SYSTEM ONE
IMPROVING PUMP RELIABILITY
R. Antkowiak
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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 pump over
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 cost $4000 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 $2000 /year over the life
of the pump
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WHY PUMPS AND SEALS FAIL
MECHANICAL
Affects Bearings, Seals and Shafts
-EXTERNAL
1. Operation off the BEP
2. Coupling Misalignment
3. Insufficient NPSH
4. Poor Suction and Discharge
Piping Design
5. Pipe Strain / Thermal Expansion
6 Impeller Clearance
7. Foundation and Baseplate
-INTERNAL
ENVIRONMENTAL
Affects Wet End Components,
Bearings and seals
1. High Temperature
2. Poor Lubrication
/ Oil Contamination
3. Corrosion
4. Erosion
5. Abrasion
1. Pump Design and Manufacturing
Tolerances
2. Impeller Balance (Mechanical and
Hydraulic)
3. Mechanical Seal Design
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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, cooling
Misapplication
Pump, seal, metallurgy selection
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RADIAL LOAD
Operation of a pump away from the BEP results in higher
radial loads ...
creating vibration and shaft deflection
H
E
B.E.P
A
D
FLOW
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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 head req’d
Buildup in pipes
Filter plugged
Automatic control valve shuts off pump flow
Change in viscosity of fluid
Parallel operation problems (starving one pump)
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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 WOULD
OPERATE 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 OF
PERFORMANCE AND PROPORTIONS REGARDLESS
OF SIZE OR SPEED
 FUNCTION OF IMPELLER PROPORTIONS
 SPEED IN RPM AT WHICH AN IMPELLER WOULD
OPERATE IF REDUCED PROPORTIONALLY IN SIZE
TO DELIVER 1 M3/h AND TOTAL HEAD OF 1 M
 DESIGNATED BY SYMBOL Ns
Ns = RPM(M3/h) 1/2
M 3/4
RPM = SPEED IN REVOLUTIONS / MINUTE
M3/h = CUBIC METERS PER HOUR AT BEST EFF. POINT
MH = HEAD IN METERS AT BEST EFF. POINT
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HEAD
HEAD, POWER
EFFICIENCY
HEAD, POWER
EFFICIENCY
HEAD, POWER
EFFICIENCY
PUMP TYPE VS. SPECIFIC SPEED
EFFICIENCY
POWER
CAPACITY
CAPACITY
CAPACITY
CENTRIFUGAL
VERTICAL TURBINE
AXIAL FLOW
SPECIFIC SPEED, ns (Single Suction)
10
20
40
60
120
200
300
500
1,000
2,000
3,000
6,000
10,000
15,000
RADIAL-VANE
FRANCIS-VANE
MIXED FLOW
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SI
US
AXIAL FLOW
RADIAL FORCES ON IMPELLER
RADIAL LOAD
BEP
CUTWATER
125%
BEP 100%
FLOW
50%
% CAPACITY of
BEP
150%
SHUTOFF 0%
Length of Line = Force
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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 3500 RPM, there will be 3500 pulses
per minute applied to the shaft and
bearing
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MISALIGNMENT
MAY BE CAUSED BY:
 Pipe strain
 Thermal growth
 Poor foundation / baseplate
 Improper initial alignment
 System vibration / cavitation
 Soft foot on motor
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NET POSITIVE SUCTION HEAD
(NPSH)
One of the more difficult characteristics to understand
In simplistic terms:
Providing enough pressure in the pump suction to
prevent vaporization of the fluid as it enters the
eye of the impeller
Two values to be considered:
NPSH available
Amount of pressure (head) in the system due to
atmospheric or liquid pressure, height of suction
tank, vapor pressure of the fluid and friction loss
in the suction pipe
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NPSH
cont.
