Black oil and bearing housing upgrades

Heinz P. Bloch, P.E. ; Westminster, Colorado, USA 

Because rolling element bearings are standard components in most centrifugal pumps, scores of authors have devoted much time and effort to bearing performance issues. Lubricant degradation is among the many topics, although not many articles have dealt extensively with the root causes of oil ring-related contamination and formation of “black oil.” The term “black oil” refers to seemingly random lubricant discoloration that can be experienced over a rather short period of time. Needless to say, any manner of repeated short-term degradation of the lube oil affects bearing reliability. Equipment outages result and time and money will inevitably be spent. That’s why the subject of “black oil” should be important to us. Black oil is the result of one of only two possible events. One of these is rooted in an overheating of the lubricant; the other blackening event is often caused by slivers of black elastomeric material abraded from O-rings. The risk of experiencing considerable wear with dynamic (or moving) O-rings in certain bearing housing protector seals will be pointed out.

Bearing housings with or without oil rings?

Figure 1: Process pump bearing housing with static oil sump and two oil rings

Figure 1: Process pump bearing housing with static oil sump and two oil rings

An estimated 45{87a03eb4327cd2ba79570dbcca4066c6d479b8f7279bafdb318e7183d82771cf} of oil-lubricated industrial pump bearing housings are furnished with inexpensive oil rings that lift the lubricant from an oil sump (Figure 1). In some applications, oil rings are used to keep the oil volume more uniformly mixed, i.e., to prevent oil stratification. The term stratification describes a layer of hot oil floating near the top of the oil sump.

 

Figure 2: Bearing housing with thrower (or “flinger”) disc. Accommodating this stainless steel lube thrower disc requires that the thrust bearings be placed in a cartridge. The outside diameter of the cartridge is slightly larger that the diameter of the flinger disc. Note bearing housing seals at shaft protrusions (Source: AESSEAL plc, Rotherham UK, and Rockford, TN)

Figure 2: Bearing housing with thrower (or “flinger”) disc. Accommodating this stainless steel lube thrower disc requires that the thrust bearings be placed in a cartridge. The outside diameter of the cartridge is slightly larger that the diameter of the flinger disc. Note bearing housing seals at shaft protrusions
(Source: AESSEAL plc, Rotherham UK, and Rockford, TN)

We might assume that perhaps another 45{87a03eb4327cd2ba79570dbcca4066c6d479b8f7279bafdb318e7183d82771cf} of oil-lubricated pumps are designed for operation without oil rings. In the typical “without oil-ring” design, the oil level should reach the centre of the lowermost ball, roller or other bearing element. This oil level centering requirement would be difficult to achieve if, as shown in Figure 1, the bearings supporting the shaft had different diameters. We estimate that, in the remaining 10{87a03eb4327cd2ba79570dbcca4066c6d479b8f7279bafdb318e7183d82771cf} of oil-lubricated pump bearings, the lubricant is probably applied by pump-around (“pressurized”) means, or the housings incorporate flinger discs (Figure 2), or oil mist is being used.

To confirm, no oil rings are needed at slow-to-moderate shaft peripheral velocities. Shaft peripheral velocities are a function of operating speed (rpm) and shaft diameter, inches or mm. It is found convenient to express peripheral velocity as “DN-values”– the product of inches of shaft diameter multiplied by shaft rpm. For example, a 70 mm (~2.75 inch) bearing operating with a shaft turning at 1,800 rpm would have a DN-value of 4,960; values below 6,000 are somewhat arbitrarily considered low-to-moderate DN numbers.

In DN < 6,000 designs, both housing geometry and constant level lubricator settings are generally selected to allow lubricating oil to reach the center of the lowermost bearing ball or rolling element. However, if the DN value exceeds 6,000 it will be risky to keep the oil level at the center of the lowermost bearing ball. Potentially excessive friction-induced temperatures would likely result. These temperature concerns led to the decision to lower the oil levels in many of the larger API-compliant pumps operating at 3,000 and 3,600 rpm. In these, oil levels are customarily set well below the periphery of even the lowermost bearing ball. Because oil then no longer reaches a portion of the lowermost part of the bearing, some other means are needed to feed, lift, spray, or splash the lubricant into the pump bearings. Oil rings are often the least expensive component used for lifting oil into the bearings. But oil rings have limitations, and some of these are either little known or not well publicized. Oil ring malfunctions may deprive bearings of oil or simply allow oil to overheat. As mentioned in our introduction, that’s one possible source of black oil. But how and why can oil rings malfunction? What application limits exist?

