49054430 Ansys Tips and Ansys Tricks

ANSYS Tips and ANSYS Tricks http://www3.sympatico.ca/peter_budgell/ANSYS_tips.html

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ANSYS TipsANSYS Tips and ANSYS TricksPeter BudgellBurlington, Ontario, Canada

1998, 1999 by Peter C. Budgell -- You are welcome to print and photocopy these pages.

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 the concepts of the finite element method that are required to obtain successful solutions. It is your responsibility to determine if you have sufficient knowlege and understanding of finite element theory to apply the software appropriately. I have attempted to give accurate information, 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.

As one writer put it, This information is free, and may be well worth the price.

Return to Home PageFEA and Optimization Introduction PageFEA Modeling Issues Page

The ANSYS manuals explain many things and give some examples, but they do not give many tips to the user. Here is a collection of things I have noted or learned. (Use at your own risk...) Necessity is the mother of invention, and I learned virtually everything here as a result of need, or as a result of trial and lots of error. I'm also thankful to my local ANSYS distributor for many helpful conversations. The comments in these pages are based on my experience with ANSYS 5.0 through ANSYS 5.3. I hope these tips will shorten your learning curve.An analyst frequently does not have a mentor for guidance, so considerable effort can be needed to deduce how to accomplish some tasks. ANSYS users need to spend a generous amount of time reading the manuals and training materials, and returning to read them again as the user's knowledge of the program increases. Don't use anything here verbatim... understand why it works, and whether my comments are in error or inappropriate for your situation, before employing any of these suggestions.

The teaching of FEA at the academic level is intended to educate the mind, teach how FEA methods are derived from first principles, and to develop students who can invent and code new elements, test their behavior, write research or industrial quality software, and apply it to difficult academic or research problems. Some professors feel strongly that the purpose of an undergrad course in FEA is further education in how applied math, engineering, continuum mechanics, energy methods, and analysis of structures come together, building on the Strength of Materials courses already taken -- I have no argument with that. A user with a comprehension of what underlies FEA work will know when to apply and how to evaluate FEA work, have more creativity, learn quickly, problem solve better, be more innovative, and make fewer serious modeling errors. The professors do not feel that the course is intended to concentrate on modeling details or learning the interface to a commercial FEA program. (Students, on the other hand, want to graduate having used an FEA package to do something significant. Assignments and projects with ANSYS/ED are a good way to get there.) I've heard the opinion expressed that with FEA technology maturing, there is less research grant money for FEA work in universities, and the supply of advanced FEA graduate students may be shrinking. The teaching of commercial FEA program use is principally focused on training people to use the interface to and commands of the particular software package, and how to perform basic analysis types. Some instructors pepper their presentations with tips, but the attendees may be drowning from information overload. Little is available to lead the user through the techniques that can be used in modeling complex structures, and around the traps that exist, except help from good vendor support people, co-workers, or other users, and substantial reading, thought, trying examples, and testing techniques on the part of the analyst. I hope that these pages will provide some helpful details.

CONTENTS:

Tip 1: Use Annotations


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Tip 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 CouplingTip 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 PathsTip 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 TemperaturesTip 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 MacrosTip 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 ProgramsTip 68: Writing Array Columns to Output or to Files


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Tip 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 more information can be included in annotations on the screen. The annotations are kept through all plots until they are deleted with the command: /ANNOT,DELE or via picking with the graphical user interface (GUI).

At the top of the Annotation dialog box, there is a list box from which the user can choose Text, Lines, etc., on down to Controls. These selections bring up different menus. The Controls selection offers a SNAP setting that makes it much easier to get the text aligned nicely. (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 of the 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 for annotations 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 parameter in an annotation, as discussed in the Commands Manual description of the /TLABLE command. When typing the annotation using the GUI, include the parameter in percent signs like this: %pname% where pname is the parameter name. The parameter can contain either numbers or text. The value of the parameter will be plotted in the annotation string. The ANSYS function NINT can be used to round a number the nearest integer, sometimes improving the appearance of the annotation for large numbers in which the fractional part is irrelevant (e.g. NINT(123.456789) = 123 ). For this, the parametric expression should be enclosed in percent signs. Annotations are usually created in the GUI, but can be entered with 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.! In this example, "pname" is a parameter with a numerical value such 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 other data 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 the annotation:

! 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, then when the next plot command is executed, the annotation will automatically update to reflect 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 get information that goes into the parameter that a /REPLOT will automatically put it on the screen. This makes it possible to automatically include far more information than can go into the 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 be decreased using theGUI. A size of 0.6 is quite readable and permits far more information to be packed into a plot. Note that there isa limit to the number of characters possible on an annotation line this is character size independent.

Tip 5: Mathematical Functions Available:


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Under the Help listing for the *SET command there appears lists of mathematical functions available in ANSYS.Another list is in the ANSYS User's Guide on APDL, Chapter 14 of the Modeling 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 the distribution of outputs based on randomized loadings and material properties. For an explanation, see a good modern engineering design textbook.

