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PUMP RELIABILITY
IMPROVING PUMP RELIABILITY
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Maintenance vs. Capital
What does a pump actually cost ?
Most plants regard the pump as a commodity...
purchased from the lowest bidder with little
consideration for:
The operation and maintenance cost of the pumpover its life cycle... which could be 20 - 30 years
Costs to be considered:
Spare parts (inventory costs)
Operation downtime (lost production) Labor to repair (maintenance costs)
Power consumption based on pump
efficiency
Environmental, disposal, and recycle costs
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TRUE PUMP COSTS
Repair costs can easily exceed the price of a
new pump (several times) over its life of 20 -
30 years
Documented Pump failures costRs.2.00.000/- or more per incident ( parts and
labor)
If MTBF was improved from 1 to 2 years for a
pump in a tough application Results in savings of Rs. 1,00,000/- per
year over the life of the pump
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WHY PUMPS AND SEALS FAIL
MECHANICALAffects Bearings, Seals and Shafts
-EXTERNAL1. Operation off the BEP
2. Coupling Misalignment
3. Insufficient NPSH4. Poor Suction and Discharge
Piping Design
5. Pipe Strain / Thermal Expansion
6 Impeller Clearance
7. Foundation and Baseplate
-INTERNAL1. Pump Design and Manufacturing
Tolerances
2. Impeller Balance (Mechanical and
Hydraulic)
3. Mechanical Seal Design
ENVIRONMENTALAffects Wet End Components,
Bearings and seals
1.High Temperature
2. Poor Lubrication
/ Oil Contamination
3. Corrosion
4. Erosion
5. Abrasion
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HOW ARE FAILURES INITIATED?
Installation Piping system & Pipe Strain
Alignment
Mechanical Seal installation
Foundation
Operational System: cavitation, dry running, shutoff
Product changes: viscosity, S.G., temp.
Seal controls: flush, coolingMisapplication
Pump, seal, metallurgy selection
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RADIAL LOAD
Operation of a pump away from the BEPresults in higher radial loads ...creating vibration and shaft deflection
H
E
A
D
FLOW
B.E.P
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Radial Forces
By design, uniform pressures exist around the
volute at the design capacity (BEP)
Resulting in low radial thrusts and minimal
deflection. Operation at capacities higher or lower than
the BEP
Pressure distribution is not uniform resulting in
radial thrust on the impeller
Magnitude and direction of radial thrust
changes with capacity (and pump specific
gravity)
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SHAFT DEFLECTION
Most pumps do not operate at BEP: Due to improper pump selection (oversized)
Changing process requirements (throttling)
Piping changes
Addition of more pipe, elbows and valves
System head variations
Change in suction pressure, discharge headreqd
Buildup in pipes
Filter pluggedAutomatic control valve shuts off pump flow
Change in viscosity of fluid
Parallel operation problems (starving one pump
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Impeller Radial Force
At Any Flow F(lbs.,Kg)
D
B
K= THRUST FACTOR
H = HEAD (ft, m)
S = SPECIFIC GRAVITY
D= IMPELLER DIAMETER (in.,cm)
B = IMPELLERWIDTH (in., cm)
FF == K x H x SK x H x S
2.312.31x D x Bx D x B
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KK
0
0.1
0.2
0.3
0.4
0.5
0 20 40 60 80 100 120 140 160
500500(10)(10)
10001000(20)(20)
15001500(27)(27)
20002000 (40)(40)
35003500(71)(71)
