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. Volume II 579 565-579 4/10/06 5:07 PM Page 579 Copyright © 1983 CRC Press LLC 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. Adynamic 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. Asealing 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 CONTACTSEALS 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. Arotary 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 Volume II 581 581-622 4/10/06 6:02 PM Page 581 Copyright © 1983 CRC Press LLC 584CRC 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. Adisadvantage is the attraction of magnetic debris to the seal faces. Abellows, 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: E =ηF f V =ηP f A f V(1) where F f = seal face normal force, η = coefficient of friction, V = mean seal face velocity, P f = average seal face pressure, and A f = projected seal face area. The P f Vterm 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 P f Vmust be controlled. Assuming zero discharge pressure, P 1 , the forces in Figure 5 tending to close the seal are hydrostatic force, F p , arising from the pressure being sealed, and spring force, F s , necessary to maintain contact between the faces at start-up and shutdown. Hydraulic force, F o , which acts to separate the seal faces, is controlled by the characteristics of the interface flow process 581-622 4/10/06 6:02 PM Page 584 Copyright © 1983 CRC Press LLC 586CRC 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 P f and 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 PVlimit for unbalanced seals is frequently taken to be 0.7 Mpa·m/sec. (200,000 psi·ft/min). The limiting PVis useful in expressing the relative merit of various face material combinations. Some definitions used for PVfactor follow (refer to Figure 5): PV = (P 2 – P 1 )V = ΔP · V (7) PV = ΔP · b · V (8) (Reference 3) PV = P f · V = (ΔP (b – β ) + P S )V (9) (from Equation 6) 581-622 4/10/06 6:02 PM Page 586 Copyright © 1983 CRC Press LLC 588CRC Handbook of Lubrication FIGURE 8. Internal pumping ring. Many seal manufacturers base their design and maximum PVrecommendations 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). 7 Occasionally, 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 P b is circulated through the outer seal chamber, usually set so (P sys – P b ) is approximately the same as (P b – P atm ). 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. 581-622 4/10/06 6:02 PM Page 588 Copyright © 1983 CRC Press LLC Volume II589 FIGURE 9. Internal-extemal seal. P b > P s for double seal; P b < P s for 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. Aclean, 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 P b greater than process pressure P s , both seals are lubricated by the buffer field. The arrangement can serve as a tandem seal when P b < P s . 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 581-622 4/10/06 6:02 PM Page 589 Copyright © 1983 CRC Press LLC Volume II591 Table 4 RECOMMENDED TEMPERATURE LIMITS FOR FACE SEALMATERIALS a Table 5 RECOMMENDED TEMPERATURE LIMITS FOR SECONDARYSEALMATERIALS a 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. 581-622 4/10/06 6:02 PM Page 591 Copyright © 1983 CRC Press LLC Volume II 593 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; D R = radial drag = n R F A ; D A = axial drag = n A F R ; n = coefficient of friction. (b) Contact forces on ring; F A = axial unbalance; F R = radial unbalance. (c) Contact forces on pressure-relieved ring. 581-622 4/10/06 6:02 PM Page 593 Copyright © 1983 CRC Press LLC 594CRC Handbook of Lubrication 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 reciprocating com- pressors. The contracting ring seal is used in hydraulic cylinders where high-pressure, high- temperatures, thermal fatigue, and reliability requirements make elastomeric packings undesirable. Split rings may be used singly or in series. Asecond 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 transmissions, 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. 581-622 4/10/06 6:02 PM Page 594 Copyright © 1983 CRC Press LLC [...]