156 Intproving Machinery Reliability Significant Differences In Bearings and Bearing Housings The eighth edition of API 610 continues to require axially preloaded 40" angular contact th
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W e have seen exceptions made when deviations from the rule-of-thumb were judged minor, or in situations where the pump manufacturer was able to demonstrate considerable experience with ANSI pumps under the same, or even more adverse conditions
Cost Justification If your company is interested in seeing the cost justification for purchasing the generally stronger API pumps instead of normally satisfactory ANSI pumps, consider statistical approximations,
Suppose that under average conditions of maintenance effectiveness and installa- tion care (foundation, baseplate stiffness, grouting, piping configuration, etc.), ASNI pumps in hydrocarbon services required service every 18 months vs three years for API-610 8th Edition pumps Assume further that each maintenance event requires the rather conservative expenditure of $6,000, including burden and overhead costs Knowing that you are likely to have a $600,000 fire for every 1,000 pmp failures means that each failure event will incur a $600 cost adder All of this translates to:
ANSI Pump repair cost, per year $4,000
Probable cost per year: $4,000 API Pump repair cost, per year $2,000
Probable cost per year: $2,200
This would yield a yearly maintenance cost advantage of $2,200 in favor of API pumps under the stated average maintenance effectiveness conditions Add to it pro- duction loss credits, and the number increases Also, how many service interventions per year equal one maintenance worker, or one reliability/technical support person,
or one planner/supervisor/purchasing specialist? Or just how much would it be worth
to your plant if your reliability engineers could spend more of their time on predic- tive maintenance and failure avoidance rather than having to struggle with equip- ment failures? Certainly a subject worth pondering
Between-Bearing Pumps There is some evidence that overhung impeller con- struction is occasionally more prone to maintenance or downtime than impeller- between-bearing construction In view of this, some contractors and pump pur- chasers have, in the past, applied rather arbitrary selection guidelines
One such guideline limits impeller diameters for overhung pumps to 15 in (38 1 mm) Another guideline calls for in-between-bearing geometries whenever the product of horsepower and rpm (rotational speed) exceeds 900,000 For example, at 3,600 rpm, motor ratings in excess of 250 hp would favor selecting between-bear- ing pumps
How rigidly should these rules be applied, if at all? W e would consider them appropriate far screening purposes Allow deviations if the pump is of double-volute design and if calculated shaft deflections d o not exceed 0.002 in over the entire operating flow range, from zero to full BEP flow Alternatively, allow deviations if the pump manufacturer can demonstrate long-term, satisfactory experience under identical (or more severe) operating conditions
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Significant Differences In Bearings and Bearing Housings
The eighth edition of API 610 continues to require axially preloaded 40" angular contact thrust bearings This requirement stems from three observations that pump users have made over the decades: Thrust bearings fail relatively often; 40" angular contact bearings have higher allowable thrust load ratings than bearings with 15" or
29" contact angles; and many failed bearings exhibit ball skid marks in the race areas Axial preloading greatly reduces the risk of incurring bearing distress due to skid- ding of rolling elements Details on this failure mode can be found in many publica- tions from major bearing manufacturers A typical analogy to skidding can be seen
in aircraft landings Upon initial touchdown, the wheels will skid until their peripher-
al speed has caught up with the forward speed of the plane Just as skidding would cause accelerated wear of tires, it would result in potentially severe metal-to-metal contact in a rolling element bearing
Axial preloading can ensure that the bearing will always be loaded With pairs of angular contact bearings, axial preloading may be necessary
Preloading or flush-grinding of thrust bearing sets will also prevent axial oscillato-
ry movement of pump rotors This motion is quite prevalent in pumps that experi- ence cavitation or low-flow induced internal recirculation The resulting instanta- neous acceleration forces must be absorbed by the rolling element bearings Again, bearing defects are much more likely to develop with bearings that operate with axial looseness than with preloaded or flush-ground bearings that operate without looseness in the axial direction
In most applications, properly installed and lubricated axial preload bearings have extended the mean time between pump repairs This is why many pump manufactur- ers, and especially the overseas manufactures of centrifugal pumps, do not take issue with the API requirement However, there are instances where preloaded 40" angular contact bearings have been unable to solve problems or have made a problem worse While we might have assumed that bearings with larger contact angles create more frictional heat, research by the FAG Bearing Corporation demonstrated 40' angular
