Mechanical seal optimization and rolling element bearing selection are critical issues for any plant. In this installment, we review case studies and associated cost justifications to learn how making the right decisions regarding these components can pay off. Each of these case studies is thoroughly experience-based—and the upgrades they describe are easy to implement, even on a tight budget.

Bearing housing seal opportunities
Most rolling-element bearings fail to reach their predicted theoretical L-10 life. One of the primary reasons for this life curtailment is contamination of the lubricant. Past practice for protecting bearings from contamination has included the use of lip seals and stationary labyrinth-type seals. These solutions, however, can suffer from short life or simply fail to exclude atmospheric contaminants.

Rotating labyrinth seals represent newer technology. Yet, by definition, even they will always have an airgap—although the airgap dimension varies with design and seal size. Therefore, rotating labyrinth seals will not preclude (at least some) air interchange. In contrast, and like mechanical seals, certain facetype bearing protector seals will prevent this air interchange entirely.

Suppose a facility had identified certain centrifugal pumps that suffered from disappointing bearing life and that these pumps were installed in a contamination- prone environment. It is not difficult to imagine bearings adjacent to steam quench injection points near mechanical seals, or in mining pumps or in areas experiencing sandstorms to be at risk here. Moreover, given source data published in Ref. 1, it is more than reasonable to anticipate a two-fold increase in bearing life with hermetically sealed bearing housings.

The term “hermetic sealing” is used to indicate no ingress of the ambient environment and no egress of the medium contained in the device or assembly so sealed. Modern mechanical seal technology has led to the development of dual-face magnetic seals (Fig. 1) that can quasi-hermetically (i.e. minimum leakage) seal pump bearing housings, gearboxes and other rotating equipment. Dual-face magnetic seals protect the lubricant and bearings by completely excluding contaminants. With quasi-hermetic seals in place, it is now possible to cost-justify the use of superior synthetic lubricants, and in fact, lower the total cost of operating equipment.

Progressive, reliability-focused equipment users who seek to improve the profitability of their operations employ Life-Cycle Costing (LCC), the conscientious application of which can help operations minimize waste. In the case of dual-face magnetic bearing housing seals applied to gear speed reducers and similar equipment, LCC often will show dramatic savings in operating and maintenance costs.

One of the simplest and most straightforward ways to assess the benefits of dual-face magnetic seals over sealing methods that allow an infl ux of atmospheric contaminants would be to look at a plant’s projected pump failure frequencies and repair costs. The cost of a set of dual-face magnetic seals would be more than offset by the avoidance of a number of bearing-related pump failure incidents. In virtually all cases so examined, upgrading to dual-face magnetic seals will show payback periods of less than six months. As explained
below, using a conservatively projected five-year life for dual-face magnetic bearing housing seals indicates a 16:1 payback.

More elaborate and more precise benefit-to-cost calculations are available and could take many forms. Only one of these, labeled a simplified benefit-tocost calculation, is given here (Ref. 1). It relates to a centrifugal pump that was originally equipped with lip seals (being replaced once per year) and now is being upgraded to five-year life dual-face magnetic seals. The five-year results are shown in Table I.

Again, the calculation in Table I covers a five-year life for hermetic sealing with dual-face magnetic seals. As shown, an incremental expenditure of ($640-$70) = $ 570 for dual-face magnetic bearing housing seals would return ($16,138-$5,084) = $11,054 over 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 air interchanges), dual-faced magnetic seals win. The potential benefits might favor dual-face magnetic seals even more if a plant were to opt not to replace its lip seals every year. In other words, not replacing lip seals would likely 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 show cost breakdowns that again favor dual-face magnetic seals—and do so by greater margins.

Lubricant pump-around (“circulating oil”) opportunities
The best example of a reliability-focused company is one that views every maintenance event as an opportunity to upgrade. The following enhancement case history is actually a hypothetical example of how, upon experiencing failure of oil ring lubricated pillow block bearings, a reliability-focused user would take steps to greatly reduce the risk of future bearing failures. Once a failure event occurs, this reliability-focused user would have a staff member (or staff members) who could immediately and authoritatively determine that upgrading is feasible and cost-justified.

