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SYSTEM ONE

R. Antkowiak

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 pump over its life

cycle... which could be 20 - 30 yearsCosts 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 applicationResults in savings of $2000 /year over the life

of the pump

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WHY PUMPS AND SEALS FAILMECHANICAL 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 1. Pump Design and Manufacturing Tolerances 2. Impeller Balance (Mechanical and Hydraulic) 3. Mechanical Seal Design

ENVIRONMENTALAffects Wet End Components,Bearings and seals

1. High Temperature2. Poor Lubrication / Oil Contamination3. Corrosion4. Erosion5. Abrasion

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Installation Piping system & Pipe StrainAlignmentMechanical Seal installation Foundation

Operational System: cavitation, dry running, shutoff Product changes: viscosity, S.G., temp. Seal controls: flush, cooling

Misapplication Pump, seal, metallurgy selection

HOW ARE FAILURES INITIATED?

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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

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 impellerMagnitude and direction of radial thrust changes

with capacity (and pump specific gravity)

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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’dBuildup in pipesFilter pluggedAutomatic control valve shuts off pump flowChange in viscosity of fluid Parallel operation problems (starving one pump)

Shaft Deflection

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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 NsNs = RPM(GPM)1/2

H3/4

RPM = SPEED IN REVOLUTIONS / MINUTEGPM = 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 NsNs = 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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PUMP TYPE VS. SPECIFIC SPEED

SPECIFIC SPEED, ns (Single Suction)

CENTRIFUGALCAPACITY

HE

AD

, PO

WE

R

EF

FIC

IEN

CY

CAPACITY

HE

AD

, PO

WE

R

EF

FIC

IEN

CY

AXIAL FLOWCAPACITY

HE

AD

, P

OW

ER

E

FF

ICIE

NC

Y

VERTICAL TURBINE

HEADEFFICIENCY

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%

FLOWRA

DIA

L L

OA

D

BEP

RADIAL FORCES ON IMPELLER

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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

THE IMPORTANCE OF ALIGNMENT

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MISALIGNMENT

Pipe strain Thermal growth Poor foundation / baseplate Improper initial alignment System vibration / cavitation Soft foot on motor

MAY BE CAUSED BY:

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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

pumpDetermined by the pump designDepends 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

A B CD

TURBULENCE, FRICTION, ENTRANCE

LOSS AT VANE TIPS

INCREASINGPRESSURE

DUE TO IMPELLER

A B C D E

ENTRANCE

LOSS

FRICTION

INC

RE

AS

ING

P

RE

SS

UR

E

PO

INT

OF

LO

WE

ST

P

RE

SS

UR

E W

HE

RE

V

AP

OR

IZA

TIO

N S

TA

RT

S

POINTS ALONG LIQUID PATH

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(friction in suction pipe)

Hf

Z

PAtmospheric

NPSH Available = P Atm. - Pvap. pressure - Z - Hf Correct for specific gravityAll terms in “feet (meters) absolute”

NET POSITIVE SUCTION HEAD

AVAILABLE

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Results of Operating Off BEPHigh Temp. Rise

Hea

dH

ead

FlowFlow

BEPBEP

Low Flow Cavitation

Discharge Recirculation

Reduced Impeller Life

Suction Recirculation

Low Brg. & Seal Life

Cavitation

Low Brg. & Seal Life

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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 chemicalsTemperature 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 curveGal. (m3) = Liquid in casingS.G. = Specific gravity of fluidS.H. = Specific heat of fluidEx.: 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

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 MM

300 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 inches250 mm

RO

TA

TIO

N

COEFFICIENT OF THERMAL EXPANSION FOR 316 S/SIS 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) 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 pumpsHave 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

Impeller AxialThrust

Impeller Radial Thrust

HydraulicallyInducedForces due to Recirculation & Cavitation

Hydraulic Imbalance

Seal

Radial Thrustdue to Impellerand Misalignment

Axial Load from Misalignmentand Impeller

Radial Thrustdue 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 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 ANALYSIS6 month period in a typical process plant