contact bearings generate less heat than thrust bearings with less a n g ~ l a r i t y ~ ~ Never- theless, by its very nature, preloading adds to the heat load, and using an interference fit between shaft and inner ring compounds the problem
Most ball bearing manufacturers consider bearings to be preloaded if an offset, or predetermined gap, exists between the inner ring and outer ring faces while a light nominal load, often called a gauge load, is applied to the inner and outer ring thrust
faces (see Figure 3-46) When axially clamped in a back-to-back or face-to-face
fashion with a second equivalent bearing, the balls and races are forced to deflect, thus creating an internal load or preload Preload can also be created by interference fits between the shaft and bearing bore, or the housing and bearing outside diameter, and by temperature differentials between inner and outer rings-most often due to the inner ring running warmer than the outer ring Interference fits and temperature differentials decrease internal clearance; this will either create or cause an increase in preload It is important to realize that shaft interference fits for back-to-back mount-
ed angular contact bearings must be near the minimum of the range normally
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/
UOUWIED PRELOAD y6 F n
rrvI ,,I.~UOANo",IRCOHI*CI UEa1(*011
Figure 3-46 Application of gage load to axially pre-loaded bearings will show gap between inner races (Courtesy MRC Bearings.)
allowed for radial or Conrad bearings An interference fit near the typical maximum allowable is almost certain to result in greatly reduced life of these bearing sets
In this regard, Figure 3-47 will prove very enlightening It shows that for a given bearing (FAG 7314, 70 mm bore diameter) a shaft interference fit of 0.0003 in will
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produce an almost insignificant preload of approximately 22 Ibs, whereas an interfer- ence fit of 0.0007 in would result in a mounted preload of 200 lbs A much more significant preload would result from temperature differences between inner and outer bearing rings Such differences could exist in pumps if heat migrated from high temperature pumpage along the shaft, or if the pump design incorporated cooling provisions which might artificially cool the outer ring and would thus prevent it from expanding By far the worst scenario would be for a pump operator to apply a stream
of cooling water to the bearing housing from a firehose It is very discouraging to see this done even today, but old habits are difficult to break, regardless how counterpro- ductive they may be
Several ball bearing manufacturers now recommend “flush ground” bearings or bearings with different contact angles (e.g., 40” on the active and 15” on the inactive side) for back-to-back installation in centrifugal pumps During operation, these bearings either become preloaded or remain preloaded This prevents (or minimizes) ball skidding problems discussed on page 172 Moreover, flush-ground bearings will operate at a lower temperature than the initially preloaded pair Lower temperatures can improve lube oil oxidation stability and life Lowering the total load (i.e., thrust load plus radial load plus preload) would allow us to anticipate greater oil film thick- ness and less metal-to-metal contact
Specification Amendments Covering Pump Bearings The “best-of-competi-
tion” plants in the processing industry use definitive component specifications when- ever reliability engineering concepts are actively pursued They somehow realize that buying on price alone is rather naive and that a better product will have to com- mand a better price
A Dutch refinery with 11,000 pumps amends API 610 by asking that:
All rolling elements should have metal rolling element retainers,
Parallel roller bearings are preferred as pure radial bearings
Roller bearings should have the roller retaining rim on the inner race
Shielded or sealed bearings should not be used except for vertical in-line pumps up
to 22 kW
A major refinery in Texas tacks a machinery component parts specification to
both inquiry and purchase documents (Table 3-13)
When the API 610 Seventh Edition became available in February of 1989, we advised a large-scale user of centrifugal pump bearings to consider replacing para- graph 2.9.1.5 with the replacement wording shown in Table 3-14
Take your pick, or combine elements of the various specification amendments into
a docuinent that suits your particular needs Seek out the world’s leading manufac- turers of rolling element bearings and ask them if they have a business team, or application engineering group dedicated to fluid machinery Get their advice, request their relevant literature-you may be pleasantly surprised to see the value of sound, experience-based recommendations But don’t expect to get better bearings, or any other precision components, without making an effort to assemble a pertinent speci- fication document
Trang 6Machinery Reliability Audits and Reviews 159 Table 3-13
Machinery Component Specification Radial and Angular Contact Ball Bearings
1.4 The bearings shall be packaged per AFBMA Standards, Section 6 The bearing box shall be
