Thumb for Mechanical Engineers 2011 Part 4 doc

Udortumtely, the seal chamber pressure varies con- siderably with different pump designs and impeller styles.. The estimated chamber pressure for this arrangement can be calculated with

Equipment Checks

One of the most important considerations for reliable seal

performance is the operating condition of the equipment

Many times, mechanical seal failures are a direct result of

poor equipment maintenance High vibration, misalign-

ment, pipe strain, and many other detrimental conditions

cause poor mechanical seal life There are also several di-

mensional checks that are often overlooked

Because half of the seal is rotating with the shaft, and the other half is fixed to a stationary housing, the dimension-

al relationships of concentricity and “squareness” are very

important The centrifugal pump is by far the most common piece of rotating equipment utilizing a mechanical seal For this reason, the dimensional checks will be referenced to the shaft and seal chamber for a centrifugal pump

Axial Shaft Movement

Axial shaft movement (Figure 28) can be measured by

placing a dial indicator at the end of the shaft and gently tap-

ping or pulling the shaft back and forth The indicator

movement should be no more than 0.010’’ TIR [ 11

Radial Shaft Movement

There are two types of radial shaft movement (Figure Mid Bearing

29) that need to be inspected The first type is called

shafi defection, and is a good indication of bearing con-

ditions and bearing housing fits To measure, install the

dial indicators as shown, and lift up or push down on the

end of the shaft The indicator movement should not ex-

ceed 0.002” TIR [ 11

The second type of radial shaft movement is called shu#

mn-out, and is a good way to check for a bent shaft con-

dition To measure, install the dial indicators as shown in

Figure 29 Checking pump for radial shaft movement

(Coudesy of Durametallic Cop.)

Seal Chamber Face Run-Out

Figure 29, and slowly turn the shaft The indicator move-

ment should not exceed 0.003” TIR [ 11

As stated earlier, because the stationary portion of the me-

chanical seal bolts directly to the pump case, it is very im-

portant that the face of the seal chamber be perpendicular

to the shaft center-line To check for “out-of-squareness,”

mount the dial indicator directly to the shaft, as shown in

Figure 30 Sweep the indicator around the face of the seal

housing by slowly turning the shaft The indicator move-

ment should not exceed 0.005” TIR [ 11

Seal Chamber Bore Concentricity

There are several stationary seal components that have

close diametrical clearances to the shaft, such as the throt-

tle bushing For this reason, it is important for the seal

chamber to be concentric with the pump shaft Addition-

ally, for gland ring designs with O.D pilots, the outer reg-

ister must also be concentric To measure, install the dial

indicators as shown in Figure 3 1 The indicator movement

should not exceed 0.005” TIR [ 11

Figure 31 Checking pump for seal chamber bore con- centricity (Courtesy of Durametallic Corp.)

CALCULATING SEAL CHAMBER PRESSURE

The seal chamber pressure is a very important data point

for selecting both the proper seal design and seal flush

scheme Udortumtely, the seal chamber pressure varies con-

siderably with different pump designs and impeller styles

Some pumps operate with chamber pressures close to suc-

tion pressure, while others are near discharge pressure

The easiest and most accurate way to determine the seal

chamber pressure on an existing pump is simply to mea- sure it Install a pressure gauge into a tapped hole in the seal chamber, and record the results with the pump running The second most accurate method for determining seal cham-

ber pressure is to consult the pump manufacturer If neither

of these two methods is feasible, there are ways of esti- mating the seal chamber pressure on standard pumps

Single-Stage Pumps

The majority of overhung process pumps use wear rings

and balance holes in the impeller to help reduce the pres-

sure in the seal chamber The estimated chamber pressure

for this arrangement can be calculated with the following

equation:

where: Pb = seal chamber pressure (psi)

P, = pump suction pressure (psi)

Pd = pump discharge pressure

In some special cases, where the suction pressure is very

high, pump designers will remove the back wear ring and

balance holes in an effort to reduce the loading on the thrust bearing In this case, the seal chamber pressure (Pb)

will be equal to discharge pressure (Pd)

