FLAT-SEALED FLANGE CONNECTION CALCULATION METHODS

Seals are calculated by checking a specific seal's suitability or by selecting a flange connection's structural features with a view to a specific gasket. This is by no means an algorithm capable of optimizing the structural features of the seal itself, which is selected based on the designer's or user's expertise. Fulfilling the calculation conditions described below assures secure strength with respect to the gasket and tightness at the water test.

It should be added that a flangescrew connection is a system of mutually interacting elements: bolts, flanges, gasket, as well as jacket or pipeline. In the case of an alteration made within a seal, it should be considered what changes it may induce in the other components. It can be assumed that the bolts' and flanges' strength and load conditions are not affected by a change in the seal, if neither the existing assembly and operation bolt loads have been exceeded, nor the average gasket diameter changed. In consideration of a flange connection design, the data independent of the connection itself must be taken into account in the first place, such as the pipeline's operating pressure and nominal diameter, and after the calculation the appropriate flange width, gasket material and the bolts' number and quality must be selected.

The following flange connections calculation methods are used in Poland:

according to DT-UC-90/WO-O [DT-UC-90/WO-O Strength Calculations]
brochure developed by the UDT Office of Technical Inspection, with consideration of the following factors:
- đm – minimum seating stress,
- đr – minimum operating stress, and
- b – relaxation factor,
according to German AD-merkblatt code, with consideration of specific calculation factors determined by DIN 28090-95 standard,
according to US ASME Code; Section VIII, Division 1; Appendix 2, with consideration of the calculation factors m and y specified by the manufacturer according to ASTM standard.

The above-mentioned methods are admitted by the UDT Office of Technical Inspection as calculation formulas for flange-bolt connections in devices subject to the technical inspection. Even if all these methods are based on the same rules, they differ in the way of calculations. Therefore, calculation factors specified by manufacturers of specific sealing materials are assigned to a specific calculation method and are not directly interchangeable.

So, to ensure a flange connection's tightness, it is necessary to properly determine the assembly bolt load. Since for the gasket sheets offered by Sinograf the calculation factors y and m were determined according to the US ASTM standard, therefore the assembly bolt load should be calculated according to the ASME Code.

The assembly bolt load "W" is determined by the following two values. The first is the so-called gasket seating load Wm2; the second is the operation load Wm1. The greater of these values is adopted as the bolt assembly load.

Seating bolt load Wm2 is the force with which a gasket should be pressed in order to close the micropores inside the gasket itself and to fill in the unevenness at the flange facing - gasket interface. Gasket is seated under atmospheric temperature and pressure conditions and with no internal operating pressure. The seating load is constant and depends on the gasket material and dimensions only. For graphite seals it amounts to 900 - 5000 psi, whereas 1000 psi corresponds to 68.94 bar.

The minimum bolt load required to seat a gasket is specified by the following formula (1):

(1) Wm2 = 3.14 bGy, where:

b – active gasket width determined from the surface enclosed between the surface facings and the actual loaded gasket width
G – so-called. active gasket diameter
y – minimum seating stress. This is a gasket material specific parameter determined by the manufacturer.

Operation bolt load Wm1 should carry the loads from internal pressure H and ensure that the gasket remains Hp compression pressure loaded even after the connection has been relaxed. The internal pressure H is the force coming from the service's design operating pressure P related to the active gasket diameter. The compression pressure Hp is expressed in ASME Code as an "m" multiple of the operating pressure.

Operation bolt load is expressed by the formula (2):

(2) Wm1 = H + Hp = 0,785G2P + (2bx3,14GmP), where:

P – operating service pressure
b – active gasket width determined from the surface enclosed between the surface facings and the actual loaded gasket width
G – so-called. active gasket diameter
m – dimensionless calculation factor specified by the seal manufacturer.

The following conclusions can be drawn from analysing the above formulas with regard to graphite seals:

Since the seals made of graphite sheets feature a large elastic return, it is recommended to use thinner seals, which improves the flange connection's tightness without the need to increase its stiffness.
Due to the fact that the graphite gasket seating loads are smaller compared to e.g. rubber-asbestos gaskets, it is possible to replace the gasket material while maintaining the existing flange connection design.

