- Process Eqpt Series Volume 3 by KS Panesar-1

CHAPTER 1

FANS AND BLOWERS

K. S. PANESAR Houston. TX

I INTRODUCTION

Pumps, compressors, blowers and fans all belong to the same family of niacl1: called TURBOMACHINERY or ROTATING EQUIPMENT. Pumps can handle I ) !

j incompressiblefluidsviz. liquids while compressors, blowers and fans, on the oti J hand, can handle only compressible fluids like air and other gases. These rnachii

can be damaged very easily if air or gases are pumped through centrifugal pumps water etc. is pumped through blowers or fans. Most of this discussion will

i limited t o the centrifugal and axial machines, which are also called the const: pressure machines as compared with the constant volume machines which are

3 aositive dirolacement machines. The oerformance characteristics of both are sho in Figure 1.1.

0 l I I 0 5 0 100

CAPACITY (%)

Figure I . 7. Pressure-volume diagram for rhe cenrrifugal and rhe positive displacement rypes of machines. Capacity is uru- ally specified in terms of rhe inlet CFM /at rhe inlet condirionr) and the head in fee1 or pressure in PSI for compressors and blowers. For fans, however, rhe pressure is usually speci-

fied in incher of water gage

Most people, othe, than those lnvolved in the design or application o f the

rotating equipment, get confused i n the definitions of fan, blower or a compressor. arid tise these terms interchangeably. Al l these machine5 can handle f luid flows ranging l r o ~ n a few hundred cubic feet pel- minute t o several mi l l ion cubic feet per . . rnrlutu. The ima111 distinctton lies i n the amount of plessure generated b y each class of inschine. Fans are supposed t o have a pressure range o f f rom a fract ion of an

inch ill water yage t o about 5 0 inches o f water gage, which is approximately 2 Psi. Bloweus, o r the other hand, develop pressures f rom 2 Psi t o about 15 Psi and pressures above 15 Psi fall in the category o f compressors. Compressors can develop pressures up t o several thousand pounds per square inch. Of course, extremely high preaures are developed i n several stages (wheels) or sometimes i n two or three

cases, each case having several stages.

FUNDAMENTALS

Most centrifugal {machines have a housing w i th an inlet and an outlet. Inside the housing is a wheel or an impeller which rotates and imparts kinetic energy t o the

and when the term (dm ld t ) is applied t o all the f luid in all the blades, i t becomes

Qplg; where

the f low through the machine in cubic feet per second; the specific weight in pounds per cubic feet: the outer radius of the impeller: the inner radius o f the impeller; the absolute velocity o f the f luid at the (lischarge: the absolute velocity o f the f lu id at the inlet; the angle between the absolute and the peripheral velocities at the exit

and the inlet, respectively; the peripheral velocity at the exit; the peripheral velocity at the inlet;

fluid. The f lu id comes i n at a pressure. P , . and leaves the housing at a higher Ipiessilre, P-. . These inacliines wil l , therefore, always generate a constant pressure

i

Now, substituting for (dmld t ) , equation number 1 becomes:

differentii l l whet1 operating at the same f low. Thaovetic.?lly, the pressrlte~volurne line is supposed t o be a straight line, when

lhc! i l i s ~ : l i ; ~ i q ! angle on the I~lades is 9 0 dcg~:ccs. The ideal pressure-volume lines for disi:h;uge a ~ g l o loss than and greater than 9 0 degrees are as shown in Figure 1.2. I n actu;~l prxt ici : , however, this is never the case. They have a curvature t o them as

already shown i n Figure 1.1. The reason for this is that there are certain losses within the housing: - disk f r ic t ion (wheel fr ict ion). blade inefficiency, circulation wi t l i in the bl.ides. cfc.

Tile k i ne t x energy is generated by the rota1.y mot ion o f the impeller and is

irnpa~tecl t o the f lu id moving through the machine. Each machine, in fact, each

~ m p l l e r is designed t o produce a certain pressure rise or (sometimes called "Head") at a given capacity w i th a m i n i ~ n u i n loss or, in other words, w i th a maximum efficiency, as shown in Figure 1.3.

I n oldel to undeustand how the head is generated by an impeller, let us look at

Figure 1.4. F g o ~ e 1 . 4 ~ shows a typical sectional elevation view o f hydraulic path o f an ~~npe I i t ? i . w l ie r~ds Figure 1 4 b shows the same impeller in plan view w i t h inlet and discharoe v d o c ~ t v t i ;males.

Le i 0 s . K S L I ~ L ' tIi;il ill t ~ ~ n e 'dt,' there i s a Inass 'dm' o f very thin layer of f lu id j;i!i, g.3~ or ltiliwcl) lc~iwfng thc impellev and a t the salne t ime an equal amount o f mass of fiutd enters the in,pellel-. Thts change i n moment o f (momentum isequal t o lnolnent of all the external foices i~nposed on the f lu id contained between the two

blades, Let T denote the moment o f external forces. Then T is given by the following equation:

C A P A C I T Y

Figure 7.2. This figure shows the rheorericaipressure-volume curves for the three differenr rype of blade discharge angler.

Horsepower Calculations

Once you calculate the polytropic head, i t is relatively easy to calculate the theoretical or actual brake horsepower as shown:

THP == w Hp/33000 (7)

where

THP and BHP are theoretical and brake horsepowers respectively. w - flow in Ibs per minute, and l ip polytrop~c efficiency in decimal.

There is another way of calculating the theoretical or actual horsepower when

you do not know or do not have to know the polytropic head. The calculation of head equation is very useful in centrifugal and axial machinery because from this you call deterniine the number of stages or wheels required to get the desired pressure o i head. For reciprocating cornpressots, on the other hand, you do not need to r.;ili:oliilu the lhcxi. There co~npressio~i is achieved in a cylinder, and if the I I I<,SSII IP I M I ~ O L ,1111 q111111 higli YOL, .~tlcl .~dtlitioo:il cylindc1.s w ~ t i l the desired pressure ttitlu is ,lcliievo~l. Thcis to calc~ilale thc horsepower, without the head, you must k ~ i o w or calculate the inlet vol~ime flow, called the ICFM (inlet cubic feet per nlinutc) or :also sometimes called the ACFM (actual cubic feet per minute). (They both (mean the same). This is the flow in the machine at the inlet temperature and pressilre. I f the process flow is given only in terms of the pounds per hour, pounds per ~niriute or [moles per hour, use the following equations to get the ICFM and then finally the horsepower:

w : (Molesihll X (M.W.)iGO lbs per minute

SCFM - (Molesihrl X (3791601 (Note: Each gas occupies the 379 cubic feet volume at standard conditions. (91

SCFM = Iw1M.W.) X 379 ICFM =: SCFM X 114.7iP, I X IT, i520)

w h r ~ e M W , is the molecular weight, and 14.7 and 5 2 0 ' ~ are the standard condit~ons. The blower and the fan manufacturers, however, use 6 8 ' ~ as the staiidard temperature, instead of the GOOF. Let's call V , as the inlet CFM or the ACFM, the horsepower is given as follows:

Just divide the THP by the polytropic efficiency to get the BHP. In fans, however. the horsepower is calculated in a little b i t different form, which is shown below:

BHP = ,000157 X ACFM X S.P./S.E. or

i BHP= ACFM X S.P.16356 X S.E.

In fans the discharge temperature i s hardly ever a problem because they are a very low pressure ratio machine. In blowers, and compressors the discharge temperature can be very high, and therefore intercooling (between stages) or aftercooling (i.e. cooling of the gas after i t leaves the machine housing) is usually required. To calculate the discharge temperature, the following equations are used:

AS is quite evident from the above equation, the discharge temperature va~lt,s linearly with the inlet temperature ( T I ) , but more so w11h the pressura );111u

P I P . That's why intercoolers and aftercoolers are used in the multistage centrifugal compressors, or reciprocating compressors. There is another way o f

calcualting the discharge temperature:

Impeller Design

Basically there are three different designs used in the centrifugal machines, as shown in Figure 1.5. In the radial design, the fluid enters the impeller along the shaft, but leaves i n a direction perpendicular to the shaft. In the mixed f low design, the f luid enters along the shaft as before, but leaves the impeller at some angle. In the axial design, however, the f luid enters and leaves parallel to the shaft. The radial blades are used i n low specific speed range, mixed f low blades are used in the medium range and axial impellers or propeller design is used in high specific speed range.

Specific speed is a kind o f a dimensionless n u m k r used in the industry to help engineers and designers get an idea of the geometry of the machine they have to design. Machines having same specific speed are geometrically similar and thercfori you can predict the performance of the new machine. Specific speed. N5, is d e f w d

as the speed which will produce one foot of head at a f low of one CFM, a:,d mathematically stated it is:

Cliuprer I . K . S. Panesar Process Equipmrrir Series Volume 3

where:

N, = specific speed Q = flow in cubic feet per second for compressors and GPM for pumps,

N = machine rotative speed in RPM H = feet of head

The above equation is applicable to pumps, compressors and blowers but for fans the equation may be rewritten and defined as follows:

where:

Q = flow in ICFM (inlet cubic feet per minute) and P, = Static Presure in inches of water gage.

For fans, the specific speed i s defined as the speed which will deliver one cubic feet per minute of flow a t one inch of water gage.

Specific speed, N,, is related with another dimensionless number called the Specific Diameter, D,, which is defined as:

where:

D = Diameter of the wheel in feet

The above equation is applicable to pumps, compressors and blowers. The equation for fans, however, i s a little different, as shown below:

D, = D P,"/Q~ (18)

where:

D = D~ameter of the wheel in inches P, = Static pressure in inches of water gage O = Flow in Cubic feet per minute

The relationship between the specific speed and the specific diameter was developed and presented by Dr. O.E. Balje, and is shown in Figure 1.6. The higher the specific speed, the higher the f low and smaller the pressure or head and vice versa. Figure 1.6a covers a wide range of turbomachinery viz. centrifugal and positive displacement pumps and compresors. Figure 1.6b, however, covers just low

30' 1 , 1 , I , , , 1 I I 11111 I 1 1 1 1 1 1 1 I 1 1 1 1 1 1 1 I I 1IIn-r DRAG PUMP -

SINGLE- STAGE . PUMPS OR COMPRESSORS-

- - - , - ,MIXED FLOW

Y -

1 3 - - AXIAL FLOW I!

0

0 I -- Y -

0.6 = OR COMPRESSOR .d D 1 N S = ~ # i / ~ 3 / 4 * 0.3 - - - oS; DHI/~/G Q FLOW, CFS o : IMPELLER DIAMETER ,FT - . H : HEAO,.FT

a l L I I I 1 1 a 1 t l I I 1111111 I I IIIIIII I I I 1 1 1 1 1 1 I I 1~

0.1 I 10 100 1,000 lo*

SPECIFIC SPEED, Ns

Figure 1.63. NS - Ds diagram for single stage pumps and compressors.

10

6

EXHAUST PRESSURE B TOTAL INLET PRESSURE

'Iod- HEAD COEFFICIENT

I

.6

FOWARD CURVED

.3 BACKWARD CURVED

RADIAL BLADES A \ k'& >4 1\dYw I I .6 .I2

.I I

1 I 100 1000 lop00

I TO TOTAL rnnauo i PRESSURE B TOTAL INLET PRESSURE

FOWARD CURVED

100 1000 lop00

! N S

T (BI 1 i

F w r e 1.66. Showing NE - DI CUNWS for single stage pumps and lowpressure blower,

and fans.

1 . 1

m 0

0.8

[L W + w 0.7 - Z 4 - 0

0 0.6 - - 5

-

0 W g 0 . 5 -

0'4-F B H P z %SEX6356 CFMX SP

0.3- FOWARD CURVED BLADE

I I I I I I I l l 10 2 0 3 0 40 5 0 6 0 7 0 8 0 90 100 2 0 0

SPECIFIC SPEED (N,) , THOUSANDS

IC)

Fiwre 1 .6~. Showing the N, - D, diagram specially adapted for the fan industv. (Courtesy of American Standard)

Substituting the actual values of the densities o f each fluid at the same temperature, we get:

Thus in order to convert inches of water to feet of Head, you multiply by 69.3

FANS AND BLOWERS

We have discussed the basic theory of the centrifugal machines and their design. Now the discussion will be limited mostly to fans, blowers and exhausters. Fans are used for a wide variety of application, in the petrochemical and chemical industry, Steel mills, paper mills, heating-ventilating and air-conditioning industry. For fan components see Figures 1.21a and 1.21b.

Basically there are two types of fans used in the industry viz, the centrifugal fans which are low to medium flows and relatively higher pressures than the axial f low fans which are mostly large volume and low pressure fans.

Centrifugal fans are further sub-classified as: (see Figure 1.7) (Depending on blade geometry).

a. Radial blade Fig. 1.7a b. Forward-Curved blade Fig. 1.7b c. Backward-Curved blade Fig. 1 . 7 ~ d. Airfoil Fig. 1 . 7 ~

Axial fans are sub-divided into the following classes: (see Fig. 1.8)

a. Vane-Axial Fans Fig. 1.8a b. Tube-Axial Fans Fig. 1.8b

Figure 1.7 Figure 1.78. Showing the radial blade desion. - Figure 1.76. Showing rhe forwardcurved blade. Figure 1 . 7 ~ . Showing me backwardly inclined blade end the airfoil design.

Procrrr Lquipmenr Series Vviume 3

Figure 1.8. Figure 1.8a. Vane-axial fan with the discharge guidevanes showing smoorh BN ROW Figure 1.86. Tubeaxial fan showing spiral air flow (no guide-vaned.

Radial Blades

This is a simple and a popular design which has many applications in systrl requiring high resistance and low flows. The manufacturing i s relatively simple. T; performance of this type blade is also simple, as shown in Figure 1.9a.

STATIC

0 5 0 100 VOLUME PERCENT

Figure 1.9 Figure 1.9a. Typical performance curve of a radial blade fan.

.: . - -.

PRESSURE OR H E A D i%) i PRESSURE OR HEAD (%I

The amount of the centrifugal force acting on a blade can be determined by:

Figure 1. 11 . Showing axial thrust an a single inlet,

,~npel ler

117 pwnp industry there are generally two ways to reduce this thrust. In the first ~ix!t l iad, l,olnwn!i rings or ribs arc providod to reduca th,: pressure (P,), but this method uses a little more horsepowel. In the second method, holes are drilled in the back shroud, which reduces the thrust but also reduces efficiency. This method is seldonl ifsed in the fan ~ndustry.

In the axtal fails, the th~us t i s approximated by the! following equation:

In c e l ? t l i l ~ l p fans, the thrust [nay be approximated as follows:

f,in total pressure in inches of water gage. f a n static pressure in inches of water gage. Ttv diameter in feet Inlet diameter in feet Constant. If the pressure in the fan housing is positive, then C = 1.0, but if the pressure in the fan housing is negative (for example, exhausters), then C = 2.0

where:

F = Centrifugal for in lbf.

4 m = Mass of the wheel in lbm. v = Tip velocity in feet per second.

1 r = Radius of the wheel in feet. g = Gravitational constant in ft-lbm/lbf-sec2

The above equation can be simpl~fied by replacing 'v' by 277 r NI60 then equatlon (22) becomes:

i The magnitude of the centrifugal force helps in the mechanical design of the blades and their attachment to the fan hub. There is another force that acts on thc housing and that is the net radial force or radial thrust, due to the volute deslg~i. There are basically two types of volutes viz. single volute and double volute. In ii single volute casing, as shown in Figure 1.12a. the radial force varies along the periphery of the impeller at off-design condition, and the net resultant radial force wil l act in a direction shown with an arrow. The radial forces are pretty much balanced when the machine is operting at the design point. In a doublr:-volut~. design, as shown in Figure 1.12b. the radial forces are balanced even at off.drsiqrl

f conditions. Actually there are two net resultants acting opposite to each other a r ~ d are almost equal in magnitude. The resultants F, and F, are shown in Fiyuri: 1.121,.

I They tend to balance out even though they are not eliminated. Any unbalancarl radial force or load i s taken up by the radial bearings which can be either anti- friction (ball bearings) or journal sleeve bearings.

I f there is any unbalance in the center of gravity and the center of rotation of

the wheel, i t would cause vibrations which are passed on to the bearings and furthci down to the foundations. That is why some manufacturers recommend that their equipment be mounted on vibration isolating pads. No matter how carefully the dynamic balancing o f the rotor is done, it is impossible to have a perfect balance. The industry practice i s t o have the operating speeds at least 20% above or below

\ the critical speed. The shafts that operate below the critical speed are called the "Stiff shaft design" and the ones that operate above the critical speed are called the "Flexible shaft." In flexible shaft design the shaft has to go through the critical speed during the starting and the shutdown of the machinery. The critical speed of a shaft is calculated by:

Chapter 1. K. S. Pa~iesar

N,, = 60 f ly 2 n

1. where: i

y = Static deflection of the shaft in inches. ! There are various other ways of calculating the critical speed of shaft, using

Rayleigh's or Dunkerley's method. Reader i s advised to refer to any text on vibrations for these and other techniques of Calculating the critical speeds. Simple cases of 'Overhung shaft' and 'Simply supported shaft' are, however, discussed below. For a single bearing overhung design, the shaft deflection can be calculated as: (see Figure 1.13at.

9 i For a two bearing design, the shaft deflection is given by: (see Figure 1.13b).

Equation (26a) i s for the case when the load acts in the center, between the two bearings i.e. when x = LIZ.

i Sometimes there are other ways when minor or secondary vibrations are excited. They can be caused by oil-whip, sleeve bearings, couplings etc. separately or in combination with each other. Let us discuss each case separately.

a. Oil Whip

Most high speed and high load machines usually use sleeve bearings which are normally oil-lubricated. The oil in these type of bearings sometimes forms what is called oil wedge, which travels around the shaft a t about half the speed of the shaft. I f this wedge speed happens to coincide with the shaft's first critical speed, the shaft will start to vibrate. Once these vibrations get started, they will not disappear even if the shaft speed is increased to about twice the critical speed. These vibrations are very difficult to eliminate, but the following procedures may help reduce them. (1) Switch to anti-friction bearings.

Process Equipment Series Volume 3

Figurn I . 12a A typical single volute or a scroll showing the radial forces acting at off-design point. Force F is the net resultant ect-

ing on the caring.

Figure I . 126. This shows the double volute design and the radial forces which are balanced. Forces F , and F , act opporire ro each

other.

fiocess Eyuipmeni Series Vuiume 3

1. Inlet Guide Vanes

(21 Decrease the bearing length to increase the load on the oi l wedge and thus reduce the oi l-whip induced vibrations.

(3) Do ,701 l ~ m the [machinery at twice the critical speed.

b. Slceve-Bearings

With sleeve-bearings small disturbances are transmitted f rom the coupling, driver or any other minor fluctuations wi l l easily produce shaft vibrations, since the shaft does not have a continuous contact with the bearings. Sometimes these vibrations disappear i f you reduce the clearances between the shaft and the sleeve bearing.

c. Coupling

These viblations are usually caused by either misalignment or are transmitted by t h ~ d l i v w Even with the elastic couplings, i f the shaft runs above the first critical speed. d slight misalignment wi l l produce periodic disturbances.

Fan Controls

Capacity control i s achieved i n several ways and some of the most commonly meil mc~rhocls of control are described below.

These are just before the air or gas enters the impeller. They are designed such that the f luid leaving these vanes would match the impeller blade angle with a minimum shock. This is an efficient way of saving energy, especially when the fan has t o operate at reduced capacities. The variable inlet guide vanes are very efficient and desirable when the fan operates at different capacities varying quite a bit.

2. Outlet Dampers

This is one of the most inefficient methods of capacity control. The inlet capacity (ICFM) stays the same and the discharge capacity is varied by letting the

extra capacity to atmosphere i f i t is safe to do so. Toxic gases are either recycled 01

disposed of in a safe manner.

3. Speed Variation

Variable speed motor drivers are sometimes used to handle reduced operatlori. Sometimes steam turbines are also used, i f the initial and the operation costs a t t i

justifiable.

Fan Construction

After the air or gas leaves the impeller tip, i t i s collected in a scroll or voluti: which converts kinetic energy imparted by the blades. into pressure rneryy. A

typical volute or scroll is shown in Figure 1.12. This conversion into pressuri: energy is achieved by increasing the area continuously around the impeller. Thr! inlet box is usually specified with the fan. As with any other air handling machinii. inlet conditions are critical, a proper selection of an inlet box wil l assure proocic performance. Even though the insulation i s installed in the field, i t should ii,, specified so that the manufacturer can design the insulation attachment method ( > t i

the fan housing.

Materials

Housing is usually made of fabricated steel, which is used up to a maximum of 900°F in a normal non-corrosive atmosphere. Above 900°F, stainless steels or high nickel alloys are employed, especially when the atmosphere is corrosive or you are pumping corrosive gases. In low temperature but corrosive atmosphere or corrosive gases, the wheels are made of FRP (fiber reinforced plastics).

Bearings

I n small sizes, single row ball bearings are used very commonly andXlhi:y 11,:

grease lubricated. Grease lubriication is satisfactory up to about 2 0 0 " ~ . 011

Chapter I. K. S. Panesar

lubrication is recommended above this temperature. Bearing cooling should be considered above 300'F and also some sort of barrier and or cooling device be- ween the fan housing and the bearing housing should be used. These air cooling/ circulating devices are called slingen or cooling disks. Usually above BOOOF, the beaiing housing or subassembly should be isolated from the fan housing. ~ a n s and the inlet and the outlet ducting should be insulated, not only t o cut down~on the heat transfer, but also to protect the personnel from injury.

In small fans with higher pressures double row ball bearings are used. In large fans, say 1000 horsepower and above, some manufacturers or users may require sleeve bearing which are oil-lubricated. The American Petroleum Institute Specifics- tions No. 610 (API-610) for centrifugal pumps for refineries has clearly stated as to when to use sleeve bearings. The same guide lines can be used for fans. The guide lines are: (for use of sleeve bearings) a. When the anti-friction bearings fail to meet the 6-10 life. b. Where D.N factors are greater than 300,000, where D = bore in millimeters, and

N = speed in rpm. c. Whenever pump rated speed times the pump rated horsepower equals or exceeds

2.7 million. Assuming the pump or fan is going to be operating at about 3550 rpm, sleeve bearings should be specified around 750 horsepower. Some users are successfully using "oil-mist" lubrication system for the ball bear:

ings for their machines. I t has a central mist generating system, which supplies tiny oil molecules carried by air under pressure, for the bearings. They claim to have increased bearing life or reduced maintenance costs on their equipment with the use of the "oil-mist."

The bearing life is usually expressed in terms of "6-10" life. Thissimply means that only 10% of all the bearings operating under identical conditions would have failed.

Special Considerations

High temperature is very critical for the rotating equipment's life and alignment. For high temperature application, therefore, i t is recommended that the fan housing be center-line mounted as opposed to base mounted or foot-mounted, for ease of alignment. In case of large equipment and also in the case of high temperature or both, use of turning gear is recommended. This would prevent the large shaft from getting a permanent set or bow.

I f you are pumping gases with some abrasive materials like sand, flyash, or cement dust etc. the use of wear plates or hard-facing is recommended. With abrasive particles in the gases, select fans at lower speeds. This will cut down on the erosion.

Couplings

There are basically two types of couplings that have been used for years, viz.

Process byu~pment Serzes Volume 3

gear type and the diaphragm type. Gear couplings for low speed applications are usually oil filled or grease packed and can operate from six morlths to a year before the oil or grease is changed. However, i f the environment is dirty, the oil or greasc may need changing sooner. The diaphragm type, on the other hand, has the advan-

e of no lubricant requirement and hence very little maintenance. They only need. periodic check to make sure that the discs, which are very thin, have not d e ~

eloped any fatigue cracks. There i s another type of coupling that should be mentioned here, and that i s the LIMITED-END-FLOAT type. As the name implies, tills type just limits the motor rotor float. NEMA (National Electrical Manufacturer:. Association) MG1-14.38, specifies that 250-HP and larger AC motors, when fitred with sleeve bearings, have '%-inch minimum total rotor end float in the bearings (or '/.-inch end float in each direction). It also specifies that the coupling should limit that end float 3116-inch maximum (or 3132-inch in each direction. Limited end float couplings restrict float by locating the rotor axially using the fixed thrusl bearing on the driven machine and prevent damage to the motor bearings.

Forces and Moments

Fan housings are usually not as rugged as those of large pumps and compressols. and the manufacturers normally do not allow any loads on the fan housing. Some times you cannot help avoid loads imposed by the piping. Especially in high - temperature cases, the loads due to thermal growth cannot be eliminated. In such cases, you must consult the fan manufacturer and he may advise you to use flexible connections (also called expansion joints) at the discharge end or at both the discharge and the inlet connection as well. In the later case, you must, then, specify that the fan be supplied with a flanged inlet, so that a proper connection with the expansion joint can be made in the field. For a picture of a flexible connector, see

Figure 1.14.

COMPANION FLANGE WlTH STUB PIPE BOLTS

-HEAVY DUTY FLEXIBLE RUBBER

"FOR FLANGED INLET SLEEVE WITH OR OUTLET CANVAS BACKING

ADJUSTABLE CLAMPS

Figure 1.74. Showing flexible connecror for blowers and fans.

t The Air Moving and Cond~tioning Associat~on (AMCA) Standards

[ . The Air Moving and Conditioning Association has published quite a few sraii-

dalds for the ari handllng equipment, but here only a few important ones will be dwxsscrl

1. Spark Resistant Construction

The! AMCA Srandard AS-401, lists the following types: a. Type "A" Construction - This states that all parts of the fan that come in

contact with the gas should be made of non-ferrous materials. Bronze and aluminum ;ire commonly used for this type.

b. Type "B" Construction - Th~s construction requires that only wheels and slewes i or rings) around the shaft openings be made non-ferrous materials.

c. Type "C'' Construction - This specifies that the design should be such that it sho~8Id pevent the ferrous parts from rubbing together.

2. Construction by Class

Tlw class dus~gndtioi, just lilnits ths pressures viz. a. Class I - This i s the lightest frame construction and limits applications to a

prcsswc o l ; ~ b o ~ , t 4 inches of water gage. b. Class II - Th~s is a mt!dlum constructiorr and limits pressures up to about 7

,,>,:l,,!s O t "";,tt31 gage. c. Class Ill -This is a heavier frame and the pressures are limited up to about 12

to 13 inches of water gage.

3. Motor Positions

The AMCA also specifies the motor position with respect to fan as shown below. The motor position i s specified by the IettersW, X, Y, or Z , (see Figure 1.15).

4. Fan and Motor Arrangements

The AMCA Standard No. 2404.66, has for the convenience of everybody come I I ~ w ~ r l l the f o l l uw i~~g arrangements (see Figure 1.161.

5. Inlet Box Positions

Thr AMCA Standard No. 2405.66. has the following positions standardized for tht3 ~ i ~ l c t box ( S L W F iwre 1.17).

6. Rotation and Discharge Orientation

The AMCA Stmdal-d No. 2406-66, i s to specify the rotation and orientation of

the discharge of the centrifugal fans. The fan is viewed from the drive side.

Process fiquipmerrl Series Volume 3

cent i f uga l -.;I

Figure 1.15. AMCA Standard 2407-66, showing rhe different motor 00s;- tions with respect to the fan.

Fans Operation

Once you understand the fan fundamentals and have some knowledge of the system in which the fan is going to be used, the selection of a proper fan is relatively easier based on the capital cost, operation cost and the maintenance cost. Fans, like any other rotating equipment, should be selected as close to its maximum or best efficiency point as possible. I f a fan i s selected to the right of its best efficiency point (bep), it means either the fan is a little too small or i s running at a higher speed. If a fanisselected t o the left too far from i ts bep, it means that the selected fan is too large or is running too slow.

Fans, like other centrifugal machinery, operate only where the system curve intersects the fan curve. A t this point the system pressure matches the fan pressure. System resistance is very simple t o calculate for a pipeline or duct and it i s propor. tional t o the square of the capacity. I f there are other obstructions or equipment in the duct system or pipeline, the pressure drop across each item should be added to the line resistance and a combined system resistance be drawn on the same grsph which has the fan curve. I f the system resistance consists mostly of the line loss, the curve will be a parabolic one as shown in Figure 1.19a and Figure 1.19b shows the system curve with some static pressure in addition to line resistance. Normally most fans are suitable to operate in a system shown in Figure 1.19a. The fan selection for systems shown in Figure 1.19b is, however, a little critical. The fan operation where the fan curve is flat or drooping will be unstable and should be avoided, as much as possible. Normally a single fan is satisfactory for a system's operation, but some-

A R l i 1 S W S i ul ,llluct i i r U ill,- I\Rn I SWSI Foi L l U I d r l w 0 6,. ARR. I OWOI F O l bell d h O I O I . ,.,,.., ,,r" ,I,,,,,,> o,, D,,"'. ,r.c, i",1,1.,.1,0,, A ,,.,,,,, U",C,,, 3 1OCI E""l l"Cl8"" nrranwment 3 ,n ,,,, N,, b,..,#~!,,>> "C, I,,,, P r # m e ,> ," ," .>>, :~", ,>4, , , ,u , ,~~,"<.~ pllr, "Ille,", pr,,numorer n ," . , , u*ru i . i i , i , l i l ? * " 8 imcgr.my dl,*, 1," ~ " , l l l , ~ l C i l

>,> <" ' ..

Figure 1.16. Showing different srrengsmentr.

times you may require two or more fans t o meet the demand which may vary from time t o time. Sometimes two fans may be required in series t o meet the higher pressures for a certain operation whereas normally a single fan wi l l do. Similarly two fans in parallel may be used to meet the higher capacity demand at certain times. The two systems are discussed below.

.> ir < .

I'rocess I: i juip~ne,i l Series Vulumr 3

Right top Hor izonta l Right bottom 0ot t b m angular in take r i g h t in take angular in take intake

Le f t bottoin l i o r i zonra l Le f t top TOP i rnta-v angular in take l e f t inrake angular intaue

Figure 1.17. Showing the different positions for the inlet box. The angularposirionr show, are at 45-degree angle.

Series Operation

For two fans operating in serles, the combined total pressure will be the sum o

total pressure of each machine at the same inlet capacity. The static pressure for th, two should not be added to get the combined static pressure. The weight flow-rat, handled by the fans operating in series is the same unless there is a loss or gain o flow in between. The performance of two fans in series is shown in Figure 1.20a.

Parallel Operation

Any two fans similar or a little b i t dissimilar can be operated in series withou, much trouble, but for parallel operation, it is difficult t o do the same withou. sophisticated controls. For parallel operation, therefore, the two fans selectec should be similar in performance if not identical. It i s important they share the total load equally, otherwise one fan wil l be overloaded and the other would bi underloaded or could even operate near shut-off conditions. This type of loadin! could even shift back and forth between the two fans and can damage the fans a, well as the drivers. So you need controls t o make sure the load is shared equally an( no driver i s overloaded. The performance of ,two fans in parallel is obtained b i adding inlet capacities at the same total pressure. To find the point of operation l o the two fans in parallel, you have again to draw a combined system curve (rag

PRESSURE (%)

PRESSURE (Yo)

0 200 400"' 600 800 INLET CAPACITY ( CFM )

Figure 1.2Oe. The poinr of operation is when, the system cvweinmr- sects me combined fan cuwe ( 1 ) + /21. The diagram shown is for two

fanr in series.