DN-value limitations and attempts to improve on troublesome oil ring lubrication
Oil rings may be unstable for a number of reasons; oil ring instability (“wobbling”) is not a new phenomenon. To ensure proper operation, Ref. 1 cites surface velocity limits around 3,500 to 4,000 fpm (~18-20 m/s) with water cooling. The stipulated water cooling serves to maintain a constant oil viscosity; oil, of course, will become thinner when heated. Without water cooling of the lubricant, Ref. 1 advises users to stay well inside the stable limit for oil rings and not to exceed peripheral velocities of 2,000 to 2,500 fpm (~10-13 m/s). Another source, a major multi-national oil corporation’s “Lube Marketing Course Book,” suggests a DN value of 6,000 as the threshold of instability for oil rings. As a precautionary rule, both of these authoritative texts warn that oil rings in field situations tend to become unstable whenever DN, the product of shaft diameter (inches) and speed (rpm), enters the region from 6,000 to perhaps 8,000. With its DN value of 7,200, a
2-inch shaft at 3,600 rpm would thus operate in the risky or instability-prone zone, whereas equipment with a 3-inch shaft operating at 1,800 rpm (DN = 5,400) might use oil rings without undue risk of instability. In just one more example, a 3-inch (~76 mm) diameter shaft at 3,600 rpm would operate with a peripheral velocity of 2,827 fpm (~14.4 m/s).

One might not wish to quarrel with pump manufacturers pointing to satisfactory experience with higher peripheral velocities. However, field situations are usually far from ideal. In real-life situations, shaft horizontality and oil viscosity, depth of oil ring immersion, bore finish and out-of-roundness are rarely perfect. The vendor’s test stand experience is of academic value and field experience trumps all else. Reliability professionals must define safe operating ranges and use a good measure of conservatism.

Over the years, conservatism was probably sought by a number of pump manufacturers. In the 1990s one major pump manufacturer decided to look into the matter of oil ring instability and black oil formation. The manufacturer’s report was later presented as “Investigations into the Contamination of Lubricating Oil in Rolling Element Pump Bearing Assemblies” at a widely respected Industry Conference in 2000 (Ref. 2). The scope of both the report and the lecture presentation was to relate an analysis of factors affecting the short-term contamination of liquid lubricants in ring-oil lubricated rolling element bearings. “Short-term” was defined as intervals ranging from one hour to several weeks. The analysis dealt with the pump manufacturer’s standard range of centrifugal pumps, and short-term oil degradation had been reported in a number of installations (Ref. 3).

The presenter described the tests conducted and discussed their results. He then made a number of recommendations (reported here under “presenter”) to which this writer would like to add relevant editorial comments in brackets (as “author’s comment”).

Unless slightly preloaded, back-to-back mounted angular contact thrust bearings allow the unloaded side to skid. The unloaded bearing then gets quite hot (Ref. 3). [Author’s comment: The pump manufacturer had supplied back-to-back angular contact bearings and had inserted a thin shim between adjacent bearing inner rings. Competent users have, since the late 1960’s, insisted on using matched pairs of lightly preloaded or flush-ground back-to-back angular contact bearings. Shims are shunned. Moreover, there are applications where the user would be best served by lightly preloaded angular contact bearings with, say, a 15-degree contact angle on one, and a 40-degree contact angle on the mating bearing. This would prevent skidding of one of the bearings.]

Much of the contamination originated from metal particles being worn off the bronze oil rings (Figure 3) which were erratically hitting the adjacent components (Presenter’s observation). [Author’s comment: Abrasive wear has been known to occur with unstable oil rings. Factors contributing to instability are discussed below.]

Instead of metallic oil rings, plastic oil rings should be used (per presenter). [Author’s comment: High-performance polymers are indeed less likely to suffer from abrasive wear than metallic oil rings. However, simply switching to non-metallic rings treats the symptoms; it does not address the cause of the problem.]

