Pump Handbook by Igor J. Karassik, Joseph P. Messina, Paul Cooper, Charles C. Heald

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2.2.5 CENTRIFUGAL PUMP OIL FILM JOURNAL BEARINGS 2.273

or the film cavitates and decreases to atmospheric pressure as the lubricant releases

entrapped air.

With respect to stability, cavitation is a more desirable condition than the development

of negative pressure. From an examination of Figure 1, it can be seen that the negative

pressure pulls the journal in an orthogonal direction and increases the cross-coupling.The

more eccentric the bearing, the larger the negative pressure or cavitated region.

In lightly loaded bearings that are pressure-fed, negative pressures can occur because

pressures in the divergent region have not approached atmospheric pressure. Cavitation

does not occur, and the bearing is prone to instability. Thus, the feed pressure and load on

a bearing are two additional parameters that affect stability.

Lobe bearings are often used because of their excellent antiwhirl characteristics. Fig-

ure 9 earlier showed two types of lobe bearings: symmetric and canted. The symmetric

lobe bearing is designed so that, in the concentric position, the minimum film thickness

occurs at the center of each lobe. Note that this permits a region of converging film fol-

lowed by a region of diverging film. Thus, depending upon the ambient pressure, it is pos-

sible to have negative pressures in a symmetric lobe bearing. Under very high ambient

conditions, a symmetric lobe bearing can go unstable.

The canted lobe bearing is designed to have a completely converging wedge and posi-

tive pressure throughout its arc length. Its stability characteristics are superior to those

of the symmetric lobe bearing. Its steady-state characteristics are also superior.

Hybrid Bearings At times, externally pressurized bearings are resorted to for stability

improvements. The philosophy is that externally pressurized bearings are not subject to

high attitude angles, as is the case with hydrodynamic journal bearings. Although this is

generally true, hybrid hearings can still be subject to considerable cross-coupling. Figure

11 showed a schematic arrangement of a hybrid bearing. Oil is fed through restrictors

from an external source into pocket recesses. From there, it exits into the clearance region

between recesses. Lubricant is also pumped into and out of recesses by the rotating shaft

by a viscous drag in the same manner as with a purely hydrodynamic bearing.

Consider recess 2 in Figure 11. Oil is pumped from the shaft via a converging wedge,

and it augments the pressure in the recess provided by the external system. The net result

is a higher pressure in the bearing domain covered by recess 2 than would occur without

rotation. Now consider recess 3. Here the journal is pumping fluid out of the recess into a

diverging film, so that the hydrodynamic action tends to reduce the pressure in this recess

domain. By similar reasoning, it can be shown that recess 4 operates at a lower pressure

than recess 1. The net result of these variations in pressure due to rotation is that cross-

coupling forces are introduced and the hybrid bearing may not prevent instability.

BEARING MATERIALS AND FAILURE MODES ____________________________

Materials

The most common material used for oil-lubricated fluid film bearings is bab-

bitt. Tin- and lead-based babbitts are relatively soft materials and offer the best insur-

ance against shaft damage. They also enable embedded dirt and contaminants without

significant damage.

Two types of babbitts are in common use. One has a tin base (86 to 88 percent), with

about three to eight percent copper and four to 14 percent antimony. The other has a lead

base with a maximum of 20 percent tin and about 10 to 15 percent antimony. The remain-

der is principally lead. The physical properties of babbitt are shown in Table 5. The pri-

mary limitations of babbitt are operating temperature (300°F [140°C] max) and fatigue

strength. The chemical composition of various babbitt alloys are indicated in Table 6.

Tin-based babbitts have better characteristics than lead-based babbitts; they have bet-

ter corrosion resistance, are less likely to wipe under poor conditions of lubrication, and

can be bonded more easily than lead-based materials. Because of cost considerations, how-

ever, lead-based babbitts are widely used. The more widely used is the SAE 15 alloy con-

taining one percent arsenic (refer to Table 6).

2.274 CHAPTER 2

TABLE 5 Properties of Bearing Alloys

Brinell Brinell Minimum Load Max

hardness hardness Brinell carrying operating

Bearing at room at 30°F hardness capacity, temperature,

material temperature (17°F) of shaft lb/in

(kg/m

) °F (°C)

Tin- 800–1500

based 20–30 6–12 150 or less (5.62  10



babbitt 10.55  10

) 300 (149)

Lead- 15–20 6–12 150 or less 800–1200 300 (149)

based (5.62  10



babbitt 8.44  10

)

OURCE

: Ref. 8

TABLE 6 Composition

Percentages of Babbitts with SAE Classifications of 11 to 15

SAE No. (Similar ASTM Spec.)