NPSH required
Pressure reduction of the fluid as it enters the
pump
Determined by the pump design
Depends on impeller inlet, design, flow, speed
and nature of liquid
NPSH available 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 the
NPSH required
Occurs when the pressure at any point inside the
pump 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
region of higher pressure is reached where they
collapse or implode with tremendous shock on
the adjacent wall
Sudden rush of liquid into the cavity created by
the collapsed vapor bubbles causes mechanical
destruction (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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RELATIVE PRESSURES
IN THE PUMP SUCTION
E
D
A B C
FRICTION
INCREASING
PRESSURE
DUE TO
IMPELLER
INCREASING
PRESSURE
POINT OF LOWEST
PRESSURE WHERE
VAPORIZATION STARTS
ENTRANCE
LOSS
TURBULENCE,
FRICTION,
ENTRANCE
LOSS
AT VANE TIPS
A
B
C
D
POINTS ALONG LIQUID PATH
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E
NET POSITIVE SUCTION HEAD
AVAILABLE
Hf
PAtmospheric
(friction in suction
pipe)
Z
NPSH Available = P Atm. - Pvap. pressure - Z - Hf
Correct for specific gravity
All terms in “feet (meters) absolute”
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Results of Operating Off BEP
High Temp. Rise
Low Flow Cavitation
Head
Low Brg. & Seal Life
DischargeRecirculation
Reduced Impeller Life
Suction Recirculation
BEP
Low Brg. & Seal Life
Cavitation
Flow
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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 fluid
S.H.
= Specific heat of fluid
Ex.: Pump w/ 100HP (75KW) @s.o. , 6.8 gal casing (.03m3)
w/ 60oF (16oC) water 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
ROTATION
10 inches
250 mm
COEFFICIENT OF THERMAL EXPANSION FOR 316 S/S
IS 9.7X10-6 IN/IN/°F OR 17.5 X10-6 MM/MM/°C
CALCULATION IS
T x 9.7 X10-6 X LENGTH IN INCHES
T x 17.5X10-6 X LENGTH IN MILLIMETERS
T° F
100 F
200 F
300 F
400 F
500 F
600 F
T° C
55 C
110 C
165 C
220 C
275 C
330 C
INCHES
0.0097 IN
0.0190 IN
0.0291 IN
0.0388 IN
0.0485 IN
0.0582 IN
EXPANSION
MILLIMETERS
0.245 MM
0.490 MM
0.735 MM
0.900 MM
1.230 MM
1.470 MM
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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) add’l 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 # 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 designs
made 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, but not
for pumps
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THRUST BEARING SNAP RING
Thrust bearings in standard ANSI pumps are held
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
Radial Thrust
due to Impeller
and Misalignment
Impeller Radial Thrust
Impeller Axial
Thrust
Coupling
Hydraulically
Induced
Forces due to
Recirculation
& Cavitation
Seal
Axial Load
from Misalignment
and Impeller
Radial Thrust
due to Impeller
and Misalignment
Hydraulic
Imbalance
Motor
8/00
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 shaft
It 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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PUMP FAILURE ANALYSIS
6 month period in a typical process plant
C AUS E
NUM B E R
% o f T O T AL
25
1 0 .5 0
B e a r in g h o u s in g
1
0 .4 2
C a s e w e a r in g r in g
Im p e lle r
2
8
0 .8 4
3 .3 6
R o ta tin g fa c e
1
0 .4 2
S c r e w s /s e t s c r e w s
1
0 .4 2
S e a ls - m e c h a n ic a l
179
7 5 .2 1
12
5 .0 4
S le e v e
9
3 .7 8
T O T AL
238
1 0 0 .0 0 %
B e a r in g
S h a ft
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OPTIMUM PUMP DESIGN
OBJECT:
Create a better environment and
greater stability for the dynamic
pump components (seals and
bearings) ….to withstand the
damaging forces inflicted upon them
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SHAFT STIFFNESS
500 Lbs.
(225Kg)
