BRAZILIAN JOURNAL

OF
RADIATION SCIENCES
BJRS 10-03A (2022) 01-17

Design considerations for rectangular bolted full face

flanged joints for surface condensers
Soavea A., Mattar Netoa M.
a
Instituto de Pesquisas Energéticas e Nucleares IPEN-CNEN, Av. Prof. Lineu Prestes, 2242 - 05508-000 - Cidade
Universitária – São Paulo – SP, Brazil
augusto.soave@usp.br

ABSTRACT

Rectangular bolted full face flanged joints are widely used in surface condensers within the power generation
industry including the nuclear one. In order to design these components, it is necessary to analyze the flanged
joint from the point of view of structural strength and leak tightness. This work presents an analytical procedure
applied to a rectangular bolted flange to determine the thickness of the flange, the bolt stresses and leak tightness
conditions. First, the proposed analytical procedure is validated by comparing its results with those from finite
element analysis (FEA) using non-linear approach considering the behavior of the materials, gasket and
contacts. In addition, the proposed procedure is applied to the design of a rectangular flanged joint of a steam
surface condenser using two different gaskets: compressed non-asbestos fiber gasket and NBR elastomer gasket.
The obtained results show a better performance of the NBR elastomer gasket in comparison with compressed
fiber gasket: better sealing condition, reduction of the flange thickness and reduction of the bolt stresses. It is
important to highlight there are practically no references of procedures for design of non-circular full face
flanged joints.

Keywords: Bolted Joint, Full face, Rectangular flange, Surface condenser.

ISSN: 2319-0612
DOI: https://doi.org/10.15392/2319-0612.2022.1989
Submitted: 2022-01-31
Accepted: 2022-11-27
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 2

1. INTRODUCTION

Bolted flanged joints are widely used within all industrial sectors, mainly because their ability to
assemble and disassemble. These joints must fulfill two primary functions: i) to guarantee structural
strength and ii) to maintain the leak tightness of the joints [1].
This work proposes a procedure applied to rectangular bolted full face flanged joints, as shown
in figure 1, which are used in several industrial equipment, like surface condensers in nuclear power
generation systems, digesters, cyclones, chutes and air ducts [2]. This procedure seeks greater
certainty in relation to the sealing requirements, since the procedure presented by the Enquiry Case
133 of PD 5500 [3], for example, underestimates the necessary bolt pretension to seating the
gaskets.
It is important to notice that full face bolted flanged joints are more popular in low pressure
applications and, also, for non-circular flanges which are difficult to seal with ring or strip gaskets.
Despite the continuous use of such flanges, no design rules are contained in the most popular design
code, the ASME VIII Division 1 [4].

Figure 1: Rectangular bolted full-face flanged joint.

Source: Author
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 3

Appendices 2 and Y of ASME VIII Division 1 [4] cover flanges with a ring gasket located
within the bolt circle and flanges with metal-to-metal contact outside the bolt circle, respectively.
The first one was developed in [5] and the second one was developed in [6]. Bolted joints used
in conjunction with soft gaskets over the full face of the flange have no specific design rules, and
the two mentioned appendices are not really suitable for such applications.
Full face gaskets are extensively used in the industry due to their simple and economical design,
and low contact stress sealing requirements. The latter is generally achieved by the use of soft
gaskets such as those based on rubber, elastomers, polytetrafluoroethylene (PTFE), and fibers. The
design of full face flanges should minimize both separation at the bore and flange rotation. In the
absence of a specific standard design procedure, full face flanges are sized by trial and error, or by
an approximate extension of the Taylor Forge method [7].
According to [8], the use of full face flanges has been applied in flanged joints to reduce the
moment applied to the flanges, especially when those have limitations in terms of thickness or
material properties (strength limits).
For this type of joint, to guarantee the leak tightness, it is required many times that the minimum
sealing gasket stress exceeds the region of the holes (figure 2), resulting in a large area of the joint
to be compressed. Thus, it is necessary that a great force must be applied by the bolts, requiring a
great bolt section area. Therefore, the use of this type of flange can become costly for equipment
with high pressures, which makes this application common and convenient for low pressure one.

