Sealing contact

Figure 1-8. Special mechanical seals (contact seals) used in hot gas turboexpanders. (Source GHH-Borsig.)...

The secondary seal contact surface(s) shall not exceed a roughness of 63 min. (1.6 (jim). Seal chamber bore comers and entry holes, such as those used for flushing or venting, shall be suitably chamfered or rounded to prevent damage to secondary seals at assembly. 

Hot-melt adhesive films (heat sealing), contact adhesives, solvent-based adhesives. 

A floating ball valve has a ball with an upper shaft to rotate the ball for operation. Upstream pressure on the ball pushes the ball into sealing contact with the downstream seat. Since the pressure acting on the ball is the diameter of the seat area and not the more limited seat area of the trunnion valve, floating ball valves normally require more torque to operate than an equivalent pressure trunnion ball valve. 

Seals are extensively used within the PCS at component interfaces to maintain the integrity of the helium flow path and associated operating conditions. These conditions lead to significant pressure and temperature differentials across the seals, as well as to substantial differential motions of the components which each side of the seals contacts. 

There are low energy loads (e.g., 50 volts or below and/or 10 mA or below) that require special contact materials or designs (e.g., hermetically sealed contacts) to eliminate oxidation buildup on contacts, resulting in unreliable operation (e.g., load dropout). This is referred to as contact wetting. When utilizing these special contacts, specific analysis is needed for these contacts to ensure that a fail-to-safe electromechanical system is being designed. 

The main part of a hydraulic cylinder seal of compact type is the soft seal element. This is usually the only part taken into consideration when calculating hydrodynamic properties in the seal contact like oil film thickness, leakage flow and friction forces. Johannesson deals in the works (5), (6) and (8) with this problem using the inverse hydrodynamic theory presented by Block (1) in 1963. The same method is used in works presented by Fazekas (2), Hirano and Kaneta (3) and Olsson (9). 

The inverse hydrodynamic theory requires a known pressure distribution in the seal contact as input data. This pressure distribution is inserted in the Reynolds equation, and then the oil film thickness, the leakage flow and the friction force can be calculated. Assumed or measured pressure distributions have been used in the works by Fazekas, Hirano and Kaneta and Olsson. 

Methods to calculate the pressure distribution in the contact zone are presented by Johannesson (4) and Johannesson and Kassfeldt (7). In (4) a semi-empirical method for the calculation of the pressure distribution in an 0-ring seal contact for arbitrary sealed pressures is presented, cind in (7) an approximate analytical method, for calculation of the pressure distribution in an arbitrary elastomeric seal contact, is suggested. In both these papers measurements verifying the calculated pressure distributions are also presented. 

The formation of oil films does not effect static pressure distributions in the seal contacts. 

At least one pressure maximum always exists somewhere in the seal contact. 

Starvation does not occur, i.e. the seal contacts are well lubricated. 

The oil film thicknesses in the different seal contacts are calculated in the same manner as in reference (5), i.e. at each point in a seal contact one of the following equations must be solved ... 

The hydrodynamic pressure, built up at the beginning and the end of the seal contacts, shall coinside with the static pressure distribution. 

The total pressure derivative distribution is also necessary to know. These quantities can be determined if the pressure distribution in the seal contact region is known. 

In the figures 3.5 and 3.6 the oil film thickness results are shown for the main seal element alone. Here the oil films are much thicker than in the main seal element part of the curves in figure 3.1 and figure 3.2. The reason is that the leakage is now governed by the slopes of the pressure curve in the main seal contact, wh ch are smaller than the maximum slopes in the whole contact. Note that the film thicknesses and shapes are the same as one would get in the main seal contact with drain grooves on the back-up rings. 

Johannesson, H. Calculation of the Pressure Distribution in an 0-ring Seal Contact. Paper XI (ii), Proc. 5th Leeds-Lyon Symp. on Tribology, Leeds, England, Sept. 1978. 

Johannesson, H.L. and Kassfeldt, E. Calculation of the Pressure Distribution in an Arbitrary Elastomeric Seal Contact. TULEA 1985 02, Lulea University of Technology, Sweden. 

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