PERCENT CAPACITY
SPECIFIC SPEED - K vs. CAPACITY
Ns (SI)
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PUMP SPECIFIC SPEED
CLASSIFIES IMPELLERS ON THE BASIS OF
PERFORMANCE AND PROPORTIONS REGARDLESS
OF SIZE OR SPEED
FUNCTION OF IMPELLER PROPORTIONS
SPEED IN RPM AT WHICH AN IMPELLER WOULDOPERATE IF REDUCED PROPORTIONALLY IN SIZE
TO DELIVER 1 GPM AND TOTAL HEAD OF 1 FOOT
DESIGNATED BY SYMBOL Ns
Ns = RPM(GPM)1/2
H3/4
RPM = SPEED IN REVOLUTIONS / MINUTE
GPM = GALLONS /MINUTE AT BEST EFF. POINT
H = HEAD IN FEET AT BEST EFF. POINT
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PUMP SPECIFIC SPEED (Metric)
CLASSIFIES IMPELLERS ON THE BASIS OFPERFORMANCE AND PROPORTIONSREGARDLESS OF SIZE OR SPEED
FUNCTION OF IMPELLER PROPORTIONS SPEED IN RPM AT WHICH AN IMPELLER WOULD
OPERATE IF REDUCED PROPORTIONALLY INSIZE TO DELIVER 1 M3/h AND TOTAL HEAD OF 1M
DESIGNATED BY SYMBOL NsNs = RPM(m3/h )1/2
H3/4
RPM = SPEED IN REVOLUTIONS / MINUTEm3/h = CUBIC METERS / HOUR AT BEPH = HEAD IN METERS AT BEST EFF. POINT
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PUMP TYPE VS. SPECIFIC SPEED
SPECIFIC SPEED, ns (Single Suction)
CENTRIFUGALCAPACITY
HEAD,POW
ER
EFFICIENCY
CAPACITY
HEAD,POW
ER
EFFICIENCY
AXIAL FLOW
CAPACITY
HEAD,POW
ER
EFFICIENCY
VERTICAL TURBINE
HEAD
EFFICIENCY
POWER
10 20 40 60 120 200 300
500 1,000 2,000 3,000 6,000 10,000 15,000
SI
US
RADIAL-VANE FRANCIS-VANE MIXED FLOW AXIAL FLOW
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CUTWATER
SHUTOFF 0%Length of Line = Force
50%
BEP 100%
%CAPACITY of
BEP
125%
150%
FLOWR
ADIALLOAD
BEP
RADIAL FORCES ON IMPELLER
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THE IMPORTANCE OF ALIGNMENT
Any degree of misalignment between the
motor and the pump shaft will cause
vibration in the pump.
Every revolution of the coupling places a
load on the pump shaft and thrust bearing
At 2900 RPM, there will be 2900 pulses per
minute applied to the shaft and bearing
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MISALIGNMENT
Pipe strain
Thermal growth
Poor foundation / base plate
Improper initial alignment.
System vibration / cavitation
Soft foot on motor
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NET POSITIVE SUCTION HEAD (NPSH)
NPSH (Net Positive Suction Head)
Pressure in terms of head above vapourpressure at the inlet / eye of the impeller isknown as NPSH (Net Positive Suction Head)
NPSH available
Pressure in terms of head above vapourpressure available at the inlet / eye of the
impeller is known as NPSHA (Net PositiveSuction Head Available)
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NPSH cont.
NPSH required
Pressure in terms of head above vapour
pressure required at the inlet / eye of the
impeller to avoid cavitation is known asNPSHR (Net Positive Suction Head Required)
NPSHavailable must always be > NPSH
required by a minimum of 3-5 feet (1-1.5m)
margin
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CAVITATION
Results if the NPSH available is less than theNPSH required
Occurs when the pressure at any point inside thepump drops below the vapor pressure
corresponding to the temperature of the liquid The liquid vaporizes and forms cavities of vapor
Bubbles are carried along in a stream until a regionof higher pressure is reached where they collapseor implode with tremendous shock on the adjacentwall
Sudden rush of liquid into the cavity created by thecollapsed vapor bubbles causes mechanicaldestruction (cavitation erosion or pitting)
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CAVITATION cont.
