Why Gaskets Are Used

Gasket are used to create a static seal between two stationary members of a mechanical assembly and to maintain that seal under operating conditions which may vary dependent upon changes in pressures and temperatures. If it were possible to have perfectly mated flanges and if it were possible to maintain an intimate contact of these perfectly mated flanges throughout the extremes of operating conditions; a gasket would not be required. This is virtually an impossibility either because of:

  • The size of the vessel and/or flanges
  • The difficulty in maintaining such extremely smooth flange finishes during handling and assembly
  • Corrosion and erosion of the flange surfaces during operations

As a consequence, relatively inexpensive gaskets are used to provide the sealing element in these mechanical assemblies. In most cases, the gasket provides a seal by external forces flowing the gasket material into the imperfections between the mating surfaces. It follows then that in a properly designed gasket closure, three major considerations must be taken into account in order for a satisfactory seal to be achieved:

  1. Sufficient force must be available to initially seat the gasket. Stating this another way, adequate means must be provided to flow the gasket into imperfections in the gasket seating surfaces.
  2. Sufficient forces must be available to maintain a residual stress on the gasket under operating conditions to ensure that the gasket will be in intimate contact with the gasket seating surfaces to prevent blow-by or leakage.
  3. The selection of the gasket material must be such that it will withstand the pressures exerted against the gasket, satisfactorily resist the entire temperature range to which the closure will be exposed and withstand corrosive attack of the confined medium.

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Effecting a Seal

A seal is affected by compressing the gasket material and causing it to flow into the imperfections on the gasket seating surfaces so that intimate contact is made between the gasket and the gasket seating surfaces preventing the escape of the confined fluid. Basically there are four different methods that may be used either singly or in combination to achieve this unbroken barrier.

  1. Compression (Figure 1 ). This is by far the most common method of effecting a seal on a flange joint and the compression force is normally applied by bolting.
  2. Attrition (Figure 2). Attrition is a combination of a dragging action combined with compression such as in a spark plug gasket where the spark plug is turned down on a gasket that is both compressed and screwed into the flange.
  3. By heat, such as in the case of sealing a bell and spigot joint on cast iron pipe by means of molten lead. Note, however, that after the molten lead is poured, it is tampered into place using a tamping tool and a hammer.
  4. Gasket lip expansion. This is a phenomenon that would occur due to edge swelling when the gasket would be affected by confined fluid, as in the case of elastomeric compounds affected by the confined fluids, such as solvents, causing the gasket material to swell and increase the interaction of the gasket against the flange faces.
collage0102 fig2

Generally, gaskets are called upon to effect a seal across the faces of contact with the flanges. Permeation of the media through the body of the gasket is also a possibility depending on material, confined media, and acceptable leakage rate.

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Gasket Seating

There are two major factors to be considered with regard to gasket seating.

The first is the gasket material itself. The ASME Unfired Pressure Vessel Code Section VIII, Division 1 defines minimum design seating stresses for variety of gasket materials. These design seating stresses range from zero psi for so-called self-sealing gasket types such as low durometer elastomers and O-rings to 26,000 psi to properly seat solid flat metal gaskets. Between these two extremes there are a multitude of materials available to the designer enabling him to make a selection based upon the specific operating conditions under investigation. Table 1 indicated the more popular types of gaskets covered by ASME Unfired Pressure Vessel Code.

The second major factor to take into consideration must be the surface finish of the gasket seating surface. As a general rule, it is necessary to have a relatively rough gasket seating surface for elastomeric and PTFE gaskets on the order of magnitude of 500 microinches. Solid metal gaskets normally require a surface finish not rougher than 63 microinches.

Semi-metallic gaskets such as Spiral Wound fall between these two general types. The reason for the difference is that with non-metallic gaskets such as rubber, there must be sufficient roughness on the gasket seating surfaces to bite into the gasket thereby preventing excessive extrusion and increasing resistance to gasket blowout. In the case of solid metal gaskets, extremely high unit loads are required to flow the gasket into imperfections on the gasket seating surfaces. This requires that the gasket seating surfaces be as smooth as possible to ensure an effective seal.

Spiral Wound gaskets, which have become extremely popular in the last fifteen to twenty years, do require some surface roughness to prevent excessive radial slippage of the gasket under compression. The characteristics of the type of gasket being used dictate the proper flange surface finish that must be taken into consideration by the flange designer and there is no such thing as a single optimum gasket surface finish for all types of gaskets. The problem of the proper finish for gasket seating surface is further complicated by the type of the flange design. For example a totally enclosed facing such as tongue and groove will permit the use of a much smoother gasket seating surface than can be tolerated with a raised face. Table 3 includes recommendations for normal finishes for the various types of gaskets.

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Forces Acting on a Gasketed Joint

The Internal Pressure: These are the forces continually trying to unseal a gasketed joint by exerting pressure against the gasket (blowout pressure) and against the flanges holding the gasket in place (hydrostatic end force). See Figure 1
The Flange Load: The total force compressing the gasket to create a seal, i.e., the effective pressure resulting from the bolt loading.

design12

Temperature: Temperature creates thermo-mechanical effects, expanding or contracting the metals, affecting the gasket material by promoting "creep relaxation" which is a permanent strain or relaxation quality of many soft materials under stress.The effect of certain confined fluids may become increasingly degrading as temperature rises and attack upon organic gasket materials is substantially greater than at the ambient temperatures (about 75°F). As a rule, the higher the temperature, the more critical becomes the selection of the proper gasket.

Medium: The liquid or gas against which the gasket is to seal.

General Conditions: The type of flange, the flange surfaces, the type of bolt material, the spacing and tightness of the bolts, etc.

