Handbook of Lubrication Episode 2 Part 10 ppt

POSITIVE RUBBING CONTACT SEALS Mechanical Face Seal The mechanical face seal, or end face seal, in Figure 1 is a device for sealing the annular space between a rotating shaft and a housi

e High-torque, high-misalignment applications are those where the centrifugal forces

are lower than 220 G’s, misalignment is larger than 3/4° (usually between 1.5 and 3°) and the shock loads exceed 2.5 times the continuous torque Many such applications also have high-ambient temperatures, and only a few greases can perform satisfactorily Besides the characteristics of a grease for “normal applications”, the grease should also have antifriction and antiwear additives, extreme pressure (EP) additives, a Timken®

OK load greater than 40 lb, and a minimum dropping point of 150°C

4 Gear coupling oils should always be of high viscosity grade, (no less than 150 SSU at

100°C) Although the viscosity cannot be too high for satisfactory coupling operation,

oils with viscosities higher than 1000 SSU at 100°C should not be used since they cannot practically be poured into a coupling

Continuous oil flow lubrication uses the oil from the system, which is seldom a high viscosity oil To increase its viscosity, the oil should be cooled before it enters the coupling

Relubrication Procedure

If manufacturers’ recommendations are not available, oil-filled couplings should be re-lubricated every six months and grease-filled couplings once a year Coupling guards should

be observed periodically for evidence of lubricant escaping from the couplings The causes for this malfunction (improper sealing) should be found and corrected, and the coupling refilled with lubricant before restarting Unless the grease used has no oil separation under the centrifugal forces present in the coupling, it is advisable to open and clean the coupling

before relubrication Without cleaning, additional soap is introduced in the coupling; and

as indicated previously, too much soap is detrimental to coupling performance

The quantity of lubricant that should be used depends on the internal volume of the coupling, which varies not only with the size of the coupling, but also with the coupling type and make The lubricant volume of every coupling can be found in the catalog or instruction manual If lubricant volume is not available, one can use the following method: the two halves of the coupling should be so assembled that the lube plugs of the halves are diametrically opposite; the couplings should be rotated until the lube plugs are at 45° to the vertical plane; both lube plugs should be removed; grease should be pumped through the lower hole until it flows out the upper hole This method may cause some overfilling, in which case some lubricant escapes past the seals on start-up Excessive overfilling should

be avoided because it generates high-thrust forces on the equipment bearings

REFERENCES

1 AGMA, Nomenclature for Flexible Couplings, Standard No 510.01, American Gear Manufacturers

As-sociation, Arlington, Va., 1965.

2 Calistrat, M M., Wear and lubrication of gear couplings, Mech Eng., 28, October 1975.

3 Clapp, A M., Fundamentals of lubrication relating to operation and maintenance of turbomachinery, 2nd

Turbomachinery Symp., Texas A & M University, October 1973.

4 Calistrat, M M., Grease Separation Under Centrifugal Forces, ASME Paper 75 PTG-3, American Society

of Mechanical Engineers, New York, 1975.

5 Filepp, L., Lubricant as a coolant in high speed gear couplings, J Lubr Technol Trans ASME, 178,

January 1970.

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DYNAMIC SEALS

W K Stair

CLASSIFICATION Fluid seals are divided into two main classes — static seals and dynamic seals Static seals are gaskets, O-ring joints, packed joints, welded joints, and similar devices used to seat static connections or openings with little or no relative motion between mating parts

A dynamic seal is any device used to restrict flow of fluid through an aperture closed by relatively moving surfaces Some dynamic seals also include static sealing elements in their design

Seals are also frequently classified as contact seals or clearance seals Some seal elements may operate as clearance seals under certain conditions and as contact seals under others The term seal may refer to a system rather than a single device A sealing system may require a mechanical seal, a viscoseal, and a labyrinth seal in order to produce the desired end result

