Improving Machinery Reliability 3 Episode 13 docx

Standard Seal for Flashing Hydrocarbon Services Flashing hydrocarbon service includes all hydrocarbon services in which the fluid has a vapor pressure greater than 14.7 psia 0.101 MPa a

514 Improving Machinery Reliability

Figure 12-14 Viscosity stability, bowl mill

had recently been reworked and tested All three bowl mills were fed the same amount of coal during the test period All three gear oils were the same IS0 320 vis- cosity grade

The average current draws were:

Synthetic hydrocarbon gear oil 68

The lower amp difference shown by the synthetic hydrocarbon is the result of the lower coefficient of friction shown in Table 12-2

In summary, the synthetic hydrocarbon gear oil has solved the original problems and provided additional benefits not anticipated The switch to synthetic lubricants has clearly improved performance and achieved significant savings in operating costs, as shown in the following tabulation

Table 12-2 Physical Properties of IS0 VG 320 Gear Oil

~~

Petroleum

Synthetic hydrocarbon

Thermal conductivity,

Lubrication and Reliability 515

The extended drain interval provides savings in three areas:

1 Lubricant consumption cost savings:

Petroleum oil cost per gal $4.00

Petroleum oil changes per yr 2

Volume of gear box, gal 300

Petroleum oil cost per yr

($4.00/ga1)(2 changes/yr)(300 gal/unit) = $2,400

Synthetic oil cost per gal $16.00

0.2 Volume of gear box, gal 300

Synthetic oil cost per yr

($16.00/ga1)(0.2 changes/yr)(300 gal/yr) = $960

Annual savings on lubricant cost-$1,440 per unit

2 Reduced maintenance cost savings

2

Maintenance cost per change $500

Petroleum oil maintenance cost per yr

Synthetic oil changes per yr

Petroleum oil changes per yr

(2 changes/yr)($500/change) = $1,000

Synthetic oil changes per yr 0.2

Maintenance cost per change $500

Synthetic oil maintenance cost per yr

Annual savings in scheduled maintenance costs-$900

Petroleum oil used per yr, gal 600

Disposal cost per gal $0.50

Synthetic oil used per yr, gal 60

Disposal cost per gal $0.50

Annual savings in disposal cost per year-$270

3 Lubricant disposal costs

The reduction in energy consumption also provides significant savings:

Average annual power cost using petroleum oil lubricant

Average annual power cost using synthetic hydrocarbon lubricant

Annual savings in power consumption-$2,067

$33,278

$31,211

The total annual savings for all of the above categories amount to $4,677 In addi- tion, savings in reduced wear and thus fewer repairs are certain to be realized

Vibration Performance Improved With Synthetics

Rolling element bearings can experience significant reductions in vibration ampli- tude and, thus, increased life expectancy when lubricated with synthetic oils The

516 Improving Machinery Reliability

higher film strengths of the synthetic oils reduce the severity of impact when the rolling elements of a bearing move across spa11 marks and other discontinuities In fact, many defective bearings have been “nursed along” by the high-strength bond- ing characteristics of these films

A major gas-transmission company documented the vibration-shock pulse activity

of a compressor turbocharger and oil supply pump before and after switching to a syn- thetic oil After the changeover, a strong, tenacious and slippery oil film reduced vibra- tion severity by “peening over” asperities on the various metal surfaces of bearings

A variety of equipment has been rejuvenated by switching to synthetics:

An external washer-filter in a Willamette, Ind., pulp mill had been operating well above 1.5 G for shock pulse activity and 0.2 in./s for vibration Upon switching to

a synthetic oil, the values dropped to 0.75 G for shock pulse activity and 0.15 in./s for vibration The change occurred immediately and continued for a month

A multistage air blower at a fiber spinning plant had its vibration decreased from 0.155 in./s to 0.083 i n h by switching oils Further, the temperature of the bearing dropped by 20°F

A IO-hp centrifugal pump had an acceptable vibration of 0.068 in./s, but the bear- ing housing temperature of 175°F was borderline After changing to a synthetic oil, the vibration dropped to 0.053 in./s, and bearing housing temperature went down

to 155°F Further, the motor amperage was cut from 5.7 Nphase to 4.4 Nphase

To obtain such improvements, however, it is important to choose a well-formulat-

ed synthetic oil As typified in Figure 12-15, there are noticeable differences in the operating temperatures of spur gear units, reactor pump bearings and bevel gear enclosures using products from different vendors