marked to identify the original bearing manufacturer and the alphanumeric bearing identifica- tion code or designation system
2.0 Ball bearing design
2.1 Radial and angular ball bearings shall have the following design and features:
type clearances retainer features manufacturers
Single row, c - 3 Riveted Conrad type, SKF/MRC, FAG
NTN, KOYO Double row, c - 3 Stamped No filling SKF/MRC, NTN,
40" angular Standard Machined Land riding, SKF/MRC,FAG,
2.2 Bearing shields, seals, snap rings, etc., shall have the configuration specified by purchaser's applicable spare parts symbol number or specified description for the stipulated application 2.3 No substitutions are allowed for the bearing manufacturers or the bearing features shown with- out the auuroval of the uurchaser's rotating equipment reliability engineer
-
able, nil prel Rollway
Regreasable vs Non-Regreasable Rolling Element Bearings
At the inception of a project, the specifying engineeer will often be confronted with the question of when sealed, non-regreasable bearings present advantages over the tra- ditional, regreasable variety It is intuitively evident that fully sealed rolling element bearings would be the right choice for a wide range of appliances, such as a vacuum cleaner or household cooling fan Conversely, experienced engineers have known for years that regreasable bearings are the correct choice for larger size or higher speed bearings There was, however, no clear-cut guideline as to when to request one or the other, or which of the two modes-regreasable or non-regreasable-would make more economic sense when the labor cost of regreasing and the cost of rectifying greasing- related errors were considered
Some time ago, a large international bearing manufacturer provided a rule of thumb that can be used under typical circumstances, The rule requires calculating the
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Table 3-14 Replacement Wording Proposed for API-610 7th Edition, Paragraph 2.9.1.5
Specification Amendment for Heavy-Duty Centrifugal Pumps
(This S.A is appropriate for attachment to inquiry or purchase document.)
2.9 Bearings and bearing housings
2.9.1 Bearings for horizontal pumps
2.9 I 5 (Exception)
a If bail-type thrust bearings are used, the vendor shall verify the adequacy of duplex, sin- gle-row, 40" (0.7 radian) angular-contact type (7000 series) bearings, installed back-to- back (DB) This verification shall include selection of proper preload values to prevent skidding of the unloaded bearing over the entire operating range of the pump
b If a., above, points to inadequacies or potential problems, the vendor shall offer suitably preloaded sets of ball-type thrust bearings with dissimilar contact angles (e.g., 40"/15")
c Sets of properly preloaded, back-to-back mounted 15" or 29" angular contact bearings are acceptable for pumps with double-flow impellers located between two bearing housings
d In each case, the need for and extent of preload shall be determined by the vendor to suit the application and meet the bearing life requirements of 2.9 I I
DN value of the bearing, with D being the bearing bore (in mm), and N representing the shaft speed (in rpm) For DN values below 80,000, we agreed to a preference for sealed, lifetime lubricated bearings The range from 80,000 to 108,000 we labeled the gray area, and from 108,000 to 300,000 we defined a clear advantage for regreasable bearings (It has been a long-standing practice for bearing manufacturers
to recommend oil lubrication for bearings with DN values greater than 300,000.) With DN values of 80,000 obtained, for example, by a 40 mm bore diameter radial ball bearing operating at 2,000 rpm, Figure 3-48 recommends regreasing intervals of approximately 9,000 hours-about one year If such a bearing is provided with seals and is therefore lifetime lubricated, we anticipate it to become unserviceable after roughly two and a half times the stipulated regreasing period These experience fig- ures have been validated by some users and are reasonably correct For a radial ball bearing with a DN of 200,000, say 100 mm and 2,000 rpm, the recommended regreasing interval shown in Figure 3-48 would be slightly less than 3,000 hours, or four months of continuous operation Using the two and a half times rule, we might expect frequent failures after about ten months of operation-clearly not acceptable for most process plants Thus, pay attention to DN values when deciding how and
when grease lubrication is applied or reapplied
Marginal Lubrication: A Factor in Pump Failures
As indicated earlier, a pump manufacturer will occasionally supply pumps with inherent design vulnerabilities In fact, we must assume that the designs of at least some pump manufacturers who give bearing temperature concerns as the reason for disliking lightly preloaded or high-contact angle bearings are really suffering from marginal lubrication In one instance, we found a manufacturer supplying pumps with excessive bearing inner race interference fit Had he chosen to furnish pre-
loaded bearings as well, early bearing distress would have been virtually certain
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A Radial ball bearings
B Cylindrical roller bearings needle roller bearings
C Spherical roller bearings, taper roller bearings, thrust ball bearings
Figure 3-48 Relubrication interval (Courtesy SKF Industries, King of Prussia, PA.)