Efficiency will be reduced as energy is
consumed in the formation of bubbles
Water @ 70oF (20oC)will increase in volume
about 54,000 times when vaporized Erosion and wear do not occur at the point of
lowest pressure where the gas pockets are
formed, but farther upstream at the point
where the implosion occurs Pressures up to 150,000 psi have been
estimated at the implosion (1,000,000 Kpa)
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E
A B CD
TURBULENCE,
FRICTION,ENTRANCE
LOSS
AT VANE TIPS
INCREASINGPRESSURE
DUE TO
IMPELLER
A B C D E
ENTRANCE
LOSS
FRICTION
INCREASING
PRESSURE
POINTOFLOWEST
PRESSUREWHERE
VAPORIZATIONSTARTS
POINTS ALONG LIQUID PATH
RELATIVE PRESSURES IN THE PUMP SUCTION
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(friction in suction
pipe)
Hf
Z
PAtmospheric
NPSHAvailable = P Atm. - Pvap. pressure - Z - Hf
Correct for specific gravity
All terms in feet (meters) absolute
NET POSITIVE SUCTION HEAD
AVAILABLE
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Results of Operating Off BEP
High Temp. Rise
Head
Head
FlowFlow
BEPBEP
Low Flow Cavitation
Discharge Recirculation
Reduced ImpellerLife
Suction Recirculation
Low Brg. & SealLife
Cavitation
Low Brg. & SealLife
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TEMPERATURE RISE
Overheating of the liquid in the casing can cause:Rubbing or seizure from thermal expansion
Vaporization of the liquid and excessive vibration
Accelerated corrosive attack by certain chemicals
Temperature rise per minute at shutoff is:
(T oF (oC) / min.= HP (KW)so x K
Gal (m3) x S.G. x S.H.
HPso = HP (KW) @ shutoff from curve
Gal. (m3) = Liquid in casing
S.G. = Specific gravity of fluidS.H. = Specific heat of fluid
Ex.: Pump takes100HP (75KW) @s.o. , 6.8 gal casing (.03m3)
water (at 16 deg C) would reach boiling in 2 min.
A recirculation line is a possible solution to the low flow or
shut off operation problems....
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CASING GROWTH
DUE TO HIGH TEMPERATURE
T F T C INCHES MILLIMETERSEXPANSION
100 F 55 C 0.0097 IN 0.245 MM
200 F 110 C 0.0190 IN 0.490 MM300 F 165 C 0.0291 IN 0.735 MM
400 F 220 C 0.0388 IN 0.900 MM
500 F 275 C 0.0485 IN 1.230 MM
600 F 330 C 0.0582 IN 1.470 MM
10 inches
250 mm
ROTATION
COEFFICIENT OF THERMAL EXPANSION FOR 316 S/S
IS 9.7X10-6
IN/IN/F OR 17.5 X10-6
MM/MM/CCALCULATION IS T x 9.7 X10-6 X LENGTH IN INCHES
T x 17.5X10-6 X LENGTH IN MILLIMETERS
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IMPELLER CLEARANCE
Critical for open impellers Normal setting .015 (.38mm) off front cover
High temperature requires more clearance
- Potential rubbing problem causes vibration
and high bearing loads
- Set impeller .002 (.05mm) addl clearance
for every 500 F (280C) over ambient temp.
Important for maximum efficiency
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IMPELLER BALANCE
MECHANICAL
- Weight offset from center of impeller
- Balance by metal removal from vane
HYDRAULIC- Vane in eye offset from impeller C/L
- Variation in vane thickness
- Results in uneven flow paths thru
impeller- Investment cast impeller eliminates
problem
- Careful machining setup can help
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TYPICAL ANSI (or DIN) PROCESS PUMP
Small dia. shaft with excessive overhang
Stuffing box designed for packing
Shaft sleeve
Light to medium duty bearings Rubber lip seals protecting the bearings
Snap ring retains thrust bearing in housing
Shaft adjustment requires dial indicator
Double row thrust bearing
Cast jacket on bearing frame for cooling
Small oil reservoir
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ANSI (ISO/DIN) STANDARD PUMPS
Industry standards for dimensions based on
requirements for packed pumps
Shaft overhang a function of no. of packing rings
and space for gland and repack accessibility Clearance between shaft and box bore based
on packing cross-section
If most pumps today use mechanical seals -
why do we continue to use inferior designsmade for packing ??