2. METHODOLOGY

The proposed procedure of this paper is based on Enquiry Case 133 of PD 5500 [3] applied to
rectangular bolted flanged joints. Two changes were done:
1) The area assumed to seating gasket is estimated as A1+A2-A3 (figure 3)
where A1 = area outside bolts lines, A2 = area of effective gasket inside bolts line and A3 =
area of bolt holes
2) A factor considering a linear stress distribution in gasket region is applied to achieve seating
stress in the limited line inside bolts lines (e.g., 5 mm near hole [9][3]).
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 4

Figure 2: Leak and no leak conditions due gasket stress values and extension.

Source: Author

The results obtained from the proposed procedure were compared with other procedures already
used in the industry, i.e., equivalent circular flange method with the Taylor-Forge full face flange
method [10] and the Enquiry Case 133 of PD 5500 [3] and, also, with the results from nonlinear
finite element analysis (FEA). Some conclusions and comments were addressed based on the
comparisons.

Figure 3: Assumed areas of proposed procedure and rectangular gasket dimensions

Source: Author
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 5

3. PROPOSED PROCEDURE

The proposed procedure is shown below where the parts in bold refer to the modifications made
in Enquiry Case 133 of PD 5500 [3] formulation. The proposed procedure is divided into 4 steps for
better understanding:
Step 1: Gasket Details

Gasket width
b'o = min [(G0L - CL); (CL - A1L)] (1)

Effective gasket width

b' = max[ ; d+5] (2)

Long side length of gasket reaction

GL = CL - b' (3)

Short side length of gasket reaction

GS = CS - b' (4)

Step 2: Forces, moments arms and moment calculation

Linear stress distribution on gasket along flange face

K= (5)

Required gasket compression force

HG= 4b'(GL+GS)mpK (6)

Gasket Reaction arm (figure 4)

hG = b'/2 (7)
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 6

Hydrostatic Force

H = p(CL - d)(CS - d) (8)

Hydrostatic force applied by shell junction

HD = pBLBS (9)

Hydrostatic force on flange face

H T = H – HD (10)

Application arm of HD force

hD = (CL-BL-g1)/2 (11)

Application arm of HT force

hT = (CL-GL)/2 (12)

Moment acting on the flange

M = HDhD + HGhG + HThT (13)

Arm of hr force outside bolts line

(14)
hR =

Reaction force outside bolt line

HR = M/hR (15)

Step 3: Bolt loads and strength

Minimum required bolts load for operation condition

 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 7

Wm1 = HG + H + HR (16)

Minimum required bolts load for gasket seating condition

Wm2 = [G0LG0S - CLCS + 2(GL+GS)b' - nd²π/4]Ky (17)

Figure 4: Forces and arms

Source: Adapted from PD5500 [3]

The bolts load required is

Wm = max(Wm1;Wm2) (18)

The one bolt load required is

w= Wm / nb (19)
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 8

And the bolt stress

Sb' = w / Ab (20)

Step 4: Flange thickness

Required flange thickness for operation condition

(21)
t1 =

Required flange thickness for maximum distance between bolts

t2 = (22)

The required flange thickness for the full face bolted flanged joint is

t = max(t1;t2) (23)

4. APPLICATION OF THE PROPOSED PROCEDURE IN A CASE

STUDY

The main rectangular flange of a surface condenser was adopted as a case study of this work.
This joint is subjected to a design pressure of 98,1 kPa and design temperature of 150 °C. The
material of the neck and flange is SA-516 70 and the bolts are manufactured of SA-193 B7, both
materials according to ASME II [11]. Two types of gasket materials were analyzed:
1) CS1: Compressed non-asbestos fiber gasket with minimum seating gasket stress (y) of 24,13
MPa and maintenance factor (m) of 2;
2) CS2: NBR Elastomer with minimum seating gasket stress (y) of 1,4 MPa and maintenance
factor (m) of 1.
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 9

Table 1 shows the complete input data for CS1 calculation, Table 2 shows CS2 gasket inputs
(the other data are the same of Table 1) and Table 3 shows the output for both cases

Table 1: Flange joint and CS1 input data.

Symbol Value Unit Description
p 0,0981 MPa Design pressure
CL 2754 mm Length of hole center line parallel to the long side
CS 1040 mm Length of hole center line parallel to the short side
A1L 2638 mm Internal long side gasket length
BL 2638 mm Internal long side flange length
BS 928 mm Internal short side flange length
d 36 mm Bolt hole diameter
n 54 Number of bolts
Ab 694 mm² Bolt tensile stress area
g1 19 mm Neck thickness
G0L 2840 mm External long side gasket length
G0S 1130 mm External short side gasket length
y 24,13 MPa Minimum gasket seating stress
m 2 Gasket maintenance factor
Pbmax 160 mm Maximum spacing between bolts
Pbmin 110 mm Minimum spacing between bolts
db 33 mm External diameter of bolt
E 195 000 MPa Flange material modulus of elasticity
SFO 138 MPa Flange allowable stress

Table 2: CS2 gasket input data.