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 from the 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, both numerical and character. The *VFUN command has a list of functions that act on an array entry. The Commands manual lists functions that act on Element Tables in the section "POST1 Command for Element Table". Creatively used, the array and ETABLE algebra commands can be surprisingly powerful.

Tip 6: Start 16-Bit Applications before Starting ANSYS under Windows NT:


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Setting Process Priority in NT

It has been my experience that some large commercial 16-bit applications will not start properly when ANSYS isalready running. If you start them before launching ANSYS, there will be no problem. If you intend to work withthose 16 bit applications in the foreground while the ANSYS SOLVE is running in the background, this will be auseful tip. I have seen other applications start up very slowly (e.g. Internet Explorer) or wait until ANSYS was done 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 do this, 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 >" comes up. Set the priority to "Low" to help make foreground applications run more smoothly while ANSYS is running SOLVE in the background. This may help more if you have a large RAM in the computer.

When ANSYS has completed the SOLVE process, return the priority to "Normal" so that ANSYS 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) seems not to be available. If any load on your structure has been applied to a geometric entity, rather than directly to elements or nodes, that load will be transferred to the elements and nodes at solution time. The transfer will overwrite any scaling of loads that you have applied. (Guess how I figured this out!)

So what can you do about this? Method 1 : Transfer the loading from geometric entities to the elements 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 stepfile. Method 3: Transfer the loading from geometric entities to the elements and nodes, then delete the relationship 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-select the geometric entities, before executingSOLVE. The element and node loading can be scaled after it has been transferred from geometric entities. An un-selected geometric entity will 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 an appropriatefile name before executing MODMSH,DETA. Method 4 is fine, as long as you do not forget and re-select the geometric entities -- ALLSEL will do this.

Scaling displacements (nodal constraint values) is also possible. One thing that has not worked for me is an attempt to reduce applied displacements to zero by using 0.0 as the scaling factor. What did work for me was to use "_TINY" as a value, which multiplied displacements by a factor of roughly 10^(-31) and reduced loads to virtually zero. Attempts to use 0.0 as the factor resulted in NO change to the applied displacements.

Tip 9: Ramping Loads Down to Zero:


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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 deformation that results from plastic yielding. If you delete the applied load, the loading will drop immediately to zero, even if you have load ramping turned on.The thing to do is to set the loading to virtually zero or the scaling factor to virtually zero, not delete the load. It is important 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 use will depend on your model.

Setting displacements to zero or near zero is, of course, very different from deleting constraints.

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't get a t=0 data point. If you leave time as 0.0 on the TIME command, you get the default 1.0 in your output. The only way to get a graph from zero that I have found is to do a first load step 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 your intent 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 that line. The loads on the nodes that the FEA program applies will be appropriate given the formulation 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 the result as the pressure.

Note that pressure on a line acts in the plane of the area that is attached to the line. If two areas are attached at 90 degrees or another angle, two loads are set up, acting in each of the area plane directions. You can use select logic on the areas to get some interesting effects as to the direction in which the applied forces act, but only if both areas are meshed, and the elements are selected. If you un-select one of the areas, pressure on the line will only be exerted in the direction of the area that is selected. The select logic must still be in place when 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 load case, 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, even under Large Displacement. This difference will govern how loads should be applied in some models.

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 the loads 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 run this as a first load step. Then set up a second load step, with ramping activated. Change those loads to be ramped from their startingvalues 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 from zero. 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 regulate your 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 is considered good practice in your industry.) Applying gravity first can give much better convergence when assessing the effect of thermal expansion moving structures across friction contact elements, where the normal load on the contact elements is caused by gravity.

I suspect that this is not possible with the Arc-Length method. I have not experimented with it, but do not see how controlled ramping of only some loads could be implemented under Arc-Length control of applied loading -- any opinions?


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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 greater than half the thickness, then membrane forces start to become significant in resisting the applied 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 inherent strength, 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 the ANSYS 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 bucklingmay need to be assessed with respect to a number of load cases. In some structures, a diagonal tension field is 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, and only nonlinear analysis will predict this.

Linear eigenvalue buckling has to assume that gap and contact elements are either closed and active, or open and inactive. Nonlinear analysis will follow the effects of these elements as they go in and out of contact, when the loading is applied.

After any Large Displacement nonlinear elastic buckling analysis (if it doesn't diverge), see whether the elastic stress limits have been exceeded (this includes the surfaces of shell elements, and be careful that nodal averagingdoes not hide anything). If significantly overstressed, the structure may not be adequate.

Combined bending and axial compression in a beam is a classic place where inadequacy in strength 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 structures undergoing elastic Large Displacement analysis withoutcontact and gap elements, the user may want to consider a Southwell plot.

