HB Tech Paper 14

8/6/2019 HB Tech Paper 14

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Copyright 2005 AESSEAL plc - AES / DOC / IN 4980 06/2005

Registered Trademarks: - AESSEAL - AESSEAL plc - All other brand and product names are trademarks or registered t rademarks of their respective holder.

Heinz P. Bloch is a Consulting Engineer with offices in West Des Moines, Iowa. Before retiring from Exxon in 1986

after over two decades of service, Mr. Blochs professional career included long-term assignments as Exxon

Chemicals Regional Machinery Specialist for the United States. He has also held machinery-oriented staff and

line positions with Exxon affiliates in the United States, Italy, Spain, England, The Netherlands and Japan.

Troubleshooting and reliability improvement missions have taken him to process plants and manufacturing

facilities in 30 or more countries on all six continents. He has conducted hundreds of public and in-plant courses

internationally.

Mr. Bloch is the author of over 300 technical papers or similar publications. His 14 comprehensive books and a

searchable CD-ROM on practical machinery management include texts on failure analysis, failure avoidance,

compressors, steam turbines, oil mist lubrication and practical lubrication for industrial facilities. These

groundbreaking books have been used for reliability improvement lectures and maintenance cost reduction

consulting worldwide. In addition, Mr. Bloch holds five U.S. and many international patents relating to high-speed

machinery improvements.

Mr. Bloch graduated from the New Jersey Institute of Technology with B.S. and M.S. degrees (Cum Laude) in

Mechanical Engineering. He was elected to three National Honor Societies, is an ASME Fellow, and maintains

registration as a Professional Engineer in New Jersey and Texas. Mr. Bloch is the Reliability/ Equipment Editor of

the monthly publication Hydrocarbon Processing.


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Field Experience with Dual-Face Magnetic Bearing

Housing Seals

CASE STUDIES COMPILED FOR PRESENTATION AT

TAMU PUMP 2005

ByHEINZ P. BLOCH, Consulting Engineer

([email protected])

West Des Moines, Iowa

and

JOHN BRIGHT, Manager, Testing ([email protected])

AESSeal, Ltd., Rotherham, Yorkshire, UK

Heinz Bloch Technical Paper 2005

Copyright 2005 AESSEAL plc - AES / DOC / IN 4980 06/2005Registered Trademarks: - AESSEAL - AESSEAL plc - All other brand and product names are trademarks or registered trademarks of their respective holder.

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Field Experience with Dual-Face Magnetic Bearing Housing Seals

CASE STUDIES COMPILED FOR PRESENTATION AT TAMU PUMP 2005

By

HEINZ P. BLOCH, Consulting Engineer ([email protected])

West Des Moines, Iowa

and

JOHN BRIGHT, Manager, Testing ([email protected])

AESSeal, Ltd., Rotherham, Yorkshire, UK

Abstract

It has long been recognized that lube oil contamination vastly reduces the life of pump bearings. Thisprompted the American Petroleum Institute (API) and pump users to seek out and recommend

preventive measures, including devices such as rotating non-contacting labyrinth bearing housingseals (bearing isolators) and contacting-face rotating magnetic seals. All of these measures haveone primary objective: the extension of bearing life through reduction, and possibly even the virtualelimination, of lubricant degradation. Two typical case histories are presented.

Various factory tests, the development of cost justifications and a thorough review of field experiencehave established the viability and effectiveness of cartridge-type magnetic dual-face bearingprotectors. As of late 2004, these devices must be considered the best available means of preventingexternal contaminants from degrading the lubricant in bearing housings for industrial machinery. Thebearings of pumps, gears, mixers, small turbines, fans, blowers, star feeders, conveyor lines, rotarydrum filters, and literally hundreds of other shaft-driven and bearing-supported types of equipmentcan be effectively sealed with dual (or double) face magnetic seals. Four examples and theirrespective cost justifications are highlighted.

