Sample 10 - Y Fitting
File: PVE-1343b
Last Updated: Dec. 9, 2008
BV
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This sample report illustrates how Finite Element Analysis is used to validate common pressure components and applications. This report format may be used to justify ASME code compliance, provide stress and displacement analysis, provide cycle life estimates, complete thermal analysis, and perform design validation and optimization studies. This format is fully CRN compliant and may be applied to many applications. This level of analysis can typically be completed within a week. See Creating NPT Connections for Piping Fittings for more information on the method used to analyze the pipe ends. |
Downloads (pdf format):
Sample 11 - Heat Exchanger
File: PVE-Sample 11
Last Updated: April 2009
CM
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This sample is of a typical fixed tubesheet shell and tube heat exchanger. The exchanger has been designed and calculated in accordance with Part UHX of ASME Section VIII Division I. ASME requires analysis of seven separate load cases on fixed tubesheet exchangers. The cases are made up using all combinations of tube side and shell side design pressures and loadings due to thermal expansion. Shell, tube and tubesheet calculations for all load cases were analyzed using PVElite software. Sample calculations and the drawing can be opened using the links below. |
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Sample 12 - Piping System
File: PVE-Sample12
Last Updated: Oct. 17, 2008
JL
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This is a basic piping system. Registration documentation consists of a calculation set and a P&ID drawing. The drawing must outline the piping layout, materials, design data, and CRN's for the fittings and valves used in construction. Specific notes regarding the installer, pipe hangers, location of pressure valves, etc. may be required as well. |
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Sample 13 - Tower
File: PVE-3602
Last Updated: Dec. 11, 2009
LB
 |
This tower is based on a combination of different vessel designs we have run. Its features are often found on towers, but the combination presented here is unlikely to be used on one tower. The wind and seismic calculations are based on the IBC calculation method but NBC or UBC methods can be used instead. Sample calculations for Advanced Pressure Vessel and the design drawings can be downloaded using the links below. |
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Sample 14 - Vertical Vessel with Bolted Cover
File: PVE-Sample14
Last Updated: June 2009
CM
 |
Three sets of calculations for a vertical vessel with a bolted cover can be downloaded below. |
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Sample 15 - Rectangular Tank
File: PVE-613
Last Updated: April 2004
LB
The Problem:
A rectangular settling tank required a sloping bottom. The application allowed the use of tie rods between the long sides, but not the two ends. Side and top supports were required, but how to account for the interaction between the tie rods and the supports?
Analysis:
Finite element analysis was used for this tank because it allowed the interaction between the different types of supports to be calculated. A plate model was created of the desired geometry. The thickness of the different components were altered during the analysis to optimize the design.
 Plate model of the tank, side and bottom supports, tie rods and base - inside view. Material thicknesses are specified during the analysis and can be easily changed to optimize the design. |
 Plate model - outside view. |
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 Outside surface stresses in the tank. All support sizes and plate thicknesses were kept the same. The maximum stresses were kept bellow allowable limits. Deflections are shown magnified by 75x. |
 Inside surface stresses in the tank. |
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 Deflection plot of the full tank. High deflection locations can be avoided for connections. Peak deflection is 0.066 inches. |
The Solution:
For ease of manufacturing, all the supports sizes and all wall thicknesses were kept the same. The support size and plate thicknesses were minimized. With the finite element method, the deflections were easily checked, and where possible, connections moved away from the highest deflection areas.
For this application, finite element analysis cost less than traditional calculations and provided savings on materials and a better design.
Credits:
This tank was built by Price Schonstrom Inc., 35 Elm Street, Walkerton, Ontario, Canada, N0G 2V0
Sample 16 - Crane
File: PVE-424
Last Updated: Dec. 16, 2003
LB
The Problem:
A 250 foot wide, 85 foot high, 90,000 lb capacity freight handling crane had cracks in its main structural girder under both the main and balancing wheels along most of its length. Repair welding was tried but the crane was re-cracking in the same locations after short periods of use.