Further, while there are many excellent products on the market today, many may not

be appropriate for use in process machine applications For example, high-film- strength oils based on extreme pressure (EP) technology and intended for gear lubrica- tion typically incorporate additives containing sulfur, phosphorus and chlorine These

EP industrial oils cannot be used as bearing lubricants for pumps, air compressors, steam turbines, high-speed gears and similar machinery, since the sulfur, phosphorus and chlorine will cause corrosion at high temperatures and in moist environments

Testing Provides Proof

The best indication of which oil-synthetic or mineral-will excel in an applica- tion can be obtained by comparing their specific performance There are numerous laboratory tests that are good indicators of how well an oil will perform in service These includes tests for viscosity, pour point, residue, strip corrosion, rust, demulsi- bility and so on

On the other hand, it should be pointed out that these tests are only predictors Realistic tests under simulated field conditions are better, while the true measure of a lubricant’s performance can only be determined in actual service For example, in

1992, Kingsbury, Inc., completed the testing of a well-compounded I S 0 Grade 32 synthetic lubricant in a thrust-bearing test machine

Lubrication and Reliability 517

- Synthetic oil “A” - Synthetic oil “B”

Figure 2 Not all synthetic oils are equal: Here oil B is superior to A

Time (days)

I I I I ’ I I , , I

Figure 12-15 Not all synthetic oils are equal Here oil B is superior to A

At low speeds and loads, there appeared to be little difference between this lubri- cant and identical premium-grade mineral oils of the same viscosity However, at high loads and speeds, above 550 psi and 10,000 rpm, Kingsbury found that the syn- thetic oil cut bearing temperature by 15”F, and decreased frictional losses by 10%

Similarly, engineers from SKF and Exxon conducted a series of tests on rolling contact bearings.13 The objective was to compare the properties of a specially formu- lated diester lubricant with those of a premium-grade mineral oil that was in service

in an Exxon petrochemical plant

Two synthetic lubricants and two mineral oils of varying viscosities were experi- mentally compared The test results indicated that the synthetic lubricant, having a viscosity of 32 centistokes (cSt) at a temperature of 100”F, offered long-term surface

protection equivalent to that of the base line mineral oil with a viscosity of 68 cSt,

without reducing bearing surface life below the theoretically estimated levels The same good wear protection could not be achieved with a reduced viscosity mineral oil The use of the lower viscosity synthetic lubricating fluid could provide projected energy savings of $140,000 per year-prorated to 1998 dollars, when the

test was conducted

Other companies, too, have been operating successfully with synthetic oils for many years, and are reluctant to publicize their experience so as not to give away a competitive advantage Suffice it to say that a forward-looking process plant needs

to explore the many opportunities for often substantial cost savings that can be achieved by judiciously applying properly formulated synthetic lubricants

Automatic Grease Lubrication as a Reliability Improvement Strategy

Experienced maintenance and reliability professionals have seen rapid progress from reactive maintenance through preventive, predictive and “hybrid” approaches toward today’s proactive maintenance methods Recognizing that there are still some cost reduction measures that could-and should-be implemented, a number of

“Best-of-Class” companies are now focusing on maintenance as part of strategy and

518 Improving Machinery Reliability

profit potential This realization has led these companies to fundamentally reassess their lubrication options for both new as well as existing machinery

Lubrication Should Not Be An Afterthought

The most advanced, truly bottomline-oriented owners/purchasers of high-speed paper producing machinery are among a growing number of buyers that base their specification and procurement decisions on life cycle cost calculations These calcu- lations have shown, in the majority of cases, the long-range maintenance and down- time cost avoidance advantage of incorporating automatic oiling or grease feed pro- visions in modern process machines Best-of-class design contractors and owner companies insist on automatic lubrication to be part of the design specification and permanent plant operating strategy

