Designing Trouble-Free Large Obround Nozzles

Treater vessels separate oil, water and solids from an oil well. A firetube heats the liquid mixture lowering the viscosity allowing better gravity separation of the oil. Horizontal treater vessels locate one or more firetubes in the head. Vertical vessels locate it in the shell. Either way a large opening is cut into the vessel to provide removable access for the firetube.

Fig 1 – Flanged obround nozzle with firetube and gasket

Large obround nozzles like those used on treater vessels can be a regular source of field trouble. The nozzle and flange can be subject to high stresses and deflections that interfere with adequate gasket seating. This can show up as a flange that cannot be kept leak free without difficulty.

The design of an ASME compliant obround nozzle is a combination of many mandatory inter-relating code passages. In our practice we are regularly asked to redesign obround nozzles that either do not meet all the required code rules, or do but still leak. To be certain our obround nozzles will be trouble free we first make sure that all applicable code rules are met. Then we use Finite Element Analysis (FEA) to check for excess stresses, flange rotation and displacement. This combination has helped us design reliable firetube nozzles. 

This article follows the design of a sample firetube nozzle through first the code design and then the FEA stages.  The specifications:

  • Materials: SA-516 70 for plate, SA-106 Gr B for firetube pipe and SA-193 B7 for bolting
  • 100% RT is assumed (all weld efficiencies are 1.00)
  • 72″ vessel outside diameter with a 2:1 semi elliptical head
  • 14″ firetube in a 16 x 32″ opening with a full face gasket
  • Operating conditions: 125 psi at 500°F
  • no corrosion allowance
  • Design to ASME VIII-1 code rules where available

This is typical for firetube vessels we design. Click for the full Code Calculations set.  The calculations are discussed throught the article.

Semi Elliptical Head

Typically 2:1 semi elliptical heads are used on treater vessels.  As calculated:

  • 72″ outside diameter
  • 2:1 ratio
  • 0.224 calculated required thickness
  • 0.390 specified minimum thickness after forming (MAF)
  • (see calculations page 3)

The difference between the required thickness (0.224″) and the specified thickness (0.390″) will be used later to help the nozzle calculations.

Fig 2 – The semi elliptical head


The U shaped firetube needs to be removable, so it has to pass through one large flanged opening instead of two smaller nozzles. A single large obround nozzle that minimizes the size of the flat cover on the firetube is the standard solution.  

ASME VIII-1 allows obround nozzles to be designed to code rules when the large to small opening dimensions make a ratio of 2:1 or less. When the long dimension is more than 2 times the short dimension, alternate methods such as FEA analysis from VIII-2 or burst testing must be used. The 16 x 32″ opening in this sample is on the 2:1 ratio limit so the VIII-1 code rules can be used.  Many obround firetube nozzles meet this criteria.  The ASME VIII-1 method calculates the oblong opening as round.  The calculation results:

  • 0.5″ nozzle thickness
  • 0.5″ weld leg size
  • 2.881″ minimum external projection
  • 0.820″ minimum internal projection
  • (see calculations page 5)

Fig 3 – The nozzle has been calculated as a round opening per VIII-1 rules.

Explanation: the ASME VIII-1 nozzle rules start simply. Material lost in the head due to the cut out opening must be replaced. Acceptable sources of this replacement material are excess material in the head (the head was designed thicker than required), excess material in the nozzle wall (the nozzle wall is thicker than required), all material in a reinforcing pad if one is used, material found in nozzle welds and the internal projection of the nozzle neck.  For this example the majority of the material comes from the extra head thickness:

  • 7.160 required area (required, but lost to the design because of the opening)
  • 5.320 excess area of shell (from difference between required and actual thickness)
  • 0.779 excess area in nozzle neck (the nozzle neck is thicker than required)
  • 0.820 area of internal nozzle projection (the internal projection is only used to provide strength to the nozzle)
  • 0.25 area of the nozzle to shell welds
  • 7.170 actual area of nozzle (sum of the above areas which is greater than the required area so area replacement rules pass
  • no repad is required
  • (see calculations page 6)

Additional rules apply, but basically this calculation is based on “area replacement” where all required material in the head opening is replaced. It is an 18th century naval calculation method for cutting openings in ship bulkheads.  It predates our modern understanding of stress concentrations, but is still the core of ASME VIII-1 nozzle calculations.  In spite of its age and simplicity, the calculation method has a long history of successful services for simple round nozzles in most vessel locations.

Obround Neck

The nozzle calculation was simplified by assuming a round opening as allowed by code rules. However, under pressure the obround shape will attempt to become round. The obround nozzle shape is not as strong as a round pipe of the same thickness. The wall can be made thicker or stiffeners added.  ASME VIII-1 Appendix 13 provides rules to design a safe pressurized obround shape.

