Some basics of fluid sealing technology

Some basics of fluid sealing technology

Some basics of fluid sealing technology B. S. Nau* The author reviews the main categories of fluid sealing applications. He discusses factors affecti...

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Some basics of fluid sealing technology B. S. Nau*

The author reviews the main categories of fluid sealing applications. He discusses factors affecting fluid flow through the sealing interface and the significant aspects of surface topography, lubrication processes and factors affecting choice of materials. Finally, he outlines heat transfer problems with particular attention to thermal crazing. Fluid sealing technology is concerned with devices for controlling and limiting the flow o f fluids between solid surfaces. For instance, the drive shaft for a centrifugal pump impeller enters the pump casing through an aperture. Some means must therefore be provided for controlling leakage of the pumped fluid between shaft and casing (Fig 1). A similar situation exists where a gearbox drive shaft enters the gearbox, where a journal enters an oil-fed bearing, where the shaft for a chemical reaction vessel mixer passes through the vessel wall, and where a ship's propeller shaft passes through the hull. The configuration is somewhat different in machinery involving axial motion which occurs in hydraulic rams used to actuate such diverse equipment as bulldozer shovels, aircraft ailerons, hydraulic presses, and positioning devices on production machinery (Fig 2). Similar axial motion is also encountered in spool valves between spool and bore as the valve is actuated. In all these devices, leakage control is required between a rod or spindle and a housing, and in a hydraulic ram there is also a need for leakage control between the piston and cylinder bore. Whether rotary or axial relative motion is involved, the sealing problem can be termed dynamic and the sealing devices used involve surfaces in close mutual proximity where one surface is moving at speed relative to the other. Much of the art and science of dynamic sealing is concerned with keeping these surfaces in intimate conformity while simultaneously avoiding the rapid failure, through wear or seizure, which can result if the sliding interface is inadequately lubricated. Here we can recognise an important distinction between a dynamic seal and other components with lubricated sliding contact. In a seal, the lubrication must usually be provided by the process fluid, whatever this is, whereas in bearings and gearbo/(es the lubricant is chosen by the designer for optimal lubrication. Process fluids are commonly aqueous liquids, which may be corrosive, carry solids in suspension, be volatile hydrocarbons or highly viscous polymer melts. The seal must be designed to live in this process fluid, however undesirable the fluid properties. Dynamic seals produce some difficult design problems for the seal technologist but, perhaps surprisingly to the uninitiated, so also do static seals. Under the general heading, static sealing problems, we can group those applications which do not exhibit more or less continuous, relative

* Head of Group D, (Seals, Tribology, and MathematicalServices] bTuid Engineering, British Hydromechanics Research Association, Cranfield, Bedfordshire, UK.

J Diffuser ring

Centrifugal impeller

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Fig I Section through supercavitating centrifugal pump showing a balanced mechanical seal fitted to the rotating drive shaft {reproduced by courtesy o f the Insfftution of Mechanical Engineers)

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Fig 2 Constant yieM double telescopic leg showing seals for reciprocating motion {reproduced by courtesy o f the Institution o f Mechanical Engineers) sliding motion. This is not to say that the situation is truly static. In the case of a gasketted flange joint, there may be limited movement within the joint assembly in response to changing pressures, temperatures and mechanical loads (Fig 3), and the seal must withstand these. The problem of designing to accommodate this type of limited movement can be complex. The key factor is not lubrication but a combination of requirements for plastic flow and elastic resilience within the components of the joint. Plastic flow is required for the initial filling of surface defects or rugosity, elastic resilience is needed to ensure long-term retention

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of adequate contact stress levels in the joint interface. In any case, the interface stresses are not uniform owing, for instance, to the finite number of closure bolts employed, radial temperature variations, or extraneous mechanical loads. There is one further category of sealing problem, namely the pseudo-static seal required for devices such as valve stems. A manually operated stop valve, for instance, may require a spindle seal which is static most of the time but exposed to a slow motion during the opening or closing sequence. The motion may be axial or helical, but in either case is slow.

