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Measuring Face Translation and Rotation in ANSYS® v14.5 Mechanical (Workbench)
ANSYS Workbench Mechanical can perform harmonic response analysis, by either modal superposition, or by full harmonic analysis. Measurement of face rotation is sometimes of interest to customers. If a remote point is associated with a face of interest, rotation of that remote point can be measured in an APDL Commands Object, and reported to the user, with the option of creating an output parameter from the rotation value.
In Figure 1 above, the far end of a plate has been fixed. The nearer end has harmonic loads applied at two edges, one pointing up, and one pointing down. The objective is to excite the plate at frequencies around a mode of vibration that has the twisting motion implied by these forces.
Associating a remote point with a face meshed with a large number of nodes can increase solution time and memory requirement—setting a pinball radius can reduce this.
Recent features in ANSYS that make it easier to find amplitude results in harmonic analysis are illustrated here.
Settings that measure face rotation in harmonic analysis will be reviewed. In the Workbench project page, a modal analysis can be linked to a subsequent harmonic response analysis:.
The modal analysis can be helpful before a harmonic run in order to discover at what frequencies a structure may respond substantially to an excitation. If a modal analysis has included enough frequencies to more than cover the range of excitation frequencies in a harmonic response analysis, then modal superposition can be used in the harmonic response analysis, as an alternative to full harmonic response analysis.
If Figure 3 below, it can be seen that the second mode involves rotation of the free end of the cantilever plate around the Z axis, at a frequency of Note: the predicted frequency can be sensitive to the mesh density.
In Figure 4 below a Remote Point has been created and automatically positioned at the centroid of the end face on the plate. In Figure 6 below, a Harmonic Response has Analysis Settings input to indicate a range of frequencies of interest, and the number of intervals between frequencies to be analyzed:. With loading as described in Figure 1, the amplitude of Y displacement calculated at Hz can be displayed as in Figure 7 below, at a frequency close to resonance:.
Insight as to the peak response around a resonant frequency can be obtained in a Frequency Response plot set to show Amplitude, as below in Figure 8 for a vertex at a corner of the free face of the plate:.
Further insight results from a phase plot for the vertex, as in Figure 9. Note the phase angle near resonance:. The frequency and phase angle as seen above in Figure 8 and Figure 9 can be used to obtain a directional deformation plot in which the values are not absolute values. Since the phase angle is entered for the frequency, we see in Figure 10 the magnitude of the displacement without the absolute value being assigned:. An APDL Commands Object can be used to load a result of interest, and read the amplitude of vibration at a remote point.
This is illustrated in Figure 11, using the node number recorded in Figure In this example, the user provides the substep number that corresponds to the frequency of interest. Users must remember to request Amplitude in the SET command:. Note that the SET command above could directly request a frequency with the Time argument:.
This angle is close to that of the remote point in Figure 11, which considers all of the nodes on the face. Workbench Mechanical can be used to perform a Harmonic Response dynamic analysis. A remote point can be created, and used in APDL commands to measure a rotation angle amplitude at a frequency of interest. A modal analysis is often performed prior to another dynamic analysis in order to characterize the system of interest.
Modal analysis (boundary condition problem)
Because a harmonic response analysis is linear, modal superposition can be used if desired, as long as enough modes are extracted to more than cover the range of harmonic excitation frequencies.The remote displacement condition is used to guide the displacement of a face or edge of a structure from a remote point.
There are several advantages compared to a classical displacement boundary condition. Also with the remote displacement a rotation can be applied to an edge or a face of a structure. There are also some limitations, mainly the fact that this is a linear boundary condition and valid only if small displacements and rotations occur in the area of the applied entity and the remote point itself.
The user defines for each of the translational directions if the displacement should either be unconstrained or predescribed.
Each predescribed value can either be defined by a scaler value, a function or a table. For function or table data the value may depend on time or frequency in case of a harmonic analysis and the spatial coordinates. For each rotational direction the rotation can either be predescribed or unconstrained. In case of a predescribed rotation the value is given in radians. The input methods and possible dependencies are the same as for the translational directions. Here the user defines the coordinates of the external point on which the displacement is applied.
The coordinates are given in the global coordinate system of the mesh. This property defines if the applied associated entities edges or faces may deform or if they are assumed to be rigid.
If the setting deformable is selected, no additional stiffness is generated on the applied entities. The remote point is connected to the entities by a RBE3-constraint. If the setting undeformable is selected the entity behaves like a rigid part. The connection of the point and the entities is a multi-point constraint which blocks all relative displacements between the affected nodes. Von Mises stress top and total displacement bottom contour plot on the deformed shape with a fixed value displcement boundary condition of 0.
Von Mises stress top and total displacement bottom contour plot on the deformed shape with a remote displacment of 0. Next Previous.I am using Workbench to get the natural frequencies for the rail as is it shown in the below screenshot. In this case, I would suggest frictionless or frictional contact instead of bonded contact.
And you probably need to change the fixed boundary condition as well. Thank you But what should I put the boundary condition?