If elastic stress limits are exceeded in the Large Displacement model, it may be desirable to do a combined LargeDisplacement and Plastic Deformation model. If the structure is overloaded, 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 or identified. 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 theArc-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 aproblem; I've encountered this with shell structures under compressive stresses. The arc-length method can sometimes cope better with nonlinear solutions, because of its ability to follow force-deflection curves that rise and fall. 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 size multipliers, 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 1/10 of the load step. Set the Arc-Length maximum multiplier MAXARC to 1.0 so that no substeps larger than 1/10 of the load step are taken. Set the Arc-Length minimum multiplier MINARC to 0.1, so that the smallest load substep is 1/100 of the full load step. I found this to help


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considerably. You may want to user a larger or smaller MINARC setting, but my experience to date suggests thatone should not get greedy with MAXARC. Obviously, you may want to play with the number of time substeps.

The solution may still diverge but it is likely that you will get more information than without arc-length analysis. You will want to set a termination condition for the analysis if buckling is expected to result.

I find it desirable to save the results at every time substep when doing this type of analysis (it helps to have a large hard drive) in order to review the process. When you review the results of 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 forward through the plots, if the load/displacement curve for the structure is falling, the decimal fraction will fall, even though some displacements are visibly getting larger.

As mentioned above, something I have not tried is to get the Arc-length solution control to ramp some loads andnot others, by having run a preliminary load step. Is this even possible? If not, then the user may face the prospect of gravity being ramped up and down, in addition to other applied loads, and the physical realism of 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 analysis loading 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 of interest to set the type of stress plot to be animated by the macro that will be run. Make the ANSYS 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 files under 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-menu of choices. Choose "Dynamic Results" to create an animation of your saved load substep results 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 viewing. It makes it easier to watch the changing stress pattern or deformation as nonlinear effects take over the model.

In animating a changing stress or other contour plot, you may wish to specify the contour levels before generating the animation file. View the load step or substep with the worst results as part of deciding where to setthe 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 results file. The following simple macro does this for me under Windows NT. There is virtually no error checking in this macro. Note that this simple macro does not update element table data at each frame. Consequently it will not work properly for plots of element table data. If stresses, strains, or other data with amplitude information are to be plotted, the user may want to fix the contour map levels ahead of time. The user will want to set the displacement amplitude scaling with /DSCALE inadvance--automatic scaling will not be satisfactory. In general, it may not be satisfactory to have /ZOOM,OFF


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active, since the view will change if plots of significant deflection are included in the animation. Manually setting a view may yield a better animation. Modify this macro as you wish.

This macro must be called from within /POST1. The file that contains the results must have already 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,NEXT command.! The number of frames needed must exist in the RST file, else errors.! NOTE: This does NOT work for plots of data in an element table.! Plotting element table results would require a macro in which! 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 last user plot executed! before this macro is called.! Scaling, etc. are all based on the last user plot. Only the SET value 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 /seg,multi,,ar11 set,first /replot *do,_iii,1,arg2-1,1 set,next /replot *enddo /seg,off anim,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 *GET command of the type *GET,NTOTAL,ACTIVE,0,SOLU,NCMSS to check the number of substeps as the SET,NEXT command isissued. The parameter NTOTAL will be re-set to 1 when the animation is complete, and the *IF and *EXIT commands can check this and break out of a do loop -- see Tip 59 below for the example of automatically plotting all mode shapes. The user would then not need to specify the number of substeps to plot, improving the automation, and letting the solver use variable substep sizes without the user having to check 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 your units) 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 graphical user interface, enter into "PreprocessorOperateCalc Geometric Items" to see the choices: "Of Keypoints, Of Lines, Of Areas, Of Volumes, Of Geometry". These items execute the "sum" commands: "KSUM, LSUM, ASUM, VSUM, GSUM" respectively. If no attributes havebeen assigned to the geometric entities, unit densities are assumed in reporting mass and center of gravity information. After the execution of these commands, the *GET command can be used to assign to a variable the implied volume of an area (based on the thickness associated 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 made


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available 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 the solution 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 pounds mass and pounds weight. There is no *GET command that directly returns the weight of selected elements. However, the volume of an element can be returned, and the volume of a set 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 by the density of that material type. Sum the masses or weights of all the material types. This will not include added mass and mass elements at nodes (check this carefully against the output 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 the minimal constraints needed to stabilize all bodies in 3-D), applying 1 g vertical, having constraints on vertical motion, running SOLVE in a linear analysis, and finding the vertical reaction force. In such a run, a combination of the FSUM (select vertically restrained nodes only, with all attached elements) and *GET commands in /POST1 might help you to get the weight into a variable. However, apartial solve will give the answer more quickly (but not put it into a variable). Depending on your system of units, remember, you may want to convert between weight and mass .