This new thinking supersedes the old notion that lubrication automation is diffi- cult to cost-justify, or that automatic lubrication provisions can be retrofitted as needed! In a highly cost-competitive environment or at a time when profit margins can vanish because of a single, unplanned outage event, buying equipment with proven lubrication provisions makes eminent sense Experience shows that procure- ment of suitable automated lubricant delivery units at the inception of a project may well be the only low-risk opportunity that presents itself to buyers who take reliabili-

ty seriously Buying critically important machinery with the thought that upgrading

to full automation at a later date will always be an option could be fallacious and may prove to be a costly mistake

Bearing Manufacturers Prefer Automatic Lube Option

The disadvantages of manual lubrication have long been recognized by the lead- ing bearing manufacturers As can be seen from Table 12-3, the service life of rolling element bearings with automatic grease feed provisions ranks well ahead of most other means of lubricant delivery This is why many European process plants much prefer engineered automatic lube application systems over traditional static oil sumps or manual regreasing These plants are often especially mindful of the short- comings of single-point automatic lubricators Although occasionally found in the United States, these spring or gas-pressurized plastic grease containers offer little control over grease quantity and grease homogeneity Greases under constant pres- sure tend to separate into their respective oil and soap constituents This is highly undesirable since the virtually all-soap matrix is now likely to enter the bearing with- out oil A properly engineered grease injection system will provide near-instanta- neous pressure pulses with adjustable rest periods interspersed to suit the require- ments of a specific application

Although static oil sumps are still used in the majority of centrifugal pumps

installed in U.S paper mills, petrochemical facilities, and general process plants, the

experience of Scandinavian pump users should prompt a fundamental reassessment and serious questioning of the role of contamination-prone static oil sumps in an era

of reduced maintenance manpower availability Moves away from “sump, occasion-

Lubrication and Reliability 519

Regular grease replenishment Sump, occasional

renewal

*By feed cones, bevel wheels, asymmetric rolling bearings

**Condition: Lubricant service life < fatigue life

a1 renewal” to “automatic grease feed” as ranked in Table 12-3 have resulted in relia-

bility increases that can no longer be overlooked This readjusted thinking has paid off handsomely in a number of Finnish and Swedish paper mills by allowing man- power reductions and resulting in plant availability extensions

Comparing Manual and Automatic Grease Lubrication Provisions

Three principal disadvantages of manual lubrication are generally cited:

e Long relubrication intervals allow dirt and moisture to penetrate the bearing seals Well over 50% of all bearings experience significantly reduced service life as a result of contamination

a Overlubrication occurring during grease replenishment causes excessive friction and short-term excessive temperatures These temperature excursions cause oxida- tion of the oil portion of the grease

e Underlubrication occurring as the previously applied lube charge is being depleted

at this time and prior to the next regreasing event

In contrast, automated lubrication has significant technical advantages Time and again, statistics compiled by SKF (Figure 12-16) and other major bearing manufac- turers have shown lubrication-related distress responsible for at least 50%, and per-

haps as much as 70% of all bearing failure events worldwide Thoroughly well-engi- neered automatic lubrication systems, applying either oil or grease, are now

520 Improving Machinery Reliability

Figure 12-16 Bearing failures and their causes (Courtesy of SKF USA, Inc also TAPPI

7995 Engineering Conference Proceedings.)

available to forward-looking, bottomline-oriented user companies These systems

(Figure 12- 17) ensure that:

The time elapse between relubrication events is optimized

Accurately predetermined, metered amounts of lubricant enter the bearing “on

The integrity of bearing seals is safeguarded

Supervisory instrumentation and associated means of monitoring are available at

the point of lubrication for critical bearings

time” and displace contaminants

How Automated Lubrication Works

Depending on the type of installation, engineered lubricant injection systems are

configured for either grease or liquid oil delivery Modular in design and easily

expandable, they are suitable for machinery with just a few lubrication points, as

well as installations covering complete manufacturing or process plants involving

thousands of points Automated grease lubrication systems are designed for the peri-

odic lubrication of rolling element bearings, as in the centrifugal pump depicted in

Figure 12-18, or for different types of sleeve bearing Also, automated grease lubri-

cation systems are used on guides (shown in the soot blower in Figure 12-19) and on

open gears, chains, and coupling devices

Depending on plant and equipment configuration, engineered automatic lubrica-

tion systems consist of a single or multi-channel control center (Figure 12-17, Item

I), one or more pumping stations (Item 2), appropriate supply lines (Item 3) tubing

Lubrication and Reliability 521

Figure 12-17 Single-header, state-of-the-art automatic lubrication system (Courtesy of Safematic, Muurame, Finland and Alpharetta, GA.)