Other than fixed from bottom of the sleepers? Modal analysis is a linear analysis. Therefore a frictional contact that is closed is automatically converted into a bonded contact to do the modal analysis. That is why you get the unrealistic simulation result you show in this image. It would be easier to match lab data to simulation data if instead of a wide flat wooden sleeper, you put a narrow rod under the rail.
Then in the simulation, you could split the face where the rod touches the rail, and make that split line a remote displacement and constrain just 3 DOF and leave the rest free. Then you don't need the sleepers in the model at all. Oops, Peter is right. Frictionless and frictional contacts are nonlinear contacts, and in modal analysis, we can only have linear contacts.
Thank you peter for explaining. But how can is it possible to replace the sleepers with a rod? I have inserted the sleepers material property in the Engineering data and if I change the geometry of the sleepers the mechanical properties will change.
Is it possible please if you can explain it with an example? I would really appreciate it. Here is a rail with the bottom surface split at the two locations where, in the lab, the rail will be supported on cylindrical steel rods instead of a wide wooden sleepers. The rail is along the Z axis.
I can not replace the wooden sleepers with cylinder rods because I want to investigate the rail behaviour when it is supported by the wooden sleepers. I suggest you study the agreement between a rail supported by rods and a remote displacement supported modal analysis. Putting a rail on a wooden sleeper without any other constraint may not be that different to the rods. If there is a significant difference, then you can no longer use modal analysis.
You need to run a full transient analysis that can have frictional contact between the sleeper and the rail. How do you excite the rail in the lab?
Do you hit it with a hammer? You will have to have an impact in the transient simulation. Extracting the data from the simulation will be similar to extracting data off the rail in the lab. How did you determine the frequency and mode shape in the lab?
But in reality, the rail does not rest on the sleeper, there are spikes that hold it down. When are you going to add the spikes?Learn more about SimuTrain or get started today by purchasing your subscription. They do not measure rotation. This article illustrates how to use APDL commands to refer to an existing Remote Point, and measure translation and rotation at the point. If the remote point interacts with a face that is set to Deformable, an averaged value of face movement is measured.
If there is no load on the remote point, movement is measured without affecting the result. The Deformation Probe supports a non-default choice of Remote Points as location for the measurement of deformation.
If a Remote Point associated surface behavior is set to Deformable, an average deformation of the surface can be measured for the Remote Point without affecting the model:. Note Deformable Behavior in Definition. Users must be aware of the Units employed in solution in reading the translation Results above, and that these rotations have been converted to Degrees.
Input Arguments could pass values to a SET command in the script. A result that is measured with a Surface created within Construction Geometry can be used to report a result where the Surface cuts across one or more bodies. The surface has to touch elements of the bodies. A Directional Deformation in Z has been measured on the Surface. This technique does not report the rotation, however. Note that the Average reported above matches the Z movement measured by other techniques in this article.
The above use of a Remote Point does not report an average Temperature in a thermal analysis. A Surface as above will report an average temperature where the Surface cuts across the elements of one or more bodies. Another APDL technique can be used to generate average temperatures for selected faces, and has been described in a previous article.
To measure an averaged rotation, a General Body-Ground joint can be attached to the face of interest. The Behavior of the selected face is set to deformable. The following figure illustrates settings and the result:. A finite value here can restrict how much of the surface is included in forming the average.
With either the Remote Point approach, or a General Joint approach, choosing a face with a large number of nodes will increase wavefront during Solve, and will result in a longer solution time. A finite pinball diameter can restrict how many nodes are involved. Workbench Mechanical supports a Deformation Probe that can scope to a Remote Point to measure movements at a face on a model. If the associated face is set to Deformable at the Remote Point definition, then an averaged value can be measured on the face without affecting the model.
Deformation Probes do not currently support measurement of rotation. APDL Commands Objects can record the node number of the Remote Point of interest, and in postprocessing can measure translations and rotations of the remote point. When a General Joint is employed for the measurement, not APDL commands are needed, and the units for the Probe results are automatically set.
Get SimuTrain Now!Between co-teaching an engineering class at nearby Arizona State University and also having a couple of customer issues regarding the concept, large deflection in structural analyses has been on my mind. So, what are large deflection effects? In simple terms the inclusion of large deflection means that ANSYS accounts for changes in stiffness due to changes in shape of the parts you are simulating.
The classic case to consider is the loaded fishing rod. In its undeflected state, the fishing rod is very flexible at the tip. With a heavy fish on the end of the line, the rod deflects downward and it is then easy to observe that the stiffness of the rod has increased. In other words, when the rod is lightly loaded, a small amount of force will cause a certain downward deflection at the top. When the rod is heavily loaded however, a much larger amount of force will be needed to cause the tip to deflect downward by the same amount.
This change in the force amount required to achieve the same change in displacement implies that we do not have a linear relationship between force and displacement. Where F is the force applied, K is the stiffness of the structure, and x is the deflection. In a linear system, doubling the force results in double the displacement. In our fishing rod case, though, we have a nonlinear system. We might need to triple the force to double the displacement, depending on how much the rod is loaded relative to its size and other properties, and then to double the displacement again we might need to apply four times that force, just using numbers out of my head as examples.