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BEARING OIL SEALS
Rubber Lip Seals Provided To Protect Bearings in
standard ANSI pumps
Have life of less than four months
Groove shaft in first 30 days of operation
External contamination causes bearing failure
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LIP SEAL LIFE
AUTOMOBILE
100,000 Miles @ 40 Miles /hr. = 2500 hrs.
of operation
PUMP
24 hrs./day x 365 days / year = 8760 hours
60% of lip seals fail in under 2000 hours
Lip seals may be fine for automobiles, butnot for pumps
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THRUST BEARING SNAP RING
Thrust bearings in standard ANSI pumps areheld in place with a snap ring
Snap ring material harder than bearing housing
Wear in bearing housing results in potential
bearing movement Difficult to remove and install
If installed backwards - potential loose bearing
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SIMULTANEOUS DYNAMIC LOADS
ON PUMP SHAFT
Impeller Axial
Thrust
Impeller Radial Thrust
HydraulicallyInducedForces due to
Recirculation
& Cavitation
Hydraulic
Imbalance
Seal
Radial Thrustdue to Impeller
and Misalignment
Axial Load
from Misalignmentand Impeller
Radial Thrust
due to Impeller
and Misalignment
Coupling
Motor
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SHAFT DYNAMICS
Radial movement of the shaft occurs in 3 forms:
Deflection - under constant radial load in one direction
Whip - Cone shaped motion caused by unbalance
Runout - Shaft bent or eccentricity between shaft sleeve
and shaftIt is possible to have all 3 events occurring simultaneously
ANSI B73.1 and API 610
Limit radial deflection and runout of the shaft to 0.002
T.I.R. at the stuffing box face(0.05mm) Solid shafts are critical for pump reliability
Eliminate sleeve runout
Improved stiffness
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SHAFT DEFLECTION
Shaft deflects because of unbalanced radial loads
on the impeller
Shaft revolves on own centerline even when
deflected
load is constant in direction and magnitude Shaft stays bent as long as operating conditions
remain the same
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Shaft Whip
Shaft changes 180o from its centerline every
revolution
Usually caused by unbalanced impeller
Heavy side of impeller on same side of shaft
Whip and deflection can occur at same time Moved to one side by the amount the shaft
deflects
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OPTIMUM PUMP DESIGN
OBJECT:
Create a better environment andgreater stability for the dynamic
pump components (seals and
bearings) .to withstand thedamaging forces inflicted upon them
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SHAFT STIFFNESS
500 Lbs.(225Kg)
500 Lbs.(225Kg)
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P = Load
E = Modulus of Elasticity
L = Length of Overhang
= PL3 I= 4 D4
3EI 64
= PL3 = L3
3E P D4 D4
64
Derivation of Stiffness Ratio
= Deflection of shaft
I = Moment of Inertia
cancel all common factors
LP
D
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Stiffness Ratio Examples
LL
DD
D L3 4 3 4
1.50" 8" L /D = 8 /(1.50) = 512/5.06 = 101
1.62" 8" L3/D4 = 83/(1.62)4 = 512/6.89 = 74
1.75" 8" L3/D4 = 83/(1.75)4 = 512/9.38 = 55
1.87" 8" L3/D
4= 8
3/(1.87)
4= 512/12.23 = 42
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Stiffness Ratio Examples
D L
LL
DD
1.87" 8" L3/D
4= 8
3/(1.87)
4= 512/12.23 = 42
1.87" 6" L3/D4 = 63/(1.87)4= 216/12.23 = 17
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Stiffness Ratio Examples
LL
DD
D L
38mm 200mm L3/D
4= 200
3/ 38
4= 8000000/2085136 = 3.84
40mm 200mm L3/D
4= 200 3/ 40
4= 8000000/2560000 = 3.13
45mm 200mm L3/D4 = 200 3/ 454 = 8000000/4100625 = 1.95
48mm 200mm L3/D
4= 200
3/ 48
4= 8000000/5308416 = 1.51
L/D
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Stiffness Ratio Examples
LL
DD
D L
48mm 200mm L3/D
4= 200
3/ 48
4= 8000000/5308416 = 1.51
48mm 150mm L
3
/D
4
= 150
3
/ 48
4
= 3375000/5308416 = .64
