Solidworks Finite Element Analysis

What Solidworks FEA Actually Is, For People Who Just Want Answers

Solidworks Finite Element Analysis is built into the standard product line, not a separate program you need to buy or install. It uses the same geometry kernel as your part files, which means you don't have to export anything or translate step files before running a simulation. The mesh runs on the solid bodies you already have. That convenience comes with real tradeoffs, and understanding those tradeoffs separates the people who get useful results from the ones who get numbers they can't trust. The solver is direct sparse. It assembles the global stiffness matrix and solves it in one pass. For linear static studies that is usually fast. A bracket with a few hundred thousand tetrahedral elements typically finishes in under a minute on a modern machine. Jump to contact nonlinearities or large displacement options and the same model can take twenty to forty minutes, and you might not even know why until you're looking at the convergence plot. I have learned to check the substep count early in the process instead of letting the solver chew through an hour only to fail at the last increment.

Solidworks Finite Element Analysis: Getting a Result You Can Actually Use

Start by cleaning the geometry before you ever touch the mesh controls. Remove fillets that are smaller than three element edges across. Delete cosmetic threads, knurling, and embossed logos. I once spent four hours debugging a model that kept generating distorted elements, and the problem was a 0.4mm decorative groove on a mounting face that added nothing to the physics but destroyed the local mesh quality. I suppressed those features and cut the preprocessing time to about twelve minutes on the same setup. Apply realistic loads. The default "Fixed Geometry" face constraint locks every degree of freedom on that surface. That looks clean in the graphics window but it creates a singular stress concentration that the solver reports as near-infinite. If you are simulating a bolted joint, use bolt pretension instead of treating the entire face as immovable. If the part sits on a flat surface, use contact with the ground or a deformation spring foundation rather than a rigid constraint. The numbers change significantly, and they usually become more useful. Meshing is where most mistakes happen. Solidworks gives you several options. The default curvature-based tetrahedral mesh is fine for quick checks on simple geometries. For anything that requires accuracy near a stress raiser, switch to second-order elements. quadratic tetrahedra capture bending and stress gradients far better than linear ones. The element count goes up, but not as much as you would think, because the solver needs fewer elements to reach the same level of accuracy. A model that gives you 80 MPa with linear elements might read 110 MPa with quadratic elements, and the second number is closer to reality. Second-order elements are the default in newer releases anyway, so if you are on an older version, check your element order explicitly.

There is a specific trick that people miss regularly. After meshing, go to the mesh statistics and look at the aspect ratio distribution. If more than five percent of your elements have an aspect ratio above ten, the results in that region are unreliable. I once ran a simulation on a heat sink with thin fins and the default mesh produced elements with ratios above fifteen in the fin roots. The von Mises stress output showed a neat gradient, but when I tightened the fin mesh size and refined the near-field controls, the peak stress moved twenty millimeters away from the fin base. The first result was wrong because the element shape was distorting the strain calculation. Material definitions matter more than most users realize. Solidworks ships with a library that includes generic steel, aluminum, and titanium. The generic steel option uses an elastic modulus of 200 GPa and a Poisson's ratio of 0.3. That is close enough for preliminary work, but if you are designing with 304 stainless versus 4140 chrome-moly, the yield strengths and toughness profiles differ substantially. Always verify the material card against the manufacturer's specification sheet. A wrong elastic modulus shifts every deflection result proportionally, and a wrong yield strength turns a passing safety factor into a failure warning or vice versa.

When the Default Setup Breaks Down

Linear static analysis assumes small deformations, linear materials, and constant boundary conditions. That covers a lot of common engineering problems. It does not cover buckling, plasticity, thermal cycling, or fatigue. Solidworks Simulation includes nonlinear options and a dedicated fatigue module, but both require careful setup. Nonlinear contact definitions alone can double or triple solve time. I worked on a press-fit assembly where the interference was 0.15mm on a 40mm diameter shaft. The initial linear run predicted a contact pressure that was in the right ballpark, but the nonlinear run with friction at 0.15 revealed that the actual pressure distribution was uneven around the circumference because of slight geometric runout. The linear solution had masked that entirely. Buckling analysis in Solidworks uses the pre-stressed state from a preceding static study. You run the static step first, then add a buckling study that references it. The output is a set of eigenvalues, each representing a critical load multiplier. An eigenvalue of 2.3 means the structure buckles when the applied load increases by 2.3 times. This is a linear buckling result. It does not account for geometric imperfections or material yielding before buckling. In practice, real parts buckle at lower loads than the eigenvalue prediction. If your eigenvalue is below three, I would treat the result as a warning flag and consider a more detailed analysis in a specialized code, or use a generous safety factor. The Solidworks buckling module is adequate for preliminary sizing, but it is not a substitute for a proper nonlinear buckling study when accuracy matters. Thermal-stress coupling works by running a thermal study first and then importing the temperature distribution into a structural study. The link between them is straightforward, but the thermal mesh does not automatically transfer to the structural mesh. You need to make sure the element sizes match closely enough, or the interpolation introduces errors at the interface. I have seen cases where a coarse thermal mesh smoothed out a hot spot that the structural mesh should have captured, resulting in a stress prediction that was ten to fifteen percent low in the critical region. Keep the mesh densities comparable and check the temperature contour plot before running the structural step.