500 Lbs.
(225Kg)
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Derivation of Stiffness Ratio
= Deflection of shaft
P = Load
I = Moment of Inertia
E = Modulus of Elasticity
L = Length of Overhang
P
= PL3
3EI
= PL3
3E P D4
64
L
I=  D4
64
= L3
D4
D
cancel all common factors
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Stiffness Ratio Examples
D
L
D
L
1.50"
8"
L /D = 8 /(1.50) = 512/5.06 = 101
1.62"
8"
L3/D4 = 8 3/(1.62)4 = 512/6.89 = 74
1.75"
8"
L /D = 8 /(1.75) = 512/9.38 = 55
1.87"
8"
L /D = 8 /(1.87) = 512/12.23 = 42
3
4
3
4
3
4
3
4
3
4
3
4
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Stiffness Ratio Examples
D
L
D
1.87"
1.87"
L
8"
6"
3
4
3
4
3
4
3
4
L /D = 8 /(1.87) = 512/12.23 = 42
L /D = 6 /(1.87) = 216/12.23 = 17
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Stiffness Ratio Examples
D
D
L
L
38mm 200mm
4
3
4
3
4
3
/ 45
3
4
3
4
L /D = 200
3
4
3
/ 38
= 8000000/2085136 = 3.84
40mm 200mm
4
3
L /D = 200 / 40
= 8000000/2560000 = 3.13
45mm 200mm
L /D = 200
48mm 200mm
L /D = 200
/ 48
4
= 8000000/4100625 = 1.95
= 8000000/5308416 = 1.51
L/D<2.0 is Adequate
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Stiffness Ratio Examples
D
D
48mm
L
L
200mm
48mm 150mm
3
L /D4 = 200
3
4
/ 48
= 8000000/5308416 = 1.51
4
3 4
L /D = 150 3 / 48
= 3375000/5308416 = .64
L/D < 2.4 Considered Adequate
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LD PUMPS REDUCE BEARING LOADS
A
100 Lbs.
6 in.
8 in.
A = Radial load on thrust bearing
B = Radial load on radial bearing
100 lb. = Impeller radial load on end of shaft
Standard
ANSI Pump
M A=0=14(100)-6B 1400=6B
M B=0= 8(100)-6A 800=6A
B
B=233 lbs.
A=133 lbs.
LD PUMP
A
100 Lbs.
6 in.
5 in.
B
M A=0=11(100)-6B 1100=6B
M B=0= 5(100)-6A 500=6A
B=183 lbs.
A= 83 lbs.
• Radial Bearing
233 lbs. To 183 lbs.
22% Reduction in Load
2.1 x Improvement in Life
• Thrust Bearing
133 lbs. To 83 lbs.
37% Reduction in Load
4 x Improvement in life
Bearing rating life varies inversely as the cube of the applied load
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LD PUMPS REDUCE BEARING LOADS
(Metric)
A
45.4. Kg
152 mm
203 mm
A = Radial load on thrust bearing
B = Radial load on radial bearing
45.4 Kg = Impeller radial load on end of shaft
Standard
ANSI (DIN/ISO) Pump
M A=0=355(45.4)-152B 16,117=152B
M B =0= 203(45.4)-152A 9,216=152A
B
B=106 Kg
A=61 kg
LD PUMP
A
45.4 Kg
152 mm
127 mm
B
12,667=152B
M A=0=279(45.4)-152B
M B =0= 127(45.4))-152A
5,766=152A
B=83 Kg
A= 38 Kg
• Radial Bearing
106 Kg To 83 Kg
22% Reduction in Load
2.1 x Improvement in Life
• Thrust Bearing
61Kg To 38 Kg
37% Reduction in Load
4 x Improvement in life
Bearing rating life varies inversely as the cube of the applied load
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MAXIMUM STIFFNESS RATIO
L3 / D4 RATIO
Less than 60 (Inch)
Less than 2.4 (Metric)
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Length (L)