Element 11 12 13 14 15

(None) (B23, alloy 2) (None) (B23, alloy 7) (B23, alloy 15)

Tin (mm) 86.0 88.2 5.0–7.0 9.2–10.8 0.9–1.2

Antimony 6.0–7.5 7.0–8.0 9.0–11.0 14.0–16.0 14.0–15.5

Lead 0.5 0.5 Remainder Remainder Remainder

Copper 5.0–6.5 3.0–4.0 0.5 0.5 0.5

Iron 0.08 0.08

—— —

Arsenic 0.1 0.1 0.25 0.6 0.8-1.2

Bismuth 0.08 0.08

—— —

Zinc 0.005 0.005 0.005 0.005 0.005

Aluminum 0.005 0.005 0.005 0.005 0.005

Cadmium

——

0.05 0.05 0.05

Others 0.2 0.2 0.2 0.2 0.2

The percentage of minor constituents represents limiting values except as noted.

OURCE

: SAE Handbook, Society of Automotive Engineers. New York, 1960, p. 201.

Tin-based babbitts do not experience as many corrosion problems as lead-based bab-

bitts. SAE 11 babbitts (containing eight percent antimony and eight percent copper) are

used extensively for industrial applications.

Babbitts are bonded to a backing shell of another material, such as steel or bronze,

because they are not a good structural material. The thinner the babbitt layer, the greater

the fatigue resistance. In automotive applications where resistance to fatigue is important,

babbitt thickness is from 0.001 to 0.005 in (0.02 to 0.12 mm). For pump applications, the

thickness varies from to in (0.8 to 3 mm). The thicker layers provide good conformity

and embedability.

Failure Modes The failure modes most commonly found are fatigue, wiping, overheat-

ing, corrosion, and wear.

Fatigue occurs because of cyclic loads normal to the bearing surface. Figure 24 shows

babbitt fatigue in a 7-in (178-mm) diameter journal bearing from a steam turbine.

Wiping results from surface-to-surface contact and smears the babbitt, as shown in

Figure 25.The usual causes of wiping are a bearing overload, insufficient rotational speed

to form a film, and loss of lubricant.

2.2.5 CENTRIFUGAL PUMP OIL FILM JOURNAL BEARINGS 2.275

FIGURE 24 Babbitt fatigue in a 7-in (178-mm)

diameter turbine bearing. (Westinghouse Electric

Corp. Photo originally reproduced in Ref. 8, p.18–19.)

FIGURE 25 Bearing wipe on a 3-in (76-mm) diameter

bearing due to temporary loss of lubricant.

(Westinghouse Electric Corp. Photo originally

reproduced in Ref. 8, p.18–25)

FIGURE 26 Corrosion on a 5-in (127-mm) diameter, lead-based babbitt bearing. (Westinghouse Electric Corp.

Photo originally reproduced in Ref.8,p.18–21)

Overheating is manifest by discoloration of the surface and cracking of the babbitt

material. Corrosion is failure by a chemical action. It is more common with lead-based

babbits, which react with acids in the lubricant. Figure 26 shows corrosion damage for a

5-in (127-mm) diameter, lead-based babbitt bearing.

Wear results from contaminants in the film and is evidenced by scoring marks that

may be localized or persist around a large circumferential region of the bearing.

REFERENCES _______________________________________________________

Fuller, D. D. Theory and Practice of Lubrication for Engineers. Wiley, New York, 1956.

Castelli, V. and W. Shapiro. “Improved Method of Numerical Solution of the General

Incompressible Fluid-Film Lubrication Problem.” Trans. ASME, J. Lub. Technol., April,

1967, pp. 211

—

218.

2.276 CHAPTER 2

Castelli, V.,and J. Pirvics. “Review of Methods in Gas-Bearing Film Analysis.” Trans.

ASME, J. Lub. Technol., October 1968, pp. 777

—

792.

Pinkus, O. and B. Sternlicht. Theory of Hydrodynamic Lubrication. McGraw-Hill, New

York, 1961.

Ng, C. W. and C. H. T. Pan. “A Linearized Turbulent Lubrication Theory.” Trans. ASME, J.

Basic Eng., Series D, Vol. 87, 1965, p. 675.

Elrod, H. G. and C.W. Ng.“A Theory for Turbulent Films and Its Application to Bearings.”

Trans. ASME, J. Lub. Technol., July 1967, p. 346.

Shapiro,W., and R. Colsher. “Dynamic Characteristics of Fluid Film Bearings.” Proc. Sixth

Turbomachinery Symposium, sponsored by the Gas Turbine Laboratories, Department of

Mechanical Engineering, Texas A&M University, College Station, Texas, December 1977.

O’Connor, J. I., J Boyd, and E. A. Avallone. Standard Handbook of Lubrication Engineer-

ing. McGraw-Hill, New York, 1968.