In order to capture the nonlinear effect we need a way for the stiffness to change as the shape of the rod changes. In our finite element solution in ANSYS, it means that we want to recalculate the stiffness as the structure deflects. This recalculation of the stiffness as the structure deflects is activated by turning on large deflection effects. Without large deflection turned on, we are constrained to using the linear equation, and no matter how much the structure deflects we are still using the original stiffness.
So, why not just have large deflection on by default and use it all the time? In other words, turning on large deflection will trigger a nonlinear solution, meaning multiple passes through the solver using the Newton Raphson method instead of the single pass needed for a linear problem.
Here is an example of a simplified fishing rod. The image shows the undeflected rod topwhich is held fixed on the left side and has a downward force load applied on the right end. The bottom image shows the final deflected shape, with large deflection effects included. The deflection at the tip in this case is 34 inches.
In comparison running the same load with large deflection turned off resulted in a tip deflection of 40 inches. Below we have a force horizontal axis vs. The fact that the curve is not a straight line confirms that this is a nonlinear problem, with the stiffness slope of the curve not constant.I know that I can apply displacement equal to zero freeze the x values on the symmetry plane.
For shell and beam elements, you must also set the nodes for Rotation-Y and Rotation-Z to 0, while leaving Rotation-X free to create the symmetry condition. If you only have solid bodies and no sheet or line bodies, then the Nodal Rotation will be greyed out.
Thanks a lot Peter.
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Harmonic Analysis Face Rotation Measurement in ANSYS® Workbench Mechanical
An example helps explain Inertia Relief. Consider a structure that has mass, and a vertical load that exceeds its weight. Without constraint in the vertical direction, the global stiffness matrix is singular, and no solution exists. If there is enough constraint to prevent vertical motion, but the vertical load does not exactly match the weight, there will be strong reaction forces where the vertical constraint is applied.
In a transient analysis, in the absence of vertical constraint, the structure would be accelerated by the force difference between weight and applied vertical force, and the structure would both move and vibrate during the transient analysis. Inertia Relief gets the FEA model to exactly balance the force difference applied force minus weight in a static analysis with acceleration body forces over the whole structure, so that the reaction on the vertical constraint is zero.
Depending on the position and direction of the applied force versus center of gravity, accelerations may exist in X, Y and Z, as well as rotational accelerations about X, Y and Z. Inertia Relief during analysis of such a structure requires that mass be properly represented, just enough constraint be applied to prevent free body translation and rotation, loads be applied, and Inertia Relief be requested.
Other conditions must be met. In the Workbench Mechanical interface, if a static analysis is requested, the Analysis Settings branch offers Inertia Relief in its Details as in Figure 2. Using Inertia Relief assumes qualifying conditions in the model are met:. Displacement constraints on the structure should only be those necessary to prevent rigid-body motions 6 for a 3D structure.
The sum of the reaction forces at the constraint points will be zero. Accelerations are calculated from the element mass matrices and the applied forces. Data needed to calculate the mass such as density must be input. Both translational and rotational accelerations may be calculated. Settings for an Inertia Relief model will be described in the following example.
As in Figure 1, a flat surface body has been created and meshed with shell elements. The body contains a circular imprint. A pressure load has been applied to the imprint, directed so that it can lift the body.
To Use Large Deflection or Not, That Is the Question
Gravity pulls down. In Figure 3 and Figure 4 a Remote Displacement has been applied to the circular imprint edge, set to zero displacement and rotation in X, Y and Z, and set to Deformable, so that the structure can deform locally. As mentioned further below, constraint at vertices may be preferred. As a check during postprocessing, both Force Reaction and Moment Reaction have been measured as in Figure 5 to see that they are virtually zero, as should happen with Inertia Relief.
Deflection of the structure takes on the form expected, as seen in Figure 6. The structure is pulled up at the center, pulled down by gravity, and pulled down by the Inertia Relief acceleration that is applied in order to make the net applied force match the inertial accelerations that are applied to provide Inertial Relief.
At the end of the Solve of the load step, the Output text listing provides information on the Inertia Relief translational accelerations and rotational accelerations that balance the model so that there are no reaction forces:.
This example has virtually zero acceleration in all directions except the Y axis, because the applied pressure is centered on the model. The accelerations are those that balance the applied loading, such that there should be not reactions at the constraints that prevent rigid body motion.
They are the accelerations that would be seen in a transient analysis if there was no vibration in the response. Force reaction is virtually zero, as seen in Figure The reactions are zero because the Inertial Relief Accelerations are balancing the applied loading. Alternative constraints that prevent rigid body motion could be used, as in Figure 9, in which UY is prevented at three vertices, UX at two vertices, and UZ at one vertex:.
Reactions are again virtually zero, and the displacement result is similar but not identical. In Figure 10 the range of vertical displacements is between 0. The difference in net displacements may be due to the remote deflection affecting shell rotation.
Constraining vertices may be preferred in general. Workbench Mechanical supports Inertia Relief in a static structural analysis, when certain conditions are met. Reaction forces of zero should result.