0 200 400 600 800

INLET CAPACITY ( C F M )

Figure 1.20b. Shows two similar fans but not identical, in parallel. The curves am not idenrical bur are continuously rising, which makes the

opera tion satisfactory.

FAN OPERATION NOT RECOMMENDED IN THIS RANGE */ /

PERCENT OF VOLUME FLOW RATE

Figure 1.20 (Continued1

Figure 1 .20~. The two fanr have idenrical curve with a hump which when combmed wtli

anorher forms a loop. Operation in rhe loop area rhould be avoided.

Figure 1.20b for the parallel operation) and the point of intersection is the pol; where the fans would operate. The parallel operation is simple if the fan curve continuously rising. The situation can become very delicate if the two fans, e w

though identical, have a hump at a reduced capacity. When you obtain the c o ~ bined performance curve, it would have a loop as shown in Figure 1.20~. The lot

is obtained by adding all the possible capacities a t a given pressure. If the syrtt. curve intersects the fan curve at point 'A,' then there will be no problem in ti satisfactory operation of the fans, provided the system does not fluctuate. But the system curve happens t o intersect in the loop area, it wi l l have two points intersection and thus two points of operation viz. '6' and 'C.' This i s not desirabl because one fan may operate at '6' and the other fan may operate a t 'C' and tht

may switch back and forth. This switching of loads back and forth is also call, 'Hunting' and can damage the fans as well as the drivers. Controls in the fan out1 areas or inlet and outlet dampers may be required to rectify the problem.

Z 1'85 01'Pb LO'OC P0'91 'M'W

Chopter 1. K . S. Paneror

Figure 1.21a. Common terminology for centrifugal fen components. /Courtesy AMCAI

Proccss Equipment Series Volume 3

GUlDE VANES STRAIGHTENING VANES

IMPELLER PROPELLER WHEEL lNTERNAL BELT GUARD

- BEARINGCASING BEARINGTUBE

IMPELLER PROPELLER WHEEL BELTTUBE

lNTERNAL BELT GUARD

- BEARINGCASING BEARINGTUBE

lNLETCONE INLET FLARE

. MOTOR WITH CODLING FlNS

INLET BELL,

,lMPELLER PROPELLER - WHEEL

U 'OUTLET CONE

BEARINGCASING

.PREFERRED

Figure 1.Zlb. Common terminology for axial and tubular centrifugal fans. (Courtesy AMCAI

Fan Selection and Evaluation

Fan selection is fairly easy once you understand the basic principles and the requirements. To get the right fan for the right job, you must send out inquiry with

- a set of specifications and a data-sheet t o several fan manufacturers. The data-sheet should be filled out by the engineer responsible for the selection of fan and should contain as much information as possible so that the manufacturer hasagood idea of what is required. It is better to have more information than not enough, otherwise you may get the wrong type of fan which may not last long enough or may have some kind of maintenance problem. Table 1.1 outlines the minimum information required to be conveyed to the fan manufacturers.

Next step in the selection process is the bid-tabulation and evaluation of each bid. A sample bid-tablulation is shown in Table 1.2. What may look like the cheap- est lor least expensive) offering, might turn out to be higher than most of the bids. Powel evaluation must be done for a t least a period of three years, i f not more. This cost should be added to the initial equipment cost, and then a proper evaluation should be made. In addition other factors like, bearings, couplings, seals ( i f any), material of construction, speed (RPM), outlet velocity (in feet per minute) etc. should be considered. I t does not pay to buy the lowest evaluated fan when it

would be down quite a bit more time for maintenance and repairs. A thorouqh evaluation of each offering, therefore, is recommended and would

pay out 111.111~ lil7ll:S mole in thc long r~111. AS a gi~ide Figure 1.21 gives common ter~ninol<xjy lor various falls.

REFERENCES

1 . Srcp.lnoff, A. J., Pumps 2nd Blowerr. John Wiley &Sons. Inc., New York. 1966. 2. Step lw l l , A. J., Cenrrifugill & Axial Pumpr. John Wlley &Sons. Inc., New York, 1957. 3. Jorgenrun. Robert. Fan Engineering. Published by Buffalo Forge Co., Buffalo, N.Y.. 1970. 4. Jenningr, B. H. and Lewis. S. R.. Air Conditioning and Refrigeration, International Text-

Imok Co., Scrimton. Pa.. 1963. 5. AMCA iArr Mowng and Conditioning Arrocialion). Fsnr & Systems. Publication 201.

Arlingiun Heighrr, 111.. 1975. 6 . Balle, O.E.. A Study an Design and Matching of Turbomachiner: Part B - Comprerror and

Pump Performance and Matching of Turbocomponena. 7. API - 610, Centrifugal Pumpr for General Refinery Services, Washington. D.C. 1973. 8. American Standard, Industrial Division, Detroit, Michigan. Fan Composite Curve. Specific

SpeedISpecific Diameter, 1967.

CHAPTER 2

COMPRESSOR APPLICATION AND SELECTION

RICHARD F. NEERKEN The Ralph M. Parsons Co.

Pasadena, CA

INTRODUCTION

Any chemical, petrochemical or petroleum process which involves the pressure rise of air or gas wil l require compression equipment. Compressor i s the term applied to the rotating machinery which produces such a pressure rise, getting its name from the fact that the volume of gas is compressed as it flows through the machine. In lower pressure ranges the term blower may be used to mean compressor; there is no uniform agreement on where or when the term blower should be used instead of compressor. In this chapter our objective will be to present basic principles of air or gas compression, an overall description of a l l types of compressors and blowers, and certain sizing and performance methods which may be used by t h e process designer or others who are concerned with compressor application, selection, and analysis. Inasmuch as compressors are often rather large, complex, expensive, long delivery machines, it is readily apparent that considerable early effort must be made in any process design involving this machinery.

TYPES OF COMPRESSORS

Compressors can be broadly divided into two basic types: dynamic and positve displacement (Figure 2.lj. Dynamic types produce a pressure rise by imparting velocity t o the gas through one or more rotating impellers. The most well-known i s

the centrifugal type machine, although axial flow compressors are also correctly categorized as dynamic types. Positive displacement types have a reciprocating piston or plunger within a cylinder, or a rotating mechanism such as mating lobes. screws or vanes within a pressure containing casing. Any of these types displace a positive volume with each revolution of the drive shaft, which after allowances for inefficiences within the machine, will produce a nearly constant output a t a given speed. Historically the most widely used positive displacement type has been the reciprocating compressor; currently there is increasing usage of the rotary positive displacement types in various designs as shown in Figure 2.1.

COMPRESSOR OPERATING CONDITIONS

Whether the reader's purpose is to select or specify a certain type of rnachtnr, to make a preliminary size estimate of a compressor, or to examine designs offered by

Process Kquipmeni Series Volume 3 Chnprer 2. Richard I". Necrken

Compressibility Factors

Compressibility factors show how the actual gas will deviate from an ideal gas. They are given or calculated for gas mixtures, usually at suction and discharge conditions, or at inlet and outlet of each stage of compression. For air or a pure gas,. charts or tables are available: for gas mixtures these factors may be determined from generalized compressibility charts (reference 1, 2, 3). or from specific data which is available for a mixture such as 75% hydrogen-25% nitrogen in ammonla synthesis. From the gas analysis and known properties of each component, the critical pressure (PC) and critical temperature (T,) of the gas mixture can be calcu. lated. Values of reduced pressure and temperature, defined as:

wil l enable the use of the generalized charts. Note that pressures and temperatures are expressed in absolute units. Figure 2.2 illustrates a typical gas analysis calculation for a compressor. Note that in this example, the value for molal specif~c heat, MC,, is taken a t 150 F, a common approximation for the average temperature during the compression cycle. For more accurate calculations, values of MC, should be used which more accurately represent the actual values during a specific cycle.

Figure 2.1. C o m p m o r types.

manufacturers for given duty requirements, it is always necessary to determine the basiccompressor operating conditions and obtain or develop certain information concerning the gas to be compressed.

m r r

Hethlnr 6 5 ,

Crllsn. , d l

Propom 6%

i-8utnnc 1,

".B"t."O 2,

Carbon Dioxide 8%

Gas Analysis

I f the air or gas to be compressed is pure, data i s available in the form of pressure-enthalpy diagrams or tables of gas properhes. I f the gas is a mixture, often unique to the given process design, a gas composition, given either as a molal analysis or a volume percent analysis, will make possible the determination of properties of the gas mixture. For air compressors, i t i s usually necessary t o know the relative humdity at the inlet conditions, or otherwise obtain data concerning the amount of water vapor in the entering air.

Comput~ l ion: . .: . 11.14

Cv KO - y.986 11.14 - 1.986 ' "'" Figure 2.2. Typical gar analysis calculation

Molecular Weight and Ratio of Specific Heats

Essential to selection and sizing of any compressor are values of molecular weight, and ratio of specific heats (k = C,/C,) either at inlet conditions, or for more accurate calculations, at the average temperature during the compression cycle or at the entrance to each succeeding stage. These properties may be obtained from published data or readily calculated from a gas analysis.

Given: Hydrocarbon gar mixrure conraining 65% methane, 14% ethane, 6% propane, 3% iso- burnne, 2% normal butane, 8% carbon dioxide, 2% warer vapor.

Find: Molecular weight (MWJ, ratio of sps~ific hears /kJ, criricalpresure (PJ, andcrirical remperarure /TJ.

Chaorer 2. Richard F. Neerken

- Process Equipment Series Volume 3

Pressure and Temperature

Pressures and temperatures are required for selection of a compressor. Pressure and temperature should be given at suction or inlet conditions and pressure at discharge conditions. Discharge temperature is calculated, based on the type o f compression cycle expected, which affects the heat rise during compression. Pres- sures are normally expressed in pounds per square inch gage (psigl or absolute (psial. where psia = psig +barometric pressure (2).

Capacity

Capacity can be stated in numerous ways, based either on weight f low or volume flow requirements. Most common expressions for capacity by the process engineer are:

b Weight flow, pounds per hour or pounds per minute. Volume flow in "standard conditions." defined in the process industries as 14.7 psia and 60 F (or 32 F). Standard units are SCFM (standard cubic feet per minute], SCFH (standard cubic feet per hour), or MMSCFD (million standard cubic feet per day].

b Volume flow may also be stated at inlet f low conditions, usually ICFM (inlet

cubic feet per minute) or ACFM (actual cubic feet per minute).

Unless capacity is given at inlet flow conditions, i t is necessary t o convert t o these conditions. Any or all o f the following relations may be used, as applicable:

molslhour X 379.46 SCFM =

60

Pounds per hour = mols per hour X MW (51

1545 Ts Specific volume, vs = Z, . -. ti3 /I b

MW 144 X P, (6)

1 Density p = - lb/ft3 (7) vs

ICFM = MMSCFD X lo6 14.7 T, 2, .-. - 0 -

1440 P, 520 1.0

PERFORMANCE CALCULATIONS

Recalling now the basic relationship of a gas:

it can be shown that the adiabatic work performed on a gas in raising i t s Pressure from P, to P, is :

i l l I

I f this calculat;on is made in English units, with w equal to the weight flow i r l

pounds per minute, the gas horsepower wil l be:

and the actual gas horsepower based on a known or assumed value of adiabatic efficiency wi l l be this number divided by that efficiency, expressed as a dec~mal.

The expression

is the adibatic head which the compressor must produce to meet the required pressure. It is developed from the adiabatic relation

P, V, = P,V, = constant (131

Certain types o f compressors closely perform according to adiabatic cycles; other types, notably the uncooled multistage centrifugal compressor, do not. Deviatio~l from adiabatic compression results in a polytropic cycle, where the value of thi: polytropic exponent, n, is substituted for the adiabatic exponent, k. For a know11 or assumed value of polytropic efficiency, Q,, the relation between the two

exponents can be expressed by:

ICFM = Iblmin X vs or lblminlp 19)

Process Equipment Series Volume 3

and

Summarizing then, either of these may be used for preliminary sizing of compressor horsepower:

CENTRIFUGAL COMPRESSORS

Although first applied in process use over f i f ty years ago, the centrifugal compressor has come into greatest popularity in the past thirty years and is today the preferred type for most process applications. Current U.S. manufacturing programs from major compressor builden include single and multistage compressors in sizes

from somewhat less than 500 cfm inlet f low to well over 100,000 cfm. High pressure centrifugals have been installed on many gas pipelines, process plants and gas injection plants, ranging up to the highest pressure installation in a North Sea gas plant at approximately 10,000 psi.

weight flow, Iblmin X head, adiabatic adiabatic gas HP = -

33,000 X adiabatic efficiency (15)

weight flow. Iblmin X head, polytropic polytropic gas HP =

33,000 X polytropic efficiency (16)

! I t is important to remember not to mix adiabatic and polytropic values in a given calcuation: either use adiabatic exponent k and adiabatic efficiency, or polytropic exponent n and polytropic efficiency. Adiabatic efficiency may be defined as:

Fundamental Principles

The centrifugal compressor consists of one or more impellers attached to a shaft, rotating inside o f a pressure-casing which has stationary diffuser passages with or without vanes, or a scroll-type volute shaped much like a centrifugal pump casing. The rotating shaft and impeller impart energy into the air or gas due to the velocity of the impeller, then the high velocity gas moves through stationary diffusers or volutes in which the velocity is reduced before exit from the stage or the cotn- pressor casing.

The basic function of the compressor impeller can be seen from examination of Figure 2.3, simplified velocity diagrams for a typical impeller wi th backward. leaning blades. Velocity triangles are shown for the entrance and the exit of t h ~ , impeller. The velocity components c (absolute fluid velocity), u (peripheral velocty of rotor), and o (fluid velocity relative to rotor) are resolved into velocity triangles at inlet and outlet. The ideal head generated by an impeller can be expressed in the form known as Euler's equation:

Compressor calculations may also be performed using enthalpy differences as a means of determining adiabatic head. From charts or tables for the given gas or gas mixture, f ind enthalpy at inlet, h, and at discharge, h z , then:

adiabatic head = (hl - h,) (778) or (Ah) (778) (18)

In polytropic compression the actual discharge enthalpy hz ' wi l l differ f rom the adiabatic value; such discharge point must be determined by including the compression efficiency:

Uz C U 2 - U I C " > Head (ideal) = (22)

9

Temperature rise during a compression cycle may likewise be found either by calci~lation or from pressure.enthalpy charts or tables.

The peripheral velocity is easily calculated from the rotor dimensions:

Thus these triangles are used only for theoretical determinations of basic

Clzapirr 2. Richard F. Neerken Process Lquipmrnl Series Volume 3

ROIATION

Figure 2.3. Impeller velocity diagram,

relationships of blade angles and prerotation at inlet. Assuming an impeller with no inlet swirl, c,, = 0 and the basic equation for head becomes:

c", 2 Since c,, = U2 - w,,? or U2 - -- tanp,

then

- In .i radl.ll-bladed impeller, wherefil = go0, the expression becomes:

and can be shown to decrease as the blade angle decreases in backward-leaning

impeller designs.

Flow

Figure 2.4. /deal and acrual heads.

The above relationships are theoretical; actual impellers produce significantly less head because o f inefficiencies due to disc friction, slip, incidence losses (at off-design conditions), and other factors. Figure 2.4 shows theoretical heads and

actual heads for radial-bladed and backward-curved impellers. I t illustrates the fact that the radial bladed impeller produces higher head but has a curve shape which tends to be flat, whereas the backward-curved impeller curve shape has greater slow and a wider range o f stable operation. Both impeller types are in wide use in

commercial machines today. Although it is helpful to the process engineer to have an understanding of the basic fundamentals, selection of each impeller for a g1vi.n

application must be left t o the compressor designer.

Head Coefficient and Flow Coefficient

H w d pcr stage (actual) . g :

u; (281

~11111 TIIB f l ow c u e f f c ~ e n t may be determined at the inlet or exit o f each impeller. A t

uthsl thdil the d r sgn rated condit ion, similar valttes o f and qI can be found.

~ c s u l t ~ n g in the display o f a given stage or an overall machine performance curve i n

several manners (Figure 2.7 ) , showing f low versus head or f low versus pressure,

both for one single speed, also f low coefficient versus head coefficient (dimension- lc:ssi ~v l i ! ch 1s LISUIUI in generating performance curves at varying speeds or on

;iItt!~ n:iw g.is condi t io~is in the same compressor. These curves also show efficiency, : I I I s s i t l t 1111~1 i117d exit f low angles f o , off-design

~ : i ~ i i i i i i ) n s i v i l br l c % t i i i i i , ~ i ~ t i f n u i n , b la i l~! ~~ i c i d~ inc t ! I ~ S S C S higher, resulting n v,iy11!1 ilvu1.1l1 c ! l f~c~uncy througlioor thi! stable apelatin(] riinge.

F i , b.~chwti~il-lu:iz,~n<] closurl impt!lli:~s which arc LMXI in conventional m(iltistagc

~:U~I~I)I~:SCOIS, irnpal lc~ peripht!lal slx:r:<l, U. will raoge thom app~.oximalely 750 tu

900 ft.lsi!c unless h iyhrr molecula~ wcight gas raquiros lowcr speed. Single stage and

~ . i r i~ ; i I t y i x wiiculs m.ly i l t i l t s h~glir?, viilucs o f U. A qo i c t check o l the relation of

Figure 2.5. Prelimii>ery selection data for cenri i luyai comirierrois.

. . .. .- .

;,o:.l !,!I:

1 1 ~ ; e l l e r :>:n::ia!, 17r.?es

.. .

1 3 - l t

. . . / - I .

'1-.,.

11. - ,! t

39- I:,

.jt .!..'

L, [., . , :,

Ir . l? t Flow Hsnpe ICFM or biS

5 0 0 - 7 . 5 0 0

1 . 5011-7, 5 0 0

~ . o o o - ~ ~ . ~ ~ o

t .C00-1E.Ooo

~ , , c w > - j ~ j . o o o

70.11W0-t,0. 0 0 0

h 0 . U 0 0 - 1 0 0 . 0 0 0

A Y r r s i F l e n d CoefflclentC

0.LH

0 . b ~ t o U.511

0 . 5 0 to 0 . 5 2

0 . 5 1 t o 0 . 5 .

0 . 5 7 - t o 0 . 5 1

rJ .5 . j L C 0 . 5 6

0 . 5 4 to 0 . 5 5

S:mplc Calculations

Fi!lurt. 2.9 shows a typical prciminary calculation for a multistage compressor

11511x1 till: owri i l l p i~lytropic head inuhod. The: gas is the same as the example in

F I ~ J U I C 2 .2 . Figure 2.10 shows how a calci~l;~tion is made on a pure gas such as c,irbof, i l~oxide, ut i l~zing a pressurcenthalpy chart to determine values of head.

i ~ w c f r : volumr. and temperature.

TYPES OF CENTRIFUGAL COMPRESSORS

( : , , I I I ilu~i.11 c : u ~ ~ ~ ~ , i ~ s s ~ ~ ~ s ;,I,: b u l t I I w,~i. : i taj l t : a11,1 ~mk~ltistiige designs. The single^ s i , i q i , ~.,,~n,,~s,ss,j~ 11,ir OIIC ~ r n , > ~ ! l l ~ y usu,r//y ~ ~ w r h u ~ i g /(om a bearing dssemMy and

, : ~ ~ l i ~ c l in .I v d i ~ ~ t ~ diffuscr asrc~nbly sim~l;!~ to a largi: centrifugal pump (Figure

2.1 1 ) Sucl? types a r c commercially available today in sire ranges from about 1000

ICFM to 150,000 ICFM. The enclosed. backward 11:aning. two-dimensional impeller

( F I ~ L ~ ~ c 212a) produces adiabatic heads up to about 12.000 f t lbllb. The semi-open

. IMPELLER TIPSPEED. FTISEC

rachal bladod design, usually with prcrol;~lion or tndut:w varws rn;lk~li$l 11 it^!,,,,

dimensional (Fiyurc 2.12b) produces hcads lo 20.000 I t ibl lb. Sim,lar dowjr , i i r , .

slightly curved blading and inducer blading (Fiyure 2.12~1. H~ghcr rtrrii<]th im;ilt..

als such as titanium may be employed in small machines, allowing hgher ~ ~ ! ~ , i ! l ~ ~ .

t ip speeds and resulting in higher heads. Small industrial high speed desigris a r i ~i ! : ,

available today, making this type in one form or another a very versatile mach~l:,.

Many process applicatiorls require higher heads or pressures than a singlc st.jiq

machine can produce. Multistage compressors thus have become the most widi:.,

used throughout industry. Many different design arrangements are available I!) i i ~ m

widely varying service conditions. The most common type is the uncoolvd

"straight through" variety (Figure 2.131 where as many as eight or ten impelrirr , t i #

arranged on one shaft, running at the same speed, and each producrny ajipiorl

mately the same head. Head per stage wil l vary to a certain extent a s the vol~imc,

flow through a given machine is reduced and succeeding stages or irnpi.Ilr:l-s . ~ I c ,

designed for lower volume flow and hence exhibit somewhat lower head ~ o i , f l ~

cients and efficiencies. Figure 2.13 shows the caslng split hor i~orr ta l ly par;,lli:i I , - the shaft, with a flat joint held tight by numerous casing studs. Figure 2 1 4 siiiiv,,.

similar type machine where the outer casing is a cast or forged barrel, h ; l v . ~ ,

horizontal joint. The rotating eemcnt is inserted from the outer i:r!d d ~ l r J , I '

vertical casing jolnt is made to withstand higher prussures ;is wi! l as in , , ~ O . J ~ I .

Flow, c f n a t i n l e t 4500

S p e c i f i c volune. f r 3 / l h ' 0.774

Weight f low, l b / m n : 5815

Adiaba t ic exponent : 0.178

Impel le r d iamete r U / 19" ~

AT RATED COIUITIOX

Head c o e f f i c i e n t , P : 0.50

i Polyf rop ic e f f i c i e n c y q ! Polyr rop ic exponent C i 0.252

P t

Rat io of compression j 3.0

Po lyf rop ic head 42070 -- I:;. I

Figure i IiHead c o e f f l c i e n t , y -~--. . -

li - - - ----l,Efficienc::,

Figure i !I ?IP i i f o i y m a p ~ z bead !i

Figure 6 :~~TO\ERLO:D COSIIITIOF i _ _ p ~

Equation 11 /j%pproXimate f i o u , ICFH

~ d / P r l inead Equation 12, us ing po ly t rop ic l lEf f ic iPnc? , qp expOneO:

~ ~~.

I Polyr rop lc head

0.72 , U/va

~ .~ - ~ ~~

, E\PL*tiATIO'. , ~~.~

2475 : 55% o f r a t e d

0.56 Llfr's d a t a

0.705 ,. ,, i

I Figure 2.9. Cenrrifugal compressor calcularioion - polyrropic head m e t h d

P." Olagram :'

Clvcn : Frame s i z e

X X YI j Impeller d i a t s r . D

p-,, D,agram :' IKead coef f ic ien t , /-

, ,. , Nunber of stager

i, Impeller t i p Speed U

61 \ .~ i l ; \ o f e I ' ' Rornfing speed N

P-,i l l > r g r a acoustic velocity u,

,, .. ' .'Preuda,' Warh NO.

I 5175 (!IS\ o f r a t e d

I 0 . 3

0.69

36180

nfr.5 d a t a

., ,,

Equatlon 12a

' a3u~ l cq lelxe ia l laq fiu!plnoicl s l l l j l pot,

Jalladw! q3ea Aq pJ3npold lsnrq l le!xe aqi aar6ap a61el e o l fiu!z!jenba s r i ~ l l ' l jeqs

awes aq l u o suo!l3a~!p al!soddo u ! 6ui3e4 slal(adw! 4 1 1 ~ p a 6 ~ ' e ~ ~ P ale ' ~ 0 1 , 2161~1s Ailensn 'sa6els 6u!paa33n~ -~a l ladu l l XIAI uo~lni is-alqnop G l o ;rpls otlo i l l i iw I L ~ I C M O ~ I

4 3 ! 4 ~ JO 43ea 'swealls OM$ 0111! p i i p ~ n ~ p ~ o l j aqi q l ! ~ '26eis ~ S J I ) ;1~11 uo AI I I 'H~I ,

' lalladw! ~ 0 1 4 - a l q n o p e 10 asn :a41 s! x1A1 a6rls1l jnlu aql 01 u o ! l r i P n ~ ,?~ I IOL IV ~ 1 ! 1 1 1 , ' I , ' ~ ~ I / 1 )\,':1 ,,,/I l / t > . 1 % 2

- In0 JeJ sa6c1uenpc aql 'i;,AaMOt( StiO!ll!3!~dd,! ii/L'lli>3 L i ( '60,dltl 11111- 5 1 ; ) ) 0 < ) . 1 l i l i l l

10 uo!lellelsu! pue Alddns l o slso3 a w snld ' sa r rou lapno plic j a i u I I ~ O ~ I I I I I ~ ~ V

q i ! ~ sufi!sap 6u!se3 (0 SOD ~ a q h ! q apnl3u! s a f i c ~ u c n p ~ ! ~ '3!lcqc!pc o i Alasolo m n i n s w r o ~ l a d uo!l3as l o a6els q3ea re @sap s!ql Aq pasealnu! s ! Anux3!l,c t~oiss;iiduios lle>anO s ~ a l l a d w ! aalql r o o ~ i l o dnorfi q3ri) 10 ialladw! i l ow ilaannlaq I i l l l jcu~,

se6 10) SJa1003Jallll q l ! ~ 'su0!13as o l u l p o p l n ~ p aq urn SI-IOSLilld1U~D ; , i i i : t s l l ( t l p J

f i s c 00s lnoqe ueql raq6iq aq II!M a.lnss.,~d i , f i~cqss~p aql 4 1 wn l x l u i s ~ l i io seb A i l i ! in1 .? r iA l

jajieq a w nsn o i Ql3Pld O O ~ I L I O S SI 1 1 'm1~:2rfi 1 0 "A,OL s! i i l n lx lu i si'n :i l l1 1 0 111111110:1 il;lfiolpAq aqi I ! Risd 052 002 :~noq? s i ~ . ~ ~ m : ~ ~ d ~ n j p:lsc aq )SI ,LLI fiu!srn i ? , I A l 1 1 1 1 1 1 ~ /

io ]!Ids A~~w!IK'A aqt l cq l s:8ij!:,nds ( P s ~ s s . w ~ ~ u ~ ~ ~r!firijrrtcr.yj r o l ( L C ) I""!'

- u e l ~ alni! lsul u in l jo l lnd u e 3 ! I . r w ~ 3 4 1 ,'firyl:w srli lil;lA,ilrl o l r:ls . 1 ~ 1 1 1 ~ 1 1 1 1 ,110111

Figure 2. 17. Modular type lowpresrure air blower.

tion by the specifying engineer or the designer to ensure a completely adequate system. American Petroleum Institute Standard 614 (Reference 5) has been written to guide users in the requirements of complicated lubricating and seal oi l systems for special applications. Numerous diagrams appear i n the appendix of API-614 which can aid the engineer in considering many optional arrangements.

Shaft Seals

Equally important in the design o f the centrifugal compressor i s the type of shaft seals to be used to prevent excessive leakage of the compressed gas to atmosphere. Seals may be classified into four basic types:

0 Labyrinth type 0 Restrictive-ring (carbon-ring) type 0 Oil-film and pumping-ring type 0 Mechanical-contact type

Labyrinth or restrictive ring types may be used only when some leakage of air or gas can be tolerated. 0i l . f i lm and mechanical contact types are normally used for

Process Equipmenr Series Volume 3

Figure 2.18. Compresson arranged in r e r k

any process gas or gas mixture. When even trace quantities of gas cannot be allowed to atmosphere, (such as in gas mixtures containing hydrogen sulfide), buffer gas is required, at a pressure somewhat higher than the sealing pressure. Buffer gas is injected into the seal and forms a barrier between the compressed gas and the,

atmosphere. Source o f such buffer gas may be a problem, even requiring a small additional compressor in some applications, to supply a gas such as nitrogen at tht: required sealing pressure. Increasing restrictions on gas leakage to atmosphere bc cause of pollution restrictions has made the use of buffered-gas seals more w~de -

spread. The specifying engineer as well as the designer of the compressor must consider this subject carefully. Figure 2.19 and reference 4 show some views of

typical compressor shaft seal types. For the oil-film and the mechanical-contact seal, oil is also required to form thi!

oil f i lm between seal faces or rotating and stationary members. Again, i t i s necer. sary t o have a complete, self-contained oil circulating system similar to the lubricating oil system. The seal oi l must be at a pressure slightly above seal operating pressure and frequently requires the use of high-pressure seal oil pumps, filters and coolers. In many designs, the seal oil pressure i s kept at a fixed static head diffcrcri tial above the seal by the use of overhead seal oil tanks, thus automatically rcgu

lating sealing oil pressure even in event of changing gas pressures. Contaminated se~ l

oil, that i s , o ~ l which has come into contact with the compressed gas, may need t o be discxried o! reclaimed before reuse. Contamined or sour oi l drain traps may be used, together wi th some external seal oi l reconditioning facility, all a part of the

- complc?~ com~~~essor installation.

Control of Centrifugal Compressors

Recalling that the centrifugal compressor develops head, not pressure, it is easy to understand how this type machine can be controlled either by throttling or by

vatying speed. Centrifugal compressors fol low the affinity laws. similar to fans or cerltrifugal pumps:

The most effective way to match the compressor to the required output i s t o vary the speed. Many centrifugal compressors are driven by steam or gas turbines making variable speed an inherent feature. Although in the past most electric motor driven machines have been constant speed applications, the variable frequency electric tnoror dr~ver has made a new appearance in the industry, and appears t o have much application for the future.

Looking again at Figure 2.7, showing typical curves for a centrifugal compressor. i t is readily seen what the effect wil l be if the speed i s varied and how the use of head coefficient and f low coefficient wil l make i t possible to find required speeds for dilferent gases in the same compressor.

I f fixed.speed operation is desired, the compressor may be controlled by f h ~ o t t l ~ ~ ~ l ihc dischxgo (least efficient), throttling the suction (more efficient), or use of vwdble inlet guide vanes (Figure 2.20). The latter method uses stationary vanes ahead of the first stage impeller to provide pre-rotation o f the gas stream flowing into the compressor. This in turn wil l cause a variation in the shape of the head-capacity characteristic as the vane angle is increased, and will result in con- slderable power savings (Figure 2.21). Inlet guide vanes are most effective on single- stage tnachines, but may have limited effect on multistage machinesdepending on the nu~nbr r o f sets of guide varies installed.

Aiwthrr asvect o f compressor control which must be considered i s that o f surge con t~o l (01 anti-surge control). Every centrifugal compressor of whatever size or design wil l l ime a minimum flow limit, known generally as its surge capacity. Below this capacity the operation is unstable and wil l cause rapid deteriation, increase in v~bratiori a ~ i d probable failure of the machine in a short time. Operation at or near the surge region is unstable and must be avoided, except as experienced during startup or shutdown cycles where the compressor may be allowed to pass through the surge region as it comes up to speed.