Flow phenomena Certain flow characteristics can be conveniently considered at this point, for example, the question of turbulent or laminar flow where the significant parameter is the Reynolds number. When inertia forces are large compared with viscous forces, Reynolds number is large and the flow becomes turbulent. Different rules apply to the calculation of flow rates for turbulent and laminar regimes. In a dynamic seal, the Reynolds number might be calculated using the sliding velocity or the leakage velocity. Generally, the former will be greater and can be used to assess the probable flow reigme. To be specific, consider a dynamic seal having a sliding velocity, V, of 5 m/s and operating with water (density p = 103 kg/m-3; absolute viscosity 7/= 10 - 3 Ns/m-2). The flow will be laminar if the Reynolds number, Re, is less than the critical Reynolds number Rec: pVh/rl < Re c for laminar flow; that is h c < ( R e c x 2 x 10 -7) m. If Re c = 200, as in pipe flow, then the clearance will have to be under 0.040 mm for laminar flow to exist. Whether this is likely can be judged by estimating the leakage flow through such a clearance, and comparing with what is experienced in practice. The radial flow through a plane annulus can be approximated by:

Fulcrum(

h 3 Ap 2rrr Q= ]-2r/~Where the flow path is assumed to be short enough to ignore the effect of curvature, this is normally justifiable. Taking the values Ap/Ar = 0.1 N m m - 2 / m m = 108 N/m 3, and r = 25 mm gives: Q_(4x

10-5) 3 1 0 8 0 . 1 5 = 8 x 1 0 - 5 m 3 / s = 0 . 0 8 0 1 / s . 12 10 -3 Such a flow rate is orders of magnitude greater than those in typical dynamic seals in this size range, where values of 1 ml/h (2 x 10 . 7 l/s) are commonly bettered. Clearly, not only is the flow laminar, but the typical clearance between the sliding faces must be an order of magnitude below the value h = 0.040 mm taken above, that is about 4/am or less. That this is indeed the case is shown by electrical capacitance measurements which show that h is typically in the range 1 - 3 / a m . There are exceptions to these conclusions, notably in large diameter or high-speed machinery where clearances and leakage may, of necessity, be much greater. Also, in those seals designed for a fixed separation of the sliding faces, by contrast with the more usual arrangement where the faces are loaded together and the load is carried hydrodynamically or hydrostatically by the interfacial fluid. Another factor which can affect the flow calculations is fluid compressibility. In the case of liquids, this is not normally significant but in very high-pressure hydraulic systems it can be important. A tyt~ical liquid has a bulk modulus of the order of 2 kN/mm z, hence for pressures of about 0.2 kN/mm 2 compressibility is becoming important. Such pressures are encountered in jet cutting applications, for example. With pressures as high as this it should be noted that viscosity can be significantly increased. This has been used to control leakage in high-pressure autoclaves. In the case of a gas, the compressibility is much more important and, of course, many seals are used to control gas leakage. Another factor achieves prominence in gas flow, namely choked flow, which arises where the gas flow velocity exceeds the local velocity of sound. This will happen in a seal when the pressure ratio P1/P2 exceeds a critical value: -Pc = 0.53 for choked air flow. Further increase in the pressure drop does not further increase the gas velocity. Hence once this state is reached the volumetric flow rate is simply Q = 27rr2 hc where c is the local velocity of sound. To illustrate typical leakage flowrates it is convenient to take the example of a cylindrical annular clearance and to plot the leakage as a function of clearance (Fig 4).

Surface geometry Having referred to sliding interfaces with clearances of the order of microns, we clearly cannot ignore surface topography of this order of magnitude (Fig 5). There are two surface characteristics of interest: surface roughness (wavelength ~1 mm or less) and surface waviness (wavelength ~1 mm or more).