Finite Element Analysis was considered to escape the continual repair cycle. Could the crane design be modified to eliminate the need for repairs and to make the crane safe?
Analysis:
A plate model of the overall crane was used to check the overall stress in the crane from gravity, the trolley main load wheels, and from the reaction wheels.
The plate model allowed the interaction between the support legs and the crane to be accounted for. The highest stresses were found with the trolley in the center position shown below.
 Crane under gravity load - deflection is greatly magnified in this and the following shots. |
 Trolley main wheel load: the trolley is at its centre travel position resulting in the highest deflection and stresses. |
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 Trolley reaction wheel load: the trolley is at its centre travel position also the location of the highest deflection and stresses. |
Brick models of one section of the crane girder were made. The original design and a series of design alternatives were checked. Various load conditions were checked with the wheels properly centered and also offset on the rails. The best design alternative was a modification where the original cracked plate material was removed and a new thicker plate was replaced. Several alternative repair plate thicknesses and heights were tried to get an optimum design.
The pictures below are for the girder under the main wheels - similar analysis was done for the reaction wheel stresses.
 Existing crane design - main wheel load on the brick model. Again deflections are highly magnified in this partial girder model. |
 Different alterations were analyzed - this one with gussets on the inside and outside did not work. |
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 The best design for the trolley wheel loads - the original cracked plate was replaced with a thicker plate. |
 A comparison of the stresses in the original design and with alterations - trolley wheel loads. |
The Solution:
The design option where the original cracked plate was removed and replaced with new thicker material was adopted and the crane was repaired. The solution chosen cost less than any of the alternative considered prior to the start of the Finite Element Analysis, and it provided lower stresses. Without Finite Element Analysis, this solution would not have been considered.
Credits:
This analysis was commissioned by and the repair was done by Top Lift Enterprises Inc., 21 Teal Avenue, Stoney Creek, Ontario, Canada, L8E 2P1, 1-905-662-4137
This analysis was done in partnership with Reed Structural Engineering, 62 South Drive, St. Catherines Ontario. L2R 4V2. 1-905-688-3895
Sample 17 - Reversed Dished Head
File: PVE-407
Last Updated: June 2, 2003
LB
The Problem:
The process in this vessel required a reverse dished head. The reverse dished head could not be fabricated thick enough to meet the ASME VIII-1 rules. The chosen solution was to reinforce the head with ribs to prevent snap through.
Various alternate methods of analysis are shown here. Only the plate analysis was used for the actual job. However, the comparison of the various methods is educational.
The head diameter and thickness and design pressure of 75 psi is the same for all of the examples bellow. The material has a yield strength of 30,000 psi and an allowed design stress of 20,000 psi. The maximum allowed membrane (tensile) stress is 20,000 psi, 30,000 at regions of discontinuities. The maximum allowed membrane + bending stress is 30,000 psi, 60,000 psi at discontinuities.
Analysis - 2D Axisymmetric with Linear Material Properties:
This is one of the simplest methods of analyzing this vessel. A cross section of the head without reinforcement is analyzed. Algor assumes that the 2D drawing is symmetrical about an axis (axisymmetric). The results show the stress distribution in the head if there is no material yielding (linear material properties).
 Cross section of the reverse dished head (from center to left side). Stresses are shown for an interior pressure in this and the following shots. |
The peak stress is 54,000 psi in the knuckle region, well above the 30,000 psi yield point. This head fails the ASME VIII-1 code calculations for exterior pressure, but the stresses in the knuckle region are less than the discontinuity stress limit of 60,000 psi. Predicted deflection is 0.15 inches (not shown). Perhaps the head is safe? The ASME code calculations provide a safe pressure of 57 psi for a regular dished head. Also, the use of regular dished head exterior pressure calculations is not proven for a reverse dished head.
Analysis - 2D Axisymmetric with Non-Linear Material Properties:
This analysis allows for material yielding. The same cross section is analyzed, but for this analysis, the pressure is applied in steps, and the material will be allowed to yield (Non-Linear). The results can be seen in this movie.