Figure 2.20. Variable inlet guide vanes for comprerror

The most simple solution t o the surge problem is use of a bypass valve which will blow off the excess capacity to atmosphere in the case of an air blower, or alr compressor application, or return the excess gas to the suctlon source or a suction vessel with or without cooling on the way. Since the temperature of the yas I S

increasing during compression, if much of this gas i s bypassed back to suction without cooling, the gas suction temperature wil l rise to an unacceptable value in a very short time. Surge control systems may be designed or purchased from firms who specialize in such designs (Reference 6. 7). The added cost of such special systems can often be justified in a very short time because of the savings in operating costs and the increased dependability o f such a custom designed system.

AXIAL FLOW COMPRESSORS

Dynamic type compressors also include axial flow machines, in which the flow of air or gas is parallel t o the shaft axis. These types utilize two-dimensional flow analysis rather than three-dimensional analysis as required by radial flow centrifugal wheels. For industrial or process use, the axial f low compressor is generally applied only for very large f low volumes (at least 75,000 ICFM or higher) and is most often used on air, although it has seen limited application on certain gases. Efficiency of

temperature rise, capacity and power determinations as previously given for centri-

fugals. Head per stage i n an axial machine is usually no t more than one-half the

amount for a centrifugal. I t is common to f ind axials applied w i th ten or fifteen stages on one shaft (Figure 2.22) when the required pressure rise demands that

rn.irlv sr.igt:s. Probably the Ialyest apl,lic;itio~l fo r th< : axial f low compressor is i n the i : ~xn l r u r l i i ~ l l gas ~ I I I ~ I I I ~ : , eithc! fo r industlial use or as an aircraft engine. Several , pjixcsses I,, the h y d ~ o c a ~ b o n p~ocrss ing industry, notably f lu id catalytic

c ~ . x k i ~ i i , make wlde use o f the axial f low type. The process or specifying cnginerr should look carefully at the axial f low compressor i f the volume o f alr or gas t o be compressed i s at or near 100,000 ICFM or above.

Axla1 compressors are readily controlled by v;irying the stationary (or stator)

the axa l f low is generally higher than the large centrifugal o f comparable size. The

cost i s usually also higher b u t can generally be justified i n view of the power saving. The axial colnpressor stage follows the same basic relationships of head, I

140

Y

g 120

$ I00 '0

8 0 3 60

,- g 40 E W

a 20

0 40 SO 6 0 7 0 M1 9 0 100 110 120 130 140 I W

PERCENT DESIGN 'IOLUME

Fiquix 2.23. Aniai l iow ocrlurmorice curve.

blade settings, analogous to the inlet guide vanes on centrifugals. Due to the two- dimensional flow, i t i s easier to use adiustable stator blades on many stages i f desired. It is common to find five stages of a ten-stage axial each with its own set of adillstable stator blades, thus giving a fairly wide capacity control range to a machine type which is somewhat narrow in i t s basic design. Figure 2.23 shows a typical axial-flow performance curve.

DRIVERS FOR CENTRIFUGAL A N D AXIALCOMPRESSORS

Three types of drivers are commonly used for dynamic type compressors. Most common is the electric motor at fixed speed. Steam turbines and gas turbines are the other two types.

Electric Motor Drivers

S~nce the centrifugal compressor operating speed is almost always 3 6 W rpm or above, the two-pole induction motor may be used for direct connection to a 3600 rpm machine, or in conjunction with a gear speed increaser for higher speeds. Four-pole or six-pole motors are generally used when gear speed increasers are requ~red, as they are somewhat less expensive and more readily available in larger slzes above about 1000 HP. Speed increaser gears are usually parallel shaft, single- ~cduct ion type, either double or single helical configuration. Epicyclic gears are also used, notably by foreign manufacturers, and this type wi l l often show a cost saving over pat allel shaft type. It i s not the purpose of this chapter to cover the subject of drivers and gears in detail. The reader is referred to available standards (references 8. 9, 10. 1 1 ) for aid i n sizing or specifying a motor or gear.

S t e a m and Gas Turbines

As ment~oned earlier, turbine drivers offer the advantage o f variable speed operation, plus sometimes a more dependable or available power source for remote installattons or where electtic supply lnay be subject to interruption. Steam turbines to drive centrifugal or axial compressors are usually multi-stage turbinesdue to siie and speed requirements. Small centrifugals, especially the modular type air blower, may utilize standard single stage turbines.

Part of any specification for a steam turbine driven compressor must include a cluai st;atrment of steam conditions, both at inlet and exhaust. I t is surprising how many process engineers ignore this until pressed for an answer. Just as the compressor cannot be sized or buil t without knowing both inlet and outlet conditions, - ~n:ither can the turbine. Consideration should be given to use o f a direct-connected turbine or a turb~ne with speed increaser gear, usually only desirable on smaller units as a means of saving on first cost. Lubrication requirements o f turbines should be considered when specifying or engineering the total installation; it is quite common to combine lubrication requirements for compressor, turbine and gear ( i f

used) into one lubrication system. Pneumatically or hydraulically controlled turbine governors provide the most dependable method for utilizing the variable speed feature which every turbine offers.

Gas turbines used to drive compressors are of two types: expansion gas turbines and combustion gas turbines. They are really both closely related except for the addition of the combustion chamber and fuel system in the latter type, which produces the hot gas at a pressure high enough for it to expand through a gas expansion turbine to produce power. The combustion gas turbine is usually a large machine and has excellent operating records driving pipeline compressors and gas compressors in gas treating, gas injection and uas ~rocessing facilities.

Expansion turbines may be the hot-gas type, as used in a combustion turbine, or may be cold-gas, radial inflow turbines, usually applied to smaller loads and often when theobtaining of a low temperature from the turbine discharge wil l serve a useful purpose i n the overall chemical process. Again, the reader is referred for further details to numerous references for gas turbines. (references 12, 13. 14).

POSITIVE DISPLACEMENT COMPRESSORS

Reciprocating Compressors

Reciprocating compressors cover a range o f capacity from the smallest flow required, such as a uti l i ty air compressor in a service station or a verysmall plant, up through sizes about 5000 ICFM. In large sizes for new installations, rotary or c r n t r l ~ fugal types wil l be utilized today to avoid what some consider to be excessive maintenance on reciprocating machines due to piston rings, valves and packinqs. Many other users have found the reciprocating type to be their standard for dependable service and continue to install this type on new applications.

Small sizes may have single-acting cylinders and be air-cooled, resulting in mimrt~, of oil vapors from the crankcase with the air or gas being compressed. This type s not recommended in process service or for instrument air service where oil in the compressed air would not be permissible. Several variations of the air-cooled design have appeared, where a double acting crosshead is inserted between the crankcase and the air cylinder, making the design "oil-free."

Small single-cylinder process types (15 to 200 HPI having water.cooled cylinder and arranged either horizontally or vertically (Figure 2.24) may be used for instrument air or process gas. The cylinder is double-acting type, with a separate packing box and distance piece, thus making it possible to have oil-free air or to conduct leaking gases away from the packing under slight pressure.

Larger reciprocating compressors for most process applications wil l usually have horizontal, water-cooled cylinders. More than one cylinder i s generally required. with machines having four, five or six cylinders being quite common in process use (Figure 2.25). Cylinders are placed on opposite sides of a frame and crankshaft, giving a close approach to a balanced design and comparatively low forces and

PROCESS REQIIIKEUEHT5 [Ci

Gas

Wlecular neighc. m

S p e c i f l c Hear Ratio, k

Flo*, MlSCFo

Suction pressure, pr ia

Temperature, *F

nischarge pressure, pr ia

. I~ORSEI'OWER CILCULIITIO\

LXPUYATIOS

Figure 26

1:

Hydrogen . Hydrocrrhon Uir iure

Cr ink lhaf t speed. rpm

lumber o f rragcr

Cylinder c learance . P I

gcr cy l inder , c fn / ilJ

Comprerrlbll iry factors: \ t IIICCiO" zs I discharge Zd

capacity at inlet conditions, ICFM V E . ~ 130) cylinder displacement. CFM

The i:ylindc:~ d~splacement can be calculated once the size of cylinder, speed of

~ota t>on, and piston stroke are known.

M,my lotmulas 101 volumetric efficiency have been suggested, determined from

tht!o~y or from actual observed results. The following is adequate for preliminary

w h u r clcarmce 1s exptrssed as a decimal.

Fr i~ rnr Load or Rod Load

The con,presso$ flame or pistol, lod has a litnit to the forces that can !d

dul i~rg the compression cycle. In its simplest form, the actual loadings can be

comouted from the known cylinder diameter and the pressures acting on the

piston. This neglects the frame loading resulting from the reciprocating weights and

imotor, oi the machine.

For ;i cioriblr acting cylinder. when the piston 1s moving inward toward the

c~ ,~nhs I~ ,~ f t , the frame load in complr:ssion is:

F g u ~ c 2.29 illustrates the baslc relationships ajid terminology. Values so calculated

niusl be checked against manufacturer's published frame limits; i t is recommended

"HEAD E N D + CRANKSHAF

PACKING OR "STUFFING B O X

PISTON ROD MOVING TOWARDS CRANKSHAFT:

FORCE - LB ICOMPRESSIONI - (Pd) (Ahe) - (,Ace)

PISTON ROD MOVING AWAY FROM CRANKSHAFT:

FORCE - LB ITENSION) = ts) Phe) - (.d) (4.)

that the actual values for a given application be not more than 75% to 85'h UI 1 1 7 , .

published maximums. I f the values calculated on a preliminary selection 81,. I.,. high, then the size (diameter) of the cylinder must be reduced, two cylirldui, I: . G

instead of one, or a larger compressor frame with greater lramr ratInq r:;iimi,~' 1

must be chosen.

Compressor Speed

Limits on rotating speed and average piston speed should be specifled lo a v w i

selecting a design that operates too fast and gives excessive wear and rna(r\tel!dllt.i.

stroke (~nchesi average piston speed I f t lmin! = rpm X X 2 ( :j '.,

12

The general l imit in use today for process service is 800 to 850 f t l m n lor l h i b ~ ~ , ~ ~ ~ ~ , ~ pistons and 700 to 750 f t lmln for non-lubricated, carbon or carbon-l~llicd ti.Ili,u - ring types.

Control of Reciprocating Compressors

Because reciprocating compressors are positive displacement type. a ~ I L W I ,

machine wil l produce an increased pressure over the rated pressure in cvrint 0 1 .r

change in downstream conditions. Closing a valve, changing anothcl CJ<OL.. '>

Chapter 2. Richard F. Neerken

variable, shutting o f f a reactor or other process item downstream of a compressor can cause the pressure to rise to a dangerous level. A relief valve or high pressure shutdowii may protect the machine from failure but may also upset the process or cause waste or tieedless plant downtime. Compressors wil l therefore be equipped with some form of capacity control, basically one or more of the following:

Extc~o;, l bypass of air or gas from discharge back to compressor suction sotrlcr, bvith or without cooling Cylinder unloaders Cylindtar clearance pockets Varii~blu speed

Most reciprocating compressors are driven by electric motors so variable speed is normally ruled out. Bypass type arrangements are necessary for close regulation and minute-by-minute process variations, but are wasteful o f power. More economical c o ~ ~ t r o l methods include the use of cylinder unloaders and clearance pockets.

Cylinder unloaders are manually or automatically operated devices on one or both cnds o f a cylinder, designed to unload or hold open the cylinder suction v.,lv,:s. Thus the compressor does not work on that portion of the stroke. For ex;~mulo, mlct valve unloaders could be placed on the head end (outer end) of a ~ : y l i O d ~ l , r e d ~ c i r q the net output o f that cylinder by approximately onehalf when the ilcvices wele actuated Unloaders are usually supplied as a means of totally ~ ~ ~ ~ l o . i ~ l i n g ;I compressor for startup, but prolonged operation with unloaders on one elld of a cylinder may cause problems of frame load, valve life, or pulsation damp-

ing. Clea~ance pockets are additional volumes of clearance built into or bolted on to

;I cylinder head or valve cap, either head end or crank end for both) to increase the cylinder clearance. Reference to equation 32 wi l l show how increased cylinder clearance results i n reduced volumetric efficiency. Figure 2.30 illustrates this relationship graphically, and Figure 2.31 shows a typical cylinder wi th both automa- t icdly operated unloaders and a manually operated valve-cap clearance pocket.

Retlief valves must always be provided in the piping immediately downstream of any rec~urocating compressor. They should be sized for the full output capacity of the compressor, set to open at 10 t o 15% above the rated compressor discharge pressure.

Piils;itson rlampe~s are generally installed with reciprocating compressors on piocrss se~.vices, to smooth out the pulsing f low generated by the reciprocating Illston. Thrsr (nay be in the form of volume bottles or special devices wi th internals designed to absorb or cancel some of the pulsing flow. It i s common t o specify that the pulsations in the piping leading to or from a compressor shall not exceed 1% of the opelatiog pressure in pressures up to about 400 psig; lower values are expected and required in pressures higher than this. These devices require a pressure drop as the gas passes through: careful calculations for sizing reciprocating compressors wi l l ~ , c l u d r ;~lli>w:mcus for such pressure losses.

1 Process Equipment Series Volume 3

1 Figure 2.30. Variation in campressor capacity with increased clearance.

Var~able Speed Drivers

t i I S r e the most coinmoll type of variable speed driver applied to ~ucip~ac,i:inij cornplessors. Figure 2.32 shows a typical integral type gas engine

C O I ~ ~ ~ S S O I . wlitxc ihc engine mtl compressol. cyli11del.s are attached to the same

I r j r n u .ii,d cr.inksh'ift. Thlr type i s widely used in gas t!ansmission and oil and gas p~oduction.

f i g u r e 2.32. Gas e t w n e driven recipra'ating comprerror

Steam turbines have been applied to reciprocating compressors by use o f single

0 1 double-reduction gears to reduce turbine speed. Although many successful in-

sr,illations have been made, caution should be used when specifying this type drive

31~3ngZlnt?llt as cons~derable additional analysis of torsional vibrations, couplings, ti l l y w l i ~ d z I S icc j~l~red.

ROTARY COMPRESSORS

l % l <,lIi,:i l w m 0 1 posltiv<: ~ l i s , , l . ~ ~ : ~ ! r ~ ~ ~ : ~ ~ ~ comcwt:ssut is thc rolrlly comprrsso! or

blowel, i l l ,nhich a rotat i~ ly element within a casing displaces a fixed volume during

e x h ~evo lu t~on. Mall" different types are-in use today, which can be grouped into

i o w basic categories:

Lobe-type 12-lobe or 3-lobe1 Vane-type

Screw-type [wet-screw or dry-screw) Liquid-ring type

Figure 2.33. Lobe-type rotary blower.

Most widely known is the lobe-type (Figure 2.33). Two figure-€ shaped I<,!cw;

(or variations from this basic concept) mesh together at relatively slow spwt i , driven by timing gears attached to each shaft. This type is available in very s n l ~ l , '

sizes. from approximately 2 ICFM. to largest sizes of about 20,000 I C F M 1 1

basically a low pressure machine, producing up to approximarely 15 USIS) ~4s

blower,and is widely used as also a vacuum pump. The three~lobe type will ~ I U O L L . .

slightly higher pressures, to abocit 20 psig. Casings are normally modc fhwn i,:.;

iron, which limits this type for certain applications where steel casrjgr ~i, , ,y I,,,

required because o f ha~ardous conditions. The lobe:-type blower hnds wdimt ,>j,lj88

cation har~clling air or ir irrt yai, although in certain a~,iilicattons it is ~ s i , w ~ r i ,

used as a gas pump on natural gas.

The second type is the sliding vane compressor, having on? offset rotol l i i l ,

slots in which vanes slide in and out during each revolution. A i r or gas is trapped within the casing and its volume gradually reduced as the pressure rises t o discharge. This type call produce up to 50 psig per stage and is also available in two-stage arrangements for pressures to about 125 psig. Volume flows range up to 2500 ICFM. A dlsadv;int.iqa of this type for some process applications is the fact that the sliil$ng vanes rerlrliro some lubricant to be injected into the casing and trace quantities of this oil wil l thus appear in the compressed air or gas.

F i y t ~ e 2.34. Screw-rvpe rotary compressor.

The rotciry screw type compressor has become more popular in the process ~ndust,ies in recent years. Although the design was introduced in Europe before World War II, U.S. ~nanufacturers did not build this type in great quantities before the 1950s. Today the rotary screw type in the "wet-screw" or oil flooded version has taken over the portable, construction air type machine almost completely, and also occuples an important segment of the refrigeration and air conditioning com-

pressor market. The dry-screw type has had some process applications and i s gradu-

ally becoming more widespread. Both types are similar in design, utilizing twln

screws which mate together and are driven thorugh timing gears. Lower speeds (to 3600 rpm) are used in the oil-flooded types, while the dry-screw types operate LIP

t o 12.000 rpm in smaller sizes. Figure 2.34 shows a sectional view of a typical screw type machine. Casing working pressures are limited to 250 psig to 400 psig, thus making this type usuable only in lower pressure applications.

The w e t x r e w or oil-filled machine wil l produce higher pressure rise per stage, with the heat of compression being absorbed by the oil which is injected into the machine for lubrication and cooling. The majority of this oil is removed after the compression cycle is over, and normally cooled and returned for re-use. The procuss designers should avoid this type however, if oil carryover into the process gas cannot be tolerated.

The liquid-ring pump or compressor i s correctly classified as a rotary, posrlve displacement type. I t consists of a circular or elliptical vaned type rotor turnny in d

circular or oval-shaped casing, in which water or other sealing liquid is also present. Centrifugal force causes the liquid to form a ring around the periphery of the casing when in operation. The air or gas travels inward towards the center of the vaned rotor, gradually decreasing in volume and increasing in pressure until i t passes discharge ports and leaves the casing. Liquid still present in the air or gas is separ ated and either recirculated or discarded. This type i s most often used as a vacuum pump, down to absolute pressures o f 3 to 4 inches o f mercury. It may also be used as d compressor for pressures up to about 100 psig (basis two stages in scrir:~). It has been successfully applied on certain difficult gases such as chlorine, hydrogeii sul fide, acid gas and so forth. Stainless steel construction is available in most slzrs.

Rotary Compressor Calculations

The rotary compressor fills only special needs such as low pressure rsr , low capacity requirements. Calculation methods are not as widespread in the industry a i for reciprocating and centrifugal machines. Adiabatic head, weight flow. inlet

capacity calculations may be applied in a similar manner to other compressor types. however a widely recognized source of information on efficiencies for these types IS

not available, and most preliminary size estimates must be based on manufacturers catalog data. A sample calculation for a lobe-type gas blower, using data from a

U.S. manufacturer, i s shown in Figure 2.35.

CONCLUSION

A great deal of material exists in published literature today relatlng to com pressors of all types. The reader is referred to certain references in this chapti!r, olio to other readily available material to investigate further almost any asvect o f corn

pressor theory, design, selection or performance. References 15, 16. 17, 18, r:t 31.

ACKNOWLEDGEMENT

Tl?r wi t e r acknowledges with thanks the information, photographs and data ~ i u i v c d f m n the following manufacturers of compressors and blowers: Allis- C/I:IIIIII~ISCO., CUDQ~, 1nd~~stres. Drt:ss~~-CIa,k, DI~SSCI-Roots, Elliott, Hoffman,

l ~ i j ~ ~ ~ s o l l - I i . i ~ i ~ i . Joy Ma~~~ i f i i c tu r~n ( ] , Nash Engineering. Transamerica DeLaval, \YmIh i~ j~ l i r t , . Al1.i~ Copco, S u l i r ~ B ~ o t h e ~ s LId.

NOMENCLATURE

Ace A, C

CFM

9 GHP h

H

ICFM k

CP

I MMSCFD MW n

N

P psia

w ig 0 R

SCFM

t

T

U, u

"s

Area, crank end

Area, piston rod

Absolute fluid velocity Cubic feet per minute

Specific heat at constant pressure

Specific heat at constant volume Diameter

Temperature, degrees Fahrenheit

Gravitational constant

Gas horsepower

Enthalpy

Head

Inlet cubic feet per minute

Ratio of specific heats

Molal specific heat

Million standard cubic feet per day

Molecular weight

Polytropic exponent

Speed, revolutions per minute

Pressure

Pounds per square inch absolute Pounds per square inch gauge

Flow

Gas constant Ratio o f comprussiori Absolute temperature, degrees Rank~ne

Stroke of reciprocating compressor

Standard cubic feet per minute Temperature, degrees F

Temperature, degrees R Impeller peripheral velocity

Specific volume

Acoustic velocity of gas

Volume; volume flow rate

Volumetric efficiency

Weight f low

Compressibility factor

Impeller blade angles

Differential

Head coefficient

Efficiency

Subscr ip ts

1 . 2 s, d

C

ad, adia

P O ~ Y

P m

U

Dens i t y

F l o w coefficient F l u i d v r l o c ~ t y re la t i ve to r o t o r

I n l e t & Discharge

D i t t o

Reduced

C r i t i ca l

A d i a b a t i c

P o l y t r o p i c

D i t t o

M e r i d i o n a l

Per iphera l

R E F E R E N C E S

1. NGSMA Data Book .9 th Ed.. 1972. 2. Coinp,erred Air :and Gar Dam. C. W. Gibbr. Editor, 2nd Ed.. 1971 publirhed b y lngenoll-

H;,Wi C". 3. " C ~ ~ ~ ~ ~ ~ e r s ~ l ~ l ~ ~ y Chmir tind Tlieir Applicarion to Problems Involving Prerrure-Volume-

Energy Helal ion for Real Garer" Research Bullel in P7637. publirhed b y Worthingron Corp.. 1949.

4. API 617 Cen t r i f ug~ l Comprerrorr for General Refinery Services 4 th Edition. 1979. 5. AP1 614 Lubrication. Shaft-Sealing. and Conrrul Oil Systems for Special Purpore App l l cb

t~ons. 1st Edition. 1973. 6. Magl iozr~. T. L. "Conrrul System Prevenlr Sulg$ng i n Cenrrifugal Comprerron. Chemical

Enq!nrcr#ng MII~JZ~OP. May 8, 1967. 7. St.,,oselsky and Liidln. " lmp,oved Surge Control for Cenrrilugal Comprerrorr. Chemical

E#~!linecrirlq Mq;~z#ne. May 21. 1979. 8. API RP 541 "Recommended Pr;aclrce for Form-Wound Squiuel-Cage Induction Mo ron 200

HP ilW Laryel, 1972. 9. ANSI!NEMA MGI-1978 Matorn and Genera~orr

10. API 613 Specla! Pulpore Gesr Un i l r fof Refinery Services. 2nd Edition. 1977. 11. AGMA 421 ' ' P i i ~ r ~ c e fur High Speed Helical and Herrin(lbanc Gear Units" 12. API 616 Combust~on Gar Turbines for General Refinery Servlcer 2nd Edition. 1980. 13. Sawyer's Gar Turbtnc Engineering Handbook. 3 volr.. 1972 publirhed by Gar Turbine

Pui,l,c.it,'>ns ir,c. 14. A P G I 2 Socci;il Pu,p,,rc Slram Twl3iner 1"s R r l r n e ~ y St!rv#cer 2nd Edltiun. 1979. I!,. Ssllrel, L.. F . "GJS .md Al l Curnpters#o,~ Machi~wry," McGrow-Hill. 1961. 16. S ~ ~ h , ~ ~ s l . I.. F. "GJS M;ich#nery" Gulf Pul,llrhirq Co.. 1977. 17. Siep.!nufl, A. J. "Tu8boblowers" John Wlley & Co.. 1955. 18. Advmced Cenl#i fugal Co~nprersurr. 4 papers b y Turbomachinery Commirree o f ASME Gar

Turbine Diwrion, published b y ASME. 1971.

19. Edm~r le r . W. C. Applied Hydrocarbon Thermodynamics. 2 volr.. 1961 and 1974. publirhed by Gulf Publishing Co.

NGSMA - Natural Gar Procesrorr Suppliers Arroc8arion. Tulsa. OK 74103 API - American Petroleum Institute. 2101 L Stieer N.W.. Washington, D.C. 20037 AGMA - American Gear Manufacturers Arraciation. 1330 Marrachurettr Ave.. Warhingron.

D.C. 20005 ASME - American Society of Mechanical Engineen. 345 East 47th Street. New York. N Y 10017 NEMA - National Electrical Manufacturers Association, 2101 L Street. N.W.. Warh8ngron. D.C. 20037.

CHAPTER 3 API Standard 671 - Couplings

CENTRIFUGAL COMPRESSORS

M. P. BOYCE Bovce Engineering International lnc

Houston. Texas

INTRODUCTION

Cciltr~l i iq,~l comwessots are an integral part of the oi l and chemical industry.

They a c used extensively because o f their smooth operation, large tolerance to

urocrss fluctuations, and their higher reliability as compared to other types of

com,~r~~ssois. Centrifugal compressors range in all sizes from pressure ratios of 1.3

kiei it:lyr to plussure ratlos as high as 1 2 : l on experimental models. We will l imit

owselves to disc~~ssion of pressure ratios below 3.5:l since these are ones used

cxtc~isively , t i the oil and chemical industry. The proper selection of these com-

ii~t!sso~s I S ;I complex and very important decision since the successful operation of

(nliu~y I~IJIIIS depmds up011 the smooth and eificient functioning o f these units. To

u ~ l s u ~ u the bust selrctio~, and consequetitly the proper maintenance of centrifugal

; r s s s , !he riic]inr:er must lhave a w ~ d e k~lowledge of many engineering

ilsc11~111,~'s.

SELECTION OF A COMPRESSOR

Dot.ril oolnpl rssor spec~f~cations c;in vary from customer to customer, some pro-

v t d ~ i i i only basic informat~on, sucii as pressure, flow rate, type o f gas, driver, and ~111, ~ : o ~ i d i t i ~ ~ i s , to ii lengthy document detailing types of bearings, rotor response,

Itrbrtcation system, acceptable tolerance on performance, etc. To do the latter, the englneer must be very conversant with comprssors and their total support systems.

A ~1.11 tlnq point for the specifications are the various American Petroleum Institute

slx!c~f~cat~ons for turbomachinery. The following are some of the applicable poblcatuns:

Al'l S l l ~ ~ ~ c l . ~ ~ ~ l G1 1 Gc~,ti~i)l PUI~USC! Slualn Turbi~x! 101 Rcfincry Services 4Pl S l ~ , d . i ~ d 612 - Specla1 Pilrpose Stealn T~lrbine for Refinery Services

API Stdtid;~id 613 - Hgh-speed Special Purpose Gear Units for Refinery Services API S t . i ~ ~ l , i ~ i l 614 .- Lube alld Seal Oil Services

API Sr;nrdard 616 - Combustion Gas Turbines for General Refinery Services API Stlimiiird 617 - Centr~fugal Compressors for General Refinery Services

API Stal ldad 670 - Non-contacting Vibration and Axial Position Monitoring

Svsrrms

These specifications are written by user engineers with the lnput of rnatiufac~

turers and engineering contractors; thus, these standards represent a weaith of eh

perience and are a very good base from which to start your turbomachinery speclf, . cations.

Many decisions, regardless of details contained in specifications, have to be mad?

by the engineers in advance. Some of these may be his company's philosophy on various units and others could strictly be job oriented.

Layout

The general topography of the plant must be known to the engineer so that tila

proper site selection can be determined. Determination of whether the urut w ~ o l u

be grade or mezzanine mounted i s very important in determining the four idat i r~~

characteristics. Enough space should exist for the ductiny so that the inlet ctirr::,

tions to each stage allows the f low to enter without large distortions of veloc8lv o!

pressure. Accessibility requirements should be kept in mind so that ~ i i p ~ ( , ~ . I V ~ :

maintenance work on the unit could be performed with relatve ease. L o c ~ ~ t u ~ , # , I

the oil system for the unit is a very important aspect of compressor installatrjm I

is advisable to locate the oil reservoir away from the base plate w ~ t h the bcirtcvr,

sloped toward the low drain point. and enough space should be provided so lhal llli

return oil lines can enter the resi!rvoir away from !he oil pump suction. Thri L V O I I ~ ~ I . greatly reduce disturbance of the pump suction and also help n kui!i;i~~<) 1 1

reservoir re ten ti or^ time to around ten minutes.

Environment

The environment in which a machine I S to operate is as imixxlant a fec tu~ 1; . 8 1 .

other. I n many cases, this factor is often uverlooked or described in a word l i h r . : , ~

such as "extreme cold climate." The vendor needs to know much more. Hi: mw,!

know what extreme cold means (usually below -25"Fl, what thri traoitll i>l,

weather is since, as a practical matter, this weather creates more problems bcxiicts~

more equipment is exposed to i t and fewer precautions are taken since opi:at~v<t

problems are either not recognized and are glossed over. Al l cold weather appli:,,

tions need to have some icing protection, especially in the air inlet suppl~es. F u i !

supply ventilation, pneumatic controls, actuators, all need to have some dry~i:r: 0 1

de-icing. Many types of de-icing systems exist. The two most commoo t v l m u . t .

exhaust gases or compressor bleed air. Tropical climate presents its own st:, i : s I ) :

problems, such as excessive corrosion, high moisture content, and h y h aml)i,r,'.

temperatures which increase the horsepower requ~red and the cooling capacity ,,I

the lube oil system. Desert locations require specjal f i l ter i~at ion systems so .I\ I S

prevent erosion of the blades and special sealing on joints to prevent ~ I I V r l l :h,

small micron sand particles from entering the lube system, etc.

Machine usage is a very important environment consideration since intermittent usage can be the most severe kind of service. The aggressiveness of the gas to be handled and its temperature are very important aspects on the choice of materials to be used.

Type of Compressor

In rnally cases, it is not obvious what type of compressor is needed for the applrcation.Therearcmany typesof compressors on the market as seen in Table 3.1. Some o f the more significant types are the centrifugal, axial, rotary and reciprocating. The positive-displacement type compressors are used for intermittent f low in which successive volumes of f lu id are confined in a closed space t o increase their pressures. The other broad class o f compressors are rotary type for continuous flow. I n these types o f compressors, rapidly rotating parts (impellers) accelerate fluid t o a Ihqh speed; this velocity is then converted into additional pressure by qrxlunl decelulation in the diffuser or volute which surrounds the impeller.

Table 3.1. Principle Types o f Compressors

,:,,,,l~l~Kss~~l:!~

I I L I Q U I D HI: I. I C A L A X I A L PISTON I . U B E FLOW

The positive-displacement type of compressors can be further classified as either reciprocating or rotary types as shown in Table 3.1. The reciprocating compressor has a piston having a reciprocating motion within a cylinder. The rotary positive. displacement compressors have rotating elements whose positive action results in

compression and displacement. The rotary positive-displacement can be further subdivided into sliding vane, liquid piston, straight-lobe and helical-lobe type compressors.

The continuous-flow type compressors, can be classified under dynamic or ejector type, entrain the inflowing fluid using a high velocity gas or steam jet and then convert the velocity of the mixture to pressure in a diffuser. The dynamlc compressors have rotating elements which accelerate the inflowing fluid, and convert the velocity head into pressure head. Partially in the rotating elements and partially in the stationary diffusers or blade. The dynamic type can be further subdivided into centrifugal, axial-flow and mixed-flow compressors. The main flow of gas in the centrifugal compressor is radial. The flow of gas in the axial compressor i s axial and the mixed-flow compressor combines some characteristics of centrifugal and axial compressors.