Figure 12-18 Centrifugal pump with automated grease lubrication (Courtesy of Safe-

matic, Muurame, Finland and Alpharetta, GA.)

522 improving Machinery Reliability

Figure 12-19 Soot blower assembly fitted with automatic grease lubrication provisions

(Courtesy of Safematic, Muurarne, Finland and Alpharetta, GA.)

(4) which links a remote shutoff valve ( 5 ) and lubrication dosing modules (6), and

also interconnects dosing modules and points to be lubricated Different sized dosing

modules are used to optimally serve bearings of varying configurations and dimen-

sions The dosing modules themselves are individually adjustable to provide an exact

amount of lubricant and to thus avoid overlubrication A pressure sensing switch (Item 7), completes the system

The control center starts up a pump that feeds lubricant from the barrel through the main supply line to the dosing modules When pressure in the system rises to a preset level, the pressure switch near the end of the line transmits an impulse to the control center, which then stops the pump and depressurizes the pipeline The con- trol center now begins measuring the new pumping interval If for some reason the pressure during pumping does not rise to the preset level at the pressure switch, an alarm is activated and the lubrication center will not operate until the problem has been rectified and the alarm subsequently reset

Special multi-channel controllers are available with state-of-the-art automatic lubrication systems These have the ability to provide lubrication to installations requiring a variety of lube types or consistencies Even different timing intervals can

be controlled from a single multi-channel controller location These systems have proven their functional and mechanical dependability in operating environments ranging from minus 35°C to plus 150°C One Finnish manufacturer tests every type

of grease supplied by usedclient companies under these temperature extremes and leaves no reliability-related issues open for questioning

Lubrication and Reliability 523

Cost Studies Prove Favorable Economics of Automated

Lubrication Systems

A Finnish paper mill, Enso Oy, has documented the production increases, labor savings, and downtime reductions shown in Figure 12-20 Downtime hours for a total of 3 1 process units encompassing over 7,500 lubrication points are illustrated in Figure 12-21 Here, the Kaukopaa mill documented the 1 1-year trend from 9,700 hours of downtime in 1985 to approximately 280 hours in 1995 In the same time period, production went from 620,000 tons (1985) to 950,000 tons (1995) Figure 12-22 shows how, from 1990 until 1995, total maintenance expenditures decreased 26%, and maintenance costs per unit of production were reduced by 46% Needless

to say, Enso Oy has realized millions of dollars in extra profits from the timely intro- duction of engineered automatic lube systems They have included these systems in the mandatory scope of every new project and Enso Oy’s mill standard (“EGO) defining automatic lubrication systems has been adopted as a National Industrial Standard in Finland, a “high-tech” country in every sense of the word

Washers, agitators, pumps, electric motors, soot blowers, barking drums, chippers, screens, presses, conveyors, and other equipment are automatically lubricated at a

modern facility A machine, which without lube automation often experiences five

lubrication-related bearing failures per year, is likely to experience none with an

engineered grease injection system This often translates into 30.40 hours of addi-

tional machine time and profit gains of $90,000-$180.000 annually

Figure 12-20 Production, lubrication labor and maintenance downtime statistics

(Courtesy of Safematic, Muurame, Finland and Alphareffa, GA.)

524 Improving Machinery Reliability

I

_ _ _ _ I _

Enso Oy - Kaukopfitl Mlll Maintenance Oown tlme M Pmductlon

Figure 12-21 Maintenance downtime vs production (Courtesy of Safematic, Muu-

rame, Finland and Alphamtta, GA.)

-1

._ -

Enso Oy - Kaukop4i Mlll

MIlnteMnce C m b oer Production

Figure 12-22 Maintenance costs referred to production (Courtesy of Safemafic, Muu-

rame, Finland and Alphretta, GA.)