... according Copyright © 1983 CRC Press LLC 581- 622 4 /10/ 06 600 6: 02 PM Page 600 CRC Handbook of Lubrication Copyright © 1983 CRC Press LLC 581- 622 4 /10/ 06 6: 02 PM FIGURE 17 Radial lip seal design variations (a) Springless, single-lip seal — economical; used to retain highly viscous materials or to exclude dust in lower speed applications, e.g., 10 .2 m/sec (20 00 ft/min) or lower When used as dirt excluder,... wide with 0.8 µm ( 32 µin.) finish [B] Maximum inside corner radius of 1 .2 mm (3/64 in.) Bore depth equals seal width plus 0.4 mm (1/64 in.) Bore finish of 2. 54 µm (100 µin.) or better Chamfer 15 to 30°, 1.75 to 2. 25 mm (0.060 to 0.090 in.) wide [C] Interference varies from 1 mm (0.040 in.) for small shaft diameters, 25 mm (1 in.), to about 3 mm (0. 120 in.) for shaft diameters of 127 mm (5 in.) or larger... Frictional drag force of approximately 12 N (2. 7 1b) without cooling would generate wear track temperatures of 550 to 650°C (1 020 to 120 0°F) To avoid severe damage to the carbongraphite seals, oil cooling jets at 9 to 12 m/sec (30 to 40 ft/sec) are directed on the runner near the wear track or under a cantilevered runner. 12 Operation in ambient temperatures of 540°C (100 0°F) has been reported .10 Friction and... characteristics of a pressure-relieved three-ring segmented circumferential seal are about as follows: diameter 165 mm (6.5 in.), speed 61 m/ sec (20 0 ft/sec), pressure differential between 316°C (600°F) air and 121 °C (25 0°F) oil, 448 .2 kPa (65 psi), wear life in excess of 600 hr with leakage of 3.5 × 10 4 m3/sec (0.75 standard ft3/min), and radial unbalance force of 306 N/m (1.75 lb/in. )of seal circumference... Over 20 0 lip seal styles have been developed for an enormous variety of applications involving shaft sizes from 5 to 1 525 mm (0 .20 to 60 in.) Retained fluids are commonly lubricating oils or liquids having some lubricating qualities, sump temperatures vary from – 60 to 20 0°C (– 76 to 390°F), and peripheral speeds range up to 20 m/sec (4000 ft/min) Sealed pressures are moderate, in the range of 20 to 100 ... of 20 to 100 kPa (2. 9 to 14.5 psi), but special seal designs have been used up to 3450 kPa at 15 m/sec (500 psi at 3000 ft/min).14 Copyright © 1983 CRC Press LLC 581- 622 4 /10/ 06 6: 02 PM Page 599 Volume II 599 FIGURE 16 Typical bonded, dual lip elastomeric lip seal [A] Alloy or stainless steel hardened to a minimum Rockwell C-30, C-45 recommended, finished to 0 .25 to 0.50 µm (10 to 20 µin.) Finish must...581- 622 4 /10/ 06 6: 02 PM Page 597 Volume II 597 RMS, minimum shaft or runner hardness 55 Rockwell C, and shaft radial runout 25 µm (0.001 in.) TIR or less While greater runout can be tolerated by increasing radial spring load, this will cause greater wear Seal housing side wall surfaces should have similar roughness and hardness values as the shaft Flatness of the seal secondary surface... a slight buffer gas overpressure of about 27 .6 kPa (4 psi) between two seal elements Gas leakage keeps the ring segments free of liquid Leakage for circumferential seals is generally about an order of magnitude less than with labyrinth seals Wear life (which depends on speed, pressure differential, temperature, material and design) ranges from about 100 hr to over 10, 000 hr To avoid shaft wear, most... Friction and temperature rise can be minimized through use of pressure-relieved seals, Figure 15 Commonly used axial dam width, E, is about 1 to 1 .25 mm (0.039 to 0.049 in.), which produces a radial force of 0.50 to 0.63 N/m/kPa (0. 020 to 0. 025 lb/in./psi) A narrow axial sealing dam reduces frictional heating, but the wear rate and fragile nature of normal seal rings must be considered Radial dam width,... usually in range of 20 to 35° [E] Oil-side angle, generally between 40 to 70° [F] Optional inner case is sometimes used for additional strength for pressfits and to protect seal lip per psi ΔP).5 Approximately 80% of high quality, carefully installed lip seals, fabricated from suitable elastomers, will leak about 0.0 02 g/hr or about 1 drop per 8-hr shift in continuous operation About 20 % of such seals . pressfits and to protect seal lip. 581- 622 4 /10/ 06 6: 02 PM Page 599 Copyright © 1983 CRC Press LLC 600 CRC Handbook of Lubrication 581- 622 4 /10/ 06 6: 02 PM Page 600 Copyright © 1983 CRC Press. the characteristics of the interface flow process 581- 622 4 /10/ 06 6: 02 PM Page 584 Copyright © 1983 CRC Press LLC 586CRC Handbook of Lubrication FIGURE 7.Approximate envelope of manufacturers recommended. (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 Volume II 581 581- 622 4 /10/ 06 6: 02 PM