A close comparison of the bearing housings of problem pumps with those of
pumps with low failure frequency can be quite revealing Low bearing failure rates are reported for the execution shown in Figure 3-49 These axially preloaded bear- ings are routinely used by a German manufacturer As shown, they elected to
achieve the proper preloading by selectively different dimensioning the width of spacers “A” and “B.” The pump manufacturer can thus control the preload by mak- ing appropriate adjustments Flinger disc “C” tosses lube oil onto the surrounding surfaces and from there it flows into trough “D’ and on towards both inboard and outboard bearing locations
The periphery of flinger “C” dips into the lube oil level; however, the lube oil level is generally maintained well below the center of the lowermost ball This reduces oil churning and friction-induced heat-up of lube oil and bearings that would
be more likely to occur in the design shown in Figure 3-50
There is a good reason to introduce the lube oil between the two back-to-back ori- ented angular contact bearings In this design, the cage inclination promotes through- flow of lubricant, whereas many other designs attempt to introduce the lube oil at points that oppose through-flows Figure 3-5 l a allows us to see how the cage incli- nation of back-to-back mounted angular contact bearings with steeper angles pro- motes a centrifugal outward-oriented flinging action from side “a” to side “b.” Con-
versely, if conventionally lubricated angular contact ball bearings are back-to-back mounted as shown in Figure 3-51b, lubricant flow may become marginal or insuffi- cient Subject to proper selection and utilization of proper installation procedures,
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Figure 3-49 Successful pump bearing housing providing optimum lubrication (Cour-
tesy KSB Pump Co.)
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b
a l : Skew o f ball seoarator
b l : Back-to-Back Mounted Angular c ) : Face-to-Face Mounted Angular
Figure 3-51 Angular contact bearings used for thrust takeup in centrifugal pumps
(Courtesy SKF Bearing Co.]
face-to-face mounting (Figure 3-5 IC) may be advantageous in those instances where lubricant flow needs to be improved It clearly promotes through-flow of lube oil and
is one of the reasons why the manufacturer of the heavy duty pump shown in Figure 3-52 opted for face-to-face orientation Oil spray generated by the flinger disc will flow in the preferred direction through the adjacent bearing and on to the next one Note, however, that this presupposes that the temperature difference between inner and outer races is minimal If the temperature were substantial, growth of the inner ring would force the bearing into a condition of high axial preload
Two additional illustrations, Figure 3-53 and 3-54, show oil rings inserted in trapezoidal ramps, which allow lube oil to move towards the bearing internals It should be noted that both of these illustrations violate API-610 8th Edition (August 1995) Paragraph 2.9.1.3, which requires bearings to be directly mounted on the shaft and disallows bearing carriers Also, snap rings and spring-type washers are not permitted Figures 3-55 and 3-57 show executions that facilitate lube oil flow toward the bearings
However, oil rings as shown in Figures 3-53, 3-54, and 3-57 to 3-59 have a ten- dency to malfunction unless the equipment centerline is aligned and installed truly parallel and horizontal This true alignment is difficult to achieve, since the machin- ist or millwright will usually place shims under either the outboard or inboard legs of the equipment while performing driver-to-driven alignment in the field It is for this reason that flinger disks represent the preferred configuration
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Figure 3-52 Oil spray generated by the flinger disc will easily flow through this face-to- face oriented bearing set (Courtesy Ochsner Kreiselpumpen Co Linz, Austria.)