L/D < 2.4 Considered Adequate
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MAXIMUM STIFFNESS RATIO
L3 / D4 RATIO
Less than 60 (Inch)
Less than 2.4 (Metric)
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FLOFLO
ZONE L3/D4ZONE L3/D4
INCHINCHA > 80A > 80
B 60 > 80B 60 > 80C 26 > 60C 26 > 60
D < 26D < 26
> 3.2> 3.2
B 2.4 to 3.2B 2.4 to 3.2
C 1.0 to 2.4C 1.0 to 2.4
D < 1.0D < 1.0
METRICMETRICAA
1515 25251010 2020
HEAD
HEAD
PUMP CURVEPUMP CURVEBEPBEP
A
B
C
D
001010202040408080PERCENT OF BEPPERCENT OF BEP
EFFECTIVE PUMP OPERATIONAL ZONES
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ALIGNMENT
EVERY TIME A PUMP IS TORN DOWN,THE MOTOR SHAFT AND PUMP SHAFTMUST BE REALIGNED
UNPROFESSIONAL OPTION TO RE-ALIGNUSE A STRAIGHT EDGE
PROFESSIONAL OPTION IS TO USE DIALINDICATORSTO MINIMIZE TOTAL RUNOUT
MODERN METHOD IS LASER ALIGNMENTWHICH IS VERY ACCURATE
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Drawbacks of Present Alignment Methods
All provide precision initial alignment Degree of accuracy varies
Cost of system, training, and time involved in theiruse is dramatic
Time consuming (possibly 2 workers, 4-8 hrs.)
Difficult to compensate for high temperatureapplications
Requires worker skill, dexterity, and training toachieve accurate results
After pump startup, cannot insure continuedalignment due to temperature, pipe strain, cavitation,
water hammer, and vibration
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SEAL CHAMBERS
Designed specifically for seals
20 Times greater fluid volume Provides superior cooling,cleaning,
and lubrication for the seal
Solids centrifuged away from seal
Eliminate seal rub problems
Designed for packing
Small radial clearances-Seal contacting bore
Limited fluid capacity
-Poor heat removal
Easy to clog with solids
OLD STYLELARGE BORE
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ELIMINATING SHAFT SLEEVES
REDUCES SEAL SIZE Sleeves are necessary for packed pumps, but with
todays new seals they serve no purpose
Add no stiffness to shaft
Run out tolerance between shaft and sleeve compoundsmotion of seal faces in addition to deflection and shaftrun out already present
Deflection must be a maximum of .002 at the sealfaces, yet faces are lapped within 2 helium light bands
Deflection or motion at seal faces is 1000 times greaterthan the face flatness
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BEARING OIL SEALS
Three basic types:
Lip seal
Inexpensive, simple to install, very effective
when new Elastomeric construction
Contact shaft and contributes to friction
drag and temp. rise in bearing area
After 2000-3000 hours, no longer provide
effective barrier against contamination
Will groove shaft
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BEARING OIL SEALS
Labyrinth seals
Required by API 610
Non-contacting and non-wearing
Unlimited life Effective for most types of contaminants
Do not keep heavy moisture or corrosive
vapors from entering the bearing frame(especially in static state)
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BEARING OIL SEALS
Face seals and magnetic seals
Protect bearings from possible immersion
Good for moisture laden environment
Expansion chamber should be used toaccommodate changes in internal pressure
and vapor volume
completely enclosed system (can be
submerged) Generate heat
Limited life
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BEARING LIFE
Bearing life calculations assume proper lubricationand an environment that protects the bearing fromcontamination
The basic dynamic load rating C is the bearing
load that will give a rating life of 1 million revolutions L10 Basic Rating Life is life that 90% of group of
brgs. will exceed ( millions of revs or hrs. operation)
Rating Life varies inversely as the cube of the
applied load Reduction of impeller dia. from maximum improves
life calculation by the inverse ratio of the impellerdiameters to the 6th power
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BEARING LIFE cont.