I base my comment, about the inability of ANSYS to directly return the weight of the model with *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 Help manual, in addition toother relevant parts of the ANSYS manuals. The *USE command help discusses the macro calling parameters and their local scope. Note a slight difference in calling 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 ANSYS command will be the only way to get the function called with picking. It may be desirable to sent 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 a set of shell elementscan all have their "top" surface face the same way. This makes application of pressure, contact elements, and review of results more feasible. This orientation should be done before running SOLVE ; the results are not re-oriented in the database 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 and a meaningless mess will result. These commands work with shell elements that are attached to areas, as well as with independent shell elements. Note: If you clear the elements attached to 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 interiors of 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 this would 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. This command will search outward from a chosen element that the user considers "correct", re-orienting a connected set of shell elements sothat they face the "same way" (this takes some interpretation), even working around corners. It searches elementsfrom the selected set of elements, 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 understand exactly what it does,


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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 elements with 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, according to their face number. A similar command for plotting areas would help, too. It could even be done 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 shell elements when PowerGraphics is ON, and "No Numbering" plus "Colors" or "Colors and 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 ANSYS will 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 of its ability to follow curved surfaces. (Shell 63 elements are flat, and can make a mess of a general curved surface under free meshing.) Shell 93 gave me good convergence for both elastic and plastic Large Displacement (nonlinear geometric) analysis. It does not like to follow too large an angle of curvature with one element, so the number of elements on an area fillet can be large. Set the angle subtended by Shell 93 elements during meshing to a value 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 structure has pressurized flat surfaces, Shell 93 often converges better when stress stiffening is activated for Large Displacement analysis. Stress stiffening for Shell 93is activated at the solution phase of the analysis, whereas Shell 63 is (apparently) only stress-stiffened by setting one of the KEYOPT values. (I have obtained different Large Displacement convergences with Shell 63 with no stress-stiffening set, with the KEYOPT stress-stiffening set, with stress-stiffening set in /SOLU, and with stress-stiffening set in both places.) Like Shell 63, Shell 93 also has the virtue of being supported by the Linear/Plus version of ANSYS 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 by concatenating perimeter lines, I have occasionally had mid-side nodes created, in the interior of the area, such that there was too much element curvature distortion in the plane of the element. One fix is to have the elements created with the sides straight, which is tolerable if the elements are flat, and if it does not cause trouble on the perimeter of the plane area being meshed. "Trouble" here means poor representation of curved boundaries--other elements on these boundaries may need to curve to follow curved surfaces, or it may be desired to have a curved 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 elements sides set straight, then clear the surrounding areas, if the surrounding areas are not intended to be meshed, or need better element shape control. This will leave the plane area of interest meshed with elements that have straight edges in the interior, and curved edges on the perimeter. This is illustrated by thefollowing images of an intentionally extreme example. In the first image, a line plot of element edges shows extreme distortion in the plane. An intended hole is meshed with triangles. All these elements are Shell 93, having mid-side nodes.


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

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

In the fourth image, the elements bordering the hole are shown, after the hole has been cleared of elements. The element curvature at the hole is visible. The interior of the plane area is meshed with straight-sided elements. Thesame problem and a similar fix can be encountered with mid-side noded SOLID95 brick elements that have 20 nodes. The surface areas of a volume can be meshed with 8-node SHELL93 elements with curvature, then the volume meshed with SOLID95 elements with the sides straight, then the shell elements on the areas removed with the ACLEAR command. This will leave the volume meshed with SOLID95 elements that are curved on the surface areas, but with straight sides in the interior. There are rare occasions when this will eliminate element distortion warning messages.


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Tip 21: Moving a Model from ANSYS Mechanical to ANSYS Linear/Plus:

Because versions of ANSYS sell for different prices, a company may own one version to be used for nonlinear models, and several licenses for linear work, or just for model creation and results review. Occasionally, a model will be moved "down" from a fuller version of ANSYS to the Linear/Plus version.

A user can run into difficulty moving a model from ANSYS Mechanical (or ANSYS Structural, etc.) to the less expensive ANSYS Linear/Plus. The Linear/Plus version limits the number of nodes allowed. Unfortunately, it implements this control by not allowing node numbers that exceed a limiting value. This means that compressionof node numbers (and element numbers) may be required in order to get larger models to be accepted by ANSYSLinear/Plus. Otherwise, the program quits without an opportunity to compress the numbering (more recent ANSYS versions may be more tolerant, but the numbering will have to be compressed at some point).

When the node and element numbers are compressed, coordination of loading with the numbering expressed in load step files is lost. The way around this that I have used is to read in the original database, read in a load step file, compress the numbering, and write the load step file. The process, reading in the original database, must be repeated for every load case (a macro could be written to automate this.) Finally, the original database is read in, numbering is compressed, and the new database is written.

Unsupported element types cannot be used in ANSYS Linear/Plus; neither can too large a wavefront (can the PCG solver get around this?). The unsupported elements need to be deleted or changed before moving the model (e.g. change SHELL181 to SHELL63). Then, if the number of entities does not exceed ANSYS Linear/Plus limitations, the database can be moved 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 the database to ANSYS Linear/Plus. This is because the ANSYS Linear/Plus program will complain 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 out the hard way.