Reliability engineers at such a facility would know that, according to findings published by several prominent bearing manufacturers, jet oil lubrication— as shown in the catalogs of all major bearing manufacturers—is considered the best possible method of applying lubricating oil to rolling element bearings.

Jet oil lubrication is unexcelled for rolling element bearings operating at high speeds and heavy loads. The oil jet is directed at the space between the outside diameter of the bearing inner ring and the bore of the cage. For extremely high speeds, means for scavenging the oil should be provided on both sides of the bearing. The oil system may be used to assure free axial fl otation of the bearing cartridge in the housing on a thin pressurized oil film. If axial fl otation is desired, a clearance of 0.001″ (0.025 mm) between the housing and the cartridge is recommended.

A pressurization source is required for the oil, however. Although circulating oil systems can economically provide this pressurization for large pumps, the economics do not favor full-fl edged circulating systems for smaller pumps. This is where small auxiliary gear pumps or inductive pumps merit consideration.

While high-velocity oil jets might be an option in some services, applications with large, heavily loaded, high-speed bearings operating at high temperatures actually may demand their use. In such cases, applying several jets on both sides of the bearing provides more uniform cooling and minimizes the danger of complete lubrication loss from plugging. The oil jet should be directed at the opening between the cage bore and inner ring’s outside diameter (O.D.). Again this is illustrated in the literature of virtually all bearing manufacturers. Adequate scavenging drains must be provided to prevent churning of excess oil after the oil has passed through the bearing. In special cases, scavenging may be required on both sides of the bearing.

Small auxiliary pumps used for jet-oil delivery. . .
Until recently, only large-scale oil mist systems were economically attractive, although smaller units are now available for a variety of stand-alone applications. Where the overall economics or unavailability of compressed air preclude oil mist, jet-oil spray systems are the best solution. As just one of many examples, a novel inductive pump in oil spray units would make this technically advantageous lubricant application method inexpensive, virtually maintenance-free and, thus, highly attractive.

Inductive pumps use electromagnetic force to drive a completely encapsulated internal piston, creating positive piston displacement within a sleeved cylinder. By using one-way check valves, both ends of the piston can be used for simultaneous suction and pumping. Since each stroke displaces a fixed volume, any increment of this volume can be delivered with a high degree of accuracy. One of several small auxiliary pump styles and sizes obtainable in the marketplace today (Ref. 2) can easily pressurize lubricant taken from the bottom of the sump to appropriate spray pressures. Lubricant rates are adjustable from 30 ml/min to 1.5 gpm (~ 6 l/min). Weighing a scant 14 lbs (~6 kg), the 3.25×9.5×4 inch (approximately 83x241x100 mm) unit can be combined with an automotive spin-on filter and an industrial spray nozzle connected to a length of fl exible tubing. Taking suction from the bottom of the bearing housing oil sump, a simple inductive pump is among those that represent a highly effective pump-around system.

Consider a user having to deal with a situation where large pillow block bearings support the shaft of an overhung blower impeller. This could well be a situation where the oil ring option described earlier in this cost justification example might be judged a reliability risk—and where objections might be raised to a conventional continuous lubrication oil system purely on economic grounds. Table II refl ects an estimated incremental cost basis to justify an attractive solution using an inductive pump.

As a result of this analysis, it can be shown that an inductive pump installation would win out on economic grounds and would be technically superior to an oil application strategy depending on oil rings.

References

1. Bloch, Heinz P. and Alan Budris; (2nd Edition, 2006) Pump User’s Handbook: Life Extension, The Fairmont Press, Inc., Lilburn, GA 30047, ISBN 0-88173-517-5
2. Technical literature, Inductive Pumps, Inc., www.inductivepump.com