Table 1 shows the dynamic seal elements which make up the bulk of industrial, com-mercial, utility, and transportation scaling applications Discussion follows of the selection and design factors involved with each of these types

POSITIVE RUBBING CONTACT SEALS

Mechanical Face Seal

The mechanical face seal, or end face seal, in Figure 1 is a device for sealing the annular space between a rotating shaft and a housing A rotary and a stationary face are forced towards rubbing contact by mechanical means and by fluid pressure acting on the rear of one of the sealing faces The two contacting faces are usually compatible materials capable

of operation with boundary lubrication Low coefficients of friction and high thermal con-ductivities are generally desirable

The face seal mating surfaces are lapped flat to within 0.5 to 1.5 µm (20 to 60 µin.) Excessive roughness causes high friction, accelerated wear, and short life However, ex-tremely smooth and flat surfaces lack the ability to generate hydrodynamic pressure in the fluid film which also leads to high friction, rapid wear, and short life

Stator and Rotor Arrangement

The stationary seal face or the rotating face may be flexibly mounted, usually with one

or more springs to keep the faces in contact This seal face is referred to as the seal head and the opposing ring as the seal seat The seal head may be internally or externally mounted (Figure 2) With pressure acting on the outer diameter of the internal seal (Figures 2a, 2e, and 2f), the seal rings are in compression This affords a wider choice of materials, many

of which are hard and brittle and should not be subjected to tensile stress

Rotary seal heads (Figures 2a, 2b, and 2c) are convenient to install on the shaft which generally is made from acceptable materials to acceptable tolerances The stuffing box housing requires minimal machining and the stationary seat can employ a wide range of designs and materials Dynamic balance of the rotating assembly is more difficult and rotary seal heads are usually employed at seal face speeds below 25 to 30 m/sec (5000 to 6000 ft/ min)

Stationary seal heads (Figures 2d, 2e, and 2f) avoid rotation of the spring assembly and are therefore often preferred for higher speeds Better tolerances and finishes are required

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584 CRC Handbook of Lubrication

FIGURE 4 Face loading devices.

may be used) As the primary seal face wears, the secondary sealing element in automatic and compression seals is pushed forward along the shaft In a bellows secondary seal, wear

is taken up by extension of the bellows made from molded elastomers, formed metal, or welded metal discs The pusher type seal is more susceptible to dirt, which may increase sliding friction, than the bellows type The bellows seal is pressure limited since high pressures may deflect the bellows radially enough to alter the effective bellows diameter Formed and welded metal bellows seals may employ loading springs in addition to the spring provided by the bellows

Seal Loading Devices

Lapped sealing faces are held in contact by sufficient preloading at assembly to keep the seal closed before hydraulic loading is developed, to withstand pressure reversals, and to overcome secondary seal friction Preload should be just sufficient to keep the seal closed

at the maximum expected axial excursion Unnecessary preload tends to increase face load and shorten seal life

The single preload spring, Figure 4a, has the advantage of simplicity and the relatively large wire cross section provides greater resistance to deterioration by corrosion The multiple spring seal, Figure 4b, requires less axial space, gives more uniform seal face load, has better resistance to centrifugal forces, can have face preload adjusted by using a different number of springs, and a large number of seal sizes can be fitted with the same springs Wave springs, finger springs, Belleville springs, slotted washers, and curved washers may

be employed in seals requiring minimum axial space These seals must be carefully designed and installed to obtain desired preload since they have a high-spring rate

The magnetic seal, Figure 4c, eliminates the need for springs and permits a compact design A disadvantage is the attraction of magnetic debris to the seal faces A bellows,

Figure 3f, may be used as a combined secondary seal and face loading device or in com-bination with a single spring

Seal Balance

Rate of energy dissipation between seal faces can be expressed as:

where Ff = seal face normal force, η = coefficient of friction, V = mean seal face velocity,