Dia. (D)
20
22
25
30
32
35
38
42
45
150 155 160
21,1 23,3 25,6
14,4 15,9 17,5
8,6 9,5 10,5
4,2 4,6 5,1
3,2 3,6 3,9
2,2 2,5 2,7
2
1,6 1,8
1,1 1,2 1,3
1
0,8 0,9
165
28,1
19,2
11,5
5,5
4,3
3
2,2
1,4
1,1
170
30,1
21
12,6
6,1
4,
7
3,3
2,4
1,6
1,2
175
33,5
22,9
13,7
6,6
5,1
3,6
2,6
1,7
1,3
25,9
14,9
7,2
5,6
3,9
2,8
1,9
1,4
16,2
7,8
6
4,2
3
2
1,5
8,5
6,5
4,6
3,3
2,2
1,7
7,1
7,6
5,3
3,8
2,6
2
6,2
4,4 4,8 5,1
3
3.2 3,4 3,7 3,9
2,3 2,4 2,6 2,8 3 3.2
1,1 1,19 1,29 1,4 1,51 1,62 1,74 1,87 2, 2,15 2,29 2,44
4,9
3,6
2,4
1,8
5,7
4,1
2,8
2,1
48
0,65 0,70 0,77 0,85 0,92
50
55
0,54 0,60 0,66 0,72 0,79 0,86 0,93 1,01 1,10 1,19 1,28 1,38 1,48 1,59 1,70 1,82 1,95 2,08
0,37 0,41 0,450,49 0,54 0,59 0,64 0,69 0,75 0,81 0,87 0,94 ,01 1,09 1,16 1,24 1,33 1,42
System one LD 17
1
180 185 190 195 200 205 210 215 220 225 230 235
ZONE 1= POOR
>3.2
ZONE 2 = QUESTIONABLE
ZONE 3 = EXCELLENT
ZONE 4 = SUPERIOR
2.4-3.2
1.0-2.4
<1.2
STIFFNESS RATIO CHART - METRIC
8/00
EFFECTIVE PUMP OPERATIONAL ZONES
PUMP CURVE
BEP
ZONE L3/D4
HEAD
INCH
A
B
C
D
A
B
METRIC
C
A
B
C
D
D
80
40
20 10
PERCENT OF BEP
0 10
> 80
60 > 80
26 > 60
< 26
15
20
25
FLO
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> 3.2
2.4 to 3.2
1.0 to 2.4
< 1.0
ALIGNMENT
EVERY TIME A PUMP IS TORN DOWN,
THE MOTOR SHAFT AND PUMP SHAFT
MUST BE REALIGNED
UNPROFESSIONAL OPTION TO RE-ALIGN
…USE A STRAIGHT EDGE
PROFESSIONAL OPTION IS TO USE DIAL
INDICATORSTO MINIMIZE TOTAL
RUNOUT
MODERN METHOD IS LASER
ALIGNMENT WHICH IS VERY ACCURATE
8/00
PRESENT ALIGNMENT METHODS
WEAKNESSES
All provide precision initial alignment
Degree of accuracy varies
Cost of system, training, and time involved
in their use is dramatic
Time consuming (possibly 2 workers, 4-8 hrs.)
Difficult to compensate for high temperature
applications
Requires worker skill, dexterity, and training to
achieve accurate results
After pump startup, cannot insure continued
alignment due to temperature, pipe strain,
cavitation, water hammer, and vibration
8/00
MOTOR ADAPTER - WHAT IS IT?
Machined component that connects a pump
power end to “C” face (D flg.) motor thru
close tolerance fits on each end
Not a new technology
Used on machine tools and gear boxes
 Operate with highest level of accuracy and
precision
Mechanical seal in a pump is a high precision
component
 Mechanical seal accounts for 75% of pump
downtime
8/00
MOTOR ADAPTER- ADVANTAGES
Provides easy, accurate, and reliable alignment
during operation
Maintains near -laser alignment accuracy despite
pipe strain, cavitation, high temperature, and
vibration
A device that reduces vibration will prolong seal
life and increase pump reliability
Reduces labor hours for initial installation
During teardown, maintenance cycle time is
reduced dramatically
vertical mounting capability
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MOTOR ADAPTER ADVANTAGES cont.