It i s not always obvious what type of compressor is needed for an application. Of the many kinds, some of the more significant are the centrifugal, axial, rotary and reciprocating. Figure 3.1 will aid in a selection of a compressor. For very high flows and low pressure ratios, art axial-flow compressor would be best. Axial-flow compressors usually have a higher efficiency but a smaller operating region thar a centrifugal machine. Centrifugal compressors operate most efficiently at medium flow rates and high pressure ratios. Rotary and reciprocating compressors (posit~vr?~ displacement machines) are best used for low flow rates and high pressure ratios.

This chapter deals with the centrifugal or mixed-flow type of compressors so the following discussion i s concentrated on these compressors. There are many applications in a centrifugal compressor as seen in Table 3.2.

The centrifugal compressors range in size from pressure ratios of 1.3:l per stage to as high as 12:2 on experimental models. In a typical centrifugal compressor the fluid i s forced through the impeller by rapidly rotating impeller blades. The velocity of the f lu id is converted into pressure, partially in the impeller and partially in the stationary diffusers. Figure 3.2 shows a section of a typical multistage centrifugal compressor used in the process industry.

A common method o f classifying centrifugal compressors is based on the number of impellers and casing design. Table 3.3 shows three general types o f centrifugal compressors. For each type of the compressor, approximate maximum ratings of pressure, capacity and brake horse power are also shown. Sectionalized casing types have impellers which are usually mounted on the extended motor shaft and similar sections are bolted together to obtain the desired number of stages. Casing material is either steel or cast iron. These machines require minimum supervision and maintenance and are quite economic in their operating range. The sectionalized

JUO!lellUa3UOJ a10 U O J I ) f iu !z !~a l~ad

u o ! l e x ) ! ~ n d lay3!u pue laddo3 AJDU!L(JBUI pue slool l o j

Table 3.3. lndurtrial Centrifugal Compressor Classification Bared on Caring Derign

Approximate Maximum Ratings Approximate Approximate Approximate

Pressure Capacity Horsepower Caring Type PSlG Inlet C F M Requirement

A. Sect~onalired Uruallv Multistage 10' 20.0001 600'

8. Horironrally split Slrlgle stage

idouble rucrionl 15" 650,000' 10.000' Mullisrage 1,000 200,000' 35,000

C. Verrically split Single stage

iringle sucrionl Overhung 30' 250,000' 10,000" Pipeline 1.200 25,000 20.000

Multirrage over 5,500 20.000 15,000

. ~ a ~ e d on nir or airnoruhsric inroke condirionr.

Figure 3.3. Multistage centrifugal compressor (Courtesy of Elliott/

Process E y u i p m e n f Series Volume 3

pressure types with overhung impellers are used for combustion processes, ventilation and conveying applications. Multistage barrel casings are used for high pressures for which the horizontally split joint is inadequate. Figure 3.4 shows the barrel compressor in the background and the inner bundle from the compressor in front. Note that in most cases once the casing is removed from the barrel i t is horizontally split.

Figure 3.4. Barrel type compressor showing inner casing in foreground and barrel io

background (Courtesy of Nliotr/.

Compressor Configuration

To properly design a centrifugal compressor, the operating conditions, the type of gas, and its pressure, temperature and molecular weight must be known. The corrosive properties of the gas must be properly identified so that proper met& lurgical selection can be made. Gas fluctuations due to process instabilities must bu

pinpointed so that the compressor can operate without surging. Centrifugal compressors for industrial applications have relatively low presuri!

ratios per stage. This is necessary so that the compressors can have a wide opcrattrq range, and stress levels can be kept at a minimum. Due to the low pressure ratios lo, each stage, a single machine may have a number of stages in one "barrel" to achevr the desired overall pressure ratio. Figure 3.5 shows some of the many config-

Parallel flow, suction in end*

Parallel flow. rucfion in center

Serial f low, two cooling poinrr

t

Series tlow, fwo cooling poinrr

I l ILL 11

series flow, one cooling Point

Series flow, one cooling point. suction on ends, cool ends

fl!jiirr 3.5. V ~ i i o i i r cunl ryur~ i r io i i r oicr?i i l i r tugal compicsrorr ro meat nceds

01 ,I,,! ,,r"c,~.s$i>l:,"~s.

urations. Some of the factors to be considered when srlect~rng a configu~a;:w ' .

meet plant needs are: 1. Intercooling between stages can considerably reuuct tne power LO' I~L~, . ,~?L

2. Back-to-back impellers allow for a balanced rotor thrust, and minimi&

overloading the thrust bearings. 3. Cold inlet or hot discha~gr at the middle ot the case roduci!s 01-sl.a! ;ind

lubrication problems. 4. Single inlet or single discharge reduces external plping problems.

5. Balance planes that are easily accessible in the field can appreciably i rduc t .

field-balanciny time. 6. Balance piston with no external leakage will greatly reduce wi:ar WI 1 1 1 ~

thrust bearings.

7. Hot and cold iections o f the case that are adjacent to each other will reduu

thermal gradients, and thus reduce case distortion.

8. Horizontally split casings are easier to open for inspection thar, vl.!rt~cal;

split ones, reducing maintenance time. 9. Overhung rotors present an easier alignment problem because shai!u;,ri

alignment i s necessary only at the coupling between the compresm a l w

driver.

10. Smaller, high.pressure compressors that do the same job will reduce found2

tion problems but wi l l have greatly reduced operational range.

Arrangements

The general configuration o f the compressors and their drive trains must ht.

evaluated to f i t the location, environment, and type of compressor selectc!d. A decision must be made as t o whether the units are to be in series or paralli:l Thj;

decision requires a knowledge o f the flow and discharge pressure required. In n l r h

cases, a number of casings are connected together to form a "compressor t r a ~ r l . "

This is nothing else but the connection of various compressors in series. The l m i ;u the number of connected casings is due to the rotor dynamics of the couplfd r w : .

In the arrangement, one must also decide what type of mounting is desirable M L ~ ! turbomachinery is mounted on structural steel platforms which are usually i r f r ~ r m u

to as base plates or skids. These platforms are then mounted on a mass of coric~,,:.,

at the job site by mounting them on sole plates or through direct grouti~l!, % I

forms should be considered part of the foundations and great care should tic: ~,I.I.G.I!

in their design. Insufficient rigidity of these platforms can allow the ro!.ilIi,;,

machinery to excite them.

Driver Selection

The three main typcs of dr~ves on compressors arc! 11 rli:om turb~r \ i : i , Zr I;.!,

turbines, and 31 electric motors. The decision o f which drvc 15 the bast i u r i : i l :i t w ~ 1

always an easy decision. The selection o f the drive depends on many factors such as location, process. and uni t size. I n remote locations, gas turbines are mostly used due t o their low maintenance and the abi l i ty t o prepackage the units. Their l ight w r g h r makos them also a must for offshore platforms. I n petrochemical plants, due to the ~ ~ o c e s s steam created, steam turbines are widely used. I n this manner, plants car> i l t ~ l l ~ ! lhuir t :r~tx(~y - t h u s operating at an overall high plant efficiency. Electric motors d ~ i v c [nost o f the smaller f low compressors, usually through a speed increasing qeat system. Figure 3.6 shows the typical ranges for the various drives. Also f rom this tigute, one can note that the higher the f low, the lower the speed. This is due t o 1h1: fact that at high flows the compressor diameter must be large; therefore, the speed must be reduced t o maintain the same stress levels.

Figure 3.6. Application chart for cerlrnfugal compressors.

AERO-THERMODYNAMICS O F CENTRIFUGAL COMPRESSORS

Thermodynamics

TI,? r :~xrqy trailsfor t o the rotor can bc obtained f rom the first law o f thermo- 11vn;lmcs. The f irst law applied betwee!) two stations can be wr i t ten as follows:

where:

is the rate o f heat transfer per uni t mass f low w i s the rate o f shaft work per uni t mass f low h is the enthalpy o f the fluid V is the velocity of the f luid

I f the energy transfer t o the f luid is considered as being reversable and adiadaba~ t ic (i.e. losses such as due t o fr ict ion in the f low passage being absent, along with no heat transfer) the work energy u t i l i l ed t o raise the enthalpy of the f luid from its initial t o final state can be wr i t ten as:

The above relationships wil l give a negative value i n the case of a compressor and a positive value in the case of a turbine thus indicating that input or output shaft work respectively is equal t o the increase in the total enthalpy rise across the turbomachine.

The mechanical energy available t o the rotor o f the centrifugal compressor from its prime mover i s given by the Euler Turbine equation and can be wr i t ten as:

w = u , v,, -U2 vo2

where UI and U2 are the peripheral velocities at the inlet (mean) and exit diameters of the rotor, and V,, and V,, are the tangential components of the absolute velocities at the inlet and exit respectively.

Efficiencies

1. Adiabatic Efficiency - The work in a centrifugal compressor under ideal

conditions occurs at constant entropy as shown in Figure 3.7. The actual work

done is indicated by the dot ted line. Therefore the isentropic efficiency o f the compressor can be writ ten i n terms o f the total changes in enthalpy:

This equation can be rewritten for a thermally and calorically perfect gas in terms of total pressure and temperature as follows:

v ; ~ ~ i i x ~ s dim~!ns~onluss paramt!trrs which art! based on the dimensions, such as Mass (M!, Lwqrh ( L ) arid T i ~ n e (T) . Based on these we cat1 obtain various independent parameters such as density (p), viscosity 1/11, speed (N), diameter (D) and velocity (V!. This leads to forming various dimensionless groups which are used in fluid tnrchanics of turbomachines.

PVD Re = - (Reynolds Number) ii

(8)

~ f i N3 = (Specific Speed) ~ 3 1 4

(9)

W11sla H is thc ; ld i ;h t ic head. Q is the volume rate, and N the speed

Whur D i s the d~amete! of the impeller.

Q @ = -7 (Flow Coefficient)

N D (111

H $ = -- (Head Coefficient)

NZ (121

The above are some of the major dimensionless parameters. In many cases for the flow to remain dynamically similar all the above parameters must remain constant, however, this i s not possible in a practical sense so one must make choices.

In seclecting turbomachines the choice of specific speed and specific diameter determines the type o f compressor which is most suitable as seen in Figure 3.9. I t is obvious from this figure that high head and low flow require a positive displacement type unit, while a medium head and medium flow require a centrifugal type unit, and high f low and low head an axial f low type unit. Figure 3.10 shows a blow-up o f the centrifugal compressor section and can be used as a reference for selection o f centrifugal compressor units.

Figure 3.9. Compressor map.

Flow coefficients, pressure coefficients can be used to determine varlous o f f design characteristics. Reynolds number effects the flow calculations as far as skin friction and velocity distribution are concerned.

It must be remembered when using dimensional analysis in computing performanoe

l i m l o l :I(/ (ii p es a r e sawn a41 '06 < 2 0 JOJ pue saum "a~s-p~eMy3eq l o panJn3

~w-,ny:,c(~;jneq 01 ples a l e 06 > zg q 1 ! ~ sJalladwl muen wpeJ aneq 01 p!es ale *06 z:! ; , l t i u ~ apelq i!xa q l ! ~ sialladw( .SalSue apelq l!xa aql 01 13adsaJ q l ! ~ pau!pp

. , o a r : q l . P L . E a~n6 ! j u! u ~ o q s se sauen Jal(adiu! $0 sadhl aaq l a x a laq l .suo!ie~n6!~uoa huew ow! ale~fialu! 0% iln'!#!p s! uo!le~n6!)uo3 aql

I.iAaMOq :~aqwnu q 3 e ~ dl1 ~aanpu! aql 6u!anpal pasn aq ues lalawe!p d! l JaJnpu!

; i leus e leql 0s MOIJ lalu! aql sanleq Alle!luassa walshs 1a3npu! A~lua-alqnop k! '&L'E a~nfi! j u! umoqs se ~a3npu! A~lua-alqnope

, t!r imnpu! Allua.al6u!s e 'Ajaweu 'swalshs Jaanpu! AJlua 40 spu!y O M 1 ale a J a U

H E A D

cot~s~sts of two basic sol~ft ions 1) the inducer which is very much like an axial flow loto,. ; i d 21 the radial blades where energy is imparted by centrifugal force. The flow enters the impeller usually in the axial direction and leaves in the radial i i ~~cc t fon . Ths velocity variations from hub to shroud resulting from these changes H I l lcm < ~ I I < Y : ~ I O ~ K complicate the desig~i procc:dure fot centrifugal compressors. C. li. W u 11.1s p~ssuntcd thu tIil.ee dimensiorlal theory in an impeller, but it is difficult i i , x,lv,, I<,( ihc! fluw III ;In impt?llr!r uslrlq the above theory without certain simpli- l i ed cu i~ i t t u i ) s . J. D. Srao~tr, T. Katsanis. F. Dallenback. M. P. Boyce. Y Senoo .ind SO on have dealt with it as a quasi-three dimensional solution. It is composed of two solutions, one in the meridional surface (hub-to-shroud) and the other in the stream surface of revolution (blade-to-blade). These surfaces are illustrated in Flgu,c 3.21

By the application of the previous mentioned method using a computer proyram. i t is possible to achieve impeller efficiencies of more than 90%. The actual flow phenomena in an impeller is much more complicated than the one calculated. One example o f this complicated flow i s shown in Figure 3.22. The stream lines observed in Figure 3.22 do not cross, but are actually in different planes. Figure 3.23 shows the flow in the meridional plane, note separation regions at the i lduci :~

I section and at the exit.

Figure 3.22. Flow map of ,mpe//er plane I

As mentioned previously, experimental studies of the flow within impellci~ passages have shown that the distribution of velocities on the blade surfaces I S

rather different from the distributions predicted theoretically. I t is likely that the discrepancies between theoretical and experimental results are due to secondary flows from the pressure losses and boundary layer separation in the blade passages. High performance impellers should be designed, when possible, with the a ~ d o i

theoretical methods for determining the velocity distributions on the blarlr S O I ~

faces. Examples of the theoretical velocity distributions in the impeller blades ol a

centrifugal compressor are shown in Figure 3.24. The blades should b,! so di!sig~,i:ci that there are no large decelerations or accelerations of flow in the impeller s111ce this would lead to high losses and separation of the flow. Potential flow solutor,~ predict the f low well in regions away from the blades where bouniary layer eff i : i . t i

are negligible. In a centrifugal impeller, the viscous shearlng forces create: a bou~)(i

ary layer with reducr?d kinetic i:ni:rgy. If the kinetic cnvrgy IS riiducnd hi:lc,vi ,, certain limlt, the flow in this layer becomes stagnant t h w icivarrcs.

Figure 3.23. Flow map as reen in meridionel plans.

Inducer

The function o f the inducer is t o increase the fluid's angular momentum without increasing its radius o f rotation. I n the inducer section the blades bend toward the direction o f rotation as shown in Figure 3.25. The inducer is very much an axial rotor and changes the f low direction from the inlet f low angle to the axial direction. I t has the largest relative velocity in the impeller and i f not properly designed can lead to choking conditions at its throat as shown in Figure 3.25.

There are usually three forms of inducer camber lines in the axial direction. These are circular arc, parabolic and elliptical. Circular arc camber lines are used in compressors with low pressure ratios while the elliptical produces good performance at high pressure ratios where the f low has tronsonic mach numbers.

Due to choking conditions in the inducer many designs incorporate a splitter blade design. The f low pattern i n such an inducer section is shown in Figure 3.26a. T h s flow pattern indicates a separation on the suction side of the splitter blade. Other designs include tandem inducers. This consists of the inducer section being slightly rotated as shown in Figure 3.26b. This modification by Boyce and Nishida gives additional kinetic energy to the boundary which is otherwise likely to

separate.

z/b = relative rneridional channel widrh

v ' / i = relative blade spacing

B L A D E - - - ?/ CAMBER L I N E I N

C Y L I N D R I C S E C T ~ O N A A

07

; t- 0 y-J + 3 0, P u ?

iz2 ZK'= w o w 3) ,i- 2 0lLz 2

11 e D

L" LO- ii

In a s~;itioriary impeller the f low could well be expected t o fo l low the blade shape and exit tangentially t o it. Here, of course, a very high adversepressure gradent dong the blade passage and subsequent f low separation are n o t considered to be genual possibilities.

Inertla ;,,xi centrifugal focces would cause the f lu id elements t o move closer t o imd ;ilong the lead~ng surface of the blade, towards the exit, and once ou t o f the bliidu passage where there is no positive impelling action present, these f luid elements tend t o slow down.

Causes o f Slip

The cause o f the slip factor phenomenon that occurs w i th in an impeller is no t known exsctly. However, some general reasons can be used t o explain why the f low is changed. These are:

1. Coriolis Circulat ion - due t o the pressure gradient between the walls of t w o adjacent blades. Coriolis forces, centrifugal forces, and follows the Helmholtz vor- ticiry law. The combined gradient that results causes a f lu id movement f r o m one wcill t o tl ie o t h e ~ and vice versa. This sets up a circulation wi th in the passage as seen il l F ~ g w c i 3.28. Because of this circulation, a velocity gradient results at the impeller exit wi th a !net change in the exit angle.

2. Boundary Layer Development - The boundary layer that develops wi th in an i~npel ler passage causes the f lowing f lu id t o experience a smaller ex i t area as shown In Figure 3.29. This is due t o small, i f any, f low wi th in the boundary-layer. For the

Figure 3.29. Boundary-layer development.

f luid t o ex i t this smaller area, its velocity must increase. This gives a higher relative exit velocity. Since the meridional velocity remains constant, the increase in relative velocity must be accompanied wi th a decrease i n absolute velocity.

3. Leakage - Flu id f low f rom one side o f a blade t o the other side is referred to as leakage. Leakage reduces the energy transfer f rom impeller t o f lu id and decreases the exi t velocity angle.

4. Number o f Vanes - The higher the number of vanes, the lowerthe vane loading and the closer the f lu id follows the vanes. With higher vane loadings, the f low tends t o group upon the pressure surfaces and introduces a velocity gradient at the exit.

5. Vane Thickness - Because o f manufacturing problems and physical necessity. impeller vanes have some thickness. When the f luid exits the impeller, the vanes no longer contain the f l ow and the velocity i s immediately slowed. Because i t is the meridional velocity that decreases, both the relative and absolute velocities dr.

crease, thus changing the exi t angle o f the fluid. T o combine all these effects, consider a backward curved blade impeller. The

exit velocity triangle for this impeller, wi th the different slip phenomenon changes is shown i n Figure 3.30. As this shows the actual conditions when the compressor i s

running may be far removed from the design condition. Several theoretical and emp i r i ~a l ' e~ua t i ons have been derived for the slip factor.

Some o f the widely used of these are the theoretically based equation derived by Stodola and Stanitz, all of which assume the inviscid f luid-f low through the impeller.

Flyurc 3.30. E f k i or8 exit velacify triar,gIes by various parameters.

6. Slip Factor Due to Stodola - The second Helmholtz law states that the

vo~t ic l ty o i a frictionless fluid does not change with time. Hence, if the f low at the

mlct to ;in irnpelle~ is irrotational, the absolute flow must remain irrotational t l~~oui r l i i x i t t he impcllrl-. As the impeller has ; I n angular velocity w , the fluid must

I I , I c y w li:lativ~: to t h ~ irnprllc,. This fluid motion is called the

I , I . Thus it thole wcta oo f low thruugh the impeller, the fluid in the

i m , x l l ~ : ~ climmels w w l d rotate wlth ;in ar~gular velocity equal and opposite to the

~mpeller's allgulai velocity. The apploxlmate the flow Stodola's theory assumes that the slip is due to the

irldtivtr eddy.

The ielatve eddy i s considered as a rotation of a cylinder of f luid at the end of

the blade passage (shown as a shaded circle) at an angular velocity of -w about its own axis. The Stodola slip factor is given by:

Whe~,? 0: is the blade angle, Z the number of blades. Wm2 is the meridional velocity

. i l ~ i U : the blade t ~ p speed. Calculations using this equation have been found to be

g c ~ l u ~ d ly lower than experimental values.

7. Stnntir Slip Factor - Stanit i calculated blade-to-blade solutions for eight

i n [ ~ c l l e ~ s and concluded that, for the range of conditions covered by the solutions,

U 1s a function of the number of blades ( 2 ) and the blade exit angle is

approxitnately the same whether the flow is compressible or incompressible.

Stanitz's solutions were for n14 < B2 < nl2. This equation compares well wlth

experimental results for radial or near radial blader.

Diffusers

Diffusing passages have always played a vital role in obtaning good perfotrnallci.

from turbomachines. Their role i s t o recover the maximum possible k i i iet~c ani!!riv

leaving the impeller at a minimum expense of loss in total pressure. The cfficii!rlw< of centrifugal compressor components has been steadily improved by advar~c~m

their performance. Significant further improvement in efficiency, however. w:!I

only be gained by improving the pressure recovery characteristics of the diffusinq

elements of these machines, since these elements have the lowest efficiency.

THROAT

(A1 S T R A I G H T - W A L L , RECTANGULAR DIFFUSER

D

THROAT

EXIT

(0) STRAIGHT - WALL. CONICAL DIFFUSER

Figure 3.37. GeornetNc classificvoon of dilfurerr

The performance characteristics of a diffuser are complicated functions o f the d~ffusel- goometry, inlet flow conditions and in some cases exit f low conditions. F w r e 3.31 shows typical diffusers classified based on their geometry. The selection of 311 olltlmuln channel diffuser for a particular task is diff icult since it must be cliosar~ f ~ n m a n almost ~nf in i te number of cross sectional shapes and wall conflguarions. In radial and mixed-flow compressors the requirement o f high performance and compactness leads to the use of vaned diffusers as shown in Figure 3.32. Figure 3.32 also shows the flow regime of a vane-island diffuser.

CHANNEL DIFFUSER

PRESSURE I SURFACE

SUCTION SURFACE /

DJUSTMENT

SEMI-VANELESS SPACE

VANE LEADING EDGE RADIUS

Figure 3.32. Flow regions of the vaned diffuser

Matching the f low between the impeller and the diffuser i s very complex because the f low path changes from a rotating system into a stationary system. This complex unsteady f low is strongly affected by the jet-wake of the f low leaving the

, impeller as seen in Figure 3.33. The three-dimensional boundary layers and srconda~y flows in the vaneless region, and f low separation at the blades also effect the overall flow in the diffuser.

The f low in the diffuser i s in many cases assumed to be o f a steady nature, in order to obtain the overall geometric configuration of the diffuser. I n a

Figure 3.33. J e t h a k e flow distribution from impeiler.

channel-type diffuser the viscous shearing forces create a boundary layer wlth reduced kinetic energy. I f the kinetic energy is reduced below a certain limit, the flow in this layer becomes stagnant and then reverses. This flow reversal cauri:r

separation in a diffuser passage, as mentioned above, which results in eddy losri!s, mixing losses and changed f low angles. The separation should be avoided or delayed to improve compressor performance.

The high-pressure ratio centrifugal compressor has a narrow stable operatfig range. This operating range is due to the close proximity of the surge and choke

flow limits. The word, surge, is widely used to express unstable operation of a compressor ingeneral. "Surge" wil l be defined as the flow breakdown period dur~ng unstable operation. The unsteady flow phenomena during the onset of surge in a high-pressure ratio centrifugal compressor caused the mass f low throughout the compressor to oscillate during supposedly "stable" operation. The throat pressure, in the diffuser increases during the precursor period almost up to collector pressure. P,,,, at the beginning of surge. Al l pressure traces, except plenum pressure. suddenly drop at the surge point. The sudden change of pressure can be explained by the measured occurrence of back f low from the collector through the impeller during the period between the two sudden changes.

Scroll or Volute

The purpose of the volute is to collect the fluid leaving the impeller or diffuser and deliver it t o the compressor outlet pipe. The volute has an important effect on

Proccss Eyuipmenr Series Volume 3

PERFORMANCE CHARACTERISTICS O F CENTRIFUGAL COMPRESSORS

The calculation o f the performance o f a centrifugal compressorat bo th design :~nd o f f design conditions require the knowledge o f the various types o f losses t!ncountrred in a centrifugal compressor.

The accurate calculation and proper evaluation o f the losses w i th in the ce~i f r i fugal compressor is as important as the calculation of the blade loading lpaliilneter, since unless the proper parameters are controlled, the efficiency drops, tht! evaluation o f the various losses is a combination o f experimental results and theory. The losses are divided in to two groups: 11 losses encountered i n the rotor .id. 2 ) losses encountered in the stator. The losses are usually expressed as a loss o f I l w t or enthalpy.

The losses are usually expressed as a loss o f heat or enthalpy. A convenient way to express them is in a non-dimensional manner w i t h reference t o the exi t blade . The theoirt ical total head ava~lable iq,o,l is equal t o the head available f rom t l v c~>crgy equ:ltion:

~ ~ l u . : the h ~ a d w h ~ c h is lost due to disk f n c t i o ~ i ( A q D F ) and due t o any recirculation ( . \ i ~ , ~ l of the tlir h c k in to the rotor f r o ~ n the diffuser..

Thai x ~ a b a t i c head that i s actually avallable at the rotor discharge is equal t o the tl i<wrctical head minus the heat due t o the shock in the rotor @qSh), the inducer loss (Aq,,>I, the blade loadings. (Aq,,, I the clearance (Aq,) between the ro to r and the shroud. and the viscous losses @qSf ) encountered in the f low passage.

Therefore, the adiabatic efficiency i n the impeller is:

I ,lit ot the over-all stage efficiency must also include the losses r~ , c i , un l r~ed I,, the dlffuset. Thus, the overall actual adiabatic head attained would be l hc actual jd~ab i i t i c head o i the impeller minus the head losses encountered i n tllc df lusur duc t o wake caused by the impeller blade (Aq,), the loss of part o f the kini!t!c head at the exlt o f the dffuser lAqe,,l, and the loss of head due t o the

frictional forces (Aqosf1 encountered in the vaned or vaneless diffuser space

Thus the overall adiabatic efficiency in impeller is given by the following relations hi^: '

The individual losses can now be computed. These losses are broken up intu two major categories: 1 I losses in the rotor, and 21 losses in the diffuser.

Rotor Losses

The rotor losses as mentioned previously are divided further into varlous cati! gories. The fol lowing i s the analysis o f each of these losses.

1. Shock in Rotor Losses - This loss i s due t o the shock occurring at the r o w inlet. The inlet o f the rotor blades should be wedge-like so as t o obtajn a weak

oblique shock and then should gradually be expanded t o the blade thickness so .IS

t o avoid another shock. I f the blades were blunt, this would lead t o a b low shock which would cause the f low t o detach from the blade wall and the loss to bc much higher.

2. Incidence Loss - A t o f f design conditions, f low enters the inducer 41 ;ill

incidence angle that i s either positive or negative, as shown in Figure 3.35. A positive incidence is that which causes a reduction in flow. Fluid approachf~ig ;)

blade wi th incidence suffers an instantaneous change of velocity at the blade irilrit t o comply w i th the blade inlet angle. Separation o f the blade also creates a loss associated wi th this phenomenon.

3. Disk Fr ic t ion Loss - This is the loss due t o the frictional torque on the back surface o f the rotor as seen in Figure 3.36. This loss is the same for a given size disk whether it is used f o r a radial in f low compressor, or a radial inf low turbine. In many cases, the losses i n the seals, bearings, and gear box are also lumped in w ~ t h this loss, and the entire loss can be called an external loss. I n this loss unless rhr! yii l ,

is o f the order o f magnitude o f the boundary layer, the effect of the gap p ? t ! 1,.

negligible. A po in t o f interest that should be indicated here i s that the disk f r c t i i ~ ! , in a housing is less than that on a free disk. This is due t o the existence o f a "Cow" which rotates at half the angular velocity.

4. Diffusion Blading Loss -This loss arises because of negative velocity gradvr:t\ i n the boundary layer. This deceleration of the f low increases the boundary l : v+

and gives rise t o separation of the flow. The adverse pressure gradient, w l i c h ;r

compressor normally works against increases the chances of separation and qvi:;

rise t o a rather significant loss.

rulccs 01, the imilcller wall which are mostly due to turbulent friction. This type of

loss i s i rsu~l ly demmined by considering the f low as an equivalent circular cross

sectioti wiih a hydraulic diameter. The loss is then computed based on the well

known 1)ipr f low pressure loss equations.

Stator Losses

1. Rccirc~datiog Loss - This loss occurs dui? to the back flow into the impeller cxlt o l ;, compfessor and is a direct function of the air exit angle. As the flow

Through the compressor reduces, there i s an increase in the absolute f low angle at the exit of the impeller as seen in Figure 3.38. Part of the fluid is recirculated from

11111 illu use^ to the tmpeller and its energy is returned to the impeller.

Figurc 3.38. Recirci,/ating loss.

2. Wake Mixing Loss - Thls loss i s duc to the impeller blades, causing a wake in 1I i~. v;lnelrsssi);ice Lxhuid the lotor. This loss is m~nimi red in a diffuser which i s s y n i ~ i w t ~ i~ J l O u l l d the axis of rotation.

3. Vaneless Diffuser Loss -This loss i s experienced in the vaneless diffuser due to

r l l i f r i c I ~ w aid the absolute f low angle.

4. Vaned Diffuser Loss - Vaned diffuser lossas x s based on the conical diffuser

r c s t 1t~su11.;. They jre 3 fw,ctiot\ of the impellet bladi! loadlng and the vanlesssspace

~ . i i i us 1.11o Thcy dso take into xcoun t the blade ~ncidence angle and the skin

11 c t ~ o i l duc I n the vanes.

5. Exit Loss - The exit loss assumed that one half of the kinetic energy leaving

the vaned diffuser i s lost.

Losses are a complex phenomena and as discussed are a function of many

parameters such as inlet conditions, pressure ratios, blade angles, flow etc. Figure

3.39 shows the loss distributed in a typical and centrifugal stage of pressure ratio below 2:l with backward curved hlades. This figure is just d guide lint! a t~d should be used as such

I,>,,

'8,)

Flgure 3.39. Losses in a centrifugal compressor.

Performance Characteristics

A Plot showing the variation of total pressure ratio across acompressoriis a function of the mass f low rate through i t at various speeds is known a s the per

formance characteristics of that compressor. Figure 3.40 shows such a plot.

The actual mass f low rates and speeds are corrected by factor (mhl and i l l -& respectively, i n order t o take into account the variation in the inlet conr j~tort i of temperature and pressure. The surye line is the line which joins the po81~ti i r r ~

different speed lines where the compressor's operation begins to be unstabli!. A compressor i s said to be in surge when the main f low throuqh the cumiirr:rsu!

Process /iqiiip,,ierii Series Volu,ne .i

reverses its directionand flows from the exit t o the inlet, for a short time interval during which the back (exit) pressure drops and then the main flow assumes its proper direction. This is followed by the rise in back pressure causing the main f low to reverse again. This unsteady process, i f allowed to persist, may result in irrepara- bit? dxn.qc to the machine. Lines of constant adiabatic efficiency (sometimes called the efflctency islands) are also plotted on the compressor map. A condition known as choke o i commonly known as "Stonewall'' is indicated on the map which shows

the maximutn mass f low rate possible through the compresor at that operating

speed.