CAUSE NUMBER % of TOTALBearing 25 10.50Bearing housing 1 0.42Case wearing ring 2 0.84Impeller 8 3.36Rotating face 1 0.42Screws /set screws 1 0.42Seals - mechanical 179 75.21Shaft 12 5.04Sleeve 9 3.78TOTAL 238 100.00%

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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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P = Load E = Modulus of ElasticityL = Length of Overhang

= PL3 I= 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

L

P

D

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Stiffness Ratio Examples

LL

DD

D L

3 4 3 41.50" 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" L3/D4 = 8 3 /(1.75) 4 = 512/9.38 = 55

1.87" 8" L3/D4 = 8 3 /(1.87) 4 = 512/12.23 = 42

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Stiffness Ratio Examples

D L

LL

D D

1.87" 8" L3/D4 = 8 3 /(1.87) 4 = 512/12.23 = 42

1.87" 6" L3/D4 = 6 3 /(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/D4 = 200 3/ 40

4 = 8000000/2560000 = 3.13

45mm 200mm L3/D4 = 200 3 / 45

4 = 8000000/4100625 = 1.95

48mm 200mm L3/D4 = 200 3 / 48

4 = 8000000/5308416 = 1.51

L/D<2.0 is Adequate

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Stiffness Ratio Examples

LL

DD

D L

48mm 200mm L3/D4 = 200

3 / 484 = 8000000/5308416 = 1.51

48mm 150mm L3/D4 = 150 3 / 48

4 = 3375000/5308416 = .64

L/D < 2.4 Considered Adequate

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LD PUMPS REDUCE BEARING LOADS

A = Radial load on thrust bearingB = Radial load on radial bearing 100 lb. = Impeller radial load on end of shaft

StandardANSI Pump

MA=0=14(100)-6B 1400=6B B=233 lbs.MB=0= 8(100)-6A 800=6A A=133 lbs.

LD PUMP

MA=0=11(100)-6B 1100=6B B=183 lbs.MB=0= 5(100)-6A 500=6A A= 83 lbs.

• Radial Bearing233 lbs. To 183 lbs.

22% Reduction in Load2.1 x Improvement in Life

• Thrust Bearing133 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

100 Lbs.A

B

6 in. 8 in.

100 Lbs.A

B

6 in. 5 in.

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LD PUMPS REDUCE BEARING LOADS(Metric)

A = Radial load on thrust bearingB = Radial load on radial bearing 45.4 Kg = Impeller radial load on end of shaft

StandardANSI (DIN/ISO) Pump

MA=0=355(45.4)-152B 16,117=152B B=106 KgM B=0= 203(45.4)-152A 9,216=152A A=61 kg

LD PUMP

MA=0=279(45.4)-152B 12,667=152B B=83 KgM B=0= 127(45.4))-152A 5,766=152A A= 38 Kg

• Radial Bearing106 Kg To 83 Kg

22% Reduction in Load2.1 x Improvement in Life

• Thrust Bearing61Kg To 38 Kg

37% Reduction in Load4 x Improvement in life

Bearing rating life varies inversely as the cube of the applied load

45.4. KgA

B

152 mm 203 mm

45.4 KgA

B

152 mm 127 mm

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MAXIMUM STIFFNESS RATIO

L3 / D4 RATIO

Less than 60 (Inch)

Less than 2.4 (Metric)

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STIFFNESS RATIO CHART - METRIC

ZONE 2 = QUESTIONABLE 2.4-3.2

ZONE 3 = EXCELLENT 1.0-2.4ZONE 4 = SUPERIOR <1.2

Dia. (D)

5,1

20

22

25

30

32

35

38

42

45

48

50

55

4,7

Length (L)

150 155 160 165 170 175 180 185 190 195 200 205 210 215 220 225 230 23521,1 23,3 25,6 28,1 30,1 33,5