High temperature applications
Motor grows with the pump
More even temperature gradient across the
pump and motor assembly
For high speed (3000/3600 RPM)
applications - Alignment more critical
Disadvantages
Not as accurate as initial laser alignment due
to inherent tolerance stackup of the various
components
8/00
SEAL CHAMBERS
LARGE BORE
• 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
OLD STYLE
• Designed for packing
•Small radial clearances
-Seal contacting bore
•Limited fluid capacity
-Poor heat removal
•Easy to clog with solids
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ELIMINATING SHAFT SLEEVES
 Add no stiffness to shaft
 Runout tolerance between shaft and sleeve compounds
motion of seal faces in addition to deflection and shaft
runout already present
 Deflection must be a maximum of .002” at the seal
faces, yet faces are lapped within 2 helium light bands
Deflection or motion at seal faces is 1000 times
greater than the face flatness
Sleeves are necessary for packed pumps, but with
today’s new seals they serve no purpose
8/00
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 cont.
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 cont.
Face seals and magnetic seals
Protect bearings from possible immersion
Good for moisture laden environment
Expansion chamber should be used to
accommodate changes in internal pressure
and vapor volume
completely enclosed system (can be
submerged)
Generate heat
Limited life
8/00
SYSTEM ONE LABYRINTH SEAL
Stationary Element
LABYRINTH
Traps liquid and
directs it to the gravity
drain
OIL TRAP
and DRAIN
Helps retain
lubrication
in bearing housing
Rotary Element
316SS for corrosion
resistance
GRAVITY DRAIN
Allows liquid
to drain
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BEARING LIFE
 Bearing life calculations assume proper lubrication
and an environment that protects the bearing from
contamination
 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 rev’s 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 impeller
diameters to the 6th power
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BEARING LIFE cont.
 90%
of all bearings will fail prematurely and
not 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 consumption
Lasts 1.5-2 times longer than conventional oils
• Maintains lube properties with water
contamination better than mineral oils
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BEARING CLEARANCES
“C3”
Normal clearance
6310 Radial Bearing (microns)
Radial .0003-.0011”(9-30)
Axial
.0016-.003” (48-90)
.0009-.0017”(27-51)
.0016-.003”(48-90)
5310 Double Row Thrust Bearing
Radial .0005-.0014”(15-42)
Axial
.0005-.0014”(15-42)
.0014-.0020”(42-60)
.0014-.0023”(42-69)
7310 Angular Contact Thrust Bearing
Axial
-.0003 to +.0003” (line to line)
NOMINAL “0”
Radial
approx .85 x Axial
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ANGULAR CONTACT BEARINGS
 Used as thrust bearing in pairs (also carry radial load)
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 internal
clearance
Thrust bearings can be either positive or
negative clearance ( 5310 vs. 7310 pr.)
Preload occurs when there is a negative
clearance in the bearing
Desirable to increase running accuracy
Enhances stiffness
Reduces running noise
Provides a longer service life under proper
applications
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BEARING CLEARANCES / PRELOAD
LIFE
Preload
Clearance
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MICROMETER IMPELLER
ADJUSTMENT
Micrometer adjusting nut allows impeller to be
set to precise clearance from the front of the
casing
Each line on the adjusting nut is a .003”
(.08mm) graduation for axial movement of the
shaft
Normal setting is .015” (.38mm) from the
casing face
For every 50 deg. above 100 deg. fluid temp...
add .002” clearance
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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
 “C” Frame motor adapter
 Centerline support for hot applications
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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 than NPSH
required
 Proper foundation and baseplate arrangement
 Absolute minimum pipe strain on suction and discharge flanges
 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)
 Shaft end play less than .0005” (.015mm)
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SYSTEM ONE PUMP WARRANTY
ONE YEAR FOR MECHANICAL SEAL
– SPARE SEAL KIT OR REBUILT SEAL OFFERED

FIVE YEARS ON SYSTEM ONE POWER END
– ANY FAILURE INCLUDING BEARINGS
– FREE REPLACEMENTS OF FAILED COMPONENT

SHAFT WARRANTIED FOR LIFE ON FRAME
S AND A PUMPS
free replacements are one time only
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