Pf = average seal face pressure, and Af = projected seal face area

The PfV term represents the energy dissipation per unit of projected seal face area for a unity friction coefficient For effective sealing and an acceptable wear rate, design factors which determine PfV must be controlled

Assuming zero discharge pressure, P1, the forces in Figure 5 tending to close the seal are hydrostatic force, Fp, arising from the pressure being sealed, and spring force, Fs, necessary

to maintain contact between the faces at start-up and shutdown Hydraulic force, Fo, which acts to separate the seal faces, is controlled by the characteristics of the interface flow process

586 CRC Handbook of Lubrication

FIGURE 7 Approximate envelope of manufacturers recommended operating

limits for inside seals (From Bernd, L H., Lubr Eng., 24(10), 479, 1968 With

permission.)

less than the unity, which allows contact pressure Pfand the energy dissipation to be reduced Note that if b equals 0.5 and one assumed a linear pressure profile, (b – β) becomes zero and only spring pressure keeps the seal faces closed The balance ratio chosen in practice

is usually in the range of 0.58 to 0.75

Ordinary Pressure-Temperature-Speed Limits

High-quality, general purpose mechanical face seals will meet a large majority of ordinary sealing requirements These involve sealing of clean, abrasive-free, safe, and only slightly corrosive fluids which provide good seal face lubrication under the following conditions:

1 Seal cavity pressure between 2.8 MPa (400 psi) and 1.3 Pa (0.01 torr)

2 Seal cavity temperature between 200°C (400°F) and −40°C (−40°F)

3 Seal face speeds less than 23 m/sec (4500 ft/min)

The approximate envelope of operating conditions for ordinary inside seals (Figure 1) is shown in Figure 7 Pressure limits for unbalanced external seals, Figure 2b, are about 20%

of those in Figure 7, while balanced external seals, Figure 2c, have pressure limits about 40% of those of Figure 7 Some inside balanced seals designed specifically for high pressure have been used at pressures in excess of 17 MPa (2500 psi) at shaft speeds of 23 to 33 m/ sec (4500 to 6500 ft/min) Stuffing box pressure limits for unbalanced seals have been set rather arbitrarily irrespective of service conditions at 0.7 to 1.4 MPa by some manufacturers, more conservatively by some users in Table 2

The upper PV limit for unbalanced seals is frequently taken to be 0.7 Mpa·m/sec (200,000 psi·ft/min) The limiting PV is useful in expressing the relative merit of various face material combinations Some definitions used for PV factor follow (refer to Figure 5):

PV = Pf · V = (ΔP (b – β) + PS)V (9) (from Equation 6)

588 CRC Handbook of Lubrication

FIGURE 8 Internal pumping ring.

Many seal manufacturers base their design and maximum PV recommendations on an es-timated life of about 15,000 hr

Temperature Extremes

Seals for temperatures above 200°C or below – 40°C often use metal bellows Elastomers become unserviceable much beyond these limits The usual temperature range for metal bellows seals is –240 to 650°C (– 400 to 1200°F), but Inconel X-750, Rene 41, or refractory alloys have been suggested for temperatures to about 1100°C (2000°F).6

Temperature control, either cooling or heating, can be obtained by bypass or circulating ring flushing, a water or steam jacketed stuffing box, or a quenching connection in the gland plate as shown in Figure 1 Clean fluid from the pump discharge can be cooled, injected to cool seal parts, and then directed through the restricted stuffing box throat back to pump suction An alternate arrangement uses a small pumping ring to circulate a small quantity

of clean fluid through a small external heat exchanger (Figure 8).7Occasionally, the material being pumped solidifies at ambient temperature In such cases, the seal region must be heated, for example, by using a steam-heated gland