Fig 3 Forces and moments acting on the gasketted johtt o f a pressure vessel

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Typical roughness amplitudes may range up to 2 - 3 / a m and wavelengths are typically 10-1 O0/am. Surface waviness (or out-of-roundness on a cylindrical surface) may have amplitudes of the same order but considerably longer wavelength. For dynamic seals, it is usual to make the surface roughness

of the sealing faces as low as economically practicable, often 0.1-0.5/am centre line average, but for certain static joint seals, a deliberately rough finish may be preferred. Surface waviness is dealt with in one of two ways. Where rigid sealing surfaces are used these are made as flat as economically possible. The alternative is to use resilient material for the seal so that the seal can conform without high local stresses; rubber rotary lip seals and soft-packed stuffing boxes are good examples of this approach. In a seal made from rigid materials, such as steel or castiron, quite small movements applied to the seal ring can produce distortions of the order of the sealing clearance, and this is an order of magnitude more important with the carbon-graphites commonly used in mechanical seals and circumferential bushing seals. Clearly, the seal designer must calculate such distortions, and either attempt to reduce them to neglibible proportions by suitable design or design the seal to operate correctly in the deformed state. The deforming movements can arise from the distributed pressure of the sealed fluid acting on the seal, from thermal stresses, or from unpredictable extraneous forces due to clamping loads.

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Hydrodynamic lubrication plays an important part in the lubrication of dynamic seals but this can be supplemented or even supplanted by hydrostatic lubrication deriving its supply of pressurised fluid direct from tire sealed fluid. During stopping and starting, there is likely to be a brief phase of boundary lubrication. As was the case in the 19th century with the original journal bearings, the hydrodynamic and hydrostatic lubrication of a dynamic seal is the result of a happy chance. In a mechanical seal, for instance, the surface waviness provides the required Oh/OOterm which, in conjunction with the I0

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Fig 6 Distribution of hydrostatic (Ps) and hydrodynamic (PD) components or film pressure along the seal length angular velocity, is responsible for hydrodynamic pressure generation. It so happens that, when lapped to a flatness of about 1 light-band per cm, the waviness is about right for the load conventionally applied to ensure adequate film thickness for lubrication, yet it is not so great as to result in excessive leakage. In reciprocating rubber seals for highpressure oil systems, the hydrodynamic effect is supplemented by a hydrostatic pressure field, and both interact with the resilient seal to provide what might be termed elasto-hydrodynamic/static lubrication (Fig 6). In mechanical seals too, there can be coupling between elastic and fluid dynamic effects at higher sealed pressures. This is due to elastic distortion of the seal ring in an angular mode about the neutral axis of the seal's cross-section. The effect of this is to allow the radial hydrostatic pressure field to develop from its initial near-triangular form to a more rectangular form, thereby providing greater load support. In the case of the rotary rubber lip seal and the soft packing, there is still an element of doubt surrounding the details of the lubrication mechanism, although the former has been the subject of considerable research effort over the years.

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Since a seal of whatever type has to start and stop, the need tbr adequate boundary lubrication capability is always important. Rubber dynamic seals are invariably used with oil and consequently there are few problems, but mechanical seals and soft packings may be called upon to operate with any imaginable fluid, and, therefore, there is a wide range of materials in current use for the sealing interface. The materials must not only be inherently good boundary lubricants

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TABLE 1 Some examples of elastomers used in seals: maximum temperatures are for dynamic applications Resistance