90% of all bearings will fail prematurely andnot reach their rated L10 life
- Calculated life by design over 20 years
- Actual life maybe 3 years
Failures:
-Fatigue due to excessive loads (20-50%
of failure)
-Lube failure - excessive temperatures &
contaminants
-Poor installation
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BEARING LUBRICATION FAILURE
OXIDATION
Chemical reaction between oxygen & oil
New compounds produced which deteriorate the
life of oil and bearings
Reaction rate increases with the presence of water
and increases exponentially with temperature
CONTAMINATION
Water breaks down lube directly reducing brg.
life - .003% water in oil reduces life of oil 50%
Oil life decreases by 50% for every 20oF (11oC)
rise in temp. above 140oF (60oC)
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SYNTHETIC
OILS Lower change in viscosity with temp. change-One synthetic can take place of several oils
Provides good lube at high temps. 300oF (160oC)
-Does not oxidize (breakdown)
At low temps.- good fluidity boosts efficiency and
reduces component wear during cold weather
Achieves full lubrication quickly
Offers longer life - less consumptionLasts 1.5-2 times longer than conventional oils
Maintains lube properties with water
contamination better than mineral oils
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ANGULAR CONTACT BEARINGS
Used as thrust bearing in pairs (also carry radialload)
Mounted back to back (letters to letters)
Provides maximum stiffness to shaft
Avoid ball skidding under light loads
Small preload eliminates potential
Line to line design clearances
Shaft fit provides preload Eliminates shaft end play
Greater thrust capacity
Required by API 610 Specification
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BEARING PRELOAD
Pump radial bearings have positive internalclearance
Thrust bearings can be either positive ornegative clearance.
Preload occurs when there is a negativeclearance in the bearing
Desirable to increase running accuracy
Enhances stiffness Reduces running noise
Provides a longer service life under properapplications
BEARING CLEARANCES /
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BEARING CLEARANCES /
PRELOAD
LIFE
ClearancePreload
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Goal: improved pump and Mechanical seal reliability
Eliminate or reduce mechanical and
environmental influences that cause pump and
seal problems.
Specify proper pump design criteria to minimize
the impact of mechanical and environmental
influences.
Specify proper mechanical seal and
environmental controls to maximize seal life
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Optimum pump design summary
Low L3D4 ratio as possible Solid shaft ( no sleeves)
Large bore seal chamber
Large oil capacity bearing housing
Angular contact thrust bearings Retainer cover to hold thrust bearing (no snap rings)
Fin tube cooling for bearing housing
Labyrinth seals
Positive / precision shaft adjustment method Investment cast impellers
Magnetic drain plugs in oil sump
Centerline support for hot applications
R i t f i i
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Requirements for proper emission
control and maximum seal life
Shaft runout at impeller within .001 T.I.R. (.03mm) Coupling alignment within .005 T.I.R. on rim & face
(.13mm)
Operation of the pump at or close to best efficiency point(definition dependent upon pump size, speed, and LD ratio)
NPSH available to be at least 5 feet (1.5m) greater thanNPSH required
Proper foundation and baseplate arrangement
Absolute minimum pipe strain on suction and dischargeflanges
All impellers dynamically balanced to ISO G 6.3 spec. Face of seal chamber square to shaft within .002 T.I.R.
(.05mm)
Seal chamber register concentric to shaft within .003 T.I.R.(.08mm)
Recommended