Instead of using mineral oils that were either too light (ISO Grade 32) or too heavy (ISO Grades 68 and higher), ISO Grade 46 lubricants should be used (per presenter). [Author’s comment: With ISO Grades 32 and lower, the resulting oil film is often too thin to prevent metal-to-metal contact in rolling element bearings for centrifugal pumps. SKF commented on the matter in Ref. 4]

The various recommendations found in Ref. 2 and the author’s comments deserve to be further examined in terms of root cause problem identification.

Examining the causes of darkened oil

Figure 3: A new oil ring (left) and a badly worn oil ring (right). Oil rings tend to abrade; the abraded particles cause bearing distress (Photo courtesy of TRICO Corporation, Pewaukee, Wisconsin)

Figure 3: A new oil ring (left) and a badly worn oil ring (right). Oil rings tend to abrade; the abraded particles cause bearing distress (Photo courtesy of TRICO Corporation, Pewaukee, Wisconsin)

We know from experience that closely observing both new and “used” oil rings will prove very revealing. Abrasive wear of oil rings is easily recognized; a previously chamfered edge is now razor-sharp or burred (Figure 3), or an originally straw-colored lubricant has recently turned dark. Many copper-containing alloys leave a grayish color; pure overheating generally produces free carbon and causes the oil to become black. A good reliability verification technique would include measuring the “as installed” oil ring width and to later compare it with its “as-found” width. Measuring will avoid the cost of analyzing for contaminant composition. Needless to say, abraded oil ring material is suspended in the oil and bearing life is cut short by the abrasion product.

Years of operating and troubleshooting experience confirm that some pump manufacturers’ “old practice” of using low viscosity “thin” ISO Grade 32 oil is flawed. While acceptable for sleeve-type (also called “plain”) bearings, ISO Grade 32 oil is risky for the rolling element bearings found in the average process pump. In many small or mid-size pumps with rolling element bearings, thin oil exacerbates the problem of premature bearing failures.

The presentation based on Ref. 2 recommended ISO Grade 46 lubricants. However, we need to look at industry experience overall; we also might examine the position taken by competent bearing manufacturers in the matter (Ref. 4). While oil viscosity is an important parameter, tweaking the oil viscosity selection and substituting ISO Grade 46 (mineral oil) for the previously used ISO Grade 32 will not make much difference. Relatively minor changes in ambient temperatures negate the effect of small viscosity changes; numerous temperature vs. viscosity charts are available to confirm the validity of this concern. Moreover, a renowned bearing manufacturer had determined decades ago that film thickness and film strength limitations rendered ISO Grade 32 mineral oils undesirable for rolling element bearings in many centrifugal pumps (Ref. 4). Years earlier, this bearing manufacturer had asked users to restrict mineral oil ISO Grade 46 lubricants to bearings operating at temperatures not exceeding 70°C (158°F) and recommended ISO Grade 68 (again assuming mineral oil) for bearing operating temperatures not exceeding 80°C (176°F). However, to simply use more viscous, thicker oils would tend to slow down an oil ring and is not usually recommended.

While black oil can form because oil rings slip, skid, or bounce, black oil can also be found randomly in slower speed bearings that are not using oil rings. In those instances black oil formation is likely the result of the oil level receding by a fraction of an inch, essentially low enough to just barely reach the “bore rim.” We use the term “bore rim” to describe the edge (at the 6 o’clock position) of the bearing’s outer ring bore.

And here is another concern: Lube oil supplied into a bearing must be allowed to return to the main oil sump. This is accomplished by the small passageway near the 6 o’clock position of the left bearing in Figure 1. Note, however, that this passageway is missing from the same position on the right bearing. This omission is important because it deprives the oil to the right of the thrust bearing set from returning to the sump and freely mixing with the main oil volume. (Note that the same return passage is inexplicably missing from both sides of Figure 2).