Polymer type

SMaximum temperature, °C

Minimum temperature, °C

Compatibility mineral oils water

wear

c o m p r e s s i o n set

Natural Nitrile Polyacrylic Polyurethane Silicon Fluorocarbon

80 130 150 100 200 175

-60 -45 -15 -50 -60 -15

poor good good fair fair good

good fair fair good poor fair

good good poor fair good good

:[: For dyanmic applications t Below 40 °C

when rubbed together, but nmst also be resistant to chemical or electro-chemical attack. One range of materials used perhaps more widely than any other are the carbon-graphites (and composites based on these). However, ceramics and metal carbides are also widely used in mechanical seals, and white asbestos, PTFE, and carbon fibre in soft packings. Where rubbers are used, fluid compatability and temperature sensitivity are recurring problems, even with oils and other hydraulic fluids. It is for these reasons that a wide range of rubber compounds are used, based on such polymers as acrylo-nitrile, polyurethane, polyacrylic, fluorocarbon, silicone, fluorosilicone, and polyimide (Table 1). PTFE can only be used with special seal designs, as it is a plastic which is subject to flow under stress and is not elastomeric. However, its good boundary lubrication properties are greatly sought after and it is used as a coating or impregnant as well as in special solid PTFE seals. Heat transfer Thermal behaviour of dynamic seals requires particular attention for the development of a successful sealing solution. There are two sources of heat, one internal, from the friction at the sliding interface, and the other external from the process fluid or environment. The heat generated at the sliding interface is conducted away through each of the two sealing surfaces and must eventually be dissipated in a heat sink. Often the heat sink is the process fluid and attention to the detail of the heat flow paths is important. For instance, it is usually better to arrange that the moving member of the seal (which may be the shaft for a lip seal or a stuffing box) is of higher conductivity since this ensures maximum exploitation of the beneficial heat transfer rate between the process fluid and the moving member, as compared with the static one. Again, where a fabricated seal is used, it is essential to ensure good thermal contact at joints, as even 0.1 mm of air gap or non-metallic adhesive may greatly increase the overall thermal resistance of the seal member. There are two particular factors to bear in mind when considering the heat transfer process. One is that the fluid in the sliding interface will be at a higher temperature than the process fluid and must not be allowed to reach its boiling point at the lowest pressure to which it is exposed, usually the ambient pressure; secondly, the resultant temperature distribution should not be such as to produce adverse thermal stresses which might upset the sealing interface alignment of a mechanical seal. In mechanical seals, the spring-loaded member is usually rotating as this gives the best heat transfer but, at high speeds, churning of the process fluid surrounding the seal

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good good poor tfair good fair

can cause unacceptable heating, perhaps as much as 50 W for a 60 mm diameter seal in a medium viscosity liquid at 3000 rpm, a figure which can be compared with 150 W for the interface heat generation of a 1000 rpm seal operating with a similar liquid at 2 MN/m 2. It follows that, if churning losses are high, then it is better if this spring-loaded member is static. We have already considered thermal stresses, but these can also give rise to another problem which is localised at the sealing face, thermal crazing. The appearance of the surface in this condition is of a fine network of more or less hairline cracks, which develop in severe cases to produce gross surface damage. There are several factors affecting the occurrence of thermal crazing. A high thermal stress is needed to initiate crazing, and low thermal conductivity plus high elastic modulus and thermal expansion coefficient promote the effect. Low tensile strength is also contributory, as are a high surface friction and high sliding velocity. It is possible to combine the relevant variables to form a parameter, B, which measures resistance to crazing. B-

~TS X E ot F V

where oTS is the tensile strength, X the thermal conductivity E the Young's modulus, ~ the thermal expansion coefficient, V the relative velocity, F the sliding friction. This parameter can be used to rank materials in order of resistance.

Conclusions In fluid seals, the flow regime is Usually laminar and incompressible but, in certain specialised classes of application, turbulent flow exists or compressibility and pressure-viscosity affects can be significant. Surface topography is important but waviness is more important to lubrication than roughness. Deformation of sealing faces in the form o f deflections due to fluid pressure or thermal stresses can significantly modify the initial surface topography on the scale of the lubricant film thickness. The hydrodynamic, hydrostatic and boundary modes of lubrication are all relevant, although not always simultaneously. With resilient materials, elasticity effects are coupled with hydrodynamic and hydrostatic effects. Materials must be chosen for fluid compatibility as well as for boundary lubricant properties. Heat transfer processes are important in minimising lubricant film temperature and both bulk and surface thermal stresses, the former interacting with the full film lubrication modes and the latter causing crazing and premature failure. Acknowledgement I would like to thank the Director and Council of BHRA Fluid Engineering for permission to publish this article.