Symbol Value Unit Description

y 1,4 MPa Minimum gasket seating stress
m 1 Gasket maintenance factor

Table 3: CS1 and CS2 outputs.

Symbol CS1 Result CS2 Result

b'o 86 mm 86 mm
b' 41 mm 41 mm
GL 2713 mm 2713 mm
GS 999 mm 999 mm
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 10

K 1,347 1,347
H 267 702 N 267 702 N
HG 160 846 N 80 423 N
HD 240 155 N 240 155 N
HT 27 547 N 27 547 N
hD 48,5 mm 48,5 mm
hG 20,5 mm 20,5 mm
hT 20,5 mm 20,5 mm
M 15 509 588 N.mm 13 860 914 N.mm
hR 30,5 mm 30,5 mm
HR 508 511 N 454 456 N
Wm1 937 060 N 802 582 N
Wm2 19 318 576 N 1 120 753 N
Wm 19 318 576 N 1 120 753 N
w 357 751 N 20 755 N
Sb' 515,5 MPa 29,91 MPa
t1 10,93 mm 10,33 mm
t2 39,42 mm 23,65 mm
t 39,42 mm 23,65 mm

5. FINITE ELEMENT MODEL

The finite element model (figure 5) used the symmetry presented by the joint and was
discretized with solid elements according to table 4. Flanges and neck materials were modeled with
elastic-plastic behavior according ASME VIII Division 2 item 3-D [12] (figure 6 shows below the
stress-strain curve), bolts material was modeled with linear elastic material (Elastic modulus are 191
GPa and 184 GPa, for ambient and design temperature, respectively), and gaskets materials were
modeled with a loading and unloading characteristic curve of each material (figure 7).
The 3D geometry used for numeric model can also be seen in figure 5, the simplified bolts were
modeled together with the nuts and without the thread detail, having a diameter corresponding to
the stress area according to ASME PCC-1[13]. The nuts were simplified by cylinders, without the
hexagonal parts. Weld details have been omitted.
Three frictionless contacts were applied to constrain the model, one on each plane of symmetry
and one on the under end of the neck. The contacts were configurated according to table 5 below.
The loads of model were inputted in two steps:
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 11

1) Installation: Bolt pretension and body temperature of 20 ºC (ambient).

2) Operation: Internal surface design pressure of 98,1 kPa and body temperature of 150 °C
(design).

Figure 5: FEA Model

Source: Author

Table 4: Mesh.

Component Mesh and Element

Hexagonal mesh with 7 elements (2,3 mm) through
neck thickness and ratio of 1:3:3 (1 thickness, 3
Neck
height, 3 width). Element type: Solid185®[14]

Hex dominant mesh with 10 elements (4,7 mm)

through flange thickness and ratio of 1:1:1. Element
Flange
type: Solid185®[14]

Near bolt holes, hex dominant mesh with 7 elements

(6,7 mm) through blind flange thickness and ratio
1:1:1. Far bolt holes, hexagonal mesh with 6
Blind flange
elements (7,8 mm) through blind flange thickness and
ratio of 1:1:1. Element type: Solid185®[14]

Bolts Hex dominant mesh, element size of 5 mm. Element

type: Solid185®[14]

Gasket Hex dominant mesh, element size of 2 mm. Element

type: Inter185®[14]
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 12

Figure 6: SA-516 70 stress-strain curve

Source: Author

Figure 7: Gasket loading and unloading curve

Source: CS1 according [15] and CS2 according [16]

Geometrically and Materially Nonlinear Analysis was performed using the software Ansys
Workbench release 2021 R1 [14].
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 13

Table 5: Contacts.