Quantifying the effects of contaminated lube oil

The fundamental reasons for favoring hermetic sealing devices for bearing housings can be derived from basicphysics (Amontons Law) and the technical literature. As regards Amontons Law we will recall that, upon cooling,the density of a gas mixture will increase. For a closed volume, a pressure reduction would result. However, if thiscooling takes place in a bearing housing and a path or opening exists to the external environment, ambient air willbe induced to flow into the housing. Should the air contained in the bearing housing heat up again later, thereverse would take place and warm air would be expelled into the surrounding atmosphere until the pressures areequalized.

It is certainly appropriate to anticipate that bearing housings that no longer have access to ambient airenvironments will preclude oil contamination from external sources. But, how clean is the oil? The variousbearing manufacturers use such qualitative terms as ultra-clean, very clean, clean, normal, contaminated, and


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heavily contaminated to describe oil cleanness. However, quantitative data are needed for purposes of costjustification or life-cycle-cost studies. We find these numbers in technical papers, articles and books that deal withthe quantitative effects of contaminated lube oil. Among the many references, only five technical data sources willbe mentioned here:

1. In Ref. 1, we find plots (provided by Dr. Richard Brodzinski, BP Oil, Kwinana, Western Australia) thatshow bearing lives with oil contamination normal to be less than one-half of the lives of bearingswith oil labeled clean.

2. The marketing literature of Royal Purple, Ltd., Porter, Texas, (Ref. 2), contains tables (Effect of FluidCleanliness on Rolling Contact Bearing Life) that assess and quantify the benefit of clean oil byassigning a life extension factor. Using the example of an ISO 4406:99 cleanliness level of initially22/19 and bringing it up to a new level of 14/11, Royal Purples experience shows bearing lifeextended by a factor of four.

3. Oil analysis experts at Tulsa, Oklahoma-based NORIA Corporation (Ref. 3) consider an ISO 4406:99cleanliness level of 23/20/17 typical for pumps. A level of 16/13/10 would be seen as world class,and >28/25/22 as evidence of serious neglect. For a cleanliness improvement from NORIAstypical to NORIAs world class and after converting its three-range numbers to the equivalenttwo-range ISO numbers 20/17 and 13/10, Reference 2 would again give a bearing life extensionfactor of four.

4. In Eschmann, Hasbargen and Weigands 1985 text Ball and Roller Bearings(Ref. 4, pg. 183),

European bearing manufacturer FAG emphasizes that the severity of the undesirable end effects ofcontamination depends on the ratio of operating viscosity of a lubricant divided by its ratedviscosity (Fig. 1). While there obviously could be an almost infinite number of combinations in theamount of contamination and ratios of viscosity, ratios of 0.5 to perhaps 1.0 are thought rathertypical. Using 0.5 for this ratio, and plotting from the mid-point of the zone labeled contaminants inlubricant (Zone lll) to the mid-point of the zone labeled high degree of cleanliness in the lubricatinggap, (Zone ll) we would find a four-fold increase in bearing life for the cleaner oil. At a viscosity ratioof 2, the projected bearing life increase traversing from contaminated to clean would beapproximately seven-fold. It should be noted that we are not here considering ultra-clean (Zone l)oil, since it would be unrealistic to find this degree of cleanliness in field-installed process pumpbearing housings.

5. But the most authoritative data on the effects of lubricant contamination might perhaps be gleanedfrom the General Catalog of one of the worlds leading bearing manufacturers, SKF (Ref. 5). For theexample shown in their catalog, SKF applied its New Life Theory to an oil-lubricated 45 mm radialbearing running at constant load and speed. Under ultra-clean conditions (nc = 1), this examplebearing was calculated to reach 15,250 operating hours. The SKF catalog text goes on to explain that,if the example were to be calculated for contaminated conditions such that nc =0.02, bearing lifewould be only 287 operating hours.