Another common technique for reducing seal chamber pressure is to incorporate pumpout vanes in the back of the impeller This is used primarily with ANSI-style pumps, and can be estimated with the equation:

The final type of single-stage pump is the double suction pump, and for this pump design, the seal chamber pressure (Pb) is typically equal to the pump suction pressure (Ps)

~~~~~ ~ ~

Multistage Pumps

Horizontal multistage pumps typically are “between

bearing” designs, and have two seal chambers On the

low-pressure end of the pump, the seal chamber pressure

(Pb) is usually equal to the pump suction pressure (Ps) On

the high-pressure end of the pump, a balance piston and

pressure balancing line is typically incorporated to reduce

both the thrust load and the chamber pressure Assuming

that the balance line is open and clear, the seal chamber pres-

sure is estimated to be:

The seal chamber pressure for vertical multistaged pumps can vary greatly with the pump design The seal chamber can be located either in the suction stream or the discharge stream, and can incorporate a pressure balancing line, with

a “breakdown” bushing, on high-pressure applications Ver- tical pumps tend to experience more radial movement than

horizontal pumps, and for this reason the effectiveness of the balancing line becomes a function of bushing wear With so many variables, it is difficult to estimate the pressure in the sealing chamber The best approach is either to measure the pressure directly, or consult the manufacturer

As previously discussed, different seal designs are used

in different seal arrangements to handle a vast array of

different fluid applications In every case, the seal must be

provided with a clean lubricating fluid to perform proper-

ly This fluid can be the actual service fluid, a barrier fluid,

or an injected fluid from an external source All these op-

tions require a different flushing or piping scheme In an ef-

fort to organize and easily refer to the different seal flush piping plans, the American Petroleum Institute (API) de-

veloped a numbering system for centrifugal pumps that is now universally used [7] The following is a brief discus- sion of the most commonly used piping schemes, and where they are used

Single Seals

API Plan 11 TO pump suction *-I+,

The API Plan 11 (Figure 32) is by far the most commonly

used seal flush scheme The seal is lubricated by the pumped

fluid, which is recirculated from the pump discharge noz-

zle through a flow restriction orifice and injected into the seal

chamber In this case, the chamber pressure must be less than

the discharge pressure The Plan 11 also serves as a means

of venting gases from the seal chamber area as liquids are

introduced in the pump This is a very important function

for preventing dry running conditions, and when at all pos-

sible, the piping should connect to the top of the gland The

API Plan 11 is primarily used for clean, cool services

Figure 33 API Plan 13 @PI-682 Courtesy of American

Figure 34 API Plan 21 @PI-682 Courtesy of American Petroleum Institute.)

Figure 32 API Plan 11 @PI-682 Courtesy of American

Petroleum Institute )

API Plan 13

The API Plan 13 (Figure 33) is very similar to the Plan

11, but uses a different recirculation path For pumps with

a seal chamber pressure equal to the discharge pressure, the

Plan 13 seal flush is used Here, the pumped fluid goes

across the seal faces, out the top of the gland ring, through

a restricting orifice, and into the pump suction This pip-

ing plan is also used primarily in clean, cool applications

API Plan 21

Figure 34 shows the arrangement for an API Plan 21 T h ~ s

plan is used when the pumpage is to hot to provide good

lubrication to the seal faces A heat exchanger is added in

the piping to reduce the fluid temperature before it is in-

troduced into the seal chamber The heat removal require- ment for this plan can be quite high, and is not always the most economical approach

API Plan 23

The API Plan 23 is also used to cool the seal flush, but utilizes a more economical approach For the Plan 2 1 the fluid passes through the heat exchanger one time before it

is injected into the seal chamber and then introduced back into the pumping stream The Plan 23 (Figure 35) recircu- lates only the fluid that is in the seal chamber In this case,

an internal pumping device is incorporated into the seal de- sign, which circulates a fixed volume of fluid out of the seal chamber through a heat exchanger and back to the gland ring This greatly reduces the amount of heat removal nec-

1 F? r"r FI an>

Figure 35 API Plan 23 (API-682 Courtesy of American

Petroleum Institute.)

arrangement does not use the pumped fluid as a seal flush

In this case, a clean, cool, compatible seal flush is taken from

an external source and injected into the seal chamber This arrangement is used primarily in abrasive slurry applications

essary to achieve a certain flush temperature (and in the

process industry, heat is always money) This flush plan is

primarily used in boiler feed water applications

API Plan 32

The last seal flush plan for single seals is API Plan 32,

shown in Figure 36 Unlike the previous piping plans, this

Figure 36 API Plan 32 (API-682 Courtesy ofAmerican Petroleum Institute.)