Pressure Extremes

Pressures greater than about 2.8 MPa (400 psi) may distort seal faces and other compo-nents Conversely, high vacuum causes elastomers to outgas and destroy the vacuum The outgas problem can be solved by using metal bellows seals High pressures require seal balance and careful design of ring geometries Cross-sectional twisting especially must be reduced, and the gland plate must ensure flatness and accurate alignment Some manufac-turers insist that the gland plate be provided as part of the high-pressure seal assembly Seal cavity pressure can be borne by two seals in tandem to accommodate high system pressures

A clean process fluid stream or buffer fluid at pressure Pb is circulated through the outer seal chamber, usually set so (Psys – Pb) is approximately the same as (Pb – Patm) The outer seal is considered a backup in the event the inner seal fails

High Speed

At seal speeds over 23 m/sec (4500 ft/min), the mechanical seal requires matched springs and careful assembly to avoid unbalance of the rotating assembly At speeds over 33 m/sec (6500 ft/min), the seal head is usually made stationary and special designs are used for speeds up to 64 m/sec (12,500 ft/min) and above

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FIGURE 9 Internal-extemal seal Pb> Psfor double seal; Pb< Psfor tandem seal.

Abrasive, Corrosive, and Hazardous Fluids

Design for abrasive and corrosive fluids can follow either of two avenues: (1) fabrication

of seal components from exotic abrasion and corrosion resistant materials, or (2) creation

of a compatible environment to isolate the seal Hazardous fluids may not be hostile to seal components, but safety considerations usually dictate seal environment control

Abrasives in the sealed process fluid may be due to (1) the inherent nature of slurries, or liquids containing foreign matter such as sand, dirt, or oxides, (2) crystalline particles which result from evaporation or from contact with atmosphere, and (3) crystalline particles which result from heating or cooling A clean, process-compatible liquid is injected through the flush connection to cool and isolate seal parts from abrasive particles in the process fluid The flush liquid flows through a close-clearance bushing at the throat back to the process Amount of liquid injected can be controlled by the supply pressure and the restriction: a plain bushing, a floating bushing, or a lip type bushing The clean injected fluid may be pump discharge from which abrasive particles have been removed by centrifugal separation

or a settling tank

For process fluids which crystallize upon contact with air, an auxiliary connection can be used to inject low-pressure water to wash away the seal leakage and prevent abrasives from forming at seal faces Such an arrangement can also be used to dilute and drain away dangerous fluid leakage Where dilution of the process fluid by flow through the throat bushing cannot be permitted, a double seal is employed with clean buffer fluid circulated

by auxiliary means between the two seal elements to provide an almost complete isolation from the process fluid Where seal housing space is limited, an internal-external multiple seal may be arranged as in Figure 9 With the buffer fluid pressure Pb greater than process pressure Ps, both seals are lubricated by the buffer field The arrangement can serve as a tandem seal when Pb< Ps

Materials

Seal components and gland ring parts for noncorrosive fluids such as gasoline, hydro-carbons, and oils are usually made from ferritic stainless steels such as 502 or 430 For moderate corrosion resistance in environments such as water, sea water, dilute acids, fatty acids and alkalis, austenitic stainless steels such as 302, 304, and 316 are widely used For highly corrosive environments such as strong mineral acids and strong alkalis, nickel-copper base materials such as Monel or nickel-molybdenum alloys such as Hastelloy B or Hastelloy

C are frequently employed Temperature range for these materials is – 100 to 400°C (– 150

to 750°F) Table 3 presents seal face material combinations for various environments Tables

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RECOMMENDED TEMPERATURE LIMITS FOR

Table 5 RECOMMENDED TEMPERATURE LIMITS FOR

a Product temperature.

From Guide to Modern Mechanical Sealing, Durametallic Corporation, Kalamazoo,

Mich., 1971 With permission.

4, 5, and 6 show recommended temperature limits for seal faces, secondary seal materials, and springs

Ring Seals

Split or segmented rings of metallic or nonmetallic material are used as piston ring, rod, and circumferential seals, Figure 10 The piston ring (expanding ring) and rod seals (con-tracting ring) are used principally in reciprocating applications, circumferential seals mainly

a Product temperature; maximum working tem-perature is higher.

b Subjecl to thermal shock fracture.