Oil cleanliness condition and quantifying the benefits of clean oil

Reasonable engineering judgment considers hermetically sealed bearing housings fully capable of maintaining

clean oil conditions. Opinions to the contrary are without foundation and do not coincide with field experience.As an example, Ref. 6, published in 1996, reaches the seemingly startling conclusion that the type of bearinghousing closure device (labyrinth, lip seal, or magnetic seal) shows no significant correlation with eitherparticulate or water contamination levels.


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However, this conclusion is easily explained by the fact that the bearing housings involved in the survey were allfitted with vented filler caps and were thus allowed to breathe. In this context, and as if to corroborate AmontonsLaw, breathing means that temperature expansion of the air/oil mixture floating above the oil levels causes thegas mixture to be expelled through the vented filler cap into the ambient environment. Upon cooling of the air/oilmixture, ambient air is again drawn into the bearing housings. This is very obviously an ongoing in-out process ofreal consequence for plants located in unfavorable environments.

As Ref. 6 states, five out of seven sample plants were located in the Houston, Texas area, which competes withLos Angeles for the worst industrial air pollution in the USA. Moreover, it is known that many, if not most, of thetested pumps employed oil ring lubrication. Oil rings are sensitive to shaft horizontality and, when slidingdownhill often contact housing-internal stationary parts. This has been shown to cause ring wear and generallyserious oil contamination (Ref. 7).

The oil cleanliness condition of any bearing housing interior that is accessed by the surrounding ambient air might,at best, be labeled normal. Rotating labyrinth seals have an open gap that allows communication between thehousing and ambient environment. While no definitive quantifiers are given that describe a refinery ambient, mostusers are aware that considerable amounts of particulates and moisture exist in the ambient air of industrialregions. A simple test situation may help in visualizing the issue.

Say, a new automobile is being washed and polished, and then left somewhere in the open near an industrial plant.Three days later, a person takes a clean paper towel and swipes it over the hood of the car. Not surprisingly, thepaper towel will no longer be clean. We can certainly envision that a considerable amount of dust is likely to findits way into bearing housings that continually breathe because of temperature expansion and contraction of the

air that fills the space above the oil level.

The severity of contamination can also be seen from the same study, Ref. 6, which had paradoxically concludedthat the amount of contamination found in pump bearing housings is independent of the type of bearing housingseal employed. It found that bearing housings (with open vents) have particle contamination levels at least 10times greater than recommended levels. Moreover, a staggering 54 percent of the more than 150 samples takenfrom industrial pumps contained contamination levels more than 100 times greater than recommended!

Recall that bearing manufacturers are using the terms clean and normal to describe the degree of lubricatingoil contamination. It can be reasoned that bearing housings with closed vents and equipped with face-type seals,or bearing protectors that employ mechanical seal principles and technology are able to keep the lubricantclean, whereas rotating labyrinth seals designed with an air gap will allow the lubricant to degrade to the point

of normal contamination. As we consider all of the above, it is simply reasonable to accept the premise thatbearings lubricated with clean oils will live at least twice as long as bearings with oil in normal condition ofcleanness. Therefore, this doubling of bearing life is often used in the most conservative and simplest costjustification calculations, as will be seen later.

Construction features of magnetically energized dual-face cartridge seals

In pumps equipped with magnetic bearing housing seals, the lubricant is totally contained while the atmosphereis effectively excluded. Shafts are sealed with magnetic seals at each end of the frame. In general, the rotaryportion of the seal is fastened to the shaft by an O-ring that performs both clamping and sealing functions. Theopposing component is O-ring mounted in a stationary, which is then fitted into the pump housing or frameadapter. Figures 2 and 3 illustrate two different configurations of the dual-face (or double-face) product.

Instead of springs to hold the faces together, one of the components of a modern dual-face magnetic bearinghousing seal is fitted with a series of small rare earth rod magnets. The two opposing parts are either a highstrength, corrosion-protected magnetic material or are made of a wear-resistant, low-friction face encased in a


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material attracted by the strong stationary rod magnets.