15,9 17,5 19,2 21 22,9 25,9

8,6 9,5 10,5 11,5 12,6 13,7 14,9 16,2

4,2 4,6 5,5 6,1 6,6 7,2 7,8 8,5

3,2 3,6 3,9 4,3 5,1 5,6 6 6,5 7,1 7,6

2,2 2,5 2,7 3 3,3 3,6 3,9 4,2 4,6 5,3 5,7 6,2

1,6 1,8 2 2,2 2,4 2,6 2,8 3 3,3 3,6 3,8 4,1 4,4 4,8 5,1

1,1 1,2 1,3 1,4 1,6 1,7 1,9 2 2,2 2,4 2,6 2,8 3 3,4 3,7 3,9

0,8 0,9 1 1,1 1,2 1,3 1,4 1,5 1,7 1,8 2 2,1 2,3 2,4 2,6 2,8 3

0,70 0,77 0,85 0,92 1 1,1 1,19 1,29 1,4 1,51 1,62 1,74 1,87 2, 2,15 2,29 2,44

0,54 0,60 0,66

,01

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

0,37 0,41 0,450,49 0,54 0,59 0,64 0,69 0,75 0,81 0,87 0,94 1,09 1,16 1,24 1,33 1,42

14,4

4,9

0,65

2,08

ZONE 1= POOR >3.2

3.2

3.2

System one LD 17

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FLOFLOWW

ZONE L3/D4ZONE L3/D4

INCH INCH A > 80A > 80 B 60 > 80B 60 > 80 C 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

METRICMETRIC AA

1515 25251010 2020

HE

AD

HE

AD

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 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

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PRESENT ALIGNMENT METHODS WEAKNESSES

All provide precision initial alignmentDegree of accuracy variesCost 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

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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 technologyUsed on machine tools and gear boxes

Operate with highest level of accuracy and precision

Mechanical seal in a pump is a high precision componentMechanical seal accounts for 75% of pump

downtime

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MOTOR ADAPTER- ADVANTAGES

Provides easy, accurate, and reliable alignment during operation

Maintains near -laser alignment accuracy despite pipe strain, cavitation, high temperature, and vibrationA 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 applicationsMotor grows with the pumpMore even temperature gradient across the pump

and motor assembly

For high speed (3000/3600 RPM) applications - Alignment more critical

DisadvantagesNot as accurate as initial laser alignment due to

inherent tolerance stackup of the various components

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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

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

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BEARING OIL SEALS

Three basic types: Lip seal

Inexpensive, simple to install, very effective when new

Elastomeric constructionContact shaft and contributes to friction

drag and temp. rise in bearing areaAfter 2000-3000 hours, no longer provide

effective barrier against contaminationWill groove shaft

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BEARING OIL SEALS cont.

Labyrinth sealsRequired by API 610Non-contacting and non-wearingUnlimited life Effective for most types of contaminantsDo 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 immersionGood 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

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SYSTEM ONE LABYRINTH SEAL

GRAVITY DRAIN

Allows liquid to drain

OIL TRAPand DRAINHelps retain lubrication

in bearing housing

LABYRINTHTraps liquid and directs it to the gravity drain

Stationary Element

Rotary Element316SS for corrosion resistance

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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 temperatureCONTAMINATION•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 Normal clearance “C3”

6310 Radial Bearing (microns)

Radial .0003-.0011”(9-30) .0009-.0017”(27-51)

Axial .0016-.003” (48-90) .0016-.003”(48-90)

5310 Double Row Thrust BearingRadial .0005-.0014”(15-42) .0014-.0020”(42-60)Axial .0005-.0014”(15-42) .0014-.0023”(42-69)

7310 Angular Contact Thrust BearingAxial -.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 bearingDesirable to increase running accuracy Enhances stiffnessReduces running noise Provides a longer service life under proper applications

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BEARING CLEARANCES / PRELOAD

LIFE

ClearancePreload

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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