2.2.6

CENTRIFUGAL PUMP MAGNETIC

BEARINGS

GRAHAM JONES

LARRY HAWKINS

PAUL COOPER

2.277

Magnetic bearings maintain the rotor of a pump in suspension through the forces of

attraction of a magnetic circuit. Thus, although they bear up the weight and hydraulic

loads of the impellers and the shaft, they are not really bearings in the traditional sense

of the rotating and stationary surfaces bearing on one another. The supporting magnet

circuit for each bearing includes stationary magnets in a stator surrounding the shaft, a

laminated rotor that fits on the shaft, and the shaft itself. The stator consists of electro-

magnets in the traditional heteropolar design, and if a homopolar design is employed, per-

manent magnets can be added. Sensors monitor the position of the shaft and signal a

controller to adjust the magnetic loads to keep the shaft to within about 0.001 in (25 mm)

of the desired position.

Magnetic bearings are found in small, high-speed turbomachinery such as high-speed,

multistage, axial-flow turbomolecular vacuum pumps

. They were introduced into large

turbomachinery in the early 1980s, mainly in gas compressors and turboexpanders. Their

use and acceptance has grown slowly but steadily since then

. Pump applications of a sig-

nificant size have appeared and have confirmed the general position that magnetic bear-

ings can provide a technically sound bearing with maintenance and operating advantages,

including zero wear. However, due to the technical complexity of magnetic bearing sys-

tems, the economies of scale associated with production quantities are required to make

these systems affordable.

Two representative magnetic-bearing-equipped pumps are summarized in Table 1.

One is a multistage boiler feedwater pump

3–6

and the other a single-stage double-suction

hydrocarbon process pump

. The multistage pump was retrofitted with magnetic bearings

(as shown in Figure 1) and is shown in Figure 2, together with another identical pump

that still contains the oil-lubricated bearings

—

both installed in an electric generating sta-

tion. The magnetic-bearing pump is not encumbered with the usual complexity of a bear-

ing lubrication system.

2.278 CHAPTER 2

TABLE 1 Example of magnetic-bearing-equipped pumps

Parameter Multistage Pump Single-Stage Pump

Power, hp (MW) 610 (0.46) 800 (0.6)

Rated speed, rpm 3,580 1780

Shaft weight, lb (kg) 520 (236) 732 (332)

Radial bearing design load, lb (kN) 800 (3.6) Thrust end: 930 (4.1)

Drive end: 1,415 (6.3)

Thrust bearing design load, lb (kN) 4,000 (17.8) 4,000 (17.8)

Number of stages 8 1

FIGURE 1 Magnetic bearing configuration in multistage pump

FIGURE 2 Multistage pumps installed in boiler feed service [610 hp (0.46 MW)]: a) pump with magnetic

bearings; b) the same pump with oil-lubricated bearings. (Flowserve Corporation)

MAGNETIC BEARING PRINCIPLES______________________________________

How Magnetic Bearings Work

In an active magnetic bearing system, a stator com-

posed of an array of stationary magnets, or electromagnetic coils, interacts with a ferrous

rotor (or a ferrous sleeve on a non-ferrous rotor) so as to suspend the shaft in a magnetic

field (see Figure 3).

The position of the shaft is maintained dynamically through a continuous feedback

system which comprises a position sensor, a controller, and an amplifier system (see Fig-

ure 4). Typically there are two radial bearings and one thrust bearing for a complete sys-

2.2.6 CENTRIFUGAL PUMP MAGNETIC BEARINGS 2.279

FIGURE 3 Typical radial electromagnetic bearing. (Axial thrust bearings have the stator coils arranged in a disk

configuration, a ferrous rotating disk being supported in the resulting magnetic circuit.)

FIGURE 4 Typical system control loop

tem. This system is tuned to the required characteristics of the pump, through a digital or

analog controller (Figure 5), with the capability of adjusting the bearing stiffness and

damping as a function of pump speed. Alarms and trips can be set at any required rotor

offset or bearing load to provide the operator with warnings or to trip the drive unit as nec-

essary. Figure 6 illustrates a typical bearing transfer function, showing a statically stiff

bearing, with a dynamic stiffness over the operating speed range designed to meet the

rotor dynamics requirements of the unit. The stiffness is then rolled off above the operat-

ing range to avoid excitation of higher modes in the rotor or stator.

Controller redundancy can be provided with the control loops switching to backup

units upon sensing a failure. Also, backup power supplies should be provided, either

through alternate sources or a battery system. Typically the power required is only one or

two kW or less.