~~~

Tandem Seals

API Plan 52

Tandem seals consist of two mechanical seals The pri-

mary, or inboard, seal always operates in the pumped fluid,

and therefore utilizes the same seal flush plans as the sin-

gle seals The secondary, or outboard, seal must operate in

a self-contained, nonpressurized barrier fluid The API

Plan 52, shown in Figure 37, illustrates the piping scheme

for the barrier fluid An integral pumping device is used to

circulate the barrier fluid from the seal chamber up to the

reservoir Here, the barrier fluid is typically cooled and grav-

ity-fed back to the seal chamber The reservoir is general-

ly vented to a flare header system to allow the primary seal

weepage to exit the reservoir

Vent

L

Figure 37 API Plan 52 (API-682 Courtesy of American Petroleum Institute.)

Double Seals

Double seals also consist of two mechanical seals, but

in this case, both seals must be lubricated by the barrier fluid

For this reason, the barrier fluid must be pressurized to 15

to 25 psi above the seal chamber pressure The API Plan

53 (Figure 38) is very similar to Plan 52, with the excep-

tion of the external pressure source This pressure source

is typically an inert gas, such as nitrogen

To External Pressure Source

The API Plan 54 (Figure 39) uses a pressurized, exter- nal barrier fluid to replace the reservoir arrangement This piping arrangement is typically used for low-pressure ap- plications where local service water can be used for the bar- rier fluid

INTEGRAL PUMPING FEATURES

Many seal flush piping plans require that the seal lubri-

cant be circulated through a heat exchanger or reservoir

While there are several different ways to accomplish this,

the most reliable and cost-effective approach is with an in-

tegral pumping feature There are many different types of integral pumping devices available, but the most common are the radial pumping ring and the axial pumping screw

Radial Pumping Ring

The radial pumping ring, shown in Figure 40, operates

much like a centrifugal pump The slots in the circumfer-

ence of the ring carry the fluid as the shaft rotates When

each slot, or volute, passes by the low-pressure area of the

discharge tap located in the seal housing, the fluid is pushed

out into the seal piping This design is very dependent on

peripheral speed, close radial clearance, and the configu-

ration of the discharge port A tangential discharge port will

produce four times the flow rate, and two times the pres-

sure, of a radial discharge tap Higher-viscosity fluids also

have a negative effect on the output of the radial pumping

Figure 40 Radial pumping ring @PI-682 Courtesy of

American Petroleum Institute.)

ring Fluids with a viscosity higher than 150 SSU, such as oils, will reduce the flow rate by 0.25 and the pressure by

0.5 Figure 41 shows the performance of a typical radial pumping ring [ 1 1

2.5 - Flow for Water

fPm rpm x ring O.D In inches x 0.262

For oils and other liquids

2000 lpm (10.2 m/s)

1000 fpm ( 5.1 mls) Feet of Head

Figure 41 Typical radial pumping ring performance

curve (Courtesy of Durametallic Corp.)