Contact Configuration
Flange-neck Bonded
Flamge-nut Bonded
Blind flange-nut Bonded
Gasket-flange Frictional (1)
Gasket-blind flange Frictional (1)

(1) Friction coefficient = 0,25

6. RESULTS AND DISCUSSION

For comparison purpose between the proposed procedure and previous existing procedures, the
results are organized in Table 6 for CS1 gasket and for CS2 gasket. The name of each column refers
to the procedure used: equivalent circular flange with Taylor forge method (ECTF) [10], the
procedure presented in the Enquiry case 133 of PD 5500 (133/5500) [3] and lastly the results
presented by the proposed procedure by this work (Proposed).
Bolt stresses were also analyzed in the FEA. First, the bolt section that presented the highest
stress value was located and then its acceptance criteria were verified (see figure 8 where bolts with
highest section stress are indicated). The acceptance criterion used for the bolts is described in
paragraph 5.2.2 and 5.7 of ASME VIII Division 2 [12]. In the proposed procedure, the maximum
bolt stress (515,5 MPa) exceeds the allowable stress and therefore is unsatisfactory for this
application. This same unsatisfactory result was obtained in FEA, where the maximum bolts stress
is above the allowable limits. For CS2 the stresses are below the allowable limits and show
satisfactory results for both approaches (proposed procedure and FEA).
The paragraph 5.2.4 of the ASME VIII Division 2 [12] was applied to check the acceptability of
the flange in the finite element analysis. The obtained stress values are within the code allowable
limits. Therefore, the thickness of the flange is acceptable according to the design by analysis and
also by the proposed procedure (for CS1 case, the required flange thickness is 39,42 mm and for
CS2 case the required flange thickness is 23,65 mm. The finish flange thickness is 47 mm).
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 14

Figure 8: Stresses in bolts sections for CS1 (left) and CS2 (right)

Source: Author

Table 6: Results Comparison.

CS1: Minimum gasket seating stress (y) = 24,13 MPa

Item ECTF 133/5500 Proposed
Bolt Load 145,67 kN 125,79 kN 357,75 kN
Bolt Stress 210,0 MPa 181,2 MPa 515,5 MPa
Required Flange Thickness 34,75 mm 39,42 mm 39,42 mm
CS2: Minimum gasket seating stress (y) = 1,4 MPa
Item ECTF 133/5500 Proposed
Bolt Load 8,4518 kN 12,799 kN 20,755 kN
Bolt Stress 12,18 MPa 18,44 MPa 29,91 MPa
Required Flange Thickness 28,30 mm 23,65 mm 23,65 mm

Regarding the gasket compression stress, the evaluated results were taken with 5 mm from the
bolt holes (as illustrated in figure 9). For the FEA carried out with bolt pretension calculated in
accordance with Enquiry case 133 of PD 5500 [3], the gasket stress observed was lower than the
required seating stress of the gasket material. The compression gasket stress for CS1 case observed
was 8,787 MPa and for CS2 case this value was 0,867 MPa (remembering the minimum values
required are 24,13 MPa and 1,4 MPa respectively). Otherwise, for FEA carried out with bolt
pretension calculated by proposed procedure, the observed gasket stress was close to the minimum
required stress (see figure 10). CS1 case reached 24,21 MPa (slightly above the required gasket
stress) and CS2 case reached 1,4 MPa (exactly the required stress).
Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 15

Figure 9: Evaluated line of gasket compression stress

Source: Author

Figure 10: Gasket Stress

Source: Author
 Soave, A., Mattar Neto, M. ● Braz. J. Rad. Sci. ● 2022 16

7. CONCLUSION

Based on the results obtained from the proposed procedure and from FEA, it can be concluded
that there was an excellent agreement between them. The bolt pretension calculated by proposed
procedure presents an exact value of the required seating gasket stress.
According to proposed procedure, it can also be observed that the bolted joint of the CS1 case
presented an unsatisfactory design result, since the obtained bolt stresses are greater than the
allowable limits. The results presented for CS1 case demonstrate that the bolted joint requires to be
modified, with a significant increase of bolt section area (increase number of bolts and increase
bolts diameter). The CS2 case, on the other hand, presented a satisfactory design result, with all
stresses values bellow the allowable limits. Therefore, as expected, full face bolted flanged joint
presents better results when using a softer gasket (as the elastomer used in CS2 case).
The pretension bolt load value calculated by the proposed procedure refers to the minimum
required by the gasket material seating. Therefore, from the installation point of view, it is
recommended to use a bolt torquing that applies a pretension of up to 50% greater than the bolt
pretension calculated by the proposed procedure, in order to increase the compression gasket stress.
It is also worth mentioning that it is necessary to follow a correct torquing procedure to achieve
uniformity in the gasket stress.

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