Figure 4: A “dynamic” (moving) O-ring in close proximity of grooves will abrade if it contacts any of the sharp groove edges shown in this sketch

Figure 4: A “dynamic” (moving) O-ring in close proximity of grooves will abrade if it contacts any of the sharp groove edges shown in this sketch

 

Figure 5: Because there are no sharp edges to contact, the two “dynamic” (moving”) O-rings (see arrows) in this bearing housing protector seal will have extremely long life (Source: AESSEAL plc, Rotherham UK, and Rockford, TN)

Figure 5: Because there are no sharp edges to contact, the two “dynamic” (moving”) O-rings (see arrows) in this bearing housing protector seal will have extremely long life (Source: AESSEAL plc, Rotherham UK, and Rockford, TN)

 

Figure 6: Traditional constant level lubricators sometimes allow the pressure inside the bearing housing to rise above ambient. In that instance, the oil level will be pushed below the Desired Oil Level and oil may no longer reach the bearing’s rolling elements (Source: Trico Manufacturing Corporation, Pewaukee, Wisconsin)

Figure 6: Traditional constant level lubricators sometimes allow the pressure inside the bearing housing to rise above ambient. In that instance, the oil level will be pushed below the Desired Oil Level and oil may no longer reach the bearing’s rolling elements (Source: Trico Manufacturing Corporation, Pewaukee, Wisconsin)

 

Figure 7: A pressure-balanced constant level lubricator (Source: Trico Manufacturing Corporation, Pewaukee, Wisconsin)

Figure 7: A pressure-balanced constant level lubricator (Source: Trico Manufacturing Corporation, Pewaukee, Wisconsin)

Black oil will also form if bearing housing protector seals incorporate “dynamic” O-rings. Notice that the design in Figure 4 has a dynamic (moving) O-ring located between its rotating and stationary components. The O-ring will be damaged if it contacts either of the two sharp-edged grooves. The same failure mode does not exist in a design where an O-ring makes only very brief contact with the relatively large and well-contoured surface of a more advanced bearing housing protector seal design (Figure 5).

Oil level and oil application concerns must be addressed
There have been many instances where the transparent reservoir bulbs of constant level lubricators showed adequate levels of oil which, understandably, led operators and maintenance technicians to assume the oil level in the bearing housing was also at the correct height. Hydraulic laws should convince us that such an assumption is not always correct. The lube oil level in a bearing housing cannot coincide with the level in the base of the constant level lubricator if the internal bearing housing pressure is different from atmospheric pressure. It should be noted that, in a typical non-balanced constant level lubricator the actual oil level in the die-cast base of the assembly (Figure 6) is contacted by ambient air. Ambient air, in most industrial locations, carries water vapor and airborne contaminants, either of which will reduce bearing life. Also, according to the laws of physics, slightly elevated pressures in the bearing housing will lower the oil level in this housing. The displaced oil volume will be pushed into the die-cast base and the occasional overflow will be seen on the equipment base.

Using a properly pressure-balanced constant level lubricator assembly (Figure 7) goes a long way towards curing the problem. A suitably sized balance line back to the bearing housing ensures this pressure equalization. Together with advanced models of bearing housing protector seals –magnetically closed or rotating labyrinth styles — a pressure-balanced constant level lubricator makes a fully enclosed oil-lubricated bearing environment feasible. (An advanced rotating labyrinth bearing protector seal will be almost as good as a dual-face magnetic seal, and we leave it to the reader to determine what constitutes advanced rotating labyrinth seals, Refs. 5 and 6). Compared to non-balanced constant level models, the typical incremental cost of an average-size pressure-balanced constant level lubricator is $40. Just think if you retrofitted 200 pumps with advanced bearing protector seals and pressure-balanced lubricators. Calculate the value of avoiding even a single unscheduled pump downtime event in each of the next five years. And imagine the peace of mind those relatively minor incremental monetary outlays would buy. That would be a tangible and verifiable reliability improvement step for your pumps.

From an operating and maintenance perspective, don’t overlook that there must always be a partial vacuum in the upper part of the lubricator bulb. If caulking is used to cement the transparent bulb to a metal body in either Figure 6 or Figure 7, it must be realized that this caulking has a finite life. Once it develops tiny aging cracks, rainwater may enter via capillary action. Finally, traditional constant lubricator assemblies are direction-sensitive and should be mounted on the correct side of the bearing housing. The manufacturer’s instruction manuals describe this requirement, as does Ref. 5. The rotation arrow in Figure 6 is important. Constant level lubricators must be mounted on the “up-arrow” side of the bearing housing.