Surge

C<~~ni i~t !ssor SLlrlJC IS a phenomeno!i o f considerable interest and is not yet fully ondcrs tw~i . tss~nt ia l ly , ~t is ,I s1tila11011 of ~ ~ ~ m a b l t : operation arid should therefore i*. avo ldcd H I both drstgn a r w ~(~erat ior r . Sutge has bee,i traditionally defined as the lowel litnlt of stable operattoll o f a compressor and involves the reversal o f flow. T h s rvcts,il of flow occurs because o f some k l l ~ d of aerodynamic instability w ~ t h i n

I \?? systum. Usu8lly i t 18 part o f the colnprcssor that i s the cause of the aerodyriamic

instability though i t is possible that the system arrangement could be capable of magnifying this instability. Figure 3.40 shows a typical performance map fol C<

centrifugal compressor showing efficiency islands and constant aerodynam~c speeo lines. The total pressure ratio can be seen to change with f low and speed. Usuallv compressorrare operated at a working line separated by some safety margln from .

the surge line. Usually, surge is linked with excessive vibration and an audible sound; vet , therr

a

have been cases in which surgu problems which are not audible have caused lailurcs. Extensive investigations have been conducted on surge. Poor quantitative umvt:~ sality of aerodynamic loading capacities of different diffusers and impi<lli!rs 2nd . I ( ;

inexact knowledge of boundary layer behav~or makes the cxact pred~ct io~ i o l I lan in turbomachines at the design stage difficult. It IS, however, quite evident that lhi, underlying cause of surge is aerodynamic stall. The srall may occur n i:tthr!l t t h

impeller or the diffuser. When the impeller is the cause of surge, the inducer is what actually causes lh i ,

stall. Either a decrease in the mass flow rate or an increase In the rotational speed 01

the impeller or both can cause stall i f the compressor is operating at the surge h i t ! .

Stalling the diffuser occurs in basically the same way as in the inducer. A diffuser usually consists of avaneless diffuser, a pre.diffuser section before thc throat containing the initial portion of the vanes, the throat and a diffusion passage. .

The pre-diffuser accepts the velocity generated by the centrifugal impeller and

turns the flow from thevaneless space to the restriction of the diffuser passage at the . throat and beyond. When the pre-diffuser stalls, the flow will not enter the thlo:it The discharge volume senses this drop in flow pressure and discharges through the diffuser. thus causing f low reversal and surge. As before, stalling of the pre-d~ffuscr

can be accomplished in two ways - by increasing impeller speed or decreas~ng the

f low rate. Whether surge is caused by a decrease in f low velocity or an increase in rota-

tional speeds, either the inducer or pre-diffuser stalls. Which stalls first is difficult to determine, but findings have shown that for low pressure ratio compressor the suryr: initiates in the diffuser section while for units with single stage pressure ratios abovi 3:l the indication is that surge is initiated in the inducer.

Surge Detection and Control

Surge detection devices may be broken into two groups: l l static surge detectlor! devices, and 2) dynamic surge detection devices. To this date, statlc sulge detecticj~, devices have been widely used and more research work i s still to be done befurti 11

dynamic detection device can be used. It i s probably the dynamic surge d e t ~ c t o i l device that will meet the requirements and hopes of many engrleers lor a cmrrr<>l device that could anticipate stall and surge and hence prevent its 0ccvrri:rtu Obviously, detect~on devices must be linked with a control dewcr which wot~l i l prevent an unstable operation of a compressor.

Static i u iq t ' i le t rc t~on devct:s are those whlch attempt to avoid stall and surge by

thc rnedsiilement o f some compressor parameters and ensure that a pre-decided

I 5 1 rxcetided. When the parameters meets or exceeds the limit, some

r o i l l r o ; i r t t i~ i , 1 li,ken A typical pressure ortented anti-surge control system i s s l i o w ~ ~ H I F$qu!t! 3.41. The pressure transmitter monitors the pressure and controls a . device wlijch inlight open a blow-off valve. A temperature sensing device corrects

lhc w x i ~ x l s f o ~ the effect of f low and speed for the effect of temperature. A typical

flow otientiid device is also shown in Figure 3.41.

F,!ww .?..<I. PW.S.W~V ;,,xi f low - o,,eort!d m # , - s w g e courro/ svsreo!. . TI,,, I!,IOL>II,III: L IU~I ,~ 101 ,ill stailc S L I I ~ ~ deteclion dwces IS that the actual

,>11~~1<)111~,1,4 01 l luw II,VIIISIII ( s c ~ r q c ) is #not b<:iiig d~recl ly nlonitored. What is being

I I ~ W ~ I ~ O I C V I ; i l l . OIIICI ,I.II;IITI~IUI~ that :11(! reI8tcd to sulge and the control limits are

set i t om p;ist cxpei ie~rce and a study of the compressor characteristics.

Dynamc swge detection and control methods are currently being researched

tuil.~y. Hue, ;in attempt is made to detect the srarf of a reversal of f low before i t

reaches the critical situation of surge. This is done using a boundary layer probe.

Boyce has obtained a patent for a dynamic surge detection using a boundary

layer probe which is presently undergoing some actual field tests. This system consists of specially mounted probes in the compressor to detect boundary layer flow reversal, as shown in Figure 3.42. The concept being that the boundary layer

would reverse initially before the entire unit would be in surge, and since it is

measuring an actual onset by monitoring the flow reversal i t is not dependent or1

the molecular weight of the gas and i s not effected by the movement of the surge

line

Figure 3.42. Boundary layer surge predfcrion technique

I l l , : i ~v~.l<lr1g. 111 F~gure 3.45e. an electron-beam techntque is used to weld on the sh~ouri or the hub. This technique i s still in its infancy and work needs to be done r i i p u ~ f ~ c t it. Its major disadvantage is that electron-beam welds should preferably bc, st!assud in tension, but for the configuration of Figure 3.45e, they are in shear. The coi~f~gur;ltions of Figures 3.459 through 3.45j use rivets. Where the rivet heads

plottude into passage, aerodynamic performance is reduced. Materials for fabricating these impellers are usually low-alloy steels, such as AlSl

4140 or AlSl 4340. AlSl 4140 is satisfactory for most applications; A lS l 4340 is used for large impellers requiring higher strengths. For corrosive gases, A lS l 410 st;iiiiless steel (about 12% chromium) is used. Monel K-500 is employed in halogen

atmospheres and in oxygen compressors because of its resistance to sparking. Tit;,niilm impellers have been applied to chlorine service. Aluminum-alloy impellers 1i;tvti bean used in great numbers especially at lower temperatures (below 300°F). W r h f ~ c w developments in aluminum alloys, this rage is increasing. Aluminum and utmriurn . I I ~ sometimes selected because of their low density. This can cause a shift H I rht, cuticlil speed of the rotor, which may be advantageous.

ROTOR DYNAMICS

TII~, ~mi~v rmrn t of the io to l and its effect on the entlre performance o f the unit i s t l i l most important aspect o f centrifugal compressordesign. Most compressors twl.ly, wliicli ale used in the petrochemical industry, are built in accordance with API 617 spt~cificarions. The natural frequency o f the rotor cannot occur in the v.11 i;lI,l<! si~m!d ra11gc 01 the comp!cssor. Maliy of thl: newer high speed compressors. O / I L ? I : i t t i ;$bow? I l i e i ~ first critical. Shafts which operate above their critical are said to la* " f luxi~lc! shafts." API specifications call for the first critical to be at least 15 IIL,II:~IOI L I I ~ O W JIIY operating r p w d and the second critical t o be at least 20 percent O V ~ I the ~n;iximutn continuous speed. I t is desirable that the first critical not be

a r o u ~ ~ d half the design speed; otherwise, a problem known as "oil whirl" may be induccri. 011 whirl is a major cause of instability in turbomachines. It may occur in 111~: jowndl beallngs ol 111 the seals in which the shaft and the stationary seal are

su~)a i~ l t?d by a f i lm of fluld. In rho case of the newer flexible rotors which operate in many cases above the

f i r s t critical and in many cases above the second or third criticals, balancing is a ~m.llot i~toblcrn. Figure 3.46 shows the various modes that the rotor shaft undergoes

i .is i l i).isses rlirougli these criticals, note that the mode shapes are also effected by

thr: bi!.$l [ng stiffness. High-speed balancing of these rotors i s sometimes a must for

i smooth orx~at ian. This inevitably means field balancing since there are only a few iigs whch can balance these rotors at design speed, and these also operate in a vacuum chamber. Figure 3.47 is a typical rotor response curve for a four stage rotol. Herr the rotor is operating above the first critical, but the steepness of the c u ~ v c near the design point i s a cause for concern. Modification of the rotor and <.ti,inil~ in bt:;iring stiffness moved the slope from design point.

MODE FORM

Figure 3.46. Mode forms for various bearing rliflners.

Ffgorr 3.47. Rotor response curve.

All rotating machines vibrate when operating, but the failure of the bearings s

due mainly to their inability to resist cyclic stresses. The level of vibration a i rn t can tolerate i s shown in the severity charts as shown in Figures 3.48. 3.49. a ~ i d 3.50. These charts are modified by many users to reflect their critical machlnc5 i r l

which they would like to maintain much lower levels. These charts should only t , ~

used as guidelines, note that the absolute levels are effected by the speed of 1 1 ~

. . 1, , ! d 10 L O O 1.000 1 . 0 0 0 1 0 . 0 0 0

ir.ii,,.i:CY, i ! i

Figure 3.48. Severiry chart.

angero our - Shutdown

~bnorms l will deterlorats IIISPBEt BBdV possible

Problems - Keep close watch

Normal

I 0- oangerou$ - Shutdown --- A- Abnormel will dererlorate

_,( -- ~~ -- Inspect sr earlv es prrssibls

P- problems - ~ e e p close

, watch

./,'- N- Normal

. .

/ . !

Figure 3.49. Severiry charr.

~ n a c h i i x The velocity chart figure i s the only one which is least effected b y this p~ob lem.

Forces Acting O n A Rotor Bearing System

There are many types o f forces which act on a rotor bearing system. The forces

! D- DangerOYI - Shutdown . A- Abnormal will dererorar ,

Inspect as early ar poraible

P- Problems - Keep close watch

N- Normal

0.01 -

Figure 3.50. Severrry chart.

can be classified in to three major categories: 1) casing and foundation. 2) focci.,

generated by rotor motion, 3) forces applied to a rotor. Tablc 5 by Rc~qer ( R i l l ? ) is an excellent compilation o f these lorces.

1. Casing and Foundation Forces - These farces can be due tu fuur ,d i i t~w instability, other nearby unbalanced mach~nery, piping strallls, rotal lon 111 ~ I ,WI, ,

tional or magnetic fields, and excitation o f casing or loundation nalural l i ,

quencies. These forces can be constant or variable with impulse loadings. The i : l f i , i : t

o f these forces on the rotor bear i~ ig system can be large for example rnplrtq S ~ I ; ~ I I , ,

can cause major misalignment problems and unwanted forces on the bearfngs.

2. Forces Generated b y Rotor Mot ion - These forces can be c l ass f~ r i l I I ~ I C I

major categories: 1 ) forces due t o mechanical and material properties, 21 forces d u

t o various loadings of the system. The forces due t o mechanical and ma te ra

properties are unbalanced which can be caused b y lack o f homogenity o f materas

rotor bow and elastic hysterisis of the rotor. The forces due t o loadings of thii

system are viscous and hydrodynamic forces in the rotor bearing system and varloui

blade loading forces which vary in the operational range of the unit.

3. Forces Applied t o a Rotor - These forces can be due to drive rorqiles. couplings, gears misalignment and axial forces such as due to balance plston a ~ i d

thrust unbalance. These forces can be very destructive and in many cases result i r , -

total destruction of a machine.

Rotor Bearing System Instabilities

Instabilities i n rotor bearing systems may be the results of different fn rc r i ,

~necl!;iu~~sms. However, one can divide these into two general and distinctly different crilrqories. The forced or resonant type is one in which the frequency o f the

oscillat~ons 1s dependent on outside mechanisms. The second category being the self t i s c : I ~ ~ i instal!ilities these are independent o f outside stimuli and are independent of

f I i < , I h m ~ ~ ~ m i ~ y . Chd~i ict~!r i~t ics o f these forces can be seen in Table 3.6.

T;8bla, 3.5. Forcer Acting on Roror.Bearing Systems (Ref. 1151.

Source of Force Dercription Application

Hydrodynamic forcer, rtatlc Hydrodynamc forcer. dynamlc.

O l ~ ~ i l l l l l a l i4iistlc 111:8111

Sriffners reaction forcer

Tr;lnlient torquer

Heavy applted rotor force

netic field Imprerred cyclic ground-or foundation-motion

Air blarr, explorion or earthquake. Nearly unblanced machinery. Blows, impact

Present in all rotatrng machinery.

M o t ~ o n around curve o l varying

radtus. Soace ap~ll icalionr. Rotarycoordinufed unalynes.

P ~ ~ w e r l y o f rowr material which a~lpei~ws when rotor I S ~ v ~ l i c a l l v rli:formrd tn bending, torrionally 11, t l ~ i d l y .

Crn~st!uct~on damplngar~r ing f rom ~cllative mot ion between shrunk l i t ted arremblier.

Dry-fricrion bearing whirl. Vircour shear of bearings. Fluid entrainment in turbomachinery. Windage.

Bearing load capacity Bear~ng rt i f fnerr and damping

properlies. Rotors with differing rotor lateral

rliffnerrer Slotted rotors, electrical

machinery, Keyway Abrupt meed change condltionr Significant in high-speed flexible

rororr with disks. Accclerarcng or constantspeed Operation

Internal comburtion engine torque and force companentr.

M#ral#gned couplings. Propellers. Fans. Internal combustion engine drive.

Gears with indexing or positioning errors

Drive gear forcer

Pwcers Equipment Series Vulu,ne 3

Table 3.5. Forcer Acting on Rotor-Bearing Systems (Ref. 1151 (Continued)

Source of Force Description Application

Misaligned 3-r-more rotor-bearng assembly.

Gravity Non-vertical machines. Nan-slutto1 applicationr.

Magnetic field, stationary or Rotating electrical machinc,ry rotating

Axial forcer Turbomachlne balance pfitor> Cvcl!c forcer from propclle~, u! fan. S e l k x c i ~ r d bemnq lorct";.

Pneumatic hammer.

(Conciudcdi

Table 3.6. Characteristics of Forced and Self Excited Vibration

Forced or Self Excited or Resonant Vibration Instability Vibration

FrequencyIRPM Relilfionshlp NF - NRpM 0' N o r

rational fraction Ampl$tudc/RPM

Relatronrh~p Peak in narrow lbanrlr 01 RPM

I n 1 o f Damping Add. dampin<] Reducc arnpliludr No change in RPM a r which

i t OCCUIS

System Geometry Lack of axial rym. external forces

Vibration Frequency At or near ahafr Critical or Natural frequency

Avoidance 1. Crittcal Freq. above running speed.

2. Axirvmetric 3. Damping

Constant end relatlvelv indrpw dent of rotating rpecd

Blorsom8ng at onrer and c < r n c < c w to increase with increar8rlq Ri'M

Add. darnrmng may d l . 1 ~ i r l ;#

hrgt~c:r RPM. W$ll no! rrm~c:ts,8il'~ affect amplitude

Independen~ly o f symmetry s,n;iIl dellsct8n t o m ax(wmetr<c system. Amplitude w d sell prapagare.

same

1. Operating RPM below onset.

2. El~minater 1nrtabll8tv Introduce damping

1. Forced or Resonant Vibration - In forced vibration the most usuai drvnq frequency in rotating !machinery i s the shaft speed or multiples of this sperd. T h s

becomes critical when the frequency of excitation is equal to one of the natiJr;ll frequencies of the system. In forced vibration the system is function of thr f t i , .

quencies that can also be multiples of rotor speed which can be exicted b y fir: quencies other than the frequency of the speed such as blade passing frequrnui:,., gear mesh frequencies. and other component frequencies. Figure 3.51 shows th:lt

for forced vibration the critical frequency remains constant at any shaft speed. T t h

SY N C H R O N S V I B R A T I O N --

I

V I B R A T I O N A M P L I T U D E

5 2 m

5 % I * " u -

O N S E T SPEED

V I B R A T I O N F R E O U E N C Y

V I B R A T I O N

. Hysttitctic Whi t -- This type of whirl has been diagnosed as occurring in t lc~xblv ~o tc~ rs duu pr imai~ ly to shrink fits.

Whw 2 ~ o i i ~ a l deflection i s imposed on a shaft, a neutral strain axis is induced

nomi11 to thv direction of flexure. From first order considerations, the neutral axis of s t l e s s I S cw~,c ida~l t with the neutral axis of strain, and a restoring force is

drv t i lo~ed perpcndiculal. to the neutral stress axis. The restoring force is then

ipalnllrl a i d oppos i~~g the ~rrduced force. In actuality, internal frictionexists in the

shaft which causes a phase shift in the stress. The result is that the neutral strain

and ntmtlal stless axis ale displaced so that the resultant force is not parallel t o the

~ l v f l i ? c t ~ ~ , ~ ~ . TIrt: t:u,qenlial mmponsnt which is ilorm,ll to thi! deflection results in a

~ I I ~ I I i l s t , ~ b ~ l ~ t y . AS whirl beyims, t h ~ ce~i t r i l i~ya l force component Increases thus

musing I'iryer defli:ctiot,s which result in larger stresses and still larger whirl forces.

Ths 1yw: o f increasing whirl m o t o n may eventually be destructive.

I t u f t m iequires some unbalance initial impulse to start the whirl motion. The

cirzu~lti;il rrffect is caused by interfaces of joints in a rotor (shrink fits) rather than

c l ~ i f t ~ t s I O the ~matclinl o f the rotor. This type of whirl phenomenon occurs only at

jot , i t~ond speeds above the first critical, i t may disappear and then reappear at a

lrglier r i~eed. To reduce t h ~ s type of whirl some success has been achieved by ~ c ~ c i i ~ i : ~ o , i 1111: IIIIIIIIW~ 01 sc:ixui~tt: p i ~ l IS. ~ ~ ~ s t ~ ~ c t i n j j the shiink fits, providing some

lock up of ,~ssarnL>Imi ulc l i la~~ts. I,. D ly Frlctlon Whirl - Dry f l ict ioi l whip is (!xpriienced when the surface of a

~ < , t . ~ t r , q sh;?ll i:oinm into cnlitact with an ~lt i l t lb~icatt:d statiorrary guide. This can

t.,kr jrl.ic:a in ;In unlub, ~cated journal, contact in radial clearance of labyrinth seals,

and loss of clearance in hydrodynamic bearings. This phenomenon occurs when the contact is made between the surface and the

rotciti~lg shaft, the coluomb friction will induce a tangential force on the rotor. This f r c r o n forcu is ;iprpoximately proportional to the radial component o f the contact

i o ~ c t i thus ci iut inj l a comlition for instability. I t should be noted that the whirl

detiictto!i is counter to the shaft direction.

c. Oil Whill - This instability begins when f luid entrained in the space between

the shaft and bearing surfaces begins to circulate with an average velocity of one-

half of the shaft surface speed. The pressures developed in the oil are not symmetric

about the rotor. Due to viscous losses of the f luid circulating through the small

clearance, higher pressure exists on the upstream side of the f low than on the

downstream s~de. Agatn a tangential force results. A whirl motion exists when the - t;mgpnti;il force exceeds any inherent damping. It has been shown that the shafting

Inust ri,t,itr .II app~oximatsly twice thc critical speed for this to occur. Thus the - ~ ; i t l ~ r 0 1 III~IIIIIIC~ to RPM IS closc to 0.5 fo, oil whirl. I t should be pointed out

hurt, tIr;~t tlris p t ~ c ~ l o ~ n r n o ~ > is ~ , o t rostiicted to tho bearing but also can wcur in

st,:,Is.

Thu most obwous way to prevent oil whirl is to rfstrict the nlaximum rotor

s ix rd to less than twice its crit~cal. Bearing designs ilicorporating grooves or tilting

pads have been found effective in inhibiting oil whirl instability. Changing the o ~ l

temperature can also sometimes get the machine out of an oil whirl condition.

d. Aerodynamic Whirl - Although the mechanism is not clearly understood, i t

has been shown that aerodynamic components such as compressor wheels and turbine wheels can create cross-coupled forces due to the motion of the wheel.

The acceleration or deceleration of the process fluid imparts a net tangential

force on the blading. If the clearance between the wheel and housing varies circum-

ferentially, a variation of the tangential forces on the blading may also be expected.

resulting in a net destabilizing force. The resultant force from the crosr.coupling of angular motion and radial forces may dr:stabilhzc the rotor causinri a whirl motc>~l .

e. Whirl Due to Fluid Trapped in a Rotor - T h ~ s type of whirl occurs w h u ~

liquids are entrapped inadvertently in internal cavities of rotors. The fluid does not remain in a radial direction but has a component in the tangential direction. The

onset of this type of instability occurs just above the first critical and below twlce

the critical speed.

BEARINGS FOR HIGH SPEED MACHINERY

Journal and thrust bearings are among the most important components to assurf!

maintenance free running of high speed turbomachines Bearings in these machirm

range from simple journal bearings and flat thrust bearing to multiwedge designs for

both thrust and journal bearings. Some of the many factors that enter into the

selection o f such bearings are:

0 Speed range of shaft 0 Maximum misalignment that can be tolerated by the shaft

Loading of the compressor inlets

0 Oil temperature and viscosity

0 Foundation stiffness 0 Axial movement that can be tolerated

Type of lubrication system and its contamination

0 Maximum vibration levels that can be tolerated

Al l rotating machines vibrate when operating, but failure of the bearings is

mainly due t o their inability to resist cyclic stresses. The level of vibration that a

unit can tolerate i s shown in the severity charts as shown in Figures 3.48. 3.49, a~rd

3.50. These charts are modified by many users to reflect the crit~cal values for their

machines.

Journal Bearings

The journal bearings for turbomachinery has a fluid film that carries the luail. Fi lm thiclmrss in most apr,lications range from 0.0003 in. for qasr:i to 0.008 i l l . fo l

hydrostatic oil lubricated bearinys.

Figuiu 3.53 shows a number of joutnal bearings in which a positive supply of

IhLmc.mt i s fed to the beari~lg at all times. The circumferential grooved bearing

i,o11n,1lly has the 0 1 groove at half the bearing length. T h ~ s provides better cooling.

but reduces load capacity by dividing the bearing into two parts. The cylindrical

I,~,;~I~IKI, ust :<l H I turhnes, has a split construction with two axial oi l feed grooves at 111c! s 1 ) I l . TI,,, I ) I ~ : S S U ~ C ! d i m bc!d,iri(j IS used whc~(! b ~ : ~ r i n g stabtlity is required.

l%t' must L.olrlnluo bed, ing i s the t l l t~oq pitd typ~!, whose inost important feature

i s st,II .i l<)~>ment whe~, the bewnq I S used wlth spht:rical pivots. This bearing offers ll ie gccr;itest i,cre;isu III fatigue life because of these advantages:

1. S1.11 d1911111g f o ~ optimum alignment and mfrlimum limit.

2. Thermal conductivity backing material to dissipate heat developed in the 0 1 '

f i lm. 3. Thin babbitt layer. centrifugally cast to a uniform thickness of about 0.005

in. Thick babbitts greatly reduce bearing life. Babbitt thickness of about 0.01 . in. reduces bearing life by more than half.

4. Oil f i lm thickness can be varied by changing the number of pads, directin~j the,

load onto or in between the pads, or changing the axial length o f the r i a d 011 fi lm thickness is critical when making beating stiffness calculations.

Thurst Bearings

The most important function of a thrust bearing is to resist the unbalaricrd lorci, developed in the working fluid of the machine, and to maintain the rotor i n 11,

Position within the prescribed limits. A complete analysis of the thrust load n>ij$i

be conducted. Compressors with back to back rotor greatly reduce thi? load o!. thrust bearings.

TAPERED-LAND THRUST BEARING

(A

F q u i r 3.54 sliows a numbel of d~f fe lent types of such bearings. When properly

dusqni!(l, thu tapered-land thlust bearing as seen in Figure 3.54a can take and

S I I I ~ ~ O ' I d l w d equal to that of a tilting pad thrust bearing. With perfect alignment. 1 1 i:,lil ~n,.ltch the Io,~d of avrn a sdf uqu;lli/ing tilting pad thrust bearing. Figure

3.5411 1s .I I I O ~ I c~ju. l l~/ ing ttlting p x l thrust be i l r i~~g that pivots on the back of the ~l.i<l . i I i , l~f ;, !xII;BI line. Fiqurc 3.54~: is 2 noc t!quali/ing liltirlg pad bearing whose

[,ails .IIC sul,l,ortcd <,I, s p h c ~ c i ~ l p ~ v o t poittts. Sillco this slows the pads to pivot in m y direction. al~gnment is not as serious a problem as in the other two types.

Figure 3.54'1 is the Kingsbury type self equalizing thrust bearing. This bearing

v i~tual ly t:liln,~;ltes the problem of misalignment. The major drawback is that

st;ii~I;iid des~grrs requtre more axial space than do a rlon equalizing type.

MISALIGNMENT

Ti l t . ;!mol,nt of inisalignment which can be tolerated depends on the types of

juit l t~dl a i d thrust bearings used. As previously mentioned, the tilting pad type

bearings greatly reduce the problem of misalignment. Figure 3.55 shows misalign-

ment in both the journal and thrust bearing. The effect of misalignment on a

jour~,.il bcafirig is that of the shaft contacting the end of the bearing. Thus, journal

Lengtli is a criteria i l l the amount of misalignment a bearing can tolerate: a shorter

Iwgt11 L~r; l ( i~ ig obviously can tolerate more misalignment. The effect on the thrust

hc,~ii,i(l i s to load up m e segment of the thrust bearing arc and unload the opposite st~r j i~wnl. T h s clft?ct i s (more pronounced at the higher loads and less flexible

i , t ' < l l l i ,( lS.

ill xi lust lot rn~saliqimcnt, beyond correctirlg by using tilting pad bearings.

vawous inisnlignnient technques have been u s i d Initially, the smplest and most

conilnan tyhx o f alignment technique is used. which is the so called "cold align-

melit'' ~i iethod. Thfs is also usually referred to as a base alignment. Once this is

accomplished hot alignment checks are required; thus, hot alignment is carried out.

Hot ,lli!]~,ment techniques measure the changes when the unit i s operational and

t < , m ~ i t , l . t ~ u ~ t ~ yrowtli stab~li/ed so th'lt accurate alignment data is done. The most

~ rco inmrndr i i technique I S to do the cold alignment. using the "reverse indicator

!j1,1pl~ic;ll +, lot t i~q" J I I ~ the hot a l i g ~ l m ~ n t by thv US,, of m~chall ical technqiues such

.is "Aci:ohgn" ot "DOD Bws" o~ o t l ie~ hot olgnment techniques such as optical or

I;isrl tecli~i~ques.

Tilt, ~cvsrsc indlcaro! gi;ipli~cal plotting trclinique is nol.mally done when the

UIIII I S cold T h s i s d o ~ ~ e by first layi~ig out the di:sired hot operating line on a

I I . This is the line which shows the final desired operating equilibrium

c o ~ ~ i i l i ~ ~ ~ i s . Then the desired c d d posltion of thu shaft i s plotted. Figure 3.56 shows sucli ;i cli;~rt. TIIC ,rctual posftlons o f tlic shaft In the field are then taken. This

~ ~ l i u l r n a t ~ o n i s plotted and the d~ l fc rencr cumputcd and the shims added to the

I Tl i~s piocudufa i s then rupcatcd aftel hilt checking the alignment. To

~nilhc t i l t : I io t :ilig$,ment checks, a mechan~cal i~ l ignmwlt procedure is recommended.

Journal aear lnq

The heart o f this technique i s a mechanical instrument with built n dial indcdti>i

t o measure displacement from precfse and reference marks. Thus, after thi! ~ r u r ; r l

cold alignment, the train is ready for start up. Bench marks arc estahlishcil i i t well end and each side of each unit of the traln as close to the couplinq: ;as posst>li:, i i ~

most units, this is the bearing housiny as seen in Figure 3.57. The t ran ir t h w

1 I COMPRESSOR 1

-q- B E A R I N G HOUSING

Figure 3.57. J y p i c , ~ l p l ~ c ~ m e n r of benchmarks on foundar,or. and brarrng hourrng.

rt,~<texl UI, JIKI upr~;ltad ;$I des~gn or near design conditions as much as possible and

~ ~ ~ ~ n p c ~ ~ l u ~ ~ ' s ~ l i e ~ ~ / I o w ~ ' d 10 stabilize. Another set o f readings are taken as shown in

FI:)(I,C 3.58. Thus, the actual thermal growth can then be plotted on thegraph and

11c:w cumectlofis can be computed. The technique outlined above i s used on new

imachinrs, and on old machines, a reverse technique can be used. First, a hot

. i l~q~~mei r t check i s made, then a cold check, followed by the mechanical reverse

~ , ~ I c ; i t o t ~<~;,dings. T h ~ s iniormation IS plortud and realignment measurements taken.

VECTOR B VECTOR A

\ - , - / FINAL

i

INITIAL SHAFT POSIT ION

Figure 3.58. Graphical deiermirlation of shaft in hor poiriior~ relative to coldpositioi~.

The above outline i s simple, but in actual practice, one must develop ths s k l .

Some of the major problems encountered in alignment arc caused by pipe stiar, This i s caused by the piping being off from the intake or exhaust from ;1 h:w

thousandths to several inches. Many engineers take the attitude that they cdo~t<,t

understand how a small pipe like this is able to move a large piece of rnachiru: y. The results are very surprising. Tension on pipe hangers can change thr vibralzw

level considerably. Another contributor to the alignment problem 1s the gear caslll(j.

Thermal growth in the new fabricated cases has been unpredictable in many a p p l ~ ~ cations. I n some cases. gear cases rise with a twist. This creates another problem

which is very hard to correct. Short couplings also present a problem and mayntfy

misalignment problems. Some users are now specifying that the coupling spacer will

be at least 18 inches long. The above has been a general cursory outlook olr alignment techniques. Alignment of high speed machinery must be accurate: othi:~

wise, major problems wil l arise. Sometimes, to prevent major shutdowns for cor~

rection, heaters are added to one or the other legs of the unit t o align them w h ~ l i

running. This technique i s not advised as a cure, but as a temporary relief wht!rt shutdown is possible.

COMPRESSOR SEALS

The internal seals that prevent leakage around the impellers are usudliy

labyrinth type, as shown in Figure 3.59. They have a series of circumferc~lt~d! knife edges that are positioned closely to the rotating impeller. If damagi!d ay rubbing erosion or corrosion. these knife edges will lose therr effect<vcncss. In sor !v

j r j ls ai-P used The oil or l i q u ~ d f i lm seal consists of two stationary bushings that

surrounc the shaft with a clearance of a few thousandths of an inch. Oil at a

nornlrlal late of 10 gpm I S introduced between the bushings at a Positive Pressure

hlqhef tharl that of the process gas, and leaks in both directions along the shaft. The c,ll i s rcrmoet i in tha seal housing by "0" rings. To limit the inward oil leakage, the

d l f t : r ~ ~ ~ t ~ i i l /> I~ssuI .~ : X IOSS the inner bushing i s only a few pounds per square inch.