Figure 3-56 shows how the oil delivery capability of oil rings varies with shaft speed Grooved rings are demonstrably superior to the traditional flat rings and should be applied whenever the less vulnerable flinger disk, shown in Figures 3-50 and 3-55, is not available
The Goulds Pump Company has experimented with different executions and found that keeping the oil level below the rotating elements results in lower bearing temperatures than would be achieved if lube oil were to reach to the center of the lowermost ball Goulds explained that their optimum design, Figure 3-59, evolved from the one shown in Figure 3-58, which incorporated a cupped oil flinger design The throw-off action of the cupped disks did not differ significantly from that obtainable with plain disks Moreover, using two separate disks occasionally resulted
in incorrect installation by inexperienced repair shops The manufacturer corrected this situation by using the single spool or spacer piece in Figure 3-59 for oil ring containment,
Goulds established a satisfactory working window for the design using IS0 Grade
68 lube oils and shaft diameters in the 2.5-in range Adequate oil flow existed with
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Figure 3-53 Oil rings inserted in trapezoidal ramps allow lube oil to move freely toward the bearing internals (Courtesy Byron Jackson Pump Co.)
I
Figure 3-54 Oil rings guided in trapezoidal slots facilitate lubricant migration toward bearing (Coun!esy Byron Jackson Pump Co.)
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Figure 3-55 Oil droplets flung against the top of a bearing housing have a chance of falling back on the shaft and migrating toward the bearing internals
Figure 3-56 Oil delivery vs shaft speed for different oil rings
oil temperatures purposely induced to range from 65°F to 220°F at speeds from 880
rpm to 3,550 rpm
This company also experimented with redesigned power ends for their line of pumps and found that a three-fold increase in oil sump capacity resulted in signifi- cantly reduced oil temperatures The value of the attendant increase in bearing life
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Figure 3-57 Oil ring and shaft sleeve outer contours are in line with bearing inner ring surface This facilitates oil flow toward bearing interior
Figure 3-58 Cupped flinger design found in some older pumps (Courtesy Goulds Pumps.)
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Figure 3-59 Optimum design for Goulds pumps A single flinger spool serves to retain two oil rings Oil pumped up by the oil rings contacts the flinger and is sprayed into the
housing walls
will be discussed in Chapter 5, Life-Cycle Costing Figure 3-60 gives a comparison between bearing submergence in oil, operating temperature, and bearing life for the old and new designs
Applying Roller Bearings in Centrifugal Pumps
Figure 3-61 shows a cylindrical roller bearing on the inboard side of the bearing housing A tapered roller bearing is used on the coupling side of a fan-equipped pump, Figure 3-62
It is certainly possible to extend bearing life by using cylindrical and/or tapered
roller bearings in many pump models Figures 3-63 and 3-64 illustrate this for bear- ings that can be used in the same bearing bracket Note that this manufacturer uses the oil ring (Figure 3-61) only to maintain a uniform oil temperature throughout the entire sump The oil ring is not expected to pump oil into the bearings, although it would probably do so in the stepless shaft-to-bearing-inner-ring geometry shown in Figure 3-62 Adequate lubrication is assured by maintaining oil levels to the center
of the lowermost rolling element At a speed of 1,750 rpm and an axial load of 12,000 Newtons, the angular contact bearing (Figure 3-61) will have a projected L-
10 life of 18,500 hours versus 90,000 hours for the tapered roller bearing shown in Figure 3-60 Many of the latter bearings are rated for long life, as indicated in the pump bearing life vs axial thrust vs speed graphs illustrated in Figures 3-63 and
3-64 Needless to say, these extended life figures are meaningless unless bearings are properly installed and unless clean, adequate lube oil flow is assured at all times The overwhelming majority of bearings fail due to lube oil deficiencies, including contamination, or due to skidding (see page 172), caused by light loads
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However, while roller and tapered roller bearings may present a viable alternative
in selected instances, they do have potential drawbacks when used in centrifugal pumps: First, proper alignment of the shaft is much more critical to ensure proper roller-to-race contact and minimize stresses Spacing or shimming between housing end plate and housing must be quite accurate This assembly method is used to ascertain proper roller-to-race contact (i.e., not too much end-play in the shaft) with-
out preload Tapered roller bearings also have much higher internal friction and thus
will not run as fast and as cool as comparable ball bearings Therefore, proper lubri-
cation, i.e., proper type of lubricant, proper amount, and proper application method
will be even more important with roller and tapered roller bearings than with ball bearings
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How Much Oil Is Enough?