Each of these factors require consideration before an effective gasket material is finally chosen. However, the proper gasket may often be rejected because failure occurred due to a poorly cleaned flange face, or improper bolting-up practice. These details require careful attention, but if complied with will help eliminate gasket blow-out or failure.

There are three principal forces acting on any gasketed joint. They are: Bolt load and/or other means of applying the initial compressive load that flows the gasket material into surface imperfections to form a seal. The hydrostatic end force, that tends to separate flanges when the system is pressurized. Internal pressure acting on the portion of the gasket exposed to internal pressure, tending to blow the gasket out of the joint and/or to bypass the gasket under operating conditions. There are other shock forces that may be created due to sudden changes in temperature and pressure. Creep relaxation is another factor that may come into the picture.

Figure 1 indicated the three primary forces acting upon a gasketed joint which we will consider for this discussion. The initial compression force applied to a gasket seating surfaces regardless of operating condition. Initial compression force must be great enough to compensate for the total hydrostatic end force that would be present during operating conditions. It must be sufficient to maintain a residual load on the gasket/flange interface. From a practical standpoint, residual gasket load must be "X" time internal pressure if a tight joint is to be maintained. This unknown quantity "X" is what is known as the "," factor in the ASME unfired pressure vessel code and will vary depending upon the type of gasket being used. Actually the "m" value is the ratio of residual unit stress (bolt load minus hydrostatic end force) on gasket (psi) to internal pressure of the system. The larger the number used for "m," the more conservative the flange design would be, and the more assurance the designer has of obtaining a tight joint.

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Bolt Load Formulas

The ASME Unfired Pressure Vessel Code, Section VIII, Division 1 defines the initial bolt load required to seat a gasket sufficiently as:
BoltLoad102

The required operating bolt load must be at least sufficient, under the most severe operating conditions, to contain the hydrostatic end force and, in addition, to maintain a residual compression load on the gasket that is sufficient to assure a tight joint. ASME defines this bolt load as:
BoltLoad2

After Wm1 and Wm2 are calculated, then the minimum required bolt area Am is determined: Bolts are then selected so that the actual bolt area Ab is equal to or greater than Am Ab = (Number of Bolts) X (Minimum Cross-Sectional Area of Bolt in Square Inches) The maximum unit load Sg(max) on the gasket bearing surface is equal to the total maximum bolt load in pounds divided by the actual sealing area of the gasket in square inches.

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Notation Symbols and Definitions
NotationSym

DEFINITIONS Except as noted, the symbols and definitions below are those given in Appendix II of the 1977 ASME Boiler and Pressure Vessel Code, Section VIII.

Ab = actual total cross-sectional area of bolts at root of thread or section of least diameter under stress, square inches.

Am = total required cross-sectional area of bolts, taken as the greater of Am1 or Am2, square inches.

Am1 = total cross-sectional area of bolts at root of thread or section of least diameter under stress, required for the operating conditions.

Am2 = total cross-sectional area of bolts at root of thread or section of least diameter under stress, required for gasket seating.

b = effective gasket or joint-contact-surface seating width, inches.

Table 2 bo = basic gasket seating width, inches.

Table 2. G = diameter at location of gasket load reaction.

Table 2. m = gasket factor.

Table 1 N = width, in inches, used to determine the basic gasket seating width b0, based upon the possible contact width of the gasket.

Table2. P = design pressure, pounds per square inch.

Sa = allowable bolt stress at ambient temperature, pounds per square inch.

Sb = allowable bolt stress at operating temperature, pounds per square inch.

Sg = Actual unit load at the gasket bearing surface, pounds per square inch. Wm1 = required bolt load for gasket seating, pounds.

Wm2 = minimum required bolt load for gasket seating, pounds.

y = gasket or joint-contact-surface unit seating load, minimum design seating stress, PSI Table 1 pounds per square inch.

*The Pressure Vessel research Council (PVRC) has developed a program to better identify loads based on gasket "sealability". Thus, new design factors are anticipated to appear in upcoming revisions of the ASME Boiler and Pressure Vessel Code. (Lamons is a sponsor of PVRC research).

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Sample Gasket Application Problem

For assistance with a particular gasket problem contact Lamons sales Department, or a technical representative.

Example Conditions: A designer wants a gasket recommendation for a special application sealing at 600 psi and 500°F.

Conditions:
  • Design pressure - 600 psi
  • Test Pressure - 900 psi
  • Design temperature - 500°F
  • Process material - steam Flange details
  • Bolting - 24 - 1-1/8 thds.
  • Bolt Material - ASTM A 193 - B7
  • Flange Material - ASTM A312 Type 316S.S.
samplegasket

Allowable bolt stress @ Ambient Temperature, according to Stress Table 1, Page 45 is only 20,000 PSI; however, to prevent leakage under hydrotest it is decided to tighten bolting to 30,000 PSI. Allowable Stress @ 500°F - 20,000 PSI Analysis The pressure-temperature conditions indicate a metallic gasket should be used. The conditions appear to be suitable for a spiral wound gasket. The flange material 316S.S., is compatible with the steam environment @500°F. Therefore, the logical choice for the metal in the gasket is 316 S.S Since Grafoil® is also compatible with the environment, it is selected as the filler material.

Analysis

The pressure-temperature conditions indicate a metallic gasket should be used. The conditions appear to be suitable for a spiral wound gasket. The flange material 316S.S., is compatible with the steam environment @500°F. Therefore, the logical choice for the metal in the gasket is 316 S.S Since Grafoil® is also compatible with the environment, it is selected as the filler material.

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