The inner leakage rate varies from 1 to 4 gph, depending on the size of the seal, but

IS independent of the gas pressure being contained.

Th~s leakage is collectcd in a chamber that is usually separated from the gas

stream by a labyrinth seal. Overflow of oil from the leakage chamber and its

si~bsncluaot wllciing the cumpressor is the biggest problem.

R O T A T I N G C A R B O N R I N G I F L O A I I N G BnBB1TT-FACED ' ROI ,>TNC. S l A L R I N G S l E E L R l N G

j S 1 , \ l I % N A R Y SI E E V E J SI:AL I Y I P E R R I N G

i 5 P 9 i N t IRETIIINER i) SEAL OIL DRAIN LINE

:> S l ' l i l N C lli R U F i i R G A S l N J E C T O N P O R l

I , <;,ili A N l l t ' i l N l , % h l l N A l 1 I1 I 1 I IYl 'A!iS UI11t ICE

<?It i l l l A l N

Mechanical contact seals have two major elements as shown in Figure 3.61

These are the oi l t o process gas seal, or carbon ring, and the oil to uncontaminated

seal oil drain seal, or breakdown bushing. This seal can maintain a lower inner leakage ratio wi th higher oil t o gas differential pressure.

In operation, the seal oil pressure i s maintained at about 25 to 50 psia over the Process gas pressure against which the seal i s sealing. High pressure oil enters the

seal cavity, completely filling it. Some of the oil (ranging from 2 to 8 gph) s forced

across the carbon seal face, and flows out the contaminated oil drain. The mecham

cal seal's great advantage over the oil f i lm seal is that i t has a minimum effect on rotor dynamics. On the other hand, when the oil fi lm bushings lose thmr lri:i,

floating feature. they can upset the stability of the rotor when operating at h q h

speeds.

BALANCING

In large rotors operating at super critical speeds, balanc~ng of the rotor bocamri

very important. This i s particularly true since stresses caused by unbalanced rotof,

are proportional to sequence of the frequency of rotation. In perfectly balanced

rotors, the vertical axis of the rotor should be located on the axis of rotation. In

reality no rotor is perfectly balanced and the unbalance is caused by various facti~cs

such as non-homogeneous material, misalignment of beari~igs and coup l~~qs . thermal gradients, hydraulic and aerodynamic faces, etc.

In the rotor, the exact locatior~ of the unbalance cannot always ba foood The. only way to locate any unbalartcc is to study the v~briltlori of the rotor. B , i l j ~ ~ c ~ i ~ i l

of the rotor is done in two stages: 1) the static balance and. 2) the dynarnlc balance.

The purpose of static balance i s to make the center of gravity of the rotol

approach the center line. Static balance is usually performed by placing the rotoi

on a set of frictionless supports; the heavy point usually has a tendency of ro l l~ng

down. Noting the location of this point helps determine the amount of unbalanci:.

Dynamic balancing i s more complex. The rotor i s placed on its supports and by

noting the vibration pattern, the location and amount of unbalance can be deter-

mined. Balancing i s then done by placing correction weights at the appropriate

locations in plates perpendicular to the rotor. A better balance is obtained by

placing weights in as many planes as possible.

It is common practice to balance individual components such as impeller wheels,

Couplings, etc. and to then mount them on the shaft and the final balancng per^

formed, i f then an unbalance exists, correction is then usually made to the last

impeller installed. The best technique for high speed rotors is to balance thrirn

not in low speed machines. but at their rated speed. This is not always posslbli: r l

the shop; therefore, i t is often done in the field. New facilities are being b u t t whtch

can run a rotor in an evacuated chamber at runniny speeds n a shop. F l y u r c 3G2 shows the evacuation chamber and Figure 3.63 the control room.

Higli speed balanc~ng should be considered for one or more of the reasons listed below:

1 Thii .ictudl field rotor operates with characteristic mode shapes significantly

different than those which occur during a standard production balance. This is a primary consideration since flexible rotor balancing must be per-

formed with the rotor whirl configuration approximating the mode in question. 2. The oeprating speed(s) is in the vicinity of a major flexible mode resonance (damped critical speed). As these two speeds approach one another, a tighter balance tolerance will bi! required. Of special concern are those designs which have a lower rotor bear~ny stiffness ratio or bearings in the vicinity of mode nodal points. 3. The predicted rotor response of an anticipated unbalance distribution is sig. nificant.

This type of analysis may indicate a sensitive rotor which should be balanced at rated speed. It will also indicate which components need to be carefolly balanced prior to assembly. 4. The avai!able balance planes are far removed from locations of expected unbalance and thus relatively inetfective at the operating speed.

The rule of balancing i s to compensate in the planes of unbalance whei, possible. A low speed balance utilizing inappropriate planes can have an adversi, effect on the high speed operation of the rotor.

In many cases, implementation of an incremental low speed balance as the rotor is assembled will provide an adequate balance, since compensatiorrs are being made in the planes of unbalance. This is particularly effective with desiqns incorporating solid rotor construction. 5. A very low production balance tolerance is needed in order to meet rigorous vibration specifications.

Vibration levels below those associated with a standard production balanced rotor are often best obtained with a multiple plane balance at the opuratrng speed(s). 6. The rotor on other similar designs have experienced field vibration problems.

Even a well designed and constructed rotor may experience excessive vibrations due t o improper or ineffective balancing. This situation can often occur when the rotor has had multiple rebalances over a long service period and thus contains unknown balance distributions. A rotor originally balanced at h ~ g h speed should not be rebalanced at low speed. Influence coefficient methods present by far the best method for balancing in

the field or even shop balancing. An additional advantage of this method is the fact that rotor can be balanced at many speeds and planes in one attempt.

The method of influencecoefficients i s based on the principle that the deflection, Zi, at plane i i s a function of the unbalanced forces or:

Z. = e.. P. I I I 1 1341

The matrix of ei,, is called the influence coefficient matrlx or co rnp l~a~~c :~~

couplings than with gear couplings on certain types of units. A rather specific

control i s placed on the disc deflection range, and the equipment has to be adjusted

axially to suit with more accuracy than with gear couplings, Some units are uery

diff icult t o move once they have been set. A gas expander, for instance. Where this

is a problem, however, arrangements can be made to permit repositioning the coupling by remachining the components for one time per coupling fitup. or by

adding a spacer plate. The application of a disc coupling therefore almost completely eliminates the

very generation reeducation of coupling users in the intricacies of the design itself ~- once the coupling is designed, becausc in order to design it. the couplinq chxacti!r

istics must be determined and adjusted to suit the connecti~ig rotols. Furthairnoii.

there are no changes or wear taking place in the coupling characteristics throughout

the life of the unit.

One of the main features of the disc coupling is its know,) axial deflectirg lo;ld value. When compared to the varying unknown slidng friction factor in the gca,

coupling, this feature eliminates the greatest concern of the user and to a, , equal

extent the designers of the rotating,equipment.

In most cases, both the disc, and the gear couplings can be applied. The d f f i? r - ence between them can be noticed, but even though one has advantayes over thr!

other, both will succeed as couplings. The one overriding difference to users is the obvious promise that the disc coupling will be less susceptible to usage than the gear

coupling, and that i t wil l require less attention.

LUBRICATION SYSTEMS

API Standard 614 covers in detail the minimum requirements for lubrication

sytems, oil type shaft sealing systems, and control oil supply systems for special

purpose applications.

The major components of a typical oil system are reservoir, oil pumps, oil coolers, oil filters, alarms and shutdowns, and thermometers

The reservoir should be separate from the equipment base plate. I t should lx:

sealed against the entrance of dirt and water. The bottom should be sloped to thc

low drain point and the return oil lines should enter the reservoir away from the oil

pump suction to avoid disturbances o f the pump suction. The working capacity

should be at least five minutes based on normal flow. Reservoir retention time

should be 10 minutes, based on normal flow and total volume below mi~limurn

operating level. Facilities for heating the oil should also be provided. I f therrno

statically controlled electrical emersion heating I S provided, the maximum watt dr:t>-

sity should be 15 watts per square inch. When steam heatinq i s used, the h i : i l t ~~ , ( element should be external to the resrrvolr.

The oil system should be equillr~cd with a main 011 pump, a standby a n d f o r

critical machines, an emergency pump. Each pump must have its own drvur aod

check valves must be installed on each pump discharge to prevent reverse flow

I r l MOIJ alqssod lsaqfilq aql apAoJd 01 sdwnd JO d w n d aql f i u l uun~ Aq suaaj3s

.',rl: i. ' l 1 1 i o q 1 1 ~ qsnlk atdl uf iag 1 1 0 pasn ueq3 l o Mau ~I!M qoniasa> aql l l ! ~

' q s n ~ aq l 6uynp . . 2 ~ . uounlrl j j o 6u?301q AlaieuJalje hq sa!113oan wnwlxew u!eiqo 01 walshs aql

. * / ? u o i I n a s 01 Aiessssau 2q hew 11 'slall!J PUP JOnlaSaJ aq1 OlU! sjlqap aql AJle3 01 t i l ; ' l s A i d i j l qfino141 sallinolan ~ o l j qB1q uo lualxa a61el e 01 luapuadap s! qsnl,a14

1 0 s w ~ n 1 1 3 a ~ ~ a a q l 'Mol l 6UlqSnl~ w n w ~ x e w ~ o l l e 01 uo !~ ! sod uado l ln j aqi u! sanjen I I I s ; l lqe!ene l ou 81 uaai3s IC311103 aq1 j ! alqeldas3e s! uaaJ3s i e j j

'' lNi 'l>ali;,jaJd s! u a a m padeqs le3luo3 a q l 'mdwn l q3ea l e suaa13s alell lal qsaw

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uo!l>alas tues l lqn l

lic.ir rhc u i to 160°F ti, tlicrm,~lly exercise the pipe. Tap the piping to dislodge

4uLms. csp!cialIy ;ilonq the horizontal sections. Flush through one complete

tmnprfalurc cycle, then shut down and install the telltale screens and flush for an

xlditronal 30 minutes. Remove screens and check for amount and type of debris. Repeat the p!eced~ng procedure unti l the screens are clean after two consecutive

it,spt,~ticlos. O b s t : ~ ~ the prcs~orc drop across the filters during the flushing opera- t i o n 110 liut allow the pressure h o p t i ) exceed 20 psig. When the system is con-

sidereil clean, empty the oil reservoir and clean out all debris by washing with a

i lcte~qcnt solution followed by a fresh water rinse. Dry ~nterior by blowing with dry .lir , i t l c l Y~CLIL I IT I I~I'u W~I~LII. R ~ p l i x r filler clemrnts. Remove jumpers and replace

orhcr!s. Return controls to thetr ~iormal settings. Refill the oil reservoir with the

s;ime oil used in the flush if lab tests indicate if is satisfactory; otherwise, refill with I,k!M!Liil.

This procedure w I I allow the fastest poss~ble cleanup of the oil system, mainly due to the high f low velocities obtained during the flush. The objective is t o carry

tlie debris into the reservoir and filters and the turbulence from the high flows

along with the thermal and mechanical exercising of the piping are the main factors

rircessary for a fast and effective system cleanup.

GEARS

c;c,.>ns ;IIV u l le o l thc inost used conf~lings between the driving and driven com-

~ l j i ~ ~ m ~ t ~ . I I ~ ~ I O ~ C ~ gc:m select~on can causc majol problems. Gears often form the

wc.ihcst link in thcwholachain, mainly due to the fact that this is the only item that 1s ~ ~ y l i ~ i ~ c d tu ,lpl,~iltl! w l lh metal pdits i l l close contact. The performance of a gear

s rlept!ndt!nt 011 many factors, some of these are:

1. Pressure Angle

Decis~ons regarding the ptessure angle need to be made early during design. The

length of time of action and contact ratlo are directly dependent on this factor. The

noise generated is dependent inversely on the contact ratio. Values selected for

ptesswe angle range from 17-22'.

2. Helix Angle

T h i Ili,li\ .II,/~II, I , U I ~ ~ , C 110111 5 1 ~ 1 20" I~II t h ~ s111(11t! hcl~cal gcal and from 20 to

.I!;' !I,,, ( l c ~ i ) ~ , , II,:I~~~,LI gt,,t~. TI,,, SI,VI,! ~ h t ~ l ~ c ~ t l ~ C : ~ I S iwc ~morc accurate and less

OIW,,. ti, 1:01111110g t l~rust . Th,y J I ? ,tlso less LIXII(?IISIVC dlthough they require expen-

slvc t l u i~s l i s I thc Iical lo;,rl o l the thlust bearings makes them less

i!lficicnt. The double helical gear are caster to manufacture, but cutting the gear

loeth i s inwe rxprnslve. They are more eftlclent and do not require expensive

thrust bearings. Gc.iis a e also available in varying hardness. Medium range gears are not too

sensitive to operational errors and wear slightly before falure. They are also ,not 2s

susceptible to scouring as hard gears.

The most common used bearings for gears are journal arid tilt pad bearnys Thrust bearings vary from ball bearing to self equalizing tilting pad bearings. The

gear housing i s usually made of steel. I t should be stress relieved before final matching. Housing should be rigid enough to prevent misalignment. Also, sufficient cleaning should be provided to prevent oil chok~ng. Lubrication in gears serves 4 s

two fold purpose that o f lubrication and cooling. The most recommended lube oil

is the AGMA No. 2. While installing the gear unit, care should be taken to allyrr the unit propecly.

Misalignment of gears can cause unequal distribution o f teeth loads and distortfor!

of gear elements. Before startup, gear face contact should be checked. T h ~ s should

be about 90%. Gears with modified helix angle may pose problems, and in such

cases, the manufacturers recommendations are useful.

The system may influence the gear operation. Particularly critical speeds

(torsional, lateral and axial) must be accounted for. Coupling lockups pose another

problem. In addition, the gear must be able to handle the maximum load expected,

and not just the design value.

Gear noise is causedsby a variety of factors. Extra heavy cast iron or double wall

housing are useful i n reducing noise. Acoustic enclosures are useful in reducing noise. When installing gears, in addition to cold alignment, hot alignment is ntxcs~

sary. High speed gearing must be treated with care or the user is in for a onq unhappy association.

CONTROL SYSTEMS

The controls for most compressor trains consist of two major systems: on<! 10,

the lubrication system, and one for the compressor. For the lubricatiori systcrn.

minimum alarms are: low oil pressure, low oil pressure trip (at some p o n t lowel than the alarm point), low oil level in the reservoir, high thrust bearing metA

temperature, and high oil temperature. Each pressure and temperature srnsjnq

switch Should be in a separate housing. The switch type should be single pole, double

throw, and furnished as open (deenergized) to alarm and close (energize) to t r p Pressure switches for alarms should be installed, with a "T" connection for presswe

gage and bleeder valve to test the alarms. Temperatures should be monitored 11,

the oi l piping to and from the coolers, and at the outlet o f each radial and thrci,.i

bearing. Bearing-metal temperature should also be measured, since problerns will

show up much faster in the metal temperature than in the oil temperature.

Pressure gages should be provided at the discharge of the pumps, bearirig hi!ati*ir. control oil line, and seal oil line. Each atmospheric 0 1 1 dra~n line should be equlpjxx

with steel nondestructive bull's eye flow indicators, pos~tioned for viewing throuqti

the sides.

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29. i i l ~ i r r , W , I<., and Johrrrm I. A,. "Pelioimance Effect o f Fuilv Shrouding a Centrifugal Sulxrcharger Impeiler." NACA ARR E5H23 or E18.

30. G!nr lw~g. A.. Rii ler, ill. K.. and Polarlcr. J., "Effecrs of Performance of Changing the ! > l v , r , i m o l Worh 8etwc:en lhcrease of Anguiar Velocity and Increase of Radlus o f Rota. , ,on r r ~ ;to Im~~el le r . " NACA T N No. 1216. 1947.

31. ~?\I>.IIV, I:., "The Fric810n LOSWI 01 Slesm Tu8bmt: Discs." Brown Boveri Review. Val. 21, 1934. 11. 120.

32. Bullock, R . 0.. and Fsnger, H. 5.. "Surging rn Centrifugal and Axial-Flow Comprerrors." SAE Quar. Tranr. Vol. 6. 1952, p. 220.

33. Shepherd. D. G.. "Pr,nciples of Turbomachinerv," The MacMillan Company, N.Y., 1057. 34. Slockmun. N. 0.. and Kramer. J. L.. "Method for Derign of Pump Impellers Ur inga High

S w r d D#g#tal Computer.'' NASA TND 1562. 35. Sranltz, J. D.. and Ellis, G. 0.. "Two Dimensional Flow on General Surfaces of Revolution

in Turbomachrner." NACA T N 2654. 36. Wu. C. H., "A General Theory of Three Dimensional Flow in Subsonic and Supersonic

Ti~#bomricl,iner of Axial. Radial and Mixed-Flow Type," NACA T N 2604. 1952. 37. Owc~wiik . J. A.. "Fundamenlals o f Gar Dynamics.'' international Textbook Company,

P~nnrylvawa. 1968. 38. Yuan. S. W., "Foundations o f Fluld Mechanics," Prentice Hall (New Oelhi). 1969. 39. Vauia. M. H.. "Aerothermodynamicr and Flow in Turbomachiner," John Wiley & Sons.

I~K.. N.Y.. 1960. 40. Petrrri iam~. H.. Unterruchunqen am Zenrripetalrad fur Kreirelverdichter Forshung, Vol.

ZAMM 8 . 1 9 2 8 . 1 ~ 460. 45. S c ~ ~ ~ ~ l t ~ - G m o w . F.. "Der E~rbiln$)rwidrrrr;ilId Rotierender Schr:iben in Gehansen" ZAMM

15.1935,1>. 191. 46. Bammrr l . K.. and Rautenberg, M . "On the Energy Transfer in Centrllugai Compresrorr."

ASME Paper No. 74-GT-121 47. Bovce, M. P., "A Plvcricai Three-Dimensional Flow V inmi ra t ion Approach lo the Com-

plex Flow Clrarac<rrirr~cr in u C e n t r ~ f u w i Impeller," ASME Paper No. 66.GT-83. 48. York , R. E., and Woodard, H. S.. "Supeironic Comprerror Cascades - A n Analyrs of the

En1r;vnce Rcgmn Flow Field Containing Detached Shock Waver." ASME Paper No. 75-GT- 33.

49. Drlnmey. R. A,. and Kavanagh. P., '7ra8lronic Flow Analyrn 8 " Axial Flow Turbo- marhlncrv Cascades by a Time-Dependent Method of Charactsrirricr." ASME Paller No. 75-<;TS.

50. tch:%udl, 0.. "lostantnnenus Ml.asu~emcnts in IIW Jet-Wake Dmzhargc Flow ol a Centri- lw.11 Cmn~~icsror impel la^ . ASME Papel No. 74-GT-90.

51. Ki.~rsun. H. A.. "E l f rc t of lnducrc lnler and Diffuser Throat Areas on Performance o f a L i w i)tc%ure t3;8tiu Swewback Ccnlr8lug;ll Comp<enror." NASA T M X-3148, Lewis Re- SL'OICII Center. Jan. 1975.

52. Rodgerr. C.. and Sapi~o, L., "Der~gn Conr8derationr for High Plerrure Ratio Centrifugal Comprersorr." ASME Paper No. 72-GT-91.

53. Bhtnder. F. S., and ingham. D. R.. "The Effect o f Inducer Shape an the Performanceof Higl, P8esrtrre Ratio Centrtfugai Campresrorr."ASME Paper No. 74-GT-122.

54. Sapiro, L.. "Preliminary Staging Selection for GarTurbtne Drlvrn Centrifugal Gar Com- prerrorr," ASME Paper No. 73-GT-31.

55. Mikolaicrak. A. A,. Weingold, H. D.. Nikkanen. J. P., "Flow Through Cascades of Slotted Compressor Blades," Journal of Engineering for Power, ASME Tranr.. Jan. 1970, P. 57.

56. Rodgerr. C.. 'Typical Performance Characterirricr of Gas Turbine Radial Campresrorf." Journal of Engineering for Power. ASME Tranr. Apr. 1964, p. 161.

57. Stahler. A. F.. "Transonic Flow Problems in Centrifugal Compressors," SAE Reprinr 268C. Jan. 1961.

58. Rodgerr. C.. "Influence of Impeller and Diffuser Characteristics and Matching on Radmi Compressor Performance." SAE Reprint 2688. Jan. 1961.

59. Kramer, J. J.. Osborn, W. M. and Hamrick. J. T., "Derign and Test a1 Mixed-Flow anr! Centrifugal Impellerr." Journal of Engineering for Power, ASME Tranr.. Series A,, V o l 82. 1960. P. 127.

60. Pfleiderer. C., "Kreirel ~umpen." Springer, F155. 61. Mechanical Engineers Handbook, "Centrifugal and Axial Fans." 6th Edition, p. 14-66 to

14-79. 62. Balie. 0. E., "Loar and Flow Path Studies on Centrifugal Comprerrarr - Part I & 11,''

ASME Paper No. 70-GT-12-A and 70-GT-17.B. 63. Eckert. 6.. "Axial and Radia1komprerroren;'Springer. 1953. 64. Wierner. F. J., Jr.. "Practical Stage Performance Correlations for Centrifugai Comprerori,"

ASME Paper No. 60-Hyd-17. 65. Ferguron, T. 8.. 'The Centrifugal Comprerror Stage," Butterworth and Co.. L id. , London.

1963. 66. Balie, 0. E.. "A Study of Reynolds Number Effects m Turbomachinery." Jounmf o f

Engineering for Power ASME Trans. Vol. 86. Series A. 1954, p 227. 67. Schlichring. H.."Boundary Layer Theory." McGraw-Hill Book Co.. N.Y., 1968. 68. Moody, L. F., "Friction Factors," ASME Tranr. Nov. 1944, p. 672. 69. Balje. 0. E., "A Study on Design Criteria and Matching o i Turbomachtncr Pal l 13."

Journalof Errginerring forPowes ASMETranr.. Vol. 84. Serer A. 1962. P. 103. 70. Lazarkiewicr. S.. and Trorkolanrki. A. T.. "Impeller Pumpr." Pergamon Prerr, 1965. 71. Kovatr, A,, "Design and Performance o f Centrifugal and Axlsl Flow Pumpr a n d Cwr!

pressow.'' The Macmillan Co., 1964. 72. Stepanoff, A. J . . "Turboblowerr." Wiley, 1955. 73. Batman. J.. "Derign and Derelopment of a Family of Natural Gar Compressorr lor a 3000

h p Gar Turbine Engine." ASME Paper No. 72-GT-10. 74.Speer. I. E., "Design and Development of a Broad Range Htgh Efficiency Cent i i iuW

Compressor lo r a Small Gas Turbine Comprersor Unit. ASME Paper No. 52-SA.14. 75. Flaheny. R.. "A Method for Estimating Turbulent Boundary Layers and Heat Transfer in

an Arbitrary Pressure Gradient." United Aircraft Research Laboratories Report OAR-G51. A u g 1968.

76. Rerhatko. E.. and Tucker, M., "Approximate Calculation of the Comprerribis Turbulrnl Boundary Layer w i th Heat Transfer and Arbitrary Presrure Gradient." NACA Tn 4154, 1957.

77. Schlichting. H., "Application of Boundary-Layer Theory 8" Turbomachinery," ASME Tranr. Vol. 81. 1959. P. 543.

78. Bovce. M. P.. and Bale. Y. S., "Duffusion Lass in a Mixed Flow Comprerror, Paper No 729061, lnterrocietv Energy Conversion Compressor Impeller." NRCC ME-220. Otlnwa. Jul. 1966.

79. Fowler. H. S.. "An lnvertigation of the Flow Procerrer in a Centrifugai Compresror I m peller." NRCC ME-220, Ottawa. Jul. 1966.

60. Bovcr. M. f'., Schllle,. R . N., and Oesa,. A. R., "Study o f Caring Treatn~ent Effectr in 4 x a Flow Compresrorr." ASME Paper No. 74-GT-89.

81. B O ~ C P . M . P.. and Derai. A. R.. "Clearance Lorr in a Cenrrifugal Impeller." Paper No. 1:191?G. IP!oc, of i h r 8th lntetsocetv Energy Conversion Engineering Confetence, Penn-

iv1v.1,ri.i A u i i 1973. 1). 638. ! 8 , M. I,,, ;and Oi.ml, A. Fi.; " A Theoiet~cal A n a l y w o f Non-Isentropic F l ow of a

C m ~ , ~ ~ i ~ : w L ~ l t ! V ~ S C O Y S Gas In Ndrruw l'arragrr," ASME Trans.. Vol. 16. N u m l x r 2. Apr. 1913. 1,. 132.

83. l ' ~ ~ w l ~ ~ ~ , I 1 S.. " E x f x ~ h r r l r r on the Flow Procerrer $ n Slmple Rorating Channel$." NRCC

ME-229. Ouawa, Jan. 1969. 84. Spnoo, Y , and Nakarr, Y . , "An Analyr~s of F low Through a Mixed F low Impellei." ASME

P;ii,rr No. I l - G T - 2 . 65. S v w o , Y., and Nakase. Y ., "A Blade Theory of a n lmpullar wtrh an Arblrrarv Suiface of

Hevolut~on." ASME Paper No. 71-GT-17. 66. K a r r a n s . T., "Urr of Arbltrarv Qua$$-Orthogonalr for Calculating F low Dmr ibu t ion in the

M e ~ i i f i m a l Pl:int. 01 a Turbuinaclxne." NASA TND-2546. 1964. ! : I I>I~,II>, V. I>. anil M8cllrl. O. J.. "An Analysis o f F low rn Rotarrng Parrage o f Large Radial

In11.t C~mwIiig;,l Con~l~rersur a1 Top Speed o f 700 Feet Purrecond." NACA T N No. 2584, 1'451

88. 81.1t. A,. " i~ i l roducuur l lo the T h r m y o f F low Machlnrr." Pergamon Prerr. 1966. 89 . Sov,;in. G., and Klomp, E. 0.. " F l u d Mechi inm o f h te rna l Flow." Elrevier Publishing

Co.. N .Y. 1967. !lo. Ku,,,l~,,, A. M . . i nd Scl?rf/cr. J. O.. ' 'Fouwi~t8onr o f Ai:~odvn;~rmcs." John Wlley &Sons.

N.Y. 1950. ill C ~ I V I P N., ,m?d Sh. i# i . S . W., "A IX~~OXI I~ I I : Mt' i lwtl 101 Pr~tllc!w?g Sepuhr,on Ploperter of

L . i r l l 1 . 1 1 1 i o ~ ~ n i l ; ~ i y I . .~vm~. ' ' A w o Wumr , ti. 257. Auil. 1957. !I? 11.11~.m>, A. ti. / t i i ~ l o l I , " S ~ r n ~ x l s # t m DII Fully Scl)arint!d Flows.'' ASME Fluids Energy.

I . : I 1 ' 1 1 . 1 1 1 I , M y 18 70. 1964. I I s , . 11.. 'TI><, 11>w.t ! O#n!~~nr ro~! :~ i Boiln<l;lry Loyl.r" N A V O R D R~!por l No. 1313,

LV.IIIII,I<IIOII o.<:., 10!>1 9 4 . Shuutn.in A. $1 . a,?d Ariderron. J. 11.. ' 7 h r Us,: 01 Cowxerror Inlet Prewhirl for the

Cntlt80 01 S m d Gar T u ~ h ~ n r r ' ' JUIIIIIII 0 1 Eiiuiiirrii,ly for Power, Trans. ASME Ser. A.. Vol. 86,1964, PP. 136-140.

95. Dean, R. C.. ' T h e Flu id Dynamic Design o f Advanced Centrifugal Compresrorr," Von Karman ln r t tu re , Lecrure Notes No. 50. 1972.

96. Harold Lown and Frank J. Wierner. Jr., "Prediction of Chocking F low i n Centrifugal Impeller." Transaction of ASME Journal of Bast Engineering, March, 1959.

97. Coppage. J. E., er a!. "Study of Supersonic Radial Comprerrorr for Refrigeration and Plerrurlrarion Svrtemr," WADC Technical Report 55-257, Art ia Document No. A D 110467, 1956.

98 Smoo. Y . , PI al, "VSCOUI Effectr on Shp Factor of Centrifugal Blowerr," ASME publica- 111111 /_ iGT.56, 1973.

!l!l. .S.i#l b r , r 'A. &,r TriiLurii E,r!priseiirig H .mibnok -- V o l I . 1 1 . I l l . 100. T l l ~ ~ o l s t x > . W. T., Mcchamcal Vlbrat8onr. Second Editcon. Prenrice-Hall, Inc.. Englewood

Cl8Ils. N.J., 1961. 101. C i t l o l w , L . J.. JI. "ROIOI Bearing S t i ~ b i l , ~ ~ . " P I O C ~ C ~ I ~ ~ S o f ihc F t r ~ l Turbomachinery

S v ~ l v o r i u m , Trxar A&M Unrv. College S i m o n , Texas. October. 1972. 102. t l a , q A. C.. "The Influences o f Oil F i lm Journal Bearings on the Stability of Rotating

M.ic11~nrr." J. o f Appl. Mech. Trans. ASME Vol. 68, p. 21 1. 1946. 1 0 J El>ucli, F . F.. "ldentilicut8on and Avoidence o f Inslab~lrt8er and Self-Exclled Vrbratmnr in

110fatnq M R C ~ ~ I ~ P ? ~ , ' ' Adopted l r om ASME Paver 72-UE-21. Lynn, Mars.. General Elec- t w c Cu., A~,cr ; , f t Enqlnr Group. Gioul, En(l~necr~rqi Dtvwon. May 11, 1972.

104. Alford. J. S., "Protecting Turbomachinery f rom Self-Excited Rotor Whrl." Journal P:

Engineering for Power, ASME Transactions. Ocr. 1965, pp. 333-344. 105..Newkirk, B. L., "Shaft Whipping" General Electric Review, Val. 27, p. 169. 1924. 106. Hourmann, J. G., Turbomachinerv Specifications. Proc. Firrt Turbomachnery Syml~.. PO.

77-78, Texar A&M Univ. College Station. Texar 1972. 107. Davis, H. M., Centrifugal Comprerrar Operation and Matntrnance. Proc. f r r r t Tulbu

machinery Svmp. PP. 10-25, Texar A&M Uoiuers~tv, College Stnrlon. Trxar 1972. 108. Wilcock. D. E. S. and Boorer. E. R., "Bear~ng D rsqnand Ap~,l~cat$on. M c G m w - H i , I+L,;

York, 1961. 109. Gunter, E. J., "Dynamic Srabilitv o f Rotor Bearing Svrtemr. NASA Sp-1 13, 1966 110. Herbage, B. S., High Speed Journal and Thrust Bearing Design. Pioc. F r r r Turlx:

machinery Symp.. PP. 55-61. Texar A & M Univ. College Station. Texar. 1972. 111. Bovce. M. P.. and Hanawa. D. A,. Development of Tcchnwues for M o r > ~ t o r ~ r ~ < i rilsih

machinery, Proc. Gar Turbine Operarionr and Maintenance Symp., Edmonlml, C;jn:3rl,, 1974.