More on the limitations of oil rings
Assume the rolling element bearing housings are routinely furnished with oil rings and the manufacturer is either unable or unwilling to offer superior lube application methods in pumps or other equipment with high shaft (or bearing bore) DN-values. In those instances, the polymer oil rings recommended in Ref. 2 will be a (minor) step in the right direction. Still, it should be kept in mind that all oil rings have limitations which include the following:

(a) a DN limitation, i.e. certain rpm-times-oil ring-bore values should not be exceeded;
(b) oil rings are viscosity-sensitive, and Ref. 2 merely confirms that oil rings slow down in thicker fluids
(c) oil rings are immersion-sensitive, and disputing it would be denying the effects of viscous drag. This drag is (approximately) proportional to the velocity
(d) oil rings must not be out-of-round or “slightly oval.” Reliability-focused users should not allow more than 0.002” (0.05 mm) ring eccentricity (Refs. 1 and 5). Oil rings bought from the lowest bidder may not have been stress-relief annealed. They will not remain sufficiently round in service
(e) oil rings will run down-hill; they tend to malfunction if the total shaft assembly is not truly horizontal. They will malfunction if allowed to touch the inside of a bearing housing
Industry now has access to and benefits from superb laser-type shaft alignment tools. However, to achieve good alignment the installer typically shims up one end of the pump, thereby jeopardizing shaft horizontality.

Also, the user will experience occasions where, in spite of lubricant delivery via “constant level” lubricators, the oil level is actually lower than the set point indicated on the constant level lubricator assembly. Recall from our earlier comments that understanding how these lubricators function is very important and can save the user much pain.

Needed: a better choice than oil rings and constant level lubricators
Most 3,000 and 3,600 rpm pumps obtain splash lubrication through the action of oil rings. Yet, oil rings have inherent shortcomings that often make them a poor choice for risk-averse plants, Flinger discs (Figure 2) securely mounted on the shaft seem to offer a better choice than oil rings. Many reputable European pump manufacturers avoid oil rings and use flinger discs instead; some have done so for many decades. Because flinger discs are secured to the shaft, they are not subject to the influence of shaft horizontality, oil viscosity, depth of immersion, shaft surface finish and ring concentricity. These five factors inevitably vary from pump to pump; they result in an infinite combination of variables. Some of these combinations tend to make oil rings prone to malfunction. These malfunctions occur more often than deemed desirable by best-of-class user companies.

Visualize also that oil rings operate at slippage conditions relative to the shaft surface and that abrasive wear often results from this slippage. Slivers or particles of oil ring material have ruined thousands of bearings. Fortunately and unlike oil or slinger rings, flinger discs (Figure 2) are securely fastened to the shaft and avoid slippage and abrasion problems. However, when metal flinger discs are made with diameters larger than the bearing housing bore, a bearing mounting cartridge (Ref. 3, 3rd Edition, pp. 162) may be required to allow assembly without interference. Mounting cartridges are machined to precision tolerances and this adds to pump cost.

In the 1990s flexible discs were developed to get around the need for this added-cost cartridge. They “fold up” while being inserted through a smaller diameter bearing housing bore and snap open after passing through this housing bore (Ref. 5). Needless to say, it is important that the designer-manufacturer selects the proper elastomer and opts for a disc geometry that gives long and trouble-free service life at predetermined permissible – safe — peripheral speeds. Suitable elastomers are costlier than originally thought, which explains why we see few plastic flinger discs in pump bearing housings.

Recommendations and future improvements
Some day, an enterprising inventor or pump manufacturer may turn his back on cost-cutting and will return to ingenuity. He or an enterprising company may then develop a better alternative to oil rings or even flinger discs for rolling-element bearing housings. This alternative may be a smart device just short of the well-proven oil mist, and less costly than the widely known conventional, pressurized, oil pump-around systems. Reconfiguring part of the shaft to serve as the rotor of a progressive cavity pump, or utilizing the principle of magnetically driving a small internal lube pump may be feasible. Direct-driving a housing-internal impeller is clearly feasible for picking up some oil, pressurizing a small stream of this oil, and then spraying it into the bearing rolling elements.