Seal / flange surface interface load

The seal/flange interface is loaded with bolt tensioning while the connection is assembled; this load is called assembly bolt load. For sealing liquids, it is recommended to adopt p_min assembly load:

p_min = a + bp0 [Mpa], where:

p0 – service pressure [Mpa]
a = 4÷ 5 [Mpa]
b = 3

For sealing gases, this load must be much higher:

p_min = a1 + b1p0 [Mpa], where:

a1 = 18 ÷ 20 [MPa]
b1 = 4,5

Service pressure relieves the connection load. This means longer bolts and lower assembly load. The relief is considered when selecting the final minimum assembly bolt load pm_min:

pm_min = p_min + p0π d2/πdśrucz [MPa], where:

d - pipeline diameter
dśr - average seal diameter
dśr= d + ucz
teach - seal width.

It should be checked that the assembly load does not exceed permissible contact load бmax for the material (specified in material properties tables).

With the assembly load and seal dimensions known, the total connection load can be determined, and thus the bolts can be selected.

Description of seal material properties

Compressibility – extent of change in the material thickness under load. This property is determined at pi = 35MPa load and is expressed as a percentage of the initial thickness. The more compressible the material, the easier is to fill-in the flange surface unevenness, but also its compressive strength is lower,
Elastic return – material's ability to recover its original dimension after load relief. Test results show that every material undergoes permanent deformation; the larger the elastic return of a material, the better its response to changes in the bolts and flanges, if there are any.
Compressive strength бmax – compressive stress allowable for the seal material and shape. The specified values correspond to the compressive stress, at which the elastic return of materials is still sufficiently large.
Tensile strength – relevant where the friction forces that keep the seal between the flanges decrease. The sealing agent's pressure on FLAT-SEALED FLANGE CONNECTION CALCULATION METHODS
the seal side stretches it. With an insufficient tensile strength and the friction force decreasing, the seal can break, i.e. get "blown out".

Specific recommendations

Seal material
- Check the selected material's compatibility with the service - see chemical resistance tables.
Seal thickness and width
- The thinner the seal, the more it can be loaded and the slower are its relaxation processes. Thin seals require rigid flanges and smooth face surfaces. 1 ÷ 2 mm thickness is recommended.
- Thicker seals are needed where the surface shape errors and roughness are larger.
- The minimum seal width should be equal to at least five times the material thickness. For sealing gases, the minimum width uczmin ≥12 mm.
- The wider the seal, the smaller the service pressure po impact on the contact load. The bolt loads must increase - to induce pm min - when the seal width increases.
- It follows from Qucz/ d = const equation that to retain leak Q in a flange with a larger diameter d,
  the seal width ucz must increase.
Flanges and bolts
- Seal's proper performance depends on the flange's stiffness, its end faces' alignment and lack of radial scratches.
- Average flange surface roughness should be Rz≤40 mm. Parameter Ra is not recommended for roughness description of the surfaces mating with static seals.
Loading static sealed connections
- The bolt loads in a connection to be pressure-tested should match the test pressure ppr. After the pressure test, the bolts must be loosened and re-tightened to match the operating service pressure p0.
- Seal increases its width and circumference while its bolts are tightened. Good results are achieved by tightening several bolts at once.
- Bolts should be loaded gradually:
  - first load ~ 50% of the full load
  - second load ~ 80% of the full load
  - third load - the full load.
- The bolt loads in a connection may differ by:
  - α= F_max/F_min=1.4÷1.6
- The difference can be reduced by lubricating the bolts and nuts with a grease. Lower coefficients α are adopted for longer bolts, higher for shorter ones.
- Bolt assembly load pmmin'>pmmin α, and the maximum load pmmax< σmax/α
- To secure a seal against blowing out, the load must be
  - pw≈po h/2ucz f≥pmmin', where:
    - h – seal thickness
    - f – friction coefficient
- In a tongue-and-groove flange connection with the seal cross-section appropriately selected the assembly bolt load will not be exceeded. With a certain bolt tightening however, the flat flange surfaces mate and from then on, the connection is not loaded further.
- Seal's inside diameter should be close to the flange's inside diameter. In this case, increasing the service pressure po and the resulting flange deformation only slightly changes the load.
- The higher the operating service temperature, the greater the assembly load should be. It should be remembered that σmax decreases with temperature.
- It is dangerous to tighten the bolts while the connection is in operation.