112. Bovce, M. P. and Hanawa. D. A,. Parametric Study of a Gar Turbine, J. Eng Power UI.L 1975.

113. ClaPP. A. M., Fundamentals o f Lutmcationr Relating to Opeiarlon and Maintenance ,>!

Turbomachinery. Prac. Firrt Turbomachinery Svmp. pp. 67--74. Texar A&bl Unlv. C c I s ~ lege Starion. Texar. 1972.

114. Cameron, J. A. and Danowrki. F. M. Some Meta!lurgrcvl Conr~derat~onr ,n C i : ~ i ,w I~~~ : . 8

Comprersorr. Proc. Second Turbomachinery Svmp., PP. 116-128. Texar 4 5 M 1111,.

College Station, T e x a r . 1973. 115. Jackson. C. J.. Cold and Hot Algnment Trchn~quer lor Turbamachncry, P ~ O C S~.CUII,I

Turbomachrnerv Svmp. PP. 1-7 Texar A & M Univerrily, College Starson. Texas 1973 116. Lesiecki. G.. Evaluation o f L iquid F i l m Seals: Asroclared Svrtemr and Procerr C o o s i ~ r ; ~

t ionr Proc. Sixth Turbomachinery Symp. pp. 145-148. Texas A & M Urliv. College SI;~L~,I I , Texar 1977.

117. Lewis. R . A.. Mechunrcal Corllilcl Shah Seal. Plot. Sixth Turbomnch~ni:ry Sumf,, 149--151. Texas A&M Un8vcrr~tv. Collcqe Slslron. T P ~ s . 1977.

118. Boyce. M. P., Morgan. E. and Whltc. G. S,mular~on o f Rolor D y r w l w r 01 l l t ~ i l ~ !;I,,.,I

Rotating Machinery, Proc. First lo t . Conf. Centr!fugol Comprusror T~chno l . , 1'1, tG- : ! : . , Madras, India, 1978.

119. Bovce. M. P.. How to Achieve Online Availability of Centrifugai Comprers~rg, Chemc;il Engineering, pp. 11 5-1 17. June 5. 1978.

CHAPTER 4

THE LIQUID RING VACUUM PUMP

A. M. D A V E Nash Engineering Co.

Norwalk. CT

INTRODUCTION

V d c u ~ m i tcchliolugy is an essential part i n most major industries l ike chemical p~occssiny. f~ l ter l l lg . power generating, pu lp & paper, text i le production. and

i i t i i ~ y olhurs. I n today's market, vacuum producing equipment includes l iqu id ring. ~ ~ : c i j ~ t u i : ~ ~ t i , i q . ~ o t a r y lobe or piston and sliding vane pumps, centrifugal blowers, ~ 1 1 d steiiln jet ejectors.

The liquid ring i)!1171/) consists of a single rotat ing element generally called "rotor." Pumping is accomplished b y means o f a rotating rotor chamber (bucket) c r~ tm i l ! j 31111 l i m i ~ i q a solid ring o f l iquid. The l iquid ring pump is no t new: it has

1)1!t,11 11s1:<1 sui:cessfully since the r u n of thc century. This chapter w i l l provide i l s ' i ~ c t ,,I i!rltm>i.,tio~r o ~ i tl ic t~ i )c l ;~ t ing p> lriciplc, si:lectlon, installation, operation and I,,.III!!,,II.~~I<:,, 0 1 ihi: li,/oicl 111!1 p1111111.

I r<lt>nI 1 1 1 1 ~ 1 p1111>ps .ilc in,mut;ictutcd i n cunicill. l lat sided, and cyl indrical design. It,,. L:LIIIII:,I/, I l i lt si<l<!(t. 01 ~ y l i ~ ~ d r i ~ a l t l ~sc r i t x s the shape o f stationary port. I n a ~.o,i~i:;iI dt,:i~qil, p011s i l ~ c Iocat~ld 011 ii COIIICIII surtace which fits in to matching I.ijwrci I,mr: i i t the lo tor . The unassumbled view of a conical t ype is shown in Fiqitlc 4. I . S ~ l l i l i i ~ l y 111 a f lat sideti design, the por t plate fits close and parallel t o ! u r o coi l sh~cx~ds, which have (matching ports between the blades as illustrated in F i g i l ~ e 4.2. The f la l sided design of a l iquid ring pump is usually less expensive t o

~ni.i~luf,+ctuie than the conical design, b u t the flexible por t design i n a conical type cxl i ibi ts .I low internal pressure loss and capability o f handling extra liquid. Liquid

ring pumps are available i n single or double acting design, as illustrated i n Figure 4.1 ;~nd 4.3. Single acting i s less expensive and usually used for vacuum pumps. The

double acting balances hydraul ic forces across the pump and thus has advantages as a compressor. The double acting completes W o inlet and discharge strokes during (~.ich i ~ w o l u t ~ n ~ , . A l l l i q ~ j i d ring pumps have essentially thesame principle o f opera-

! t i in , 1ty.11~11tss ot lIit?ir design.

PRINCIPLE O F OPERATION

i i > c p011,p consisls of a circular casing and a rotor wi th a series o f blades

p io jeotny f rom a hol low cylindrical hub through which a shaft is pressed. Roro, l~lades are shrouded at the sides t o form chambers (buckets) such as (A).

Figure 4.1. Single acrjng conical design liquid ring pump (Photo couriesy of Ndsh Enyrri-

eering Companyl.

Figure 4.2.

Rotor chambers contain a l iquid compressant, usually water (4). As a result o f

centrifugal force, water follows the contour o f the casing as it rotates wi th the rotor.

Starting at "A," the rotor chamber is fu l l o f water. As the rotor moves c l o ~ k wise, water recedes unt i l the chamber is empty (5) . As the rotor continues fu r t l l f :~ the converging casing forces water back into the rotor chamber unt i l i t is dydln i t i l l at (6). This cycle takes place once each revolution in a single acting pump. ~ITh,:

Figun, 4.3. U,iarrembled view of double aciing liqufd nng compressor.

Figure 4.4. The liquid Nng principle.

operation is similar i n a double acting as shown in Figure 4.5. I t completes two inlet and discharge strokes during each revolution o f the rotor and, therefore, i t includes

j two internal inlet and discharge ports.

I BOD)

SCHEMATIC SECTION A T lNLtT AND DISCHARGF SECTORS

Figure 4.5. Typical cross recriori of doublr acting compressor.

I n a l iquid ring pump, the compression is a direct result o f the energy mparte i i t o the liquid ring as it is thrown out centrifugally by the rotor blades. This energy 1 5

a function of the rotor velocity (RPMI and the specific gravity of the liquid corn.

pressant. A por t ion o f this imparted energy is recovered as i t compresses the gas H I

the discharge cycle. The unrecovered energy is lost in pumping of the seal l iqwc and f lu idf r ic t ion loses. I f energy in seal l iquid is not adequate to )meet the c o ~ r i Pression ratio, the l iquid ring velocity slows down prior t o the discharge p o ~ t opening, and the pump loses its solid liquid piston iring1 in the rotor bucket. T h s creates instability i n the pump due t o collapse of l iquid ring. This is known as

"stall," and is usually accompanied b y an increase in power and noise and reducton in pump capacity.

Liquid ring pumps can be operated as a vacuum pump or as a compressor. In

fact, vacuum pump is actually a compressor, compressing from a sub-atmospheric condition t o atmospheric pressure.

CAPABIL ITY O F L IQUID R ING PUMP

C i e i i e ~ i i l ~ i ~ t i capabilir~es o f the l iquid ring pump is indicated i n Table 4.1. L iquid tinq p!t~iws art? available in capacities LIP t o 14,000 CFM (396 M3/min) as a vacuum pm,i> 0 1 ;is :I i:m,[,rcsso~. The vacu!)m range can be expanded up t o 5-10mm Hg Ails I l y o s ~ ~ ~ q Iuw ~11por-prfssure sea l iquid (o i l or hydraulic f luid). A n air ejector u n 1x1 !,r.lgixl w ~ t l i l iq~t ic l iniig V U ~ I I S fo r V ~ C L I U I ~ levels UP t o 20-25mm Hg Abs. Twu s~~ i< l l t i \ t , i y licltilci r ing pumps ciin be staged in series when there are capacity l~ l~: i t i i t tc~ l is wirh i t i t e g r ~ l t w o stage unit , i f necessary. Typical performance of a l i<l t l i i l rln<l IIIIIII~ is i l lust r t~twl in Figure 4.6.

AI'PIIOXIMATE CAPACITY OF LIQUID R I N G PUMPS

0 A i r c a p a c i t y 50% R . H . S t d . Barometer & 60 F ~ ~ 3 1 Water.

Si!i):lc S tngc

V.!<~ULUII LIP t o 2.0" Hg Abs. C.lpacity up t o 11,500 ACFM (326m3/min)

Two S tage

Vacuuni up t o 0.78" Hg Abs . Capaci ty up t o 11,500 ACFM (326m3/min)

Vncuuiil r ange can be extended by u s e of low vnpor p r e s s u r e l i q u i d ( o i l o r h y d r a u l i c f l u i d ) . Also an a i r e j e c t o r can be s t a g e d w i t h L.R. pump.

Compressor

S i n g l e A c t i n g - P r e s s u r e u t o 35 p s i g 5 (2.5 Kdcm g) c a p a c i t y up t o 14 ,000 CF'M

Double A c t i n g - P r e s s u r e ug t o 125 p s i g (8 .8 Kg/cm g ) c a p a c i t y up t o 300 CFM ( 8 . 5r113/min)

Vol m e

Figure 4.6. Typical pcrlormance curve of liquid nny pump.

OWNER BENEFITS

Liquid ring pumps have only one moving part (rotor) and pumping is accomplished w i thou t pistons, valves, sliding vanes, and wi thout any metallic contact between rotating and stationary element. I n addition, they provide ;1 continu;il

source of vacuum or pressure wi thout pulsation. L iquid ring pumps provide significant advantages over other vacuum proriuc~rn;l

equipment when wet. contaminated, corrosive. and explosive gases are handled

Proper Selection o f a vacuum pump system can reduce environmental poilutior! and energy consumption. Following is an explanation o f some of the owner benefits w i th l iquid ring pumps:

1. Liquid Compressor Cools the Gas - The gases being handled in a liquid rtng pump are cooled while the heat o f compression is absorbed b y the liquid com-

pressant and discharged into a gas l iquid separator. The separator removes liquid from the gas. I t is obvious that the cooling of a l iquid ring compressor is dlrrcr

rather than through the walls o f the casing as i n the other types of compressors. Because of intimate contact o f gas and liquid, the final discharge temperature can

be held close t o the temperature o f the inlet compressant hquid and, thus. l i q i ~ i

ring pumps eliminate the need for after-coolers that may be required wi th o t l i e~

types of compressors. 2. Wet Gas wi th Possible L iquid Carryover - The liquid ring pump can handle

saturated vapor gas mixtures w i th l iquid carryover wi th no adverse effect on pomp performance, while this type of service can be detrimental t o other types o f vacuIfrn pumps.

3. Saturated Vapors - Saturated vapors are conder~srrl I I I 2 l iqu~cl n n y rmrnI>

when l iquid compressant i s coder than the saturated vapor illre to fhc r : o m j ~ r i i \ s i r ~ ~ ~

F~gure 4.9. Overs11 rhermal efficiency of vacuum pumps.

determining the required cost t o produce steam vs k i lowat t usage of an equivalent vacuum pump. Table 4.2 illustrates the comparison of induced heat load between vacuum pump and jet. Table 4.3 relates operating cost per hour of steam jet vs electrical cost o f an equivalent vacuum pump. Table 4.4 shows an average cost comparison for sream at $3.011000 lbs and electricity at 2.5cIKWH. It shows the operating cost of a vacuum pump is approximately one.fifth that o f steam jet ejectors. Generally, the pay-back period w i th an installation o f a l iquid ring vacuum pump can result between 3-20 months.

9. Reduce Pollut ion - I n applications where steam jet ejectors are used t o

pcovide vacuum, o t t e ~ i i t may contaminate motive steam by mix ing w i th a pol luted process steam. Gas is heated and this may result i n thermal pol lut ion. The necessity of tleating contaminated condensate makes l iquid ring pumps an attractive replacement. The m e o f l iqi l id ring vacuum pumps i n Inany applicarions can reduce pol lu- t ion and save the cost o f waste water treatment by closed-loop recirculation of cornpressant l iquid.

Tabie 4.2. 4.3 and 4.4 Cost Comparison o f Steam Jet Ejector vs Liquid R i n y Pump

HERE'S HOW OPERATING COSTS COMPARE

Table 4.2. Steam Jet Eiector vr. Mechanical Vacuum Pump

Vacuu~! leve I requi red 20- i r~. l lg 24-in. l ig 2 L - I ,).I19 2 k - i 1 ~ . ~ , ' ~ !

Steam j e t e jec to rs :

100O psi stuam i n lb /h r 'JLIU I,65C i . li!ll ; , , , , ) I#

Heat load i n Btu/hr . . 924x10~ 1 , 6 9 0 ~ 1 0 ~ 3,380x102 2.!17~1xl\!'

Vacuum Pump:

ACFM required a t vacuum . , . . . . . . 326 543 81 5 1 ,6X

Motor bhp required. . . 20 36 45 I 3 1 Ki lowat ts required. . . 16.6 29.9 37.2 83.5

Heat load i n Btu/hr . . 50x10' 93x10' 114x10~ 2 5 5 x 1 ~ ! '

How Steam and Electrical Costs Relate

Table 4.3. Cost per Hour to Remove 500 Iblhr Dry Air

Steam j e t e jec to rs :

Cost per 1,000 l b 20-in.Hg 24-in.Hg 26-in.Hg 28-in.tlg

Vacuuw puulp:

Cost p e r k w h

1 . 0 ~ SO.17 $0.30 5 0 . 3 7 S O . i:1 1 5 c 0.26 11.45 0 . :5C I . ? ' ; 2.0~ 0.3'1 0.60 0. 7'1 I , i i i

2 . 5 ~ 0.63 0.75 0 . 4 3 > , O J , 3 . 0 ~ U.51 0.90 1.11 ?.:.f,

And How You Can Save With a Vacuum Pump

Teblr 4.4. Costs with Steam @ $311.000 ibr and Electricity @ 2.ScIkWh

2 0 - i n l l g 24-in.Hg 26-in.Hg 28-in.Hg

L q u d ~ i ~ i q pumps have pressure I i rn i ta t io~l up to 125 psig (8.8 ky /cm2g i . I n nit xi^ i i s t ~ o c u s they consume Inore energy than blower and reciprocating type

ijorllris. Ilur less than the stealn elector. A por t ion o f the energy consumed b y a l ~ ~ l c ~ i c l 1111:) p1111l11 is i ~ : c w e ~ e d i l l the omp press ion cycle. The rest of the energy is co~,su~nt!tl u i p~in11111y the seal h q ~ l i d and overcoming f lu id friction. Also, i t requires

.j const.mt w p p l y o f seal l iquid, either on once-thru or recirculated seal basis.

L ~ c l t ~ ~ c i ~ i , l < ) piinips x e considered less efficient than the blower or mechanical p111111>, h ~ l t llli: SLIPCI~OI. gils ha~ ld l iny capabilities o f l iquid ring pumps greatly com-

i,,,~,s.il$, lui the relatively lower efficiency.

OPERATION O F L I Q U I D R l N G PUMP

1. Vs. Steam Jet - The typical performance o f a two stage condensing jet vs a

l , ~ ~ i ~ l I o g t w o stage is i l lust~ated in Figure 4.10. I t indicates significant capacity

~i i f ler i 'nces at YacLlwn below design po in t or 24" Hg Vac. The l iquid ring pump is a c.<mst:lnl volume machine while a jet is designed on a constant weight basis. As a

~ ~ w l t . rhc pro11Ic:ol lo excess leakage can reduce jet capacity significantly. L iquid

~ l n g p i m p performance increases at lower vacuum. Liquid ring pumps show notice.

able capacity difference below 24" Hg Vac. The increased capacity at lower vacuum cuts dawn evacuation t ime b y 213 as compared t o a steam jet. I n an emergency

cundirion, the ttme difference can be of value in either start-up or regaining lost

vacuwn The f o l l ow~ng advantages are summarized:

1 . Automatron. Liquid ring v a c u u r pump systems can be easily automated

com i~a r rd t o special control valve arrangements required for steam jets.

2 Sin,/~/ i f icd Pip inq The l iquld l ing prtmp does no t require high pessure insu-

l.,l,Yl ,~,uIl~:,. 3 . i l l < , ! , , , C~l, lsul". i t l l i l

.I l i l :dln~c<l w;ltul / lo l lu to l l

5 l . 0 ~ 11\1;111.111011 c :0 \1

ti Lmw ~ lu isv p u l l u r ! o ~ ~ .

7 . Briltcr gas handling capability. Saturated vapors can be handled w i thou t pre- cwidonse~

I Two S t a g e L ~ o u i d B l n q Pwnn

Figure 4.10. Typical performance of a two stage liquid ring Dump vr two stag^

condensing steam jef.

The l iquid ring pump can be used for backing up steam jet ejector ax1 stilt

provide substantial savings when compared w i th a multi-stage steam ejector systt:,n

Reliabi l i ty o f equipment is quite important to plant personnel. The liquW r l n q pump has been used successfully for years in condenser exhausttny in power ~ , l ~ r ~ i .

where reliability is of most importance.

2. Vs. Blower and Mechanical Pumps - Liquid ring pumps have pressure I i r n t ; ~ t ion of 125 psig (8.8 kg/cm2) and it is l imited b y the energy o f l iqu id ring. Blower' and mechanical pumps are capable of achieving higher compression ratios and thi,v

are considered more efficient than l iquid r ing pumps.

The l iquid ring pump is non-pulsating, which eliminates vibration p rob le~ i~s t i 1 1 1

the need for mandatory receiver and heavy foundations. Since there is ~ r ! y , i l l .

moving element, rotating wi thout metallic contact, there 1s nothing to we:?,. . ,

align, or to be adjusted as may be required wi th othei types. As discusse,I uadt,.~, l iquid r ing pump does n o t need an after-cooler. I n addition, supcrioi y;is I ~ ; i t ~ r l i i ~ ~

capabilities o f l iquid ring pump greatly cu~nperisate for lower et f ic ie~lcy.

THERMODYNAMIC CHARACTERISTICS OF L IQUID R lNG

PUMP & RELATED GAS LAWS

I n a l iquid ring pump, the cornpressant l iquid and fncominqgas orva(Jo,s I J ~ I , ~ ~ : :

go intimate mixing. Gas reactions can tie ~e r f o rmed . i f required. by ai,l,!i,l,r~,~~,

selection o f l iquid.

COOIIII!~ is w l u o i the most obvious therrnodynalnlc functions of the l iquid ring

I . I t I S i~i<leui i a l ~ q ~ l i d cooled p ~ i m p . 111 t h ~ s type of pump, heat o f compression I 1 1 o i condensation i s removed by direct contact o f compressant liquid, 111mefi>1t' ~ i i i i k i rnom heiit transfer is obtained. This is in contrast t o water jacketed i :~,f~l l~ca~ss<,~s whc!re lhcilt t lmsfers through the cylinder wall is less effective. I n a r ! : i ; ~ t i i type of compressor, it's possible t o have h o t spots w i th in cylinder w;~l l w1ud1 i-.ii> dtico~nposc or polyrneri/e the gas. The l iquid r ing pumpcan obtain iou, i~ isot l ler~nal compression as compared t o adiabatic compression of reci-

;)toi:atin<l w m p s 01 blowers. T I S I i so t l i r rn~a l compression o f single stage l iquid r ing vacuum pump.

Ih~?ill 20001111 H!] ALls t o .itmospheric discharge, gives a temperature rise o f 3OC

uornpaicri t o ~jr icooled adiabatic machines wi th temperature rises o f about lll°C. The isuther~nal compression characteristics makes the l iquid ring pump quite ideal i?11~,1n cool . i ~ r or gas is req~l i red, and eliminates the necessity o f an after-cooler.

CAPACITY OF L I Q U I D RING PUMP

Tlh? sllape o t thu revolving rotating l iquid r ing is dependent upon speed of

rot3t1011, tl le con to^^!.^ o f casing and rotor blades, and the compression work. I t ~lo,.s in<,t dcsc~ibc 2 pcjrtern that can be clearly defined or measured, therefore i t is ,,(nr,ll.isiitiil 111i1t t!it: I ~ q u i t l ring pump cannot be related similarly t o positive dis- I , I I t y ~ x ~ c c i l m ~ c i i t i n g 01 rotary ~ L I I T I ~ S . Gerierally, l iqu id r ing pumps are

~ust tx l .~ctu,il cic!livered capacity. Solnu suppliers test all pumps; others test none .I st.itistictil sanplc o f production. Standard performance curves published b y

~h i~ lo l , i c to ! i? rs are based om air 150%, RH.1 at 760mm Hg Abs barometer. and 1 5 ' ~ tv;itl,~. A i ~ y u,iriatior~ on these corrdit io~is wi l l change the pump performance. The

n ~ . ~ x i m i l ~ n v;icLiwn that car1 be obtairiec b y the l i q i ~ i d ring pump is l imited by the u:lpol ji~c!ssilre of wmplessimt l ~ ( l i l i d .

C a p m t y lot SO~IIC vendors is subject t o a 5% tolerance, whi le for others as high ;is 10%: and horsepower varies f rom 0% to 10% depending upon manufacturer. This

IS .in i r npo~ t imt consideration in evaluating published curve.

PERFORMANCE

The l i i l l ~ ~ w ~ ~ i g cotir!itiwns play a significant lola i n pump selection and have

~ I i ~ f l l l l t ~ ! ~litL'i:ls 011 ,I l i ~ l l l l t l ring ~~~1111p's perforlnallce: 1. Opci..>tin:l Pressurc - 01,tv>,t1ng I,ressu!c ~l6:tcrniines whether sintjle stage, two

I Lrrv v . i l w ~ ~ I C S L I I ~ ' s~:.ll,:tl, 111 wLiter s ~ i ~ l u l p m ~ p s a ~ c req~iired. The operating , ~ c w t , i ~ ! i s vfltv:tt:r! b y ~ h y s ~ c a l propt!vties of l iquid colnpressant (vapor pressure, sp. < ) I . 5p. l l t . d l 1 1 "ls<:osity).

2. Gas Composition - The types of gases and vapors wi th its composition detei~mines i f pump is required t o handle any condensiblevapors. During condensa-

t o n o i vapors, the latent heat o f vaporization is given o f f t o the compressant l iquid.

~ , i s ~ i g its temperature and affecting pump performance.

3. D ry Gas - When dry gas enters the rotor bucket of a liquid ring pump, some

o f the seal l iquid vaporizes and occupies some space, reducing ideal pumping capa-

c i t y . I f compressant water entering the pump is greater than 60°F il5'Ci, actual capacity is less than published curve. Table 4.5 illustrates general seal water correc- t ion factors w i th respect t o various seal temperature for a vacuum pump handling dry air only.

Table4.5. Typical Liquid Ring Vacuum Pumps

I 'I

NOTE: This data applies t o Vacuum Pumps selected for handling relatively dry

air, that is, up t o a maximum of 50% relative humidity, These factors do not apply for pumps handling saturated air and vapor mixtures. (For Reference

On ly ) EXAMPLE: Wanted, a vacuum pump to operate at 22" vac w ~ t h a capacity ot 1000 c fm at 22" vac. Seal water available at 100°F, barometer 30" mercury. Correction factor from table is 1.27. Select pump f rom standard Nash curves for

capacity o f 1270 c fm at 22" vacuum. 4. Wet or Saturated Gas - Liquid ring pumps sealed w i th water act as a heat sink

capable o f cooling and condensing vapors. If seal water in the pump is cooler thmi the incoming vapors, pumping capacities wi l l be greater than standard published curve. As the vapors come in contact wi th cool seal water, i t reduces volurne by condensing vapors before entering the rotor chamber, increasing pumping capacity A typical performance curve (Figure 4.1 1) shows the gain in capacity when satw rated air-water vapor mixture is handled. The net pumping capacity o f a liquid rinil pump increases w i th greater temperature difference between vapor and a seal water.

5 . Liquid Cornpressant Temperature - High seal temperature no t only causes thc incoming gas t o expand, bu t also raises the compressant l iquid vapor pressure, thus decreasing pumping capacity. Conversely, lower temperature reduces volume o f

incoming gas t o allow more weight f low for a given pump. The higher the seal l iquid temperature, the more vaporization occurs and occupies more space i n the bucket. reducing pumping capacity.

6. Solubi l i ty of incoming gases into l iquid compressant. The gassolubility affects

the p l l lnp selection as some o f the dissolved gases i n the liquid ring flashes back a t

the inlet o f the pump, occupying some space i n rotor bucket, reducing its net

lhandl~n(l capacity. Soluble gases, l ike sulfur dioxide, carbon dioxide, and some

,,tlir~i. lh;ivc I ,wn succossfidly handling i n a l iquid ring pvmp w i th proper considera. I i O l l o l c,,,,;,cty loss.

7. Effect o f Al t i tude on Pump Performance - The capacity o f the l iquid ring

pump l e n i a n s constant for a pressure rat io unless the inlet pressure is above atmos-

phruc . Thc uperatlng pressure o f a vacuum pvmp is l i m t e d b y exerting barometer. i t is . i t i so l~ t t ~ i y ilecessa~.y TO convert opeiatirlg pressures glven at alt i tude t o the

r~!lcrence seii level c o ~ ~ d i t ~ o n s us~ng Eq (1) . then compare required capacity w i th

i P l i i P l 1 ( P I ) at sea level : iP21 at altitude

Fiyuic 4, 1 I . increase in capaclry of liquid ring pump handling raruraredmixruie.

Calculation of Vapor Pressure Effect For:

a. Miscible Seal L iquid - when two or more liquids are miscible (example:

methanol and water), the total vapor pressure exerted b y the l ~ q u i d compressant (seal l iquid) fol lows Roulte's Law, and matnematically stated in Eq-5.

PVt = PA X mFa + PB X mF, + ~. Eci-6

b. Immiscible Seal L iquid - when two or more liquids are immiscible lexam~)lc benzene and water), the total vapor pressure exerted by the liquid comuresiijnt 1 3

an algebric sum o f the individual vapor pressure as stated in Eq-7 .

Volume required at the pump inlet i s computed, considering vapol pleisuli. effect at the pump operating tefnperature and assuming gas m x t u r e iollows l~ii.;rl behavior.

Temperature Rise Across The Liquid Ring Pump

I n a l iquid ring pump. ideally assumed, all heat of compression and heat ol

condensation is transmitted t o the l ~ q u i d compressant due to intimate m i x ~ n y o f gdi and liquid. Temperdtur(? rise c;ui h,! compute(i by simijl if ied hcat-mass t r ,~r i \ l i i~ calculations. Typical tulnlmdtl i rc ~ i s c ! acloss the ~ ~ ~ l i r l p with i ju t ,>rlv i : ~ m t l i : n ~ ~ i t ~ w ~ is between 10- 15°F ( 6 8°C). T u r ~ l x r d l t ~ r e c$se across the pu$nl j c a n lx l ~ ~ w t ~ r ~ ~ ~ l IQ increasiny l iquid cmrrprcsialit f low.

The curve presented i n Figure 4.14 provides the data to dcterr r lnr t h ~ . L ~ I ~ O L ~ ~ H I

of free dry air i n saturated air-vapor mixtures following Dalton's Law. As an example: Given 30 lbs o f saturated mixture at 29 in, vacuum and 60°F

how much free d r y air is contained in the mixture? Referring t o curve below rrx. across the line corresponding t o 60°F mixture temperature unti l i t intersects the curve line corresponding t o one inch absolute vacuum (equivalent to 29 in. q;"~gi':

vacuum). Reading down f rom the curve iine we have 0.68 lbs of water vapor per I0 of air. That is t o say, an air-vapor mixture at 6 0 ' ~ and one inch mercury absoluli: vacuum wi l l consist o f one pound o f air and .68 l b of water vapor; therefore $ 1 30 Ib o f saturated mix tu re under these conditions o f temperature and pressure t he~v

wil l be 3011.68 = 17.9 lb o f free dry air.

COMPRESSANT SELECTION

Water i s the m m t common compressant or seal l iquid used due to its ava~l ;d i l~ l '+ and suitable properties. However, other liquid compressants can be used wi th t i l t .

following desirable characteristics.

1. Low Vapor Pressure - The vapor pressure of a l iquid compressant increases

w ~ t h incru~sin!) temperature. An increase o f the vapor pressure reduces available

bucker space and thus reduces net pumping capacity. I f the vapor pressure o f the

compressant is lower than water at operating Dressure and temperature, it increases

pcln~l,in!l c,ip;~city. Capacity gain using oi l as the seal l iquid is illustrated i n Figure

4.15. For opurating pressure lower than 20mm Hg A (0.78" Hg Abs), oi l or hy -

d rad ic f lu id can be used as the compressant. I f the l iquid phase o f vapor condenses

in the l iquid ring pump, proper vapor pressure correction for miscible or immiscible

Iiq~Bids must be made.

2. High Specific Heat - Gcner;,lly i t is assumed that all energy (power) required t o oi,c~,itc ;i l ic l i~c l 11119 ~LIITI,) is ~ t l t i ~ n a t e l y trarlsrnittetl t o the l iquid cornpressant i n

the t o ~ i n of heat. The telnperature rise across the l iquid r ing pump increases wi th

~ l e c w a s ~ ~ ~ ~ ) specific heat of the compressant. The greater the temperature rise, the

yeatc l the vapor pressure o f l iquid compressant, and a corresponding reduction in

caoacity. Therefore, i t is preferable t o select compressant l iquid w i th higher specific

lieat, if possible.

3. Specific Gravity - The l iquid r ing energy i s a function o f ro to r velocity and a speclfic gtavity o f the liqrlid cornpressant. In a vacuum pump, when the specific

! l~ ; i v~ ty n l the seal l iq~r i r i is greater than 1.0, the standard curve capacity remains I HOWUVCI, this COIISLIIT~~S more power than shown in the standard cuwe.

Whi,!~ sji~.i:ific !ql;ivlty is less t l i x ~ 1.0, c;wocity imd vower remains unchanged. bu t

p o ~ > i p spc!?cI i n ~ j s t Ibc i~ lcre<~sed to p o v i d e adequate energy to the less dense com-

~~1c:ss,~111. 1 0 ,i ~OII ,~I~CSSU\ the pc t f o r~na~ l ce changes w i th a specific gravity o f the

Low Vanor Pressure Sealed Pwnp - -

! \

GAS l K E T - 2 ! !

SOLENOlO VALVE

L l W l D a

W R Y

( Q I L l W l O GAS - SEPARATOR - I

LlDUID WTLET

Figure 4.1% Vacuum pump once through seal.

BY-PAS5 VALVE

GAS I NET- -

HD( VALVE % L l O J l O OJTLET

Figure 4. 756. Compressor once through rcsl.

Fii)urc 4. 15. Caoaciiv gain of a liquid ring vacuum pump sealed with /ow ,vapor pressure liquid compressant

INQUIRY INFORMATION SHEET

Engr.