In dual-face magnetic seals two seal faces are magnetically held against their respective mating faces. Modernseals (Fig. 3) use stationary nickel-plated samarium-cobalt magnets, a tungsten-carbide rotary, an antimonycarbon face on the side towards the housing-interior, and a bronze-filled Teflon face on the atmospheric side.While single-face aerospace magnetic seals date back to the late1940s, the new and improved dual-facemagnetic seals have only been fully operational since about 2002.

Properties of magnetically energized dual-face cartridge seals

Of the two types of dual-face magnetic seals shown in Figures 2 and 3, the one illustrated in Figure 3 has beenindependently third-party tested with regard to its electrostatic and electromagnetic properties to ensure safeoperation in hazardous areas and potentially explosive atmospheres.

Dual-face magnetic bearing protector seals were subjected to independent testing to ensure ATEX compliance. ATEX is the European safety standard for explosive environments and test results are contained in an ATEXdocument which qualifies the double-face magnetic bearing protector seals that are the subject of this write-up.These bearing housing seals have been certified for use in hazardous areas where a large numbers of Group II,Category 2 equipment is required. The temperature classification is dependent upon the specific application. Theprocedure and methodology contained in this document enable the end user to determine the safe use of thedual-face magnetic seals in a given application.

Are there limitations?

The application options for dual-face magnetic bearing housing seals are almost limitless. They are presently usedon many pump configurations, including horizontal and vertical pumps, rotary lobe, progressive cavity and gearpumps. Gear speed reducers and a wide variety of different machines found in pulp and paper mills, corn millingequipment, different pillow blocks and rotary valves have also been successfully sealed with these dual-facecartridge magnetic seals, which incorporate neither clips nor set screws. Their dimensional envelope fits manylocations where lip seals were originally installed. Yet, their clamping locations avoid contacting the shaft surfacesworn off by lip seal contact.

The external faces of both of the modern dual-face magnetic seals show in Figs. 2 and 3 operate under dryrunning conditions. They are designed for dry running and will to do so without distress as long as the internal

face is not too hot. With marginal (oil splash) lubrication on the inner face of the seal in Figure 3 and dry runningconditions on the outer face, allowable peripheral shaft velocities were 22 m/s. Assuming a 3-inch (~76 mm)diameter shaft, it would rotate at slightly over 5500 rpm. At an 18C (65 F) ambient, the resulting facetemperature would stay within the stipulated ATEX limitation of 85C (185F), see Fig. 4.

The internal face is made of antimony-infused carbon, selected for its desirable properties, including that ofoptimal heat dissipation. Additionally, and regardless of actual need, the lubricant application conditions in pumpbearing housings are typically known to provide a small, or incidental, amount of oil to at least a segmented regionof the inboard sealing faces. The inboard faces of the aforementioned dual-face magnetic seal models (Figs. 2and 3) are designed to operate quite well with this incidental thin film lubrication.

Still, to safeguard against blatant misapplication, the manufacturer warns against dry running. As a matter ofgeneral policy, either continuous monitoring or other appropriate inspection and examination methods areadvocated so as to ensure correct equipment oil levels. Fortunately, the bearing housings of properly designedpumps will always incorporate lube application methods that generate an oil fog that results in not only adequatebearing lubrication but also a thin coating of oil for the seal faces and only the complete loss of oil could cause


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an unacceptable temperature increase at the inboard/outboard faces. Then again, thats an academic concernsince deprivation of lube oil would invite catastrophic bearing failure in any case.

In general, the manufacturers of different bearing housing seals are reluctant to publish leakage data on theirrespective devices. Occasionally, a test report finds its way into the hands of consulting engineers. However,dual-face magnetic seals embodying the features illustrated in Fig. 3 have always won in laboratory tests againstevery other bearing housing seal thats presently on the market.