The MRC Bearing Division of SKF USA calculates the theoretical oil flow required for cooling from the expression:
.000673 x Fr x P x PD x rpm
H, x (To -Ti)
where Fr = coefficient of friction referred to PD
= pure thrust ball bearings
= angular contact ball bearings
.00076 = cylindrical roller bearings
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P = imposed equivalent load, lb
PD = pitch diameter, in
To = outlet oil temperature, O F
Ti = inlet oil temperature, “F (To - Ti, generally about 50°F)
H, = specific heat of oil in Btu/lb/”F (usually -46-.48)
rpm = operating speed
= .195 + .000478 (460 + Ti)
Conversion: (lbs of oilhin) x (0.135) = gallmin
From this equation it would seem that larger quantities of oil are needed than could reasonably be expected from either oil ring or flinger methods This discrepan-
cy takes on several orders of magnitude if we realize that with oil-mist lubrication three rows of bearings with a 3-in bore diameter use 3 grams of lube oil per hour and can operate for years
Observing the bearing manufacturer’s coefficient of friction numbers we note that the value for angular contact bearings exceeds the one for radial ball bearings by a fac- tor of 1.5 It stands to reason that angular contact thrust will either require more cool- ing oil than pure radial bearings, or will run warmer than pure radial bearings Also, getting the right amount of oil to an angular contact thrust bearing will both be more critical and more difficult than getting the right amount to a typical radial bearing
Bearing Selection Can Make a Difference
API-610, 8th Edition, requires ball-type thrust bearings, if used, to be dual single row, 40” (0.7 radian), light preload, angular contact type (7000 series), installed back-to-back A lot of controversy revolves around this specification clause, with some bearing manufacturers expressing concern that the term “light preload” does not adequately quantify the desirable preload Also, users occasionally report more failures with preloaded bearings than with conventional bearings However, neither observation has presented a dilemma to this writer Here is why
The overall intent of providing preload is to prevent axial shuttling of the rotor and skidding of the rolling elements in a bearing Skidding can be extremely detri-
mental to rolling element bearings and we have often observed the unloaded half of
a duplex-mounted bearing generate more heat and fail before the loaded half showed any distress However, using a preloaded bearing may mandate lowering the custom- ary interference fit between shaft and bearing inner ring Unless this is done, addi- tional heat may not be carried away by certain oil ring and/or flinger arrangements often found in centrifugal pumps Also, if marginal lubrication was provided to begin with, preloading may be “the straw that breaks the camel’s back.” It is in those instances that flush ground bearings may offer adequate operating preload to assure quiet operation, minimize ball skidding, and reduce the risk of thermal runaway The introduction of PumpacO thrust bearings by the MRC Company has allowed many users to extend pump mean-time between failure (MTBF) MRC’s thrust bear-
Trang 20Machinery Reliability Audits and Reviews 173 ing system eliminates skidding by mounting a 40” angular contact bearing back-to-
back with a 15” angular contact bearing
Air Cooling Provisions for Bearing Housings-How Good?