On rare occasions, a model from a more recent version of ANSYS may be moved back to an earlier ANSYS version. If IGES is not satisfactory, a user could use CDWRITE to write out the 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 includes modifying commands for element creation, after deducing what format is needed. Writing the element data with EWRITE then cutting and pasting with the CDWRITE file may be easier -- I haven't tried it. A user-written program can expedite cleanup for a large model.


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Tip 22: Deleting Nodes that have Nodal Coupling:

When deleting a set of nodes for which some were members of coupled node sets, delete the coupling equations BEFORE deleting the nodes. Otherwise, unwanted coupling equations may be active if you create more nodes. The coupling equations are not automatically deleted when the nodes are deleted--is this a bug? Select the nodes to be deleted, then delete node coupling equations for which any nodes are selected, then delete the nodes. (You will have had to first delete the elements.) Clearing solid model entities is the same as deleting the elements and nodes simultaneously.

I find it very helpful to turn on the symbols for nodal coupling when checking for proper use of 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) whennonlinear elements such as gap elements, hook elements, and surface contact elements are used. Because of (3) itis clearer to refer to a "linear" analysis as "small displacement elastic", since "linear" may be perceived as meaning that there are no nonlinear elements present. A nonlinear analysis will take longer, usually considerably longer, than a linear analysis. For a large finite element model, it helps to have a computer with an extremely fastCPU, large RAM, large hard drive, and fast hard drive data 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 material nonlinearity (plasticity). Shell 43, Shell 143, and Shell 181 are capable of both Large Displacement 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 both Large Displacement and material nonlinearity. Shell 93 has the advantage that it can follow a curved surface. There are also shell elements for composite materials and for P-element solutions. 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 quad elements are unavoidable during meshing, it may be desirable to use either small triangles, or the Shell 93 element. Note that within the ANSYS manuals, high order elements are not considered to be ideal for nonlinear work. However, I seem to have had some success with the Shell 93 element (can't say if the results were ideally accurate). You can evaluate the model quickly by doing a partial solve (Partial Solu in the GUI), only generating the element matrices, 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 Stress Stiffening before running the solve process. Stress stiffening for elements 63, 43, 143, and 181 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 elementtypes with the GUI). Some beam elements are like this, too. It apparently (I find the manuals difficult to interpret on this) can NOT be set within the Solve module, even though the GUI has a selection box for Stress Stiffening. However, I seem to have had convergence differences with Shell 63 with stress-stiffening set and not set in the solve module. For Shell 93, stress stiffening IS set within the Solve module, by choosing it under Analysis Options in the GUI (SSTIF). The use of stress stiffening for convergence improvement is contraindicated by some conditions such as the substantial use of nodal coupling or nodal constraint equations... see the ANSYS manuals 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) to converge when substeps are being used is to activate the Predictor (PRED) in the Solve module. (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. I usually stick to letting the program decide how to use Newton-Raphson iteration and adaptive descent in the Solve module. Under the Nonlinear settings of the GUI, the user can modify 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 the number of Equilibrium Iterations to 15 when doing shell models, preferring to use smaller substeps instead. However, in a model with gap or contact elements it may be desirable to have 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 to converge. If shell models


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of flat plates subjected to pressure or perpendicular forces are included in the analysis, the shell will at first act as a flat plate in bending. Once the shell has curved, by movement as small as half its thickness, the shell will start to carry the applied load 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 small first 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 of the force application ramp. Where there is no pressure or perpendicular force on flat shells, I would start with a substep such as 1/10 or 1/4 of the applied load, but allow a minimum substep as small as 1/100 of the full load. If these approaches will not work, it is likely that convergence 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 can follow arising and falling force-displacement relationship. I find PlotCtrls/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 atevery substep when doing this (Output Ctrls). When using Arc-Length analysis, it is usually desirable to set a criterion to stop an analysis (NCNV). I usually use maximum displacement as the criterion for shell work.

Remember to ramp up your loads, permit automatic time stepping, and in the NSUBST command, allow the program to bisection by setting the maximum number of substeps greater 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, and constraints, (2) a Large Displacement elastic solution, and possibly (3) a plastic small displacement solution. If these work without significant warning messages, you should be making some progress. If gap or contact elements are being used, consider (4) softening the normal 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, a coarsely meshed model may improve speed enough for you to study what helps get an answer. These preliminary studies may help you to find what settings help you to get convergence ordiscover modeling problems before you do more time-consuming accurate analysis. 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 learn the peculiarities and pitfalls of a particular time-consuming method.