Case histories and cost justification

There are a number of areas that merit being considered in justifying the incremental cost of dual-face magneticbearing housing seals over the cost of simple lip seals or one of the various forms of straight and rotating labyrinthseals.

Progressive, reliability-focused equipment users that seek to improve the profitability of their operations employLife Cycle Costing, or LCC. The conscientious application of LCC concepts will help reliability-focused plantsminimize waste. In the case of double-face magnetic bearing housing seals, LCC will often show dramaticsavings in operating and maintenance costs.

Life cycle component cost comparisons would compare the total lifetime cost to purchase, install, operate, andmaintain equipment with, vs. equipment without, magnetic bearing housing seals. The comparison would have to

include associated downtime, plus certain imputed values of having fewer failure events. Reliability-focusedplants include here the avoided cost of plant fires occasionally brought on by catastrophic bearing failures andthe implicit value of utilizing work force members that previously spent time on remedial tasks and are now freeto pro-actively work on preventive tasks.

A simplified mathematical expression for life cycle cost (LLC) might be:

LCC = Cic + Cin + Ce + Co + Cm + Cdt + Cenv + Cd + Cf - Cv

Where:

LCC = Life Cycle Cost

Cic = Initial cost, purchase priceCin = Procurement and installation, incremental costCe = Energy costs (pump, driver & auxiliary services)Co = Operation costsCm = Maintenance and repair costsCdt = Down time costsCenv = Environmental costsCd = Decommissioning and/or disposal costsCf = Fire damage costCv = Gain due to pro-active use of re-assigned workforce

Suppose a facility had identified certain centrifugal pumps that suffered from disappointing bearing life and

suppose further that these pumps were installed in a contamination-prone environment. It is not difficult toimagine bearings adjacent to steam quench injection points near mechanical seals, or bearings in mining pumps,or in areas experiencing sandstorms to be at risk here. Moreover, given the earlier source data, it is more thanreasonable to anticipate a two-fold increase in bearing life with hermetically sealed bearing housings.


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One of the simplest and most straightforward ways to assess the benefits of dual-face magnetic seals oversealing methods that allow an influx of atmospheric contaminants would be to look at a plants pump failurefrequencies and costs. The cost of a set of magnetic seals would be more than offset by the projected number ofbearing-related pump failure incidents. In virtually all cases so examined, upgrading to dual-face magnetic sealswill show payback periods of less than six months.

Simple payback calculation

Suppose a plant had centrifugal pumps with an average MTBF (mean-time-between-failures) of 2.5 years andsuppose further that the average repair is costing the plant $6,400, including burden, overhead, field and shoplabor, replacement parts, etc. The bearing housings are not hermetically sealed and there is clear evidence oflubricant contamination. Sets of magnetic seals cost $640 and hold the prospect of extending pump MTBF to 5years. The plant would avoid a $6,400 repair and, over a five-year period, would realize a payback of $6,400/$640---a 10:1 ratio.

More elaborate benefit-to-cost calculations could take many forms, and only one of these, labeled a simplifiedfive-year benefit-to-cost calculation, is given here. It relates to a centrifugal pump that was originally equippedwith lip seals and is now being upgraded to dual-face magnetic seals:

Lip Seal-Equipped vs. Dual-face Magnetic Seals________________________________________________________

Cic ----- (5 yrs)(2)($35) = $350 (2pc)($320) = $640

Cin ---- (0.08)($350) = $ 28 (0.08)($640)= $ 51[Procurement cost is ~8% of component cost. Assume no modifications needed to mount cartridge-typedual-face magnetic seals in pump bearing housing]

Ce ---- = $ 0 = $ 0[Assumes no significant change in frictional energy magnetic seals vs. lip seals]

Co ---- = $186 = $ 45[Cost of mineral-type lube oil, replaced once per year, vs. synthetic lube,replaced once in 5 years. Includes labor for oil changes]