Up to 64 psi, the head can be seen deflecting linearly under pressure. At 69 psi snap through is beginning (and the deflection is greater than the material thickness). At this point the head has started permanent deformation - it will not return to the original shape after the pressure is removed. Pressures beyond 72 psi show rapid snap through. The final frame shows the fully snapped through shape at 72 psi. This shape is kept permanently after the pressure is removed.
 Defection of the center of the head vs pressure. Snap through starts around 66 psi. |
 Original and final shape of head. Loaded to 75 psi and Pressure released. |
The head snapped through before the 75 psi operating pressure was reached. The test pressure at 1.3x the operating pressure would also not be reached. Actual snap through pressures tend to be lower than calculated pressures due to manufacturing tolerances and material variations. A factor of safety would also be needed.
Analysis - 3D Plate Analysis:
Reinforcing ribs were put on the head to prevent snap through. 3D analysis is required to calculate the stresses. A surface model was created in SolidWorks. The material thickness is specified at time of analysis in the Algor FEA program.
 Plate model - top view - created in SolidWorks. |
 Plate model - bottom view. |
The FEA analysis of the head in Algor showed that the stresses were acceptable. The maximum allowed membrane (tensile) stress is 20,000 psi, 30,000 at regions of discontinuities. The maximum allowed membrane + bending stress is 30,000 psi, 60,000 psi at discontinuities. Peak stresses around stress concentrations can be larger.
 Membrane Stress - limited to 20,000 psi except in areas of discontinuities. At areas of discontinuities, membrane stress can be 30,000 psi. This plot shows maximum membrane stresses at 42,000 psi at a concentration which is acceptable. |
 Total Stress (Membrane + Bending) - limited to 30,000 psi except in areas of discontinuities. At areas of discontinuities, membrane stress can be 60,000 psi. The total stresses are acceptable. |
Analysis - 3D Solid Analysis:
A solid model was created in SolidWorks including the reinforcing ribs and all weld fillets. The actual material thickness was modeled. This was not done for the original analysis, but is included here for educational purposes.
 Solid model - bottom view |
 Solid model detail - meshed at 1/8" mesh size |
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 Top side stress |
 Bottom side stress detail |
The solid model maximum calculated stresses are found in the same location as for the plate model, but are much lower. The solid model accounts better for the stresses at connections, and allows the effect of weld fillets to be included.
The maximum stress is 28,000 psi, found in small peak areas. This value could be used with a fatigue analysis if required. All of the general stresses are below the 20,000 tensile limit, so no stress linearization is required to separate membrane and membrane + bending loads.
 Snap through analysis results for the solid bottom head. pressure at 1 sec is 75 psi. At 3.5x operating press the head starts to yield. |
 Displaced head at 5x operating pressure - displacement magnified 2x. |
The Solution:
The design with the reinforcing ribs was successfully used. A report interpreting the results according to ASME VIII-2 rules allowed the vessel to be registered. A later modification to the process allowed a less expensive double wall head to be used instead.
Comparison of Methods Shown:
The Solid and Plate analysis methods here produced almost identical stress results except at attachments. The Solid model with the weld fillets gave more realistic and lower stress results. The solid model was also easier to make than the plate model which required each surface to be split at all intersections. If the stresses were higher in the solid model, stress linearization would have been required to separate the membrane and membrane + bending stresses. The solid model stress linearization is more difficult than reading the stresses off of the plate model.
Credits:
This tank was built by Price Schonstrom Inc., 35 Elm Street, Walkerton, Ontario, Canada, N0G 2V0
Sample 18 - Flat Head Section
File: PVE-3259
Last Updated: Jan 14, 2009
BV
 |
This sample report illustrates how FEA is used to validate simple designs and code compliance. Pressure Vessel Engineering Ltd. utilizes this short report format to provide a cost savings to our customers while still meeting all CRN report requirements. This format may be applied to many applications and can typically be completed within a few days. |