If gap or contact elements are the only nonlinearities in a model, consider substructuring the linear regions of the model. This can result in a tremendous increase in solution speed. If only a sub-region of a model will behave in a nonlinear manner, it may reduce solution time to substructure the region that can be regarded as acting in a linear manner. This speedup effect 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 on a structure. A load step file contains loads on elements and nodes. It does NOT contain loads on geometric entities. Consequently, a load step file can be generated after all loads from geometric entities have been transferred to a model. After all loading on geometric entities has 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 may contain the KUSE instruction telling ANSYS to re-use the TRI file if the constraints have not changed. If the user deletes a load step file, changes the order of their execution, or manually modifies their contents, invalid analysis might result.

If the model is re-numbered after load step files are generated, the node and element numbers in the load step filewill no longer be synchronized with the model, and will be invalid. A way around this is mentioned elsewhere in these notes (See Tip 21).

The reader should take note of the ANSYS user guides comments on the LSCLEAR command. This deletes all


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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 LSCLEAR command when a load step file is read in. If it did, then ANSYS would have to implement substantial 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 when the user employs load step files in a manner for which the method was not designed. It may help to read the contents of the LSSOLVE.MACmacro in predicting what will happen, and to see what LSSOLVE does to avoid trouble. The LSSOLVE.MAC macro at ANSYS 5.3 includes 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 in load step files. It selects all DOF labels, sets xCUM labels to "replace", and does a few other things. 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 clear ALL current loads onthe 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 BFE commands (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 ambient it may be necessary to use the BF and BFE command to set the nodes and elements to the reference temperature (by default 0) rather than just deleting the loads using BFDELE and BFEDELE and using BFUNIF to input the uniform temperature. It may help to use commands such as "nsel,s,bf,temp,-999,99999" and "esel,s,bfe,temp,-999,99999" to select all of 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, for example, four locations on an element with BFE, and in a later load step set the value at only two 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 a series 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 additional load step file, the old constraint may still be present when the series of load step files is read in under LSREAD; check for this in your results. Be careful with this. It may compromise the use of load step files, or require some intervention like writing an input file that calls load step files in using LSREAD, implementing fix-up commands as needed -- be careful that a TRI file is not re-used because a load step file contains "KUSE,1" when your changes to constraintsmean that a new TRI file should be generated. Statements in the LSSOLVE.MAC macro can provide guidance on using LSREAD effectively. You may need to look inside the load step files with a text editor. Be warned that changing the contents of load step files with a 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 and displacements from one load step do not appear inappropriately in later load steps. This is true when generating the load step files in the first place, and may apply when reading in load 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 consistent with getting smooth ramping of loads and displacements, for those cases when this is desired, either for transient analysis, or for goodnonlinear 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 geometric entities and elements andnodes 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 the solid 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 command generated in subsequent load step files. Because ANSYS does not permit the kind of analysis to be changed when applying a series of load steps, error messages will be result if the user changes the value of NLGEOM in the middle of a set 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


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shell elements (63, 43, 143, 181, and 93) are reported at the top and bottom surfaces of the shell element. The user can has four options in ANSYS 5.3 for plotting shell stresses (and strains). Three of them are selected with the commands: "SHELL,TOP" , "SHELL,MID" or "SHELL,BOT". These will cause 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 stress is 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 on the element's orientation when it was defined (see elsewhere in these pages). It is possible to have adjacent elements, one with a "top" surface pointing upward, andits neighbour with the "top" surface pointing downward. In complex structures it happens all the time. If nodal averaged plots are done, for example with "PLNSOL,S,EQV", when either top surface or bottom surface plottingis chosen, then with such adjacent elements, the plotted top surface and bottom surface results will get blended, causing a misleading mess to be displayed. (See Tip 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 no hidden 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 astructure.

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 render local stress plot information meaningless atthe intersection. This is true of both surface and mid-plane stress plots. This is one way in which excessive stresses will be unintentionally missed.

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 of their much cleaner appearance (I do them myself). The fourth option in plotting shell stresses is 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 /EFACET commands 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 there are certain discontinuities in the material or geometry in the model. I'm going into this detail, because a high stress that is washed out by nodal averaging could be a stress that causes serious 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 user has few options here. Sometimes it is important to select only regions of a model when doing nodal averaged mid-plane stress plots (using "SHELL,MID", without Powergraphics) so that the 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 look messy, but at least it doesn't hide an extreme stress. An alternative I used is discussed elsewhere inthese pages: I wrote a macro to get the mid-plane averaged stress (all components) at every node of every element (a given node has different results with reference to each of the elements to which it is attached, so a given node will be looked up as many 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, with discontinuities. CAUTION: This ruins the results database. The macro is extremely slow to run. The method (under Powergraphics) does, however, give far better looking plots than using the "PLESOL,S,EQV" command to plot mid-plane element stresses without nodal averaging (without Powergraphics).