Cm ---- = $ 10,625 = $2,125[Pumps being partially dismantled 5 times for lip seal installation, vs. once for magnetic seal installation.The alternative of NOT replacing lip seals yearly would incur bearing failures and more costly repair incidents]

Cdt ---- = $ 0 = $ 0[Assumes a facility with redundant, or installed spare pumps]

Cenv -- = $ 5 = $ 1[Assumes Kyoto Protocol values in a signatory country]

Cd ---- = $ 0 = $ 0[Assume waste oil will be mixed with furnace feed; hence, no disposal cost]

Cf ---- = $ 4,444 = $2,222[Incremental cost due to a $4,000,000 fire occurring once per 1,500 pump failures per year in refinery


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pumps. Assume one-third of these are bearing-related. Therefore,$4,000,000/1500X33.3%X5yrs = $4,444]

Cv ---- = $ 500 = $ 0[Imputed loss due to not being able to assign technical work force to perform proactive or preventive taskselsewhere]

Five-Year [One-Year] Totals: $ 16,138 [$3,228] $ 5,084 [$1,017]

Again, this calculation covers a five-year life for hermetic sealing with dual-face magnetic seals. As shown, anincremental expenditure of (640-$70) = $570 for dual-face magnetic seals would return ($16,138-$5,084) = $11,054over a five year period. The payback would be approximately 19:1.

Compared against any other means of sealing, i.e. housing seals that would allow breathing (ambient airinterchanges), dual-face magnetic seals win. The potential benefits might favor dual-face magnetic seals evenmore if a plant were to opt not to replace its lip seals every year. In other words, not replacing lip seals wouldlikely result in more bearing replacements or even total pump overhauls and, in certain cases, unit downtime costs.Likewise, it is noteworthy that studies with lip seals on centrifugal pumps being replaced twice every year showcost breakdowns that again favor dual-face magnetic seals, and do so by greater margins.

Case History 1: Axially split case centrifugal pumps in boiler feed service

Although supplied by a well-known major manufacturer, five hot condensate pumps at a petrochemical plant hadbeen furnished with lip seals at their respective bearings. These lip seals leaked immediately upon startup, whichrisked depleting the oil in the rather small volume (1 quart, or ~1 liter) oil sump.

Steam quench escaping from mechanical seals adjacent to the bearing housing seal area undoubtedly contributedto the disappointing performance of the factory-supplied elastomeric lip seals. Water intrusion caused rapid andrepeated bearing failure, with estimated repair costs in the vicinity of $10,000 --- a per-event cost approximationfor multi-stage pumps in this size range (Ref. 7).

Since three sealing locations are involved per pump (Fig. 5), and with the incremental cost of three dual-facemagnetic seals remaining well below $1,000, avoiding even a single repair incident per year would equate to abenefit-to-cost ratio in excess of 10:1. Indeed, the facility advised payback in the vicinity of one month.

Case History 2: Rotary valves and Star Feeders

In late 2003, the manufacturer of dual-face magnetic seal was asked to look into an ongoing problem of sealingthe stuffing boxes on a rotary valve (Fig. 6) at a bulk processing unit for powdery material. The valve, located atthe bottom of a bin that feeds highly abrasive powder to a pneumatic conveyor system, has upper and lowershafts that penetrate both sides of the valve body. The shafts go through stuffing boxes that are packed withbraided packing.

The product caused unacceptably high shaft and sleeve wear in the area of the stuffing box. In fact, the stuffingboxes had to be repacked every three days, causing many man-hours of labor and downtime associated with unitoperations.

The problem was solved by applying simple, basic sealing principles. The product had to be kept out of thestuffing box and a suitable sealing device had to provide a positive seal to the atmosphere. A three-step process


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was implemented to accomplish just that:

A throat bushing with a diametral shaft clearance of 0.010 inches (0.25 mm) and an O-ring placed inan O.D. groove was fitted at the process side of the stuffing box (Fig. 7).