Lubrication and Reliability 525

Payback for these systems, both originally supplied as well as retrofitted, typically ranges in the six-months to three-year timeframe This might be one of the explana- tions why Scandinavian paper producers, whose workers have higher incomes than most of their American counterparts, are profitable and able to compete in the world markets Automated lubrication has consistently yielded increased plant uptimes ranging from 0.1 % to 0.5%

What Lessons Can Be Drawn from All This

Engineered automated lubrication systems are no longer a luxury Instead, they are one of the key ingredients of any well-thought-out maintenance cost reduction and failure avoidance strategy Procurement specifications for a wide range of machinery used in process plants, chemical manufacturing plants, petroleum refiner- ies, and paper mills deserve to be updated wherever life cycle cost studies demon- strate the reliability and general maintenance advantages of this mature lube applica- tion method Retrofitting of engineered automatic lubrication systems, although somewhat more costly to implement than initial installations, is nevertheless justified

in the numerous critically important machines that operate in adverse temperature and moisture environments Since the bearing failure history of these machines is generally painfully evident, cost justifications are easy to develop with often remark- able accuracy More importantly, these studies may point the way toward greater profitability

References

1 Bloch, H P., “Criteria for Water Removal from Mechanical Drive Steam Tur- bine Lube Oils,” ASLE Paper No 80-A-1E-1, Presented at the 35th Annual Meeting in Anaheim, California, May 5-8, 1980

2 MacDonald, J W., “Marine Turbine Oil System Maintenance,” Lubrication

Engineering, Volume 21, No 10, 1965

3 ROC Carbon Company, Houston, Texas 77224, Technical Bulletin

4 Wilson, A C M., “Problems Encountered with Turbine Lubricants and Associ-

ated Systems,” Lubrication Engineering, Volume 32, No 2, 1976

5 Wilson, A C M., “Corrosion of Tin Base Babbitt Bearings in Marine Steam Turbines,” Transactions of Institute of Marine Engineering, Volume 73, No 11,

196 1 (discussion)

6 Appeldoorn, J K., Goldman, I B and Tao, F F., “Corrosive Wear by Atmos-

pheric Oxygen and Moisture,” ASLE Transactions, Volume 12, No 140, 1969

7 Schatzberg, P., “Influence of Water and Oxygen in Lubricants on Sliding Wear,”

Lubrication Engineering, Volume 26, No 9, 1970

8 Schatzberg, P and Felsen, I M., “Effects of Water and Oxygen During Rolling-

Contact Lubrication,” Wear, Volume 12, 1968, pp 33 1-342

9 Bloch, H P and Geitner, F K., Machinery Failure Analysis and Troubleshoot-

10 May, C H., “Separation of Water from Oil by the Principle of Coalescence,”

Lubrication Engineering, Volume 19, No 8, 1963

526 Improving Machinery Reliabiliry

1 1 Allen, J L., “Evaluating a Waste-Oil Reclamation System,” Plant Engineering,

April 29, 1976

12 Halliday, K R., Why, When and How to Use Synthetic Lubricants, Selco, Fort

Worth, Texas, 1977

13 Morrison, F R., Zielinski, J and James, R., Effects of Synthetic Industrial Fluids

Louisiana, Feb 1980

14 Eschmann, Hasbargen and Weigand, Ball and Roller Bearings, John Wiley &

Sons, Ltd., New York, 1985

Chapter I 3

Government agencies in the United States are imposing increasingly stricter limits

on emissions of Volatile Organic Compounds (VOCs) from rotating equipment These fluids change from the liquid to the vapor phase upon leaking to atmospheric pressure and thus become airborne pollutants Leakage of VOCs from valves, fit- tings, and pump seals is known as “fugitive emissions.”

Emission readings in the U.S are done per EPA Method 21.’ An organic vapor

analyzer draws in a continuous sample of air from the space within 1 cm of the seal

end plate and shaft interface VOC emissions are measured in parts per million

(ppmv) concentration in air Emission readings are affected by factors such as wind

speed, shaft rotation rate, and proximity of the probe to the emission source Since the measurement is not in mass units (e.g., gramshour), the EPA considers Method

21 to provide only an emissions screening value

The U.S Environmental Protection Agency (EPA) has established a national emissions limit of 1000 ppm for most pumps in organic hazardous air pollutant ser- vice.* In response, mechanical seal manufacturers have developed extremely low- emission, reliable seals