I n q u l ~ y No. Customer D;its Ruc'cl. Due Complete

GAS H A N D L E D ~

A i l Other Mol.Wt. Sp. Gravity Temp. 'FI'C.

CAPACITY REQUIRED -~

Ac:FM [i,~easut cd JI i ~ i l e t l L b s . / H r . S C F M (Corrected t o 14.7 psia

s r ~ l 60°F) Other Saturated ("'01 D r~

OPERATING CONDITIONS ~

~ , ,~111~,~~uus I n t c ~ m t t e n t Cycles iH~

A l t i t u i l ~ ' (F t 'Meters I Ambient Temp (°F/oC) Inlet P iessu~ l "Hg AbS " ~ g Vac Psia P s i g

Othet Min. Max. Inlet Temp. Discharge Pressure (Psia) P s i g -Other Min . Max.

SEAL OR COOLING L IQUID . -

Warel Other Temp "F/OC Autiil:~l)le at Psig. Once thru Recirc P;iriml Reclrc Vapor Pressure P s i a l " ~ ~ Abs a t -'FI0C

Psi:a!"Hg Abs at -'FIQc ~ s i a / " ~ g Abs at 'F/"c ~,,,:<.ll!c G ~ a v i t y L p n c i f i c Heat Mol . Wt.

Viscosity Te~nperature (at ten temperatures)

ELECTRICAL

Availal,le power supply, Volts Phase H z . - Control Circuit Voltage

MOTORS

Enclosure ODP TEFC S e v e r e Du ty E x p l o s i o n Proof - Se~vi<.r? Factor Ir isdation List Special Feature-

CONTROL .~ .

Fu tn~sh t m w,d m o w i t Moil ill Mount & Wire T)csc~ ~bt? I d l y

SPECIAL REQUIREMENTS

blater id of Const: A l l Bronze A l l I ron B r o n z e Fi t ted - Si~mloss - Std. Cast I ron O t h e r $:,-,.h S~,.ds (Sioylc:l idnt lhle) T y p e T a n k Mount -

2. Partial recirculation system - may be employed where liquid compressant I S

i n short supply or desired t o reduce l iquid consumption. Some of the liquid corn pressant is recirculated through a specifically sized orifice wi thout use of heat exchanger. This may result i n a 10-67% saving o f l iquid compressant, dependins upon the temperature of non-recirculated seal and operating vacuum. Partial recir culated seal arrangement is shown in Figure 4.16. Generally i t operates a t highe! temperature than once.thru-seal and may affect pump performance.

GAS DISCHARGE

CHECK VALVE (PI1 V A C W RELIEF

INLET

I GAS LlQUlO SEPARATOR LlOUlC OUTLET

Figure 4.76. Parrral recirc. real for liquid saving.

VENT P-t

ORIFICE SMENOID VALVE

3. Ful l recirculation real system - this type of arrangement is employed whew l iquid compressant is i n short supply and/or the cost o f the liquid comrJress;riil (other than water) necessitates the recirculated seal system. I t is often required 1 c

reduce l iquid contamination or pol lut ion and reduce cost o f waste treatment 111

certain applications. Typical recirculation systems w i th and wi thout recirculat io~l pumps are illustrated in Figure 4.17. The l iquid compressant is cooled through this heat exchanger whereheat o f compression and condensation is removed and the11 recirculated back into the pump. Installation o f a compressor and vacuum punll> w i t h inlet seal sometimes d o n o t require a recirculation pump. However. a vacuum system wi thout a recirculating pump requires a careful piping arrangement I n lower pressure drops.

-Y +I*- - - - ; kJ Y

UNIOI*

L I W I D w

Tank Mounted Un i t -Vacuum Pump

6

This arrangement requires no external heat exchanger or circulating pump. Coii l

ing o f seal l iquid is accomplished by natural heat transfer through the wall of th.: tank, piping, and pump. Typical piping arranyement is shown in Figure 4.18.

R.EvuTIC [R

CECTRCNI C)

INIICATCR

IRA IN VALVE

Figure 4. 773. Full wcirc. real vacuum pump with centrifugal pump.

GAS INLET

GAS L l W l O SEPARATDR I I I .

HEAT EXCHANGER OUTLET

C U F R E S S m BY-PASS VALVE-

(PI 1 0

GAS 0 I S M G E

- PKESSL'li KEL I t i

LEV& VN~VI CONTRa

RIEU1I\TIC I37 ELECTRONIC

L l O J l O CUTLET

*C.W. GAS L l W l D ORAlN VALVE

HEAT M W G E R

Figure 4 .17~. C o m p r e ~ ~ o r (generally reek. pump not req'dl..

M E M VALVE, - --VACU

Figure 4. 78. Tank mounted vacuum pump.

GENERAL NOTES

1) Vacuum and pressure rellef valves suggested for pump and system prmectron

21 Liquid level controls often iecommended to protect pump froln operarlnq dry or ovrr f loodir~g

31 B v ~ a r r control valve ir neccsrsrv when requiaflon o l r v s l e m pressure r e q u i r e d

Water Conservation

Variations or1 recirculated seal systems are numerous. Seal water y I>, . conserved by using pre-used water or by re-using water after pump dischair),: FI,,

example:

1. Feeding pump wi th fresh water and discharging t o another system 2 . Feivijny punlp w i th pre-used water and discharge t o drain or discharge t o

allother sys1e111.

3 . Fe~ul ing seal water in a cascade manner, feeding discharge f rom one pump of<> ;~~wrh t? ! p ~ ~ n l ) .

4. i icciicularirqJ cooling towel warel f rom il pl lrnp t o tower, or recirculating cooling rower water through heat exchanger for cooling circulating seal l iquid.

Typical List of L iqu id Ring Applications Are as Follows:

1 . Vacuurn f i l t rat ion 2. Condenser exhauster 3. Centrifugal pump priming

4. Central vacuum and compressair f o r hospitals and laboratories 5. Deaeration 6. Recarbonation

7. Vapor recovery 8. Process gas compressor 9, Vinyl-chloride monomer or butadiene recovery

10. Chlorine conpressing 1 I. Tail gas coinpressor 12. l ~ , s l r i ~ n ~ c n l : i l l C I I ~ ~ M ~ S S ~ I ~

Some o f the '+ppi~cations are show,, schematically and br ief ly discussed on the tollowing pages.

TI?,. reference chart indicates applications o f l iquid ring pumps i n various indus1,ies.

F ILTRAT ION: In this arrangement, the l iquid ring pumD evacuates the fi ltrate

tecc~ver.The constant, non-pulsating vacuuni i t maintains results in the deposit o f an

wen. ~ m f o l - m filter cake. A sinall compressor supplies air for cake blowing. (Figure 4.19i11

SOLVENT RECOVERY: Vacuum tunible drying draws solvent vapors o f f solids.

I n the vacuur~i pump, solvent vapors come in contact w i th cooled solvent l iquid. The pump itself sewas as a partial condenser. Non-condensibles are separate o u t i ~ i t : d i ; ~~ i i cd l y , a ~ d the recovered solvent is stored for re-use. (Figure 4.19b)

MC)ISTlIRE EXTRACTION: Texti le f d i c containing mosture comes ou t o f a iv.irli IIO\ ;I I I~~ p~isscs O V ~ I ;I slotted tube. The V I I C ~ L ~ ~ put1111 draws moisture, along iv i t l i B c ~ n s ~ d c ~ a b l e amount o f air, f l o w the extractor so as t o maintain a continu- LNIS V,ICLILIII, in the s y s t e ~ n Any textile fibers that colne through i n the mixture I ! I I . Tl icy can be rlisposu~l of when the separator is cleaned ou t

periodically. (Figure 4 . 1 9 ~ ) CONDENSER EXHAUSTING: Maintaining vacuum depends on removal of air

Reference Chart

AIR FOR CAKE

COMPRLSSANl

FILTRATE REMOV

FABRIC

, j ( ~ t i OII~III !jim?s that cannot be condensed. The l iquid ring vacuum pump normally opt?l-,ites in series w ~ t h : an air ejector at condenser vacuumsabove 27" o f mercury. At l i w u ~ v.ic:i~cwns, thc ejector is I~y-pass~111 i ~ ~ ~ t o m a t i c a l l y . and the vacuum pump

I 1 1 y Ihuni thc c o ~ l d r n s e ~ dt i~icrrascd capacity. Automatic changeover l k ' t w ~ w i l i i q .111(1 l l i ~ l d i ) j ~ t ?~ r i t l on is a major bencfi t that the Nash exhausting system

i l l l k 3 ~ s .is cu!~iuari!i l wi th s t ~ m 1 ejectors which 111ust be iidjusted manua l l y (Figure

4 ImI

P I I I M I N G A CUIIL~I~II~~II p1111111 lucated above 11s suction level IS kept ful l of

I I ~ L I I C :itid icady tu puinp on startup. Instead u f depending on a f oo t valve, a V.ICLIIIII~ i s d~r iwr i 111 the pump casitiy. The pr t~n ing system, served b y a vacuum

ixj inlr , is cur~nccted l o each centrifugal pump th ro~ lgh a priming valve. I t is a f loat

Figure 4. 19d

valve - closes when l iquid rises in i t . Several connections may be needed ta i!v;ir.,,

ate air f rom all high points in the centrilugals pump's suctiori eye and voluti: T i . priming valve is tapped at three places for those connections. (F~gure 4 . 1 9 ~ )

In chlorine production, vacuum puml, removes byproduc t hydrogen a r ~ i l i i l v

centrated sulfuric acid is used as the liquid cornpressant in the chlnrmt c ] , , , .

comoressor.

i.:.ij~:.~ DRY i l i COMPRESSED Cle TO 1 lOULiACTlON STORAGE

1 !:YrIC W W t f i . , COMPRFSSOR

r ) and for all practical pur-

IIC,~~.% m tx t <,I l l ~ i ? ~ r , ;ire v1l~ ldt1011 f lee. Sinrlle foilndations can be designed t o sup-

1,o I t1,t- wm!11,1 L X I L I ~ I ~ ~ , : I I ~ iati<I t o stilt local conditions. Foundation layout and rjt,ic,~.il i u is I . i l I .~ I~~)~ l 11~sI1~11:1ioris i l l e 11sli:illy I ~ l l r l i ~ l l i ! d i l l rnr~r iu l i~ct~ l rer 's instruction

i,1.1i1t~.il rl,u ~ u ~ s l . i l / . i l ~ o ~ ~ ,$IN ;~11/1111n~:111 01 l l lc dive IS similal to other balanced.

IOI . I I I I I~ I,IIIIII)II,CIII. l l i let i m ~ l d ~ s c h i i t ~ ! ~ ~ i)ii)iri!i for the I ~ U I I I ~ must be f reeo f any

!,~,itn, IIIIWILYISI, 1 1 ~ ~ 1 1 1 ~ N S U ~n i i sd i ! j~ i lne~ l t o f the clrive, rub internal parts, and

I ! : I s I i I . 011 il nl!w ii l isti~llatioo, iolet screens, clean-

, . iotl i l ~ t ~>ud,cts should lx insliilIi:il l o i11wei71 c;ilryover o f any harmful

l r l l t ' $ ! j l l 111,11I1?1 l l i l o thc p ~ l i l l l l at i l l l t i i l l Stal t l l v . Piping Arrangement Well designed piping Inust provide for the unique character-

Is1Ic.s ill l l lu liclilitl ~ i l l ( l i-)~111111.

Start-up. PLIII,~ C J S C ~ ~ ] ~niust b e :; to 5: fill1 o f seal l iquid. I f underfil led, there is

i i~s i> l f i i . iuot I q w c l t o inate a rlng, and ~ l n i t wi l l be unstable or wi l l seize. Overfil l ing

lc.ids to I looi lhig and broken shaft or rotor blades. S,,III,, ,I i~s~!j~is .ill! var t ~ c a ("12 o'clock") discharge. Theseusually require an

. i iu\ l i . i~ y i . . i s i~ i ! l dl ;?it1 t,, I iIcvcnt floodud stdits.

Shu t r i uw~~ . /\ ~ . i i . ~ ~ o i n ~NIIII~ ~~1111 ~ i i l c t i:hi:ck valve wi l l suck l iquid back in to

1)1111111 i . . i ~ w i ~ ~ . A t ~ g l i t check valve wi l l lhold l i q ~ ~ i c l in upper part o f casing and give

ti,,. c l l , ~ ~ of a l loodetl start-up. Small vacu im breakers are available for installation

011 pt imp ,:.isln!] to cou~ltei-act this undesiiable condit ion.

Operating. Piping between pump discharge and the liquid-vapor separating device

is critrcal. This should be short, direct, and w i thou t excessive elevation. For pump

dos~~lr is w i th lhoi i iontal discharge, the piping is usually 0-12'' above shaft

centerline. Designs w i th vertical (12 o'clock) discharge should likewise minimlie

discharge elevation as discharge pipe floods w i th l iquid on pump shutdown.

Refer t o Figure 4.20.

Excess elevation hurts efficiency and may flood pump at shutdown.

GAS

SEAL LIQUID SHADY PRACTICE

Pump has little excess discharge head to overcome and flooding danger is minimized.

SEAL LIQUID ACCEPTABLE PRACTICE

Figure 4.20b

SHAFT

CHAPTER 5

JET EJECTORS AND CONDENSERS AS

VACUUM PRODUCING DEVICES

ROGER G. MOTE AMETEK

Schrittr & Kowting Div.

Corllwells Heights, PA

INTRODUCTION

R;lpld, ol high-capacitv vacuum pumping o f gases or vapors to 10-"orr is

g~ner.,lly pr i fo lmed most economically by a jet-ejector vacuum system. Mechani-

cally, i t is the simplest of all present day types of vacuum pumps or compressors with theconsequent benefit of minimum maintenance demands. Jet vacuum systems

Ii.tvi, Ilia ;idv;mtagu of very low initial cost as well as low installation cost. Jets

cont~iw 111) lnioving plirts a x ! therefore have no lubrication or oil problems. They do

11~11 i c i lo i~< ' I , x ~ I L I I ~ ~ : ~ ~ close t~ ie~a ,ces . ;]lid they can be made from practically any

i .rl8 I t > > ~ t l i > ~ i , \ s t ~ ~ ~ , t ~ n ~ i l t c ~ !;11. lo i~l)plicarions which ,oquire unusual materials o f con-

s t , uct~w,, !]I.iss 1 0 lh(j7ly col iusivu procms vapors as one example, a jet-ejector may

I h I ;lie I 0 1 u . Those fabrication materials lacking sufficient st!e~,gth of their own, can be used as liners with steel backing.

A je t vacuum system frequently has higher rrtility Isteamlair and water) require- imwnts t l i m conventional rnechariical type vacuum pumps. Also, in the case o f direct

con tx r condens~ng systems, the motive gas and condensing water contact the pro-

cess vapors w ~ t h the result that the effluent may require special treatment i f the

s i i ~ r i o ~ ! vapus should contai l any ob~ectionable conta~ninants. The use of surface

coniiemrrs does liot eliminate the problem but greatly reduces the volume of

t i f f lut~nt to I)< 11e;itcd. F o applications having great fluctuations of suction load at

,i ( ~ v e ! suction plessurr, the use of parallel rlnlts (called dual or multiple element) ;ilr xiv;n,rogeous i l l keeping i!ti l ity consumption to a minimum.

A I.,! iqcctoi i 5 is I i d , I i l l w thermodynamic device, which rl!rc?t,qh ~ ~ ~ ~ ~ ! i o l ~ r ~ f eup;ir,sjorr 111 thr motivt, ,liu7lr:s, plwluces a jut of motive gas ~ m i > v i i ~ j j ~I I sk~pr(sotlic swvd, 1Ii11s c o n ~ c ~ t i n g the static energy of the motive gas ~ , l ~ u u i ~ , to kn i : t~c zmelqy. I i ~ t ~ o d c ~ i ! r l circu!nfr~c!r~t~a!ly i l r o ~ ~ n d this jet, is the

li,l.ilivt,lv sihv n iou\ i~g 5111!;1m of suctio~, v.ipm. This cenI!ally locdtud, hiyh velocity

I L ! ~ <>I inotlvi? g x picks 1111 dnd rnlxes with the slower suction stream as i t passes

!h~uu!ili thi: converging sectioli of the venturi diffuser. I t then enters the throat

st icl~w> o f t l ir diffust!~. completely mixed, at thr? critical (sor!ic) velocity of the mtxlult!. Th,: mixture f l om the t h r i m passus through the diverging section of the

,RESSURE CONNECTION

iUCTlON CONNECTlON A l R OR OTHER GASES!

DISCHARGE CONNECTION (MIXTURE, -_d

Flgure 5 1. Typ&ai jet ejecrorr

- venturi diffuser, and as the cross sectional area increases the velocity decrrdsiii. converting its high kinetic energy into static pressure. The discharge area ir di:r~rl~iwd

t o recover some 90-97% of the kinetic energy in the throat. The pressutc I ! ilv

throat corresponds to the critical value of the discharge pressure of the let. I l u r ~ ~ ~ ,

this conversion o f pressure energy at the inlet to the motive nozzle, to vrlocttv

energy in the throat, to pressure energy at the discharge. work is performed r ? ihr.

form of entrainment, mixing and compression so that the discharge l p r r s i h l l i 8 .

always intermediate to the motive and suction pressures. Critical jets are those that involve a compresston ratio greater than or e q u d i . I

the critical pressure ratio of the suction vapor. For steam (water vapor] t h ~ s rsuo , 1.81 and for air the value is quite close to the same. That is, i f the suction p~r,sti.

is less than 0.55 times the discharge pressure, then the jet i s a crlrxal let. l i 11,

compression ratio is less than this critical pressure ratio, then the ji!t is a i i o ~ ~ c : ~ ! ~ : . , ~

jet.

In the same sense, the motive gas f low through the rior?lf! may t x c r t i : , i ( 8 1

non-critical. However, this f low does not detcrrn~ne whether the jet i s caI1i:il C I I I ~ . , . ~

or non-critical. Usually, this motive gas flow is cr~tical and must be i f the l e t 111<!I! I.

critical.

With critical jets, those designed for compression ratio5 of 4 to 10, the e f f ) ~ ~ t . ~ ! ( : j -

i s greatly impaired i f the motive gas pressure is not a t least sewn tlmer thi: d~,cl~;irq

pressure; efficiency i s greatly improved wher this same ra to s eleven or yir:atl:l. *

With extremely low compressions, the above ratlos are riot iidvantaqiioa:. C I t r . ~ , !

jets can be designed for lower motive pressures but with ri ir iuciid compr i :ww . . ; ! i w

riiwsi si.l!i'!r ;m~: I ( ! ~ / L J I I C : ~ 10 i~ccompl~sh the same total compression. This can be

i:r:w!~~~nic,~lly ddvailtagiioos, especially ic the recycling of low pressure waste steam

ICI m lucc~ c ~ ~ i ~ g y costs ~i generating supply steam. The quantity of waste steam ~ w i ~ ~ ~ ~ i i ~ i l indy lrt, s~,v~!~.d limns r l i ? rc?q~ii~<!nrent o f 10.55 kq/crn2 supply steam with I : I I i ! ~ . LCII!II:I jcls ill,! Inect?ssafiy because the capacity of each

\I.iili, I \ l i)wc!~ wlilll! 1111. i ie i : tc : iw in 8ntlximum compression available, especially in I I.r.1 lwir sI.iw!s, c<!<~uil<!s a glt!att!l iwmber of stages. Further, the increased v d u n i ~ , 0 1 stem ine.uls lalgar condensers. whethur direct contact or surface type, JOCI rnwt, c n o l i ~ ~ g water. Nonetheless, depending upon the application, the volume

t ~ l w,isw S!C~I~II ,~v,~ili~l)It!. ;$ntl am malysls o f costs, critical jets designed for the lower ~m~i~ ivc : ~xiissri,ii 0 1 wiiste ztt:am can be iwt only a cost effective equipment acquisi-

t ~ o i i but one which contributes greater savings as energy costs inevitably rise. A ji't cioctn~ is composed of three basic components - body, nozzle, and dif-

f i i s t , l . Thc body is essen1ially a hollow vessel, whose prime purpose i s t o support the ilu/rliv, the soctlon port, and the diffuser in their proper physical positions. The < . o ~ ~ v i i ~ , t i u , ~ ~ l I~III~VP gas rno~zlc? is of the converging, diverging type and is used to mi,.r!,ult! 'lnd expand the motive gas to a predetermined pressure, volume and vel- < i i : t~v Thc inlet of thc convelging section must be well rounded to obtain a high I I l l I ! f f i c i ! i t . Thf? olll ice is sized 10 give the correct quantity

t11 >I,,.II,, i !su~ l ly c.ilcul:iled st ct ,tical flow. The diverging section allows the motive I l t i i , x ) , . ~ ~ , < i . l i l t1 :111115 1 1 i : cn t~ i~ l ly through the diffuser inlet into the diffuser 1 1 , . I I s 15 .i C O I I Y ~ : ! ~ ~ O , ifiviv(]ing wnlur i with a constant diameter

1 l11 t~ . i i s,-<:l(i,n I1111wtwl t l l c IWII COIIIC;II s c , i : t i ~ ~ l s . Figulc 5.2 shows the relationship I 1l1i:sc II.ISIC C O ~ ~ ~ O I I ~ ~ , I S 2nd tll<'ir co~~mpondi r rg pressure, velocity, and work

vat x n c t c ~ s .

For ~ L J I the, i r fwencc to standard rnomcnclature, configurations, fundamentals, avlc ~ u l i ' ~ 111 HE1 (1) ( I i c i ~ l Exchalge lnst~tule) "Standard for Steam Jet Ejectors"

dlid ASME (21 (American Soc~ety o f Mechanical Engineers) standards. Several ex-

culle~it tcchnic;il papers to supplement certain sections of this text are noted in R C ~ ~ ~ ~ I I ~ I I C I ~ S G ti, 13.

vacuums, o f about one inch of mercury or less, i t is customary to state the vacuum

in "inches of water vacuum or draft."

The three standard methods of measurement are, of course, equivalent w ~ t t l

proper transposition constants. Figure 5.3 shows that a standard atmosphere can bc:

represented by 760 Torr (or m m of mercury), 29.92 inches of mercury (both at

O'C) or 14.696 pounds pet square inch absolute. Equivalent pressures are rcpri rented by horizontal intersects on Figure 5.3.

An example of the importance of utilizing either the absolute pressure or refer- encing the local barometric pressure to absolute units when desfgnlng a vacuurn system can be illustrated as follows:

Two companies, one i n Philadelphia and one in Denver, want equlp. ment t o produce 635 mm of mercury vacuum. The mm mercury barometric pressure in Philadelphia is about 762 mm mercury, so that company requires equipment which will produce a 137 mm mercury abso-

. lute pressure. However. in Denver where local barometric pressures are 635 m m of mercury. the company has unknowingly requested zero suction pressure which i s technically unobtainable.

STANDARD ATMOSPHERE

NOTE- 1MM HG. ASS = 1 TORR

, ABSOLUTE ZERO PRESSURE

SUCTION LOADS

In addition to those parallel pressure measurement systems already described. flow rates are often specified in varying ways. The relation between these several approaches is summarized as follows:

Capacity loads are specified in either weight or volumetric rates (i.e. weight or volume per unlt o f time). For this text weight rates in kilogram per hour (kglhr) will be followed. Other commonly used weight rates are pounds per hour (PPH). pounds per second (lbslsec), tons per day, and so on. Simple conversion constants are available for these various rates from many published sources.

Volumetric rates for incompressible fluids are very similar to weight f low rates in tli:jr convaislorls from volumetric to weight flow iates are simple and direct. HOW- W C I , convcrslon for compressible f lu~ds may often depend on actual operating p ~ r ! s s i i ~ c ;ind tcniperatures. Very often the chemical and allied industries refer to SCfM (st. i~icia~d cuhc f e t i t p c ~ minute). Gases measured in SCFM means gas at s t .$m, l .~~~ i co~~r l i t ions and pri:ssulcs (70°F a x l 14.696 PSIA). If thi! late is listed as ACFM (actual cublc feet per minute) the load bein$ considered is at the actual f luw~ng p~essum and temperature. Note that a volumu flow rate such as CFM (cubic feet oer minute) i s def\nitcly incomplcte: ~t must be further amplified by stating

"actual" or "standard" conditions; a similar amplification must be made for temperature i f other than the 2 1 " ~ which is standard in the jet vacuum industry.

The conversion from either ACFM or SCFM, cubic meters per second, barrels per day, gallons per minute, liters per hour, and so on to kilograms per hour can also be obtained from published conversion charts. Though gas mixtures require a

slightly more complex calculation to determine volumetric to weight rate conver. sion these too can be obtained from marly published conversion charts.

The basis for selecting proper jet vacuuin equipment is determned by the requiredsuction pressure inTorr, and suction loads in weight f low rates of kilograms per hour of dry air equivalenf (DAEI at 21°C. The method of converting we~ght flow rate i n kg/hr to DAE at 2 1 " ~ involves corrections for temperature i f other than 21°C, molecular weight i f other than air, and water vapor inclusion if suction

load i s saturated. Figures 5.4. 5.5 and 5.6 show "Temperature Entrainment Ratio Curve." "Molecular Weight Entraniment Ratio Curve." and "Air and Water Vapor Mixture Data (Dalton's Law)." These curves are extracted from HE1 Standards for Steam Jet Ejectors, Third Edition.

l l i i t l l tht!s~~ Lmlc , ~ n ! j ~ ~ i e o l m ] ~o~d' : tst ; i~ l i l i~~$]s, a ii!vitiw o f operat i~~g principle, c i t l i ? o s t c t 1 . nlay ix! made for the following vdi:uum ploduclng di:viccs:

Stram jot syphori

W31e1 jet rxhar1stt:r Multl-jet and multi-jet spray direct contact condenser Low level eductor and low level multi-jet condenser Sii~glc stage jct-ejector with ail or str!am !motive

Multi-stage jet ejector, both non-condensing and condensing

JET SYPHON

When low initial cost and low available motive pressure are of major importance, t l ~ c j e t syphon offers considerable advantages. This type unit is comrnonly utilized

In sampling, priming, evacuating, and exhausting applications and operates well i f the possibility ot flooding exists. I t is, however, uneconomical for continuous

o v ~ r 3 t i o ~ 1 l~cause of high motive uti l i ty requirements. The jet syphon (Fig. 5.7) is simple in construction with only two components,

body and nozzle. I t can also be supplied in any castable material and is normally

stocked in cast iron, bronze, and stainless steel. Performance ranges, suction capacities, and evacuation curves are shown in

Figure 5.6. Enrrainment rario curves for airmream minrurer.

Figures 5.8. 5.9 and 5.10 for two common nozzle sizes. These curves are based on steam motive. Suction capacities utilizing air as the motive stream can be approxi. mated by using one half o f the listed capacities for steam motive.

WATER JET EXHAUSTER

The water jet exhauster operates on the jet principle utilizing water as l t i i

motive force. Because of i t s condensing capabilities it is ideal for handling mixturi::

of condensibles and small quantities of non-condensible gases. Jet exhaustels ari:

EVACUATION CURVE 1 ' 1 FIG 53 7 STEAM JET EXHAUSTEH

CAP. FACTOR

MOTIVE STEAM CONSUMPTION

SIZE

3/4"

1"

I%',

2"

2%''

30 NOZZLE 3 30 PSIG.

115 NOZZLE @ 60 PSIG.

PRESSURE WATER

DISCHARGE

/ CONVERSION FROM FT. WATER TO 1

DIRECT CONTACT CONDENSERS

Co-Current or Parallel Flow Type Units

The direct contact condenser IS employed in a variety of industr~es as d n

economical means of removing ail. i:xhaust str!am, and other vapors f r r j r n vawlurr!

equipment. I t is found in almost every area of the chemical and process ndusties which use vacuum stills, calandria pans, multiple-effect evaporators, and vacuun: crystalizers. I n addition they have been used for decades in the food industry . (edible oil, milk, and sugar) as well as the distilling, pulp and paper, and refinery industries.

A principle feature of the direct contact condenser is that injection water may be discharged through a tail pipe by gravity, without requiring a pump. Another advantage is its capability of withstanding flooding in the event of priming or Iiqwd entrainment. I n most plants, the vapor exhaust connection of the process vessel

under vacuum is located at considerable elevation above ground level. The use of a direct contact condenser at these higher elevations permits shorter exhaust vapor lines which reduces leakage hazards and vapor line pressure drop, thereby reducilq initial cost.

There are two basic types of direct contact condensers: (1) Cocurrent, orparallel flow, in which the vapor to be condensed enters at thi,

top of the uni t and flows in the same direction as the condensing water, and (2 ) Counter-current, or counrer flow, in which the vapor enters near the bottom

of the equipment and passes upward against the water flow, with the non-conrle~~s,i bles discharging near the top. Injection water is delivered to the condenser ( $ 1 ~ t l i ,

form of a solid head, jets, sprays, water curtains. or a combination theri!i,f. Titis section wil l consider the co-current or parallel flow type unlts. C o u n t e ~ - c i j ~ ~ u ! ~ : type direct contact condensers will be discussed in the multi-stage jet ejector, cor l

densing system section 182.). Direct contact condensers o f rhe co-current type can be further cateyorr/ed .$i

multi-jet. multi-jet spray, or multi-spray type units. These are depicted i n F>~gutiis 5.14, 5.15 and 5.16 respectively. The denotation of "jet" in the classiOcat~o~ indicates that the condenser is capable of handling some non-condensbles, whili' performing its primary condensing function. The denotation of "spray" in dri:i;r contact co-current condensers indicates l i t t le non-condensible handling capabihry

The performance of any condenser is described by a simple heat balance. Thi: heat added to the system is the quantity of sream being condensed muit ipl lrd 1,"

the latent heat of vaporization at that pressure and temperature. This must be u q u ~ i to the heat removed by the condensing water which is the quantity of warel multiplied by the temperature rise from inlet to outlet times the specific heat.

It is obvious that the larger the allowable temperature rise of the cordensr~~j water, the smaller the amount of water required and the higher the contlensalc discharge temperature. Under theoretically perfect conditions, a condenser could operate undera vacuum corresponding to its tail and discharge water femperarure. but no higher.

Under normal operation this, of course, can never occur. Air entering w ~ r h the ' injection water and non-condensibles entering with the vapor load exert a parral pressure. The operating pressure of the condenser then, is the sum o f the vapo

Figi,nr 5. 74. Typical multi-,el direci contact condenser,

LIICSSIIII~ . i t tlic tail tempt:rature plus the partial pressure of the "on-condensibles

prcsent. The difference between the temperature of the tail water and the tempera-

ture corresponding to saturated water vapor at the actual condenser pressure i s kllown as the "terminal diffeience."The efficiency of a condenser can be measured

by ti>? ''tc>min;,i d i i e ~ u n c e , " aithough i t also varies with the percentage o f non-

~ : ~ > ~ l c l ~ ~ ~ l s i l ~ l ~ ~ s p r c s e ~ , l .

0 k u 1 . 1 1 1 0 1 , \ ) I I l i t f ~ p c allit ~nitll l i-jct S ~ I C I Y tvpc diwct co~,tact con-

~lc,llst,li IS Y L ' I ~ swn~ i i$~ . COI I~~ ! I>S I I?~ wiit? is d ~ I ~ v i : ~ c d i l l to ll ic