This obround nozzle is reinforced with stiffeners – in this sample the obround flange acts as the stiffener. The flange is made large enough until it can support the straight sides of the obround nozzle.  The extra head material has already been used for nozzle calculations, and is not considered here as additional support material.  


  • 16 x 32″ inside dimension (17 x 33″ outside dimension)
  • 0.5″ wall thickness (same as nozzle calculation)
  • 2.125″ high x 1.1875″ thick reinforcing ring
  • 5.0″ maximum pitch distance between reinforcing ring and the head (dimension will vary around the opening)
  • (see calculations page 10)

Fig 4 – The obround nozzle as calculated by Appendix 13.  The calculated distance between stiffeners cannot be less than the maximum head to flange distance at any point around the nozzle.  Only one section of the shape is used for the nozzle.

Obround Flange

The flange also has bending loads transmitted through the cover bolts. ASME code calculations exist for the round end portions of the flange. No ASME VIII-1 code rules exist for the straight sides so it is customary to also use BS PD 5500. This is an iterative design process also involving the previous Appendix 13 and nozzle area replacement calculations. The results:

  • 0.5″ nozzle thickness (as used in all previous calculations)
  • 21.25″ short direction outside flange dimension (as used for App 13 above) 
  • 16″ straight side length (as above)
  • 52 bolts @ 5/8″ UNC-11 located on a 20″ bolt circle with 16″ straight sides
  • 0.5″ flange weld size
  • (see calculations page 20 for the ASME flange, last 2 pages of calculations for the BS PD 5500)

Fig 5 – ASME and BS PD 5500 has been used to calculate the flange bending and bolt loads.  Flange dimensions match the reinforcing dimensions from the previously step. No increases in thickness were required.


The firetube experiences the internal pressure of the vessel as a 125 psi exterior crushing load. The pressure from the burner fan is negligible. The calculated requirements:

  • 14″ nominal pipe size (14″ OD)
  • 0.375″ nominal thickness (0.321″ after undertolerance has been removed)
  • 120″ maximum length inside the vessel
  • (See calculations page 24)

Fig 6 – The firetube

Cover and openings

The cover is calculated as a non-round plate according to ASME rules.  The results:

  • 20″ x 36″ bolt pattern (20″ round with 16″ straight sides) same as above
  • 1.1875″ thick
  • (see calculations page 27)

Fig 7 – The cover calculated on its own

The cover has two closely spaced openings into which the 14″ diameter firetube is welded.  The openings are too close together to use area replacement rules, but it is possible to calculate the round ends of the cover as reducing flanges.  No allowable code rules are available for the design of the area between the two nozzles. The results for the flange calculations on the round ends:

  • 20″ bolt circle (same as above)
  • 1.1875″ thickness (same as above)
  • 34 x 5/8″ UNC-11 located on a 20″ bolt circle (52 bolts – bolts on the straight sides)
  • full face soft rubber gasket
  • (see calculations page 29)

Fig 8 – The round end of the cover is calculated as a reducing flange. No code calculation is available for the area between nozzles.

Drawing of the Nozzle As Code Calculated

Finite Element Analysis (FEA)

Many of the above code calculations are in isolation, and the nozzle calculations are for a round not an obround shape. This is the largest limitation of the VIII-1 approach. Also no rules are available for the space between the two nozzles on the cover plate. This is a practical limit for design by rules as found in VIII-1 and other sources. FEA to the rules of VIII-2 is used for the remainder of the analysis.

This is what we are looking for from the FEA analysis:

  1. Excess deflection: The code calculations do not check deflections. Do excess deflections exist, especially in the flange area that can lead to excess bolt bending and/or interfere with gasket operation and seating?
  2. Gasket contact stress: Is the gasket adequately compressed all the way around every bolt hole?  Lack of contact stress greater than the operating pressure is a predictor of gasket leakage through bolt holes.
  3. Excess stress: The calculations passed when examined individually, but is this assembly including its fasteners overstressed once component interactions are considered?