Nevertheless, there are additional benefits that can be attributed to these govern- ment-initiated moves Higher reliability means reduced product losses, fewer down- time events, and reduced maintenance expenditures These are issues that are equally important to modern process plants, and reliability professionals must be thoroughly acquainted with the topic Suffice it to say that the average refinery has 1,000 cen- trifugal pumps that undergo, on average, 300 seal replacements per year “Best-prac- tices” refineries replace as few as 150 seals, and the resulting savings easily exceed

$ I,OOO,OOO each year

This chapter* starts with a brief review of API Standard 682 specifications for seals in flashing hydrocarbon services Next, it gives an in-depth presentation of design guidelines for low-emission single seals This latter work is based on more

*Contributed by Bill Key, BW/IP International, Inc., Temecula, California

527

528 Improving Machinery Reliability

than ten years of development using mechanical seal testers and employing propane

as the operating fluid Low-emission single mechanical seals are shown to be an effective means to comply with government emission limits

It also reviews dual seal options to control emissions to very low levels Included are liquidlliquid dual seals and their auxiliary equipment, dry gas secondary seals, and dual gas seals using pressurized barrier gas

API Standard 682

API Standard 6823 covers the minimum requirements for sealing systems for rotary and centrifugal pumps in refinery services The objective of the API 682 Committee is reflected in their Mission Statement, “TO produce a reliable sealing system that has a high probability of operating 3 years of uninterrupted service, meeting or exceeding environmental emission regulations.” The API 682 document

3 covers seal standards for most refinery applications This discussion will only pre-

sent highlights of the standards for seals in flashing hydrocarbon services Seals designed to these standards should perform well in most other volatile fluid services

Standard Seal for Flashing Hydrocarbon Services

Flashing hydrocarbon service includes all hydrocarbon services in which the fluid

has a vapor pressure greater than 14.7 psia (0.101 MPa) at pumping temperature The API 682 standard seal for flashing hydrocarbon services is a pusher seal with special features to maintain adequate vapor suppression Other specifications include:

Inside mounted cartridge seal

Balanced seal

Flexible element rotates

This seal is specified as a type “A” seal (Figure 13-1) When requested, the stan-

dard alternate to the type A seal is a totally engineered sealing system with an engi- neered metal bellows (type “B” seal) A standard single seal (one rotating face per

seal chamber) is classified as Arrangement 1

The operating window for type A seals in flashing hydrocarbon service is:

Temperatures from -40°F to +500”F (-40°C to +26OoC)

Pressures to 500 psi (34.5 bar)

Seal face surface speed < 5,000 ft/min (25 rnetedsec)

A totally engineered sealing system is required for operation outside the above parameters

Providing Safety and Reliability Through Modern Sealing Technology 529

Figure 13-1 Type “A” seal Arrangement 1-single seal

Dual Seal Arrangements

For arrangements of two mechanical seals, the terminology “tandem” and “double” are no longer in use A seal configuration composed of two mechanical seals is now known as either an unpressurized dual seal or a pressurized dual seaL3 Furthermore, the lubricating fluid between unpressurized dual seals is a buffer fluid and it has a pressure less than that of the pump process fluid Pressurized dual seals employ a bar- rier fluid between the seals that is at a higher pressure than the pump process fluid The standard unpressurized dual (Arrangement 2) mechanical seal is an inside balanced type cartridge mounted mechanical seal with two rotating flexible elements and two mating rings in series, Figure 13-2 The inner seals of Arrangement 2 are designed with a positive means of retaining the sealing components and sufficient closing force to prevent the faces opening to pressurization of the buffer fluid to 40

psig (2.75 barg) Also, the outer seal must be capable of handling the same operating pressure as the inner seal

The standard pressurized dual mechanical seal is an inside balanced type cartridge mounted mechanical seal with two rotating flexible elements and two mating rings in

series (Figure 13-3) It is classified as Arrangement 3 The inner seal must be

designed to stay in place in the event that pump process fluid is lost This criterion ensures that the faces do not open up in the event that the highest pressure is at the

ID of the inner seal

Vapor Pressure Margin

API 682 requires the seal chamber pressure to be a minimum of 50 psi (3.4 bar) or

10% above the maximum fluid vapor pressure at seal chamber fluid temperature? This margin can be achieved either by raising the seal chamber pressure or lowering the seal

530 Improving Machinery Reliability

Figure 13-2 Arrangement 2-standard dual seal with buffer fluid pressure lower than process fluid pressure