From Guide to Modern Mechanical Sealing,

Dura-metallic Corporation Kalamazoo, Mich., 1971.

With permission.

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as rotary seals A single-stage ring seal may employ one, two, or three split or segmented rings in a single groove or housing for rubbing contact with either the shaft or bore The pressure of the sealed medium forces the ring into axial and radial contact and the initial or static contact load is provided by the elastic properties of the ring (expanding or contracting seals) and/or by auxiliary springs (circumferential seals) Ring seals have three potential leakage paths: (1) between ring and bore or ring and shaft, (2) between ring and side wall

of groove or housing, and (3) the ring gap

Contact loads and drag forces caused by fluid pressure on a ring are depicted in Figures 11a and 11b These loads increase with increasing pressure with an accompanying increase

in wear High pressure may also prevent the ring from following dynamic excursions of the shaft, piston, or rod Improved dynamic response and wear reduction in one-directional seal rings can be obtained by pressure relief grooves (Figure 11c) for a “balanced” ring The grooving tends to increase leakage by reducing the leak path length across the seal dam Good design requires a compromise between wear, factional heating, and leakage, Ring

FIGURE 11 Pressure-induced forces on a seal ring (a) Drag forces on rod seal ring; a = face dimension;

b = wall dimension; DR = radial drag = nRFA; DA = axial drag = nAFR; n = coefficient of friction (b) Contact forces on ring; FA = axial unbalance; FR = radial unbalance (c) Contact forces on pressure-relieved ring.

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and sealing dam dimensions are usually selected such that axial and radial forces are about the same Auxiliary springs are required in low-pressure applications to maintain contact in both axial and radial directions

Split Ring Seals

Expanding split rings (piston rings) and contracting split rings (rod seals) are used as piston head, rod, rotary, butterfly valve, and static seals to control leakage of hot combustion gases or fluids The most common expanding ring application is to seal between the recip-rocating piston and cylinder wall in internal combustion engines and reciprecip-rocating com-pressors The contracting ring seal is used in hydraulic cylinders where pressure, high-temperatures, thermal fatigue, and reliability requirements make elastomeric packings undesirable

Split rings may be used singly or in series A second step joint ring will reduce leakage

by approximately 15% Although a third ring will provide little additional leakage improve-ment, it may extend the overhaul period by coming into operation when the first and second rings are worn

Expanding rings are manufactured with free-ring dimensions to produce uniform radial pressures of about 70 to 550 kPa (10 to 80 psi) when installed Auxiliary springs are normally unnecessary The contracting ring, however, can provide only limited tension and frequently employs auxiliary springs to insure conformation to the rod surface as shown in Figure 12 Expanding split rings (Figure 13) are frequently used as rotary seals on hydraulic transmis-sions and clutches, torque converters, hydrostatic transmistransmis-sions, crackcase seals on large engines, and turbosuperchargers Step seal rings have a face dimension greater than the wall dimension Relative motion and wear take place at the side contact area; this prevents wear grooves in the housing which would prevent removal of the shaft from the housing during overhaul

Lubrication is necessary when using metallic ring seals In oxygen compressors, oil-free air compressors, food processing plants, and certain chemical processes where lubricants cannot be tolerated, nonmetallic materials may be employed for seal rings Table 7 gives recommended temperature limits for commonly used split ring materials Typical perform-ance ranges of split ring seals are shown in Table 8

Circumferential Seals

Adaptation of split and segmented rings to prevent leakage of high-temperature air and combustion gases into the bearing cavities of aircraft gas turbine engines led to development

of high-performance, highspeed, elevated temperature circumferential seals, Figure 10c The basic arrangement and balance considerations in Figures 10 and 11 also apply to circumferential seals.12

FIGURE 12 Two-piece rod seal.