As time progresses and if we are really fortunate, motivated engineers serving on the various API Standards Committees might address the various oil application issues and carefully listen to all sides of the story. Meanwhile, owner-purchasers of process centrifugal pumps should consider paying close attention to well-documented user experience and depend not only on the routine answers offered by some pump manufacturers. Let your equipment manufacturers know that your plant is dealing with real-world issues and that you are interested in lasting failure avoidance measures for chronic maintenance problems. Try to discourage temporary patch-ups and the flawed solutions that don’t cure root causes. Consider rejecting action steps that require maintenance expertise bordering on perfection. Don’t get into the habit of rewarding the lowest bidders by purchasing only from them. If you don’t heed this advice, the long-term experience will almost certainly prove disappointing.

Figure 8: Oil ring with evidence of damage due to “wobbling”

Figure 8: Oil ring with evidence of
damage due to “wobbling”

It is important not to miss the bottom line of this article. If the safety and reliability of a plant were dependent on upgrading bearing lubrication of vulnerable pumps, trust neither wobble-prone oil rings (Figure 8) nor lubed-to-center-of-lowermost ball applications. Be a bit more reluctant to put faith in most, if not all, constant level lubricators for really trouble-free lubrication (and trouble-free takes into account human imperfection of operator and mechanics). So, in most cases, try to get away from 18th century oil rings, which would also eliminate questions on oil ring material selection preferences. If at all possible, favor dry-sump oil mist for lubrication and for preservation of standby equipment, whenever this modern oil application method is cost-justified.

If oil mist cannot be justified, the author would — without the slightest hesitation — simply take oil from the drain location at the bottom of the bearing housing (Ref. 5). This oil would be piped to the suction of a small (yet highly reliable) fractional horsepower canned motor pump and generate a few psi of pressure. About two or three gpm (8 to 12 liters) of pressurized, filtered ISO Grade 68 PAO/Diester-based synthetic oil from the bearing housing drain would be forced through two spray nozzles into the inboard and outboard bearings of the most vulnerable process pumps. Implementation of such a strategy would close the book on issues involving black oil. Spraying oil into the bearings would terminate endless arguments over the merits of maintaining oil viscosity at tight limits, the adequacy of metallic versus plastic oil rings, the maintenance requirements of constant level lubricators, and the ramifications of hot bearings.

References:

  1. Wilcock, Donald F., and E.R. Booser, (1957) “Bearing Design and Application,” McGraw-Hill Book Company, New York, NY 10121
  2. Bradshaw, Simon; “Investigations into the Contamination of Lubricating Oil in Rolling Element Pump Bearing Assemblies”, Proceedings of the Texas A&M Pump Users Symposium (2000).
  3. Bloch, Heinz P.; “Improving Machinery Reliability,” (1982 and later Editions), Gulf Publishing Company, Houston, TX, ISBN 0-87201-376-6; ISBN 0-87201-455-X; ISBN 0-88415-661-3
  4. SKF USA Inc, “Bearings in Centrifugal Pumps, Application Handbook” Publication 100-955 (1995), pp. 20
  5. Bloch, Heinz P. and Allan Budris; “Pump User’s Handbook—Life Extension,” (2006) 2nd Edition, The Fairmont Press, Lilburn, GA 30047, ISBN 0-88173-517-5
  6. Bloch, Heinz P.; “Pump Wisdom: Problem Solving for Operators and Specialists,” (2011), John Wiley & Sons, Hoboken, New Jersey, ISBN 978-1-118-04123-9

About the author
hp blochHeinz P. Bloch resides in Westminster, Colorado. He holds B.S. and M.S. degrees in Mechanical Engineering and is an ASME Life Fellow. His professional career spans five decades and includes long-term assignments as Exxon Chemical’s Regional Machinery Specialist for the United States. He has authored well over 520 publications, among them 18 comprehensive books on practical machinery management, failure analysis, failure avoidance, compressors, steam turbines, pumps, oil mist lubrication and practical lubrication for industry. His “Pump Wisdom” (2011) explains issues and reasons why process pumps world-wide suffer from many elusive repeat failures.

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