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LOAD PATHS: The macro I mention above could be modified to multiply the mid-plane averaged stress components by the local shell element thickness at each node. The resulting values 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 a complex shell structure. "PLVECT,S" plots that would now show arrows corresponding to the 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 any other use. Before plotting "load-per-unit-length" data, the user needs to decide how to orient the results data coordinate systems with RSYSfor information such as Sx or Sy that contains direction information (stress and strain with EQV does not containdirection information).

Note: The Output Data section on Shell63, Shell43, and Shell93 includes In-plane element X, Y, and XY forcescalled TX, TY, and TXY. Consequently, shell "force per unit length" data can be obtained directly in an Element Table very quickly, though with a resolution of one value per element. (For Shell63, 43, and 93, use SMISC setting 1, 2, or 3 when generating the element table data.) The Theory Manual uses the term In-plane forces per unit length while the elements manual refers to just forces as above -- a simple test I ran shows the data to be force per unit length. The elements manual ought to clarify this. The Element Table data 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 the double-headed arrows for an ETABLE as in a principal stress vector plot.) The Elements manual 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 not explicitly 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 that if the command "SHELL,MID" is followed by "ETABLE,SEQVMID,S,EQV" that the column 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 the update command "ETABLE,REFL" is executed. Warning: When plotting ETABLE shell element 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 information DOES NOT reflect the SHELL surface setting conditions in effect when the ETABLE data was stored, and could be misleading. For this reason, the label used for the column should indicate 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-fill columns with results data. A change of the SHELL layer setting can change stress results that are loaded in an update. Consequently, loading shell element data must be handled very carefully in order that the layer choice is controlled. Element table data from the CALC module (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 as taken from the element in Element Coordinates. Unlike SX or SY, the values of TX, TY, and TXY appear to be insensitive to the RSYS setting. The Element Coordinate System will vary orientation from element to element, particularly under free meshing, and affects the usefulness of TX, TY, and TXY data. The element table data can be processed by the user to yield 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 shell mid-plane. The plots


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will usually be more informative without nodal averaging. Section 2.3.4 of the ANSYS Theory manual discusses Forces and Moments per unit length on shell elements -- the suggestion is that internally, at least for some shell elements, the mid-plane stress 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 coordinate systems before meshing areas (or otherwise generating shell elements) and use ESYS to get all 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. It may be desirable to have a local coordinate system aligned with each flat area to be meshed with shell elements so that all shell element coordinate systems can be aligned in the plane of the 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 the images below. The firstshows 3 elements that were created during free meshing. The elements are plotted using vector graphics, with theelement coordinate systems shown. Each element has its coordinate system oriented differently. The image below it lists the elements and their node numbers. Look at the sequence of node numbers for the three elements to see why the element coordinate systems point in such different directions.

The next two images show a plot of TX done from an element table. The element table was filled by the TX values for the elements (this is the load-per-unit-length in the element coordinate system X direction). The valuesdiffer so much from element to element because of 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 shown in the global coordinate system.


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The final images in this section show a group of Shell63 elements that have had their element coordinate systemsaligned with local coordinate systems at the time of the creation of 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 be aligned in the plane of the elements creation if RSYS,SOLU is active when plotting stress results. Knowlege of the alignment of the loads and stresses can make plots more useful in understanding load paths, reduce the total number of plots required in model assessment, and help facilitate an evaluation of loading on welds. The first plot with vector graphics shows the elements with their element coordinate systems. Note that they are aligned. There are two local coordinate systems at work in this example -- they are numbered 11 and 12 and their symbols are plotted. Elements have been created aligned with number 11 in one plane, and aligned 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 shown in all elements. If RSYS,0 were active when the Sx plot was done, the plane of elements that is perpendicular to the global X axis would show zero stress in the X direction in this example.


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There is an alternative to using ESYS and RSYS,SOLU to align element coordinate systems for the purposes of


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stress plots like Sx, Sy, and Sxy. During postprocessing in /POST1, a local coordinate system can be aligned withthe plane of shell elements of interest, and RSYS set to that local coordinate system, before plotting Sx, Sy, or Sxy. However, this would do nothing for TX, TY, and TXY which depend on the element coordinate system andare generated in an Element Table.

I leave the topic of whether to plot surface or mid-plane shell stresses to the reader to determine. 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 and fasteners, 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 equations should have been employed. (The nodes concerned were not at the same location in space.) Rigid region constraint locks together aselected set of nodes so that they translate AND rotate in space as if they were locked together by an infinitely stiff structure. Nodal coupling locks together selected degrees of freedom (translation and/or rotation) individually, so that the same degree of freedom value will result for the nodes in the coupled set. Nodal coupling will not combine the rotations and translations that are necessary to imply rotation as a rigid body in space.

Note that rigid region constraint may not be appropriate for Large Displacement, when the displacement rotationsare significant (sin(theta) differing from theta, etc.). This is because ANSYS uses a linear approximation to the rigid body rotation matrix. A rigid region grouping 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 are attached.) 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 could result. If the beams are of very widely varying lengths, then some may be too stiff, others too flexible -- remember that flexibility is proportional to length cubed.