The first two questions ask if the design is reliable.  The last asks if the design has an adequate factor of safety. The FEA analysis shows that all three are issues for this sample in the AS Code Calculated condition:

  • Excess deflection is found in the head along the straight sides of the oblong opening, and on the straight length of the flange.  (See Fig 9)
  • Excess rotation of the flange and cover reduces the chance of a good gasket seating. FEA measures zero gasket contact stress on the inside of the bolt holes predicting difficulty obtaining or maintaining a leak free condition. (See Fig 10)
  • Excess bending stress exists in the nozzle to shell junction and in some of the flange bolts. (See Figs 11 and 12)

This predicts poor serviceability and reduced safety margins. A Revised Design resolves these issues. The changes made are:

  • A 0.5″ thick by 4″ wide repad is added to the nozzle. The maximum nozzle projection remains at 5″, but now measured from the repad, not the head.  Effectively the nozzle is 0.5″ longer.
  • The nozzle wall thickness is increased from 0.500 to 0.750″ (VIII-1 code rules also require the matching increase in the nozzle to flange weld size from 0.5 to 0.75″)
  • The nozzle inside projection is increased from 0.820 to 2.137″
  • The flange OD increases from 21.25 to 24″
  • The flange Bolt circle increases from 20″ to 21″

The reasons for the changes and their effects are explained below.  Click for a drawing of the Revised Design.

Deflection Check

In the As Code Calculated design, both the straight and curved sections of the obround flange show areas of high deflection. No VIII-2 guidelines exist for acceptable deflection limits, but excess displacements lead to high stresses. Increasing the interior projection, nozzle thickness and adding a repad reduces the deflection. Making the nozzle longer reduces the amount of flange rotation caused by the remaining displacement. Increasing the flange width reduces the straight side deflection and rotation. Determining what works is an iterative process. Compare the drawings of the As Calculated Design and the Revised Design to see all of the changes.

Fig 9 – Deflections in the As Code Calculated design (left illustrations) are compared with the Revised Design (right illustrations). The revised design reduces deflections and flange rotation. The closeup section views are taken from the location indicated by the arrows.  The same deflection scale is used for all images, all displacements are shown magnified 80x.

Gasket Seating

Large flange deflections and rotations lead to difficulties getting gaskets to seat and operate leak free. To be leak free, this full face gasket has to have operating compressive gasket stress greater than the operating pressure on all sides of all bolts. The As Code Calculated design has areas around the bolt holes with no seating load which predict a leakage path and field service difficulties. The Revised Design has reduced flange rotation resulting in adequate compressive loads on all sides of all bolt holes.

Fig 10 – Gasket contact stress in the As Code Calculated and Revised Design.  The As Code Calculated has areas indicated by the arrows where the contact pressure is less than the operating pressure indicating a likely leakage path around the bolt holes.  The Revised Design shows a desired seating stress on all sides of all bolts.

FEA Stress

The As Code Calculated design shows stresses around the small end of the flange neck which are too high to pass VIII-2 rules at the allowable VIII-1 stress levels. The multiple design changes in the Revised Design reduce this stress. This is one of the many cases where design by VIII-2 FEA is more restrictive than by use of VIII-1 rules and it shows up in many code compliant nozzles, not just the oblong ones shown here. (This is also discussed in our article comparing hemi, SE, F&D and flat head types.)

Fig 11 – Stress results for the As Code Calculated and the Revised Design.  Stresses are plotted from 0 (blue) to to the 30,000 psi local membrane stress limit (red). The red arrow shows an extensive area in the head to nozzle region where the membrane limit has been exceeded through the full shell thickness.  This region fails VIII-2 FEA stress requirements.  The Revised Design has acceptable stresses through all areas of the full model.

Bolt Stress

ASME VIII-1 code design rules cover allowable bolt tension loads, which when checked using VIII-2 FEA rules pass in both the As Calculated and the Revised Design (Refer to the blue lines in Fig 12 below).  VIII-1 design rules do not provide a way of checking bending stresses in bolts.  The red lines in Fig 12 shows that some bolt bending stress are excessive in the As Code Calculated design, but acceptable after the Revised Design.  Again design to VIII-2 FEA rules is often more conservative than the rules found in VIII-1.

Fig 12 – Flange bolt stresses for the As Code Calculated and the Revised Design.  All bolt tension stresses are acceptable (solid blue lines below the dashed blue line).  The flange motions and rotations as seen in the displacement results (Fig 9) show up here as bolt bending stresses.  Some bolts in the As Code Calculated design have excessive stress (solid red line above the dashed red line).  The Revised Design is acceptable.

Clean Up

The design calculations for the As Code Calculated design initially used to prove the design to the ASME VIII-1 code rules no longer match the Revised Design that passed the VIII-2 FEA requirements.  For an Authorized Inspector or reviewer to check this design, the code calculations need to be updated to match the final Revised Design dimensions. For this sample the result will be code calculations that comply with the VIII-1 rules, but look conservative.

Two sets of documents now exist for code review.  Some reviewers and Authorized Inspectors require the FEA approach, some require the code calculations. Either way, the required documentation is available.

In our experience, using the both the code rules and FEA methods outlined in this article results in reliable and safe obround nozzles.


[PVE-11110 Ben Vanderloo and Laurence Brundrett, July 2017]

ASME Code Design at PVEng

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