Figure 13-3 Arrangement 3-dual seal with barrier fluid pressure higher than process fluid Inner seal is double balanced to withstand reverse pressure without opening

chamber fluid temperature If the temperature is above 140°F (60°C), an internal circu-

lating device and an API Plan 23 closed-loop cooling system should be used

Distributed Seal Flush

A distributed seal flush system such as a circumferential or multiport (see Figures

13-1 and 13-12) arrangement should be provided for all single seals with rotating flexible element^.^ Ports with a minimum diameter of fig" (3 mm) must be used in

multiport systems Distributed flush systems are not mandated for stationary single seals and dual seals

Providing Safety and Reliability Through Modern Sealing Technology 531 Throat Bushings

API 682 requires that throat bushings be provided unless otherwise ~pecified.~ A

throat bushing provides a restrictive clearance around the shaft (or sleeve) between the seal and the impeller It is used to increase seal chamber pressure and control seal

flush flow rate A floating bushing allows a tighter clearance and, hence, higher seal chamber pressure For Plan 23 flush systems, a throat bushing acts to isolate the seal chamber fluid from the pump process fluid

Seal Face Materials

One of the seal rings must be premium grade, blister resistant carbon graphite; the mating ring should be reaction-bonded silicon carbideU3 Self-sintered silicon carbide should be furnished, when specified

Seal Manufacturer Qualification Testing

To assure the end user that a seal will perform reliably, each seal/system must undergo qualification testing on an appropriate test rig.3 Testing is done on two seal sizes, 2" (50 mm) and 4" (100 mm) The test sequence consists of dynamic, static, and cyclic phases run consecutively without disassembly of the seal Seals for flash- ing hydrocarbon services are tested on propane

Dynamic testing is for a minimum of 100 hours, and the static (0 rpm) phase is for

at least four hours The cyclic phase includes startup, dropping pressure to vaporize seal chamber fluid, turning off flush for one minute, and shutdown A total of five

cycles are run

Low-Emission Single Seal Design

Single mechanical seals provide reliable sealing for most VOC services when the following conditions are satisfied:

* Fluid specific gravity > 0.45

Vapor pressure margin in the seal chamber > 25 psi

* Flush fluid provides good lubrication of the faces

A dual seal with a barrier fluid is recommended if any of these conditions are not met

532 Improving Machinery Reliability

Low-emission single seals are designed to run with contacting faces To minimize heat generation and wear, contact loading must be light Face contact promotes longer seal life by minimizing the possibility of abrasive particles getting between the faces Contact occurs on face high spots (asperities) A small amount of fluid migrates across the sealing dam in the spaces between asperities Typically this leak- age is on the order of 1 gm/hr for low emission single seals

Seals on flashing hydrocarbon services often operate with seal chamber pressure close to vapor pressure Seal face surfaces may be hotter than the fluid boiling point

In these cases, the entire fluid film between the faces is a vapor Well-designed seals can run successfully with vapor between the faces provided that careful attention is paid to the following:

Face deflections

Seal balance ratio

Materials

Flush

Vapor pressure margin

Recoverability from an upset

Face width

Face Deflections

Seal faces are lapped to about 1 light band flatness In operation, however, the faces distort (cone) due to pressure and thermal loading Generally faces are designed so that pressure deflects the faces toward OD contact and thermal loading deflects the faces toward ID contact Optimized seals run with total coning (sum of

pressure and thermal deflections for both faces) close to zero light bands Figure 13-

4 shows faces that are in ID contact and slightly open to the high pressure fluid

Wear of the carbon face under steady running conditions usually produces full face contact, from OD to ID across the sealing dam (Figure 13-5) Contact occurs on

the asperities Asperity size is greatly exaggerated in Figure 13-5 Roughness aver- age (RJ is on the order of 0.10 pm (4 win) for a well-running seal A fluid film, with

a pressure drop from OD to ID, exists between the asperities The fluid film provides hydraulic load support and allows some leakage in spite of face contact