A catcher (also known as auxiliary, backup, or touchdown) bearing (indicated in Figure

1) is required to protect the rotor stator interface during maintenance and in the event of

2.280 CHAPTER 2

FIGURE 5 Controllers for magnetic bearings, containing rectifier and amplifiers: a) digital controller;

b) analog controller

FIGURE 6 Bearing transfer function

loss of power or a severe transient beyond the force capability of the bearing.Typically, the

catcher bearing is designed for 5 to 20 lifetime drops from full speed, and will be a readily

replaceable rolling element or sleeve bearing.The pump of Figures 1 and 2 has rolling ele-

ment catcher bearings. The radial clearances G1 (see Figure 7) between the magnetic

2.2.6 CENTRIFUGAL PUMP MAGNETIC BEARINGS 2.281

FIGURE 7 Clearance arrangement: Seal ring clearance G3 is greater than catcher-bearing clearance G2, but is

less than magnetic-bearing clearance G1.

bearing stator and rotor are of the order of 20 to 40 mil (0.5 to 1 mm), and those in the

catcher bearings (G2) are about half of that value.

Reasons for Using Magnetic Bearings There are several reasons to use magnetic

bearings in pumps. While any one of these reasons may not be sufficient justification on

its own, together they can provide a strong justification.

Reliability is a key incentive.The components of a magnetic bearing are essentially the

same components as are found in an electric motor: laminations and coils. Because no

wear is involved due to the lack of contact, these components will generally last the life of

the equipment involved. Thus maintenance of a magnetic bearing system is transferred

from mechanical components inside the pump to the external controller, which has plug-

in card replacement maintenance. Pump reliability is therefore improved, whereas repair

times and costs are reduced.

Reduced power consumption is a second advantage, with the elimination of all losses

associated with fluid film bearings and oil pumping equipment. This is replaced by the

smaller power requirements of the bearing controller. Further, if the lifting force is sup-

plied by a permanent magnet, supplemented by an active control circuit, this power

requirement can be even smaller.

The ability to submerge the bearing in the pump fluid is a major advantage that

allows the outboard mechanical seal to be eliminated, thereby eliminating maintenance

and replacement of this seal

. [This was not done for the pumps of Table 1 (and Figures

1 and 2), because in both cases magnetic bearings were retrofitted to existing

machines.]

More indirect savings are also possible in two other areas. Rotor dynamics can be con-

trolled through the ability to adjust stiffness and damping as a function of pump speed,

allowing higher imbalance without the need for shutdown. The diagnostic output inherent

in the information provided in the controller can be fed into the overall plant operating

system and the short-term and long-term health of the pump and the system can be mon-

itored. This is done by inferring seal wear, transient hydraulic loads, and so on.

The actual figures for the savings possible due to the previous advantages are very

pump- and system-specific, and general numbers are not very useful. Reference 9 has devel-

oped methodology for considering the economic effect of the types of advantages given.

Main Types of Magnetic Bearings and Their Selection There are two main types of

magnetic bearings: passive and active. Passive bearings rely only on permanent magnets

in repulsion and provide low stiffness, low damping, and no ability to control either of

these parameters. Passive bearings are not applicable to pumps for this reason.

2.282 CHAPTER 2

FIGURE 8 Heteropolar and homopolar bearings

Active bearings using the feedback system previously described are essential for pump

applications. Within the active bearing systems are the options of heteropolar and

homopolar and of electromagnetic and permanent magnet bias. The principles of the het-

eropolar and homopolar approaches are shown in the Figure 8.

The main difference between the two types is that in the heteropolar design, the bias

and control flux flow in the same magnetic circuit radially through the rotor, whereas in

the homopolar design the bias flux flows axially along the rotor and only the control flux

flows radially through the rotor.

The homopolar design has two options for providing the bias flux for the bearing sys-

tem

. One is to use an electromagnetic effect, and the other is to use a permanent mag-

net. The permanent magnet generation in the homopolar configuration results in a more

linear relationship between force and distance. In a simple magnetic circuit, the attraction

force of a magnet on a ferromagnetic target decreases as the square of distance (the tar-

get cuts as many flux lines at twice the distance). With the permanent magnet in a

homopolar circuit, the effect of the air gap is therefore reduced.

DESIGN CONSIDERATIONS ____________________________________________

Design Loads

The specification of a magnetic bearing requires a different approach to

that required for a conventional bearing system. This section identifies the key areas

where these differences occur.

The rotor can move considerably within the magnetic bearing and catcher bearing

clearances, typically up to 10 mil (0.25 mm), before any contact is involved. Thus the clear-

ances G3 in the internal ring-seal system (see Figure 7) become the key controlling para-

meter in setting the bearing clearance design limits and in the degree of motion permitted

during transients.Thus, very early in the design process, the magnetic bearing clearances,

catcher bearing clearances, and sealing-ring clearances must be optimized with due con-

sideration for manufacturing tolerances and assembly concerns.

Synchronous filters or open-loop control methods can be used to handle imbalance

loads so that the degree of imbalance acceptable (based on allowable bearing loads and

shaft motion within the catcher bearing clearances) can often be significantly greater than

in conventional bearings.

Pump Handbook by Igor J. Karassik, Joseph P. Messina, Paul Cooper, Charles C. Heald - 3rd edition

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