As explained later in this text (see pages 434-440), it can be demonstrated that cooling water can be deleted from virtually all centrifugal pumps with rolling ele- ment bearings It was shown that sizeable maintenance cost credits resulted from this deletion and that these cost credits could be attributed to several factors:
* Cooling water pipes did not have to be maintained
* Utilities requirements and the attendent operating costs were reduced
Bearing failure incidents decreased substantially
The last observation was most striking because it was perhaps least e ~ p e c t e d ~ ~ , ~ ~ A
decrease in bearing failure incidents can be explained by reduced water contamination and by lowered risk of incurring distortion of bearing races Contamination originates with cold cooling water, which promotes condensation of water vapors contained in the oil/air mixture inside the bearing housing Distortion comes from non-uniform cooling through water jackets These are sometimes partially surrounding the bearing outer race and can force the bearing to assume an out-of-round shape
Several pump manufacturers have implemented air cooled bearing housings in efforts
to provide a suitable temperature environment They often execute the bearing housings with cast-in cooling fins and advertise that no fan is required For high-temperature pumpage, typically 500°F and higher, a shaft-mounted fan is often used We found some of these fans thoroughly engineered for low noise and high efficiency Figure 3-62 depicts a typical example of an intelligently engineered fan In contrast, the “fan” offered by another manufacturer (Figure 3-65, item 375) leaves much to be desired
bed
Figure 3-65 “Bearing cooling fan” on a centrifugal pump
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However, we are certainly not ready to give a blanket endorsement to all fan- cooled bearing housings Air-cooled bearing housings can cause bearing distress by allowing uneven cooling of bearing rings The inner ring is mounted on the shaft and the combined assembly represents a poor heat sink In comparison, the bearing hous- ing is a relatively effective heat sink, especially if cooled externally The heat transfer rate is influenced by the properties of the housing material, by housing geometry and
by the temperature difference between pumpage and external ambient conditions Fan cooling is forced convection and thus increases heat transfer If inaterial properties and housing geometry are assumed constant, cooling the housing by any of several possible methods will increase the rate of heat transfer and thus cools the bearing
outer rings Very little heat is transferred through rotating elements from the inner ring The inner ring, therefore, runs hotter than the outer ring
The result, of course, is differential thermal growth, with the inner ring expanding more than the outer ring If bearings are initially flush ground (Le., no preload and
no end-play) or ground for preload, cooling the housing creates or, respectively, increases radial and axial preload and negative clearance exists The resulting tem- perature excursion may or may not be self-limiting In any event, there would now
be increased demands on the lubricant to effectively prevent metal-to-metal contact (See also Figure 3-47)
Stuffing Box Cooling Is Not Usually Effective
Many pumping services require that the mechanical seal environment be kept at moderate temperatures This is generally not difficult to achieve if external flush injection is used In this case, a flush cooler can perform the task, but, of course, at some utility expense (fan power in air-cooled systems, cooling water in conventional heat exchanger circuits) The cost of recirculating the flush fluid may have to be added as well
Another option for achieving a moderate seal environment is stuffing box cooling
In conventional pumps, Figure 3-66, the stuffing box cavity (“A”) is rather remote
from the seal faces (“B”) that we wish to cool An experiment conducted around
1970 showed a disappointing 1” to 2°F decrease at the seal faces when cooling water was introduced into a previously empty stuffing box jacket
A superior design, from the point of view of effective cooling, is shown in Figure 3-67 The manufacturer recognized that heat migration from the casing is primarily responsible for elevated stuffing box temperatures He, therefore, designed the pump with an air gap “A” ahead of the cooling water cavity “B.” Equally important is the fact that the throat bushing “C” is made extremely long and that cavity “B” contacts the throat bushing over a good portion of this length It should be intuitively evident, however, that this configuration will lead to directionally higher L/D ratios than competing designs Accordingly, shaft deflections must be compensated by control- ling the forces acting radially on pump impellers and by more careful design of com- ponent clearances
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B
Figure 3-66 Conventional stuffing box cooling is quite ineffective
environment
for control of the seal
Figure 3-67 Effective control of mechanical seal environment is achieved in pumps with this cooling jacket execution (Courtesy Sulzer-Weise, Bruchsal, Germany.)