I ran a model in which about one thousand beam elements were used to position gap elements. These beam elements would ideally have been infinitely stiff. I needed elements, instead of nodal coupling or constraint equations, because of thermal expansion considerations. The beam elements were widely varying in length. This created solver trouble, until I wrote a macro that assigned each beam element a unique REAL value, which set values for each BEAM4's Ixx, Iyy, Izz, and Area as a function of the element's length. I found it sufficient to set their stiffness a couple of orders of magnitude stiffer that contact stiffness 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 of the structure under ANSYS. There is a "PRESTRESS" flag to set under modal analysis. This is available in the dialog box for ModalAnalysis Options. First, do a static analysis with the prestress flag set. Exit Solution (click Finish or enter "/fini"). Re-enter Solution, and do a modal analysis with the prestress flag set again. This does not seem to work when the stress run 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 few things to do. First, select the elements to be copied and get their nodes with NSLE. Copy or reflect the nodes, noting the nodal number offset that will be used -- write it down. Copy or reflect the elements, using the nodal offset number that you wrote down. ANSYS should default to a nodal offset number equal to your highest numbered node. If you make it smaller, you run the risk of changing the location of nodes that already exist, resulting in a lovely mess. If you are running something like ANSYS/ED you may want to compress your node 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-fatal warning message and quit some time later if you don't clean up). You could compress the node numbers, and then make the offset number equal one plus the difference between the


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maximum 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. (Remember that compressing node or element numberswill destroy synchronization with Load Step files.)

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" for which you will need the same nodal offsetnumber. Do a replot to see the newly created nodal coupling. Caution: Be sure that if nodes were deleted earlier, that nodal coupling equations that in the past included those deleted nodes were also deleted. If you forget, you may get a pretty mess.

Remember that if there are nodes on the plane of reflection, new nodes will overlay them. Merge commands maybe 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 the node order reversed, because of the mirror imaging. They Will Not Merge with the element from which they were reflected. They may have to be deleted, depending on what you are trying 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, in order to reflect the elements that will join them to the remainder of the reflected structure, so the 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 section replaced, or requires the addition of more structure. The way to attach new geometry onto existing nodes and elements is to: (1) Place keypoints on the nodes onto which new geometry is to be built (i.e. grafted). (2) Join these keypoints with lines. (3) Set mesh density along 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) Select the nodes (new and old) along the interface between the old nodes and the nodes of the new geometry. (7) Merge ONLY these nodes along the interface using the NUMMRG,NODE command. Alternatively (much more work unless a macro is written or the CPINTF command is used correctly), fully couple the PAIRS of nodes with the CP command. In the event of elements with mid-side nodes, lines will have to be created curved so that a single line 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-side nodes.

This attaches the new geometry and mesh to the old elements and nodes. Be sure to double check that the merging has been done correctly and according to your intentions -- I have found 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 the surrounding structure) in order imply a rigid region grouping of nodes, which works under Large Displacement (a CERIG group does not work with large displacement). This is an old FEA trick -- it is not perfect. A separate material should be created for these beams, and be given zero mass (set the material density to zero) so that no gravitational or other inertial load acts on the material. A thermal expansion coefficient should be input if appropriate -- it would 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 of nodes to be connected is selected. The lowest numbered of the set of nodes is attached to each of the other nodes in the set by a beam element. The beam element to use has to be set up in advance, and the appropriate MAT, REAL, and TYPE set by the user. A macro like this is very fast to run. Caution: Such a macro would become complex if it checked for duplicate 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 anythingthat plots a symbol, all symbols become active, and plots become extremely slow. Turn off symbols with /PBC,ALL,,0 to speed things up. I put this command in the Toolbox for convenience. I have found that plotting can become slow with very large models when loads have been applied, and even when applied and deleted. Presumably ANSYS is checking to see if any symbols should be shown. The plotting speeded up considerably when symbols were turned off with "/PBC,ALL,,0" even though there were, in fact, no symbols to be plotted.


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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 applied load. If a modal vibration is excited, small amplitude vibrations may result in very little response from the support, and vibration similar to a structure that is free in space may result (this is obviously very problem dependent). If so, it may be desirable to run a modal vibration analysis with no constraints. More than six modes must be requested, since the first 6 represent the free translation and rotation, and give Zero eigenvalues. A better approach would be to characterize the flexibility of the constraint points. With some structures, you may get a few surprises, as torsional 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 choice for 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 your company, claiming that the engineers are incompetent (stupid, uninformed, can't spell, and so on) and making a huge mistake. If the boss is not an engineerand cannot understand the issues, 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 athorough manner. Write down some criteria and your assessments. Also, be careful that you cannot be accused ofleaking information unfairly from one vendor to another -- date your correspondence carefully, and work throughyour purchasing department if 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 th

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49054430 Ansys Tips and Ansys Tricks