Vapor temperature margin is often less than 20°F on flashing hydrocarbon services Laboratory tests on propane show that seal interface temperature is typically 40°F (or more) hotter than the seal flush.4~~ Thus, vaporization of the fluid film usually occurs

near the face OD (Figure 13-5) due to face temperature rise Some researchers have

stated that seals will not be adequately lubricated unless at least 50% of the fluid film

radial width is in the liquid state However, flashing hydrocarbons have a liquid vis-

cosity on the order of 0.1 cP, which is less than 1% the viscosity of lubricating oils Due to the very low viscosity of light hydrocarbon liquids, there cannot be any signif-

icant hydrodynamic lubricating effect Lubrication may be enhanced by a tribochemi-

cal reaction of the fluid film hydrocarbon vapor and the seal faces.6 Laboratory test- ing and field experience show that five-year life, with low emissions, is attainable for optimized seals running with a vapor film from OD to ID

Providing Safety and Reliabiliv Through Modern Sealing Technology 533

Figure 13-4 Faces usually contact at ID or OD at startup Vaporization often occurs near the OD on flashing services due to face temperature rise

~ 1 V B P O I

Figure 13-5 Wear results in full face contact from OD to ID

Competent seal designers use finite element analysis and fluid mechanics models7

to predict seal face deflections Computer aided seal analysis has proven to be a valu- able tool in designing seals for optimal performance Due to model assumptions, however, the predictions need to be adjusted based on lab tests of seals running under simulated field conditions Testing on propane is used to guide the design process for seals running on flashing fluids For an optimized seal, a post-test surface trace of the carbon shows a smooth profile, with worn-in coning less than four light bands

Seal Balance Ratio

The force acting to push the faces together is proportional to the seal balance ratio

(ignoring spring forces) Reduced leakage can be achieved by using a higher balance ratio to increase the closing forces A very high balance ratio must be avoided to pre- vent heavy contact loading and rapid wear Too low a balance ratio results in exces- sive leakage

An opening force, tending to separate the faces, is generated by the pressure distri- bution in the fluid film When the sealed fluid changes from liquid to vapor as it

534 Improving Machinery Reliability

moves across the faces, the opening force can be significantly greater than if the

fluid remains in the liquid phase.* Figure 13-6 illustrates typical pressure distribu-

tions in the fluid film for three different fluid conditions: all liquid leakage, liquid flashing to vapor half way across the face, and all vapor leakage The pressure pro- files are for faces that are running with full-face contact It is clear that two-phase flow results in higher film pressure and thus higher opening force Higher balance ratios are thus required for seals operating on volatile liquids If the balance ratio is not sufficiently high, the faces can “blow open.”

Ideal balance ratio for two-phase seals can be estimated using theoretical models.* Actual seal performance is usually somewhat different than that predicted Laborato-

ry testing on propane shows that optimum balance ratio is in the range of 77% to

85% for flashing hydrocarbon services

Face Width*

Selection of face width involves contrasting criteria A narrow face generates less

heat, but a wider face is more resistant to pressure deflection and provides a longer

leak pathway A narrow face is more sensitive to O-ring drag effects

For narrow faces, O-ring drag is a larger fraction of net closing force Net closing force consists of hydraulic closing force, spring loading, and O-ring drag O-ring drag can either add to or subtract from the axial closing force.9 Since hydraulic clos-

ing force is proportional to face area, O-ring drag has a larger relative effect on clos-

ing force for narrow faces

Providing Safety and Reliability Through Modern Sealing Technology 535

(CASA) predicts a large amount of pressure deflection for the narrow face, causing

OD contact The divergent gap results in low opening forces developed within the fluid film and, thus, high contact loading For this identical geometry case, the com- puter model predicts a face AT of 58°F for the standard width face, and a AT of 178°F for the narrow face However, highly satisfactory experience is reported with narrow face seals and optimally configured seal components in light hydrocarbon services lo

Optimum face width to achieve both low emissions and long service life is thus a function of both seal design and service conditions Properly designed and operated seals with face widths ranging from 0.140” to 0.280” can comply with strict emis- sion limits and attain long life Narrow faces typically perform satisfactorily at pres- sures below PO atmospheres (150 psi), but may require extensive redesign in higher pressure applications

Reaction-bonded silicon carbide is preferred over sintered S i c because it is per- ceived to have higher chip resistance The use of large chamfers improves the chip

resistance of both types of silicon carbide.12 Sintered S i c has superior chemical resistance and it performs well in corrosive applications such as HF acid