Pitfalls That Waste Time

The most common mistake I see is applying a force to a face and interpreting the average stress as the maximum stress. A 1000N load on a 50mm by 50mm face gives an average pressure of 0.4 MPa. The stress results will show something higher near constraint boundaries or geometric discontinuities. Average stress over an area is not useful for failure prediction. Look at the peak stress value, and if it is on a line or point rather than in a volume, understand that it is a numerical artifact. Stress recovery in finite element analysis is computed at integration points and extrapolated to nodes. Near a singularity, the extrapolated value keeps climbing as you refine the mesh. There is no finite value that is the "true" stress at a sharp re-entrant corner. You need to either fillet the geometry or use a nominal stress approach based on a defined area away from the singularity. Another issue is the default safety factor display. Solidworks calculates it as yield strength divided by von Mises stress. That is correct for ductile materials under static loading. It is meaningless for brittle materials, where you should be looking at principal stress criteria. It is also misleading if the part operates at elevated temperature, because the yield strength drops and the software does not automatically adjust for that unless you define a temperature-dependent material curve. I had a bracket made of 6061-T6 aluminum that passed the default safety factor check at room temperature, but the operating temperature was around 120°C, where the alloy loses roughly thirty percent of its yield strength. The part failed in service because the safety factor shown in the report was based on the room-temperature value. Always define temperature-dependent properties when the thermal environment is outside standard conditions. Convergence diagnostics are another area where users skip too quickly. When a nonlinear solve fails, the software tells you which increment did not converge and gives you the residual forces. The usual fixes are reducing the minimum substeps, increasing the maximum, tightening the convergence tolerance, or improving the contact definition. I once had a model with self-contact on a thin-walled snap fit. The initial runs kept diverging at increment 12 out of 100. The contact detection was bouncing between open and closed states because the initial gap was smaller than the element size. I switched the contact type from "bonded" to "no penetration" with a refined mesh in the contact zone, set the initial gap tolerance to match the element size, and the solve completed in three increments without oscillation. The total time went from repeated failed attempts over twenty minutes to a single successful run in eight minutes.

Results Interpretation and Reporting

Displacement results are the easiest to validate visually. If a cantilever beam deflects downward under a downward load, the direction is correct. The magnitude should scale linearly with load in the elastic range. If doubling the force roughly doubles the deflection, the solution is behaving as expected. If the relationship is nonlinear without any nonlinear options enabled, there is likely a boundary condition or constraint error. I once found a model where the deflection was off by a factor of five. The issue was that I had applied a fixed geometry constraint to a face that shared edges with the loaded region. The constraint was effectively clamping part of the beam that should have been free to deform. Removing it brought the deflection into agreement with the hand calculation within three percent. Stress results require more care. Von Mises stress is a scalar quantity derived from the deviatoric component of the stress tensor. It is used for yielding criteria in ductile materials. It does not tell you about shear failure, tensile cracking, or delamination in composites. If you are working with composite laminates, Solidworks Simulation has a composite material module, but the available layup options are limited compared to dedicated composites codes. For most general-purpose metal parts, von Mises is sufficient. Always compare it to the correct yield strength, not the ultimate tensile strength. Using UTS as the limit gives a falsely optimistic safety factor. The post-processing tools allow you to create cut planes, isolate regions, and calculate reaction forces. Reaction force reports are useful for checking equilibrium. The sum of reaction forces should balance the applied loads within a small percentage. If the imbalance is above one or two percent, there may be an error in the model setup or a missing constraint. I use this check routinely before trusting any stress output. A model that does not satisfy global equilibrium cannot produce trustworthy local results.

What It Cannot Do Well

Solidworks Simulation is not designed for explicit dynamics, crash simulation, or highly nonlinear material behavior. If you need to simulate drop impacts, ballistic events, or metal forming processes, you should use a dedicated code like Radioss, LS-DYNA, or Abaqus/Explicit. The import-export workflow between Solidworks and those tools is possible but adds time and potential for error. For the vast majority of static structural problems encountered in mechanical design, the built-in tool is adequate. It is faster than exporting to an external solver and it keeps the simulation tightly coupled with the CAD model, so design changes propagate automatically. Mesh quality control is another area where the built-in tool has limitations. You do not have the same level of manual control over element sizing gradients, boundary layer meshes, or structured hexahedral generation that you would find in a professional preprocessor. For complex geometries with multiple tight clearances, the automatic mesher can produce poor quality elements in hard-to-reach regions. In those cases, simplifying the geometry strategically or using local mesh controls with smaller element sizes on critical faces is the practical workaround. There is no substitute for knowing where the high gradients will be and directing the mesh there. The frequency and modal analysis module produces natural frequencies and mode shapes, which is useful for avoiding resonance. But the results are sensitive to how constraints are applied. A fully fixed base gives higher frequencies than a realistic support condition. I have seen designs that passed the modal analysis with all faces fixed and then vibrated excessively in the field because the actual mounting allowed some flexibility. Define the constraints to match the real assembly, or use spring elements to approximate support compliance. The difference in predicted natural frequency can be twenty to thirty percent, which is enough to miss a resonance condition entirely.

If you need to get started with a basic study, open your part or assembly in Solidworks, navigate to the Simulation tab, and create a new static study. Select the materials, apply fixtures that reflect the real boundary conditions, add loads, and generate the mesh. Review the mesh statistics before solving. Run the analysis and check the displacement and stress results against a hand calculation or published data point if you have one available. The whole process for a typical bracket takes about fifteen to twenty minutes from start to validated result on a decent workstation, assuming the geometry is clean and the loads are straightforward.