A standard double-face magnetic bearing housing seal was placed into a machined gland followerand the entire unit bolted to the stuffing box. The gland follower had a gasket on the nose that wentinto the box so as to create a positive seal when tightened down. All leak paths for the product to exitthe valve stuffing box area were eliminated.

An air purge at 3 psi (0.2 bar) was connected to each stuffing box so the box area maintained ahigher pressure that the valve body area. This relatively low pressure was sufficient to keep thepowdery product away from the box area and low enough to allow use of the double-face magneticseal without employing any lubrication provisions.

After eight months of running the valve has required no maintenance in the area of the stuffing boxes. No producthas leaked from the double-face magnetic seals and no shaft or sealing maintenance has been required to date.

The owners of the plant estimated the value of downtime avoidance--three days at $26,000/day--as $78,000.Plant maintenance and the reduction in parts consumed were valued at $30,000. The plant estimated a combinedsavings of $108,000 in a period of only eight months.

Case History 3: Vertically oriented axially split pumps at a water works

From Figures 8 through 10 it can be seen that the pillow block bearings are installed in close proximity to theproduct sealing areas of these medium-pressure, 8,300 gpm (1,925 cubic meter/hr) water pumps. Water intrusionand oil contamination required replacing the pillow blocks every six months.

With dual-face magnetic seals now having operated flawlessly for over 15 months and no water intrusion orrelated distress foreseen, it has been said that a $400 upgrade has avoided in excess of $10,000 worth of repairs.Investing in dual-face seals achieved a payback of two weeks.

Case History 4: Cooling tower fan drive gears*

Cooling tower gears are exposed to one of the most adverse environments encountered in modern industry. Thesefan drive gearboxes are typically surrounded by clouds of water vapor that often contains a mix of treatmentchemicals. In turn, the chemically loaded vapors enter the gearboxes through casing vents and shaft protrusions.Once inside, the water vapors may cause further damage by interacting with the lube oil additives.

The contaminating mixture caused severe rust formation on gears as well as bearing distress in eight of thegearboxes at a major power plant in the Southern United States. Lip seal life did not exceed six months. Withrestricted access to cooling tower internals, even early preventive measures seemed expensive. To gain access,the cooling tower shrouds have to be opened and special cranes (cherry pickers) are brought to the site forlifting duty. Temporary scaffolding is needed for some fan cell configurations; also, in some locations, safety andhealth regulations require the wearing of cumbersome breathing apparatus. This would explain why gearboxrepairs costing around $30,000 per event are in the minority and repair costs involving gear-internal replacementsare very often in the $60,000 league.


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Fig. 2: State-of-art magnetically energized dual-face cartridge bearing housing seal

(Source: AESSeal, Inc., Rotherham, UK, and Knoxville, TN)

Fig. 3: Dual-face magnetic seal and O-rings (2, 4, 6, 10), snap ring (11),rotating face (1), stationary faces (3, 9b),stationary magnets (8), magnet carrier (7), and outer body (5). (Source: AESSeal, Inc., Rotherham, UK, and Knoxville, TN)

Fig. 4: MagTecta ll ATEX Temperature Graph (@18 deg. C/ 65 deg. F ambient for oil splash (marginal lubrication)


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Fig. 7: Rotary valve stuffing box with double-faced magnetic seal and suitable purge gas (plant air) purge willprevent both abrasive product escape and entry of ambient air

Fig. 6: Rotary valve fitted with double-face magnetic seals

Fig. 5: Multi-stage, axially split centrifugal pump in boiler feedwater service


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Fig. 8: Axially split pump in vertical orientation at a municipal water facility in the United States

Fig. 9: Lower bearing region of vertically oriented pump requires effective protection against water intrusion

Fig. 10: Dual-face magnetic seal installed near lower bearing in vertically oriented water pump at a water treatment facility

HB Tech Paper 14

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