being printed is the force with which the nodes react tothe forces input from outside -- if a moment is input, a nodal "force" moment isoutput in reaction.)

2. If an input force has been applied to a constrained node, the nodal force and thereaction force magnitudes will differ. When I tested this, the reaction force thatANSYS listed was modified by the presence of a force applied directly to a

constrained node, whereas the nodal force (that is based only on elementdeformation) was not affected.

3. IMPORTANT: All the elements to which the selected node is attached must beselected in order to get the total force with which the node pushes on the outsideworld (use ESLN after selecting the nodes). The generation of Nodal Force (andReaction Force, if I remember correctly) is determined from the deformation andstiffness of attached elements. If elements attached to a node of interest are notselected, then the contribution of those elements to the force at the node is notincluded and will be missing.

4. Caution: If you have used a rigid region with the node of interest, the lack of elementdeformation means that you will NOT get the Nodal Force or Reaction correctly --you may need to work from the set of nodes where the rigid region attaches to theflexible part of the structure. I'm not sure what kind of effects nodal coupling wouldhave.

There are various other uses to which you can put Nodal Force. You can plot the NodalForce vectors along with your model (see the /PBC command), after SOLVE, giving visualcues during your review. You can use NODAL FORCE to find out about the load beingcarried in certain Load Paths:

• Determine where to position a "cut" in the model. Locate it where you want todetermine the force carried across the cut. The "cut" should follow a path along theedge of a set of adjoining elements. Select all the elements on ONE side of the "cut".

• Select the nodes on the "cut" side of those elements.• Printing the Nodal Force (forces only) will tell you about the forces that your selected

elements apply to those nodes. The sum that is printed tells you the total force carriedacross the "cut" in the X, Y, Z directions, based on the selected elements.• Caution: Getting the moment across the cut is not so easy, because moment is

determined with respect to an axis. You would have to do extra work to pursuemoment across a cut, determining your "neutral" axis, and using other commands.See, for example, ANSYS manuals information on the SPOINT command.

Note my earlier comment Tip 25 on making load paths visible in shell models. For furtherinformation, read the ANSYS manuals on the FSUM and NFORCE commands.


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Tip 37: Inputting Temperatures with BF, BFE, and TUNIF in Structural Analysis

As discussed by ANSYS in Chapter 2.6 of the Elements Manual, Body Loads (temperaturesfor structural analysis that cause thermal strains and affect temperature dependent material

properties) may be input in a nodal format or an element format. " Either the nodal or theelement loading format may be used for an element, with the element format taking

precedence. Body Loads are designated in the "Input Summary" of each element." This

means that if both BFE and BF are applied to an element and its nodes, and the inputs differ,the BFE setting will govern. If temperature is input on a nodal basis, the temperature input ata node will influence all the attached elements. If temperature is input on an element basis,the temperature(s) input will influence only the element to which it was applied. Thecommands TUNIF and BFUNIF can be used to set all nodes to one default temperature thatdiffers from the reference temperature. Then, BF or BFE commands can to used on specificregions of the model to put in other temperatures.If you use piping commands to

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