Axial Pumping Screw

The axial pumping screw, shown on the outboard seal of

Figure 42, consists of a rotating unit with an O.D thread

and a smooth walled housing This is called a single-act-

ing pumping screw Double-acting screws are also avail-

able for improved performance and utilize a screw on both

the rotating and stationary parts Unlike the screw thread

of a fastener, these screw threads have a square or rectan-

gular cross-section and multiple leads The axial pumping

screw does have better performance characteristics than the

radial pumping ring, but while gaining in popularity, the

axial pumping screw is still primarily used on high-per-

formance seal designs

Figure 42 Axial pumping screw @PI-682 Courtesy of American Petrobum Institute.)

Piping Considerations

Integral pumping features are, by their design, very in-

efficient flow devices Consequently, the layout of the seal

piping can have a great impact on performance The fol-

lowing are some general rules for the piping:

Minimize the number of fittings used Eliminate elbows and tees where possible, using long radius bent pipe as

a replacement

Where possible, utilize piping that is one size larger than

the seal chamber pipe connections

Slope the piping a minimum of K" per foot, and elim-

inate any areas where a vapor pocket could form

Provide a minimum of 10 pipe diameters of straight pipe

length out of the seal housing before any directional

changes are made

SEAL SYSTEM HEAT BALANCE

Excessive heat is a common enemy for the mechanical

seal and to reliable seal performance Understanding the

sources of heat, and how to quantify the amount of heat,

is essential for maintaining long seal life The total heat load

(QTotd) Can be stated as:

QTotd = Qsgh + Qhs

where: Qtod = total heat load (btu/hr)

Qsgh = seal generated heat (btu/hr)

Qhs = heat soak (btu/hr)

Seal-generated heat is produced primarily at the seal

faces This heat can be generated by the shearing of the lu-

bricant between the seal faces, contact between the differ-

ent asperities in the face materials, or by actual dry running

conditions at the face Any one, or all, of these heat-gen-

erating conditions can take place at the same time A heat

value can be obtained from the following equation [2]:

Qsgh = 0.077 X P X v X f X A

where Qsgh = seal generated heat (btu/hr)

P = seal face pressure (psi)

V = mean velocity (ft/min)

f = face friction factor

A = seal face contact area (in2) and

The face friction factor (f) is similar to a coefficient of fric- tion, but is more tailored to the different lubricating condi- tions and fluids being sealed than to the actual material properries The following values can be used as a general rule:

f = 0.05 for light hydrocarbons

f = 0.07 for water and medium hydrocarbons

f = 0.10 for oils For a graphical approach to determining seal-generated heat values, see Figure 43 [2]

TYPICAL SEAL GENERATED HEAT VALUES

t

Figure 43 Typical seal-generated heat values (Courtesy of Du-

rarnetallic C o p )

Heat soak (Qhs) is the conductive heat flow that results

from a temperature differential between the seal chamber

and the surrounding environment For a typical pump ap-

plication, this would be the temperature differential between

the chamber and the back of the pump impeller Obvious-

ly, seals using an API Plan 11 or 13 would have no heat

soak But for Plans 21 or 23, where the seal flush is cooled,

there would be a positive heat flow from the pump to the

seal chamber Radiant or convected heat losses from the seal

chamber walls to the atmosphere are negligible

There are many variables that affect heat soak values, such as materials, surface configurations, or film coeffi- cients In the case of a pump, heat can transfer down the shaft, or through the back plate, and can be constructed from several different materials To make calculating the heat soak values simpler, a graphical chart, shown in Figure 44, has been provided which is specifically tailored for mechani-

cal seals in centrifugal pump applications [SI

-SEAL SIZE, INCHES

Figure 44 Heat-soak curve for 316 stainless steel

(Courtesy of Durametallic Cow.)

FLOW RATE CALCULATION

Once the heat load of the sealing system has been de-

termined, removing the heat becomes an important factor

Seal applications with a high heat soak value will typical-

ly require a heat exchanger to help with heat removal In

this case, assistance from the seal manufacturer is required

to size the exchanger and determine the proper seal flush

flow rate For simpler applications, such as those using API

Plans 1 1,13, or 32, heat removal requirements can be de-

termined from a simple flow rate calculation Using values

for seal-generated heat and heat soak, when required, a flow

rate value can be obtained from the following equation [2]:

(gpm) = Q~otal

500 x C, x S.G x AT

where: C, = specific heat (btdlb-"F)

AT = allowable temperature rise ("F)

S.G = specific gravity

The allowable temperature rise (AT) will vary depending

on the fluid being sealed For fluids that are very close to

the flashing temperature, the temperature rise should not ex-

ceed 5-10°F For nonflashing fluids, the maximum allow- able temperature rise is 20°F Once the flow is determined

Figure 45 can be used to obtain the proper orifice size [ 5 ]

Figure 45 Graph of water flow through a sharp-edged ori- fice (Courtesy of Durametallic Corp.)

1 Durametallic Corporation, “Guide to Modern Mechan-

ical Sealing,” Dura Seal Manual, 8th Ed

2 Durametallic Corporation, “Sizing and Selecting Seal-

ing Systems,” Technical Data SD-1162A

3 Durametallic Corporation, “Dura Seal Pressure-Veloc-

ity Limits,” Technical Data SD-1295C

4 Durametallic Corporation, “Dura Seal Selection Guide,”

Technical Data SD-634-9 1

5 Durametallic Corporation, “Dura Seal Recommenda-

tions for Fugitive Emissions Control in Refinery and

Chemical Plant Service,” Technical Data SD- 1475

6 Durametallic Corporation, “Achieve Fugitive Emissions Compliance with Dura Seal Designs,” Technical Data

7 API Standard 682, “Shaft Sealing Systems for Cen- trifugal and Rotary Pumps,” 1st Ed., October 1994

8 Adams, Bill, “Applications of Mechanical Seals in High

Temperature Services,” Mechanical Seal Engineering Seminar, ASME South Texas Section, October 1985

9 Will, Thomas P., Jr., “Mechanical Seal Application

Audit,” Mechanical Seal Engineering Seminar, ASME

South Texas Section, November 1985

SD-1482B

Pumps and Compressors

Bhabani P Mohanty Ph.D., Development Engineer Allison Engine Company

€ W McAllister P I , Houston Texas

Pump Fundamentals and Design 93

93 Pump Design Parameters and Formulas 93

Types of Pumps 94

Centrifugal Pumps 95

Net Positive Suction Head (NPSH) and Cavitation 96

96 Recirculation 97

97 97 Performance Curves 98

Series and Parallel Pumping 99

Design Guidelines 100

Reciprocating Pumps 103

Pump and Head Terminology

Pumping Hydrocarbons and Other Fluids

Pumping Power and Efficiency

Specific Speed of Pumps

Pump Similitude 98

Compressors 110

Definitions 110

Performance Calculations for Reciprocating Estimating Suction and Discharge Volume Bottle Sizes for Pulsation Control for Reciprocating Compressors 114

Compressor Horsepower Determination 117

Generalized Compressibility Factor 119

Centrifugal Compressor Performance Calculations 120

Estimate HP Requkd to Compress Natural Gas 123

Estimate Engine Cooling Water Requirements 124

Estimate Fuel Requirements for Internal Combustion Engines 124

References 124

Compressors 111

92

Pumps convert mechanical energy input into fluid energy

They are just the opposite of turbines Many of the basic

engineering facts regarding fluid mechanics are discussed

in a separate chapter This chapter pertains specifically to

pumps from an engineering point of view

Pump and Head Terminolosv

Symbol Variable and Unit

barrel (42 gallons) barreldday barrelslhour brake horsepower water horsepower acceleration due to gravity (32.1 6 ftlsec)

torque (ft Ibs) temperature rF)

specific gravity of fluid impeller diameter (inch) pump efficiency (in decimal) revolution per minute (rpm) specific heat

total head (ft)

velocity (Wsec)

area (sq in.) net positive suction head (ft of water) efficiency

density specific weight of liquid

Pump Design Parameters and Formulas

Following are the pump design parameters in detail:

Flow Capacity: The quantity of fluid discharged in unit

time It can be expressed in one of the following popular

units: cfs, gpm, bph, or bpd

gpm = 449 x cfs = 0.7 x bph (4)

Head This may also be called the specific energy, i.e., en-

ergy supplied to the fluid per unit weight This quantity may

be obtained through Bernoulli’s equation The head is the

height to which a unit weight of the fluid may be raised by

the energy supplied by the pump

The velocity head is defined as the pressure equivalent of the dynamic energy required to produce the fluid velocity

Power: Energy consumed by the pump per unit time for

supplying liquid energy in the form of pressure

bhp = Q x H x ~43,960 x e) = Q x P/(1,715 x e) (6)

Efficiency: The ratio of useful hydraulic work done to the actual work input It consists of the product of three com- ponents: the volumetric efficiency, the hydraulic efficien-

cy, and the mechanical efficiency

The overall efficiency varies from 50% for small pumps to 90% for large ones

Types of Pumps

Pumps fall into two distinct categories: dynamic pumps

and positive displacement pumps

Dynamic pumps are of two types: centrifugal and axial

They are characterized by the way in which energy is con-

verted from the high liquid velocity at the inlet into pres-

sure head in a diffusing flow passage Dynamic pumps

have a lower efficiency than positive displacement pumps

But their advantages lie in the output of relatively high flow

rates compared to their sizes, and their low maintenance

costs They also operate at relatively higher speeds

Positive displacement pumps are of several types, in-

cluding reciprocating, rotary screw, and gear pumps These

pumps operate by forcing a fmed volume of fluid from the inlet pressure section to the discharge section of the pump

In reciprocating pumps, this is done intermittently; and in others this is done continuously These types of pumps are physically larger than the dynamic pumps for compa- rable capacity, and they operate at relatively lower speeds

Table 1 shows major pump types, their characteristics, and their applications

Source

Cheremisinoff, N P., Fluid Flow Pocket Handbook Hous-

ton: Gulf Publishing Co., 1984

Table 1 Major Pump m s

splk asing

Impeller antilevffsd beyond bearings

2 impellers cantilevered beyond bearings

lmpsller between bearing; casing Capcing patterns &signed with thin

b w flow pamgol erosion n ~ w l

Pump and motor endosed in pressure

Nozzle usually in bottom half of casing

Outer casing confines inner stack of

rsdially or axially split

snctlonr for high c a t alloys; sniall

sizes

fsabrrsr shell: no muffing box

m P h m * Verdcal orientation

Many rmgcr low headhge

Anmgsd for i d l n c Innallation, like a

HslVR

speeds to 380 rpr head to in0 m

Qring immned in sump for lrntdlmien

convenience m d primlng (LBIB

Vow long shafts Propeller rhaped impeller, usually

l a w size

Fluted impeller; flow path like m a w

amund porlphery

slow speeds: valves, Mindem stuffing

Smqll units with precision flow control

bo== Wblfft to F

t y m m

or hydraulicaIly rtuat.d

1.2or 3 screw roton

lntarmahing gaar wheds

Low Medium

LOW

Medium Low Medium Low

COlnnnnO

Cawcity varies with haad

Low to medium speciris speed

For heads dbwc single stage capebilii

For high flow to330 m head

Low pressurn and lempenture mings

Low speed; adjustable axial c l c w a r a

Low h e a w i t y limits for modelr usnd

For modeme temperaarngrr~up Rtinor

For high tempersturepmm~re rabngs

*le used primarily to exploit low NPSH requirement

Hi& head capnbility, low NPSH requirement

piping rymmr

Low cost inmllation

Waler well service with driver at grade

Low flow-hii head perfomanm

CapaSw v i m d y independent of head

Mineries

D r i i bv stwm enghe cylinders or moton

Diaphragm through and ~ k c l r s c packed plunger tvpa

Used for chmid durries: dlaphrapr

prone to failurn

For high v i m , h i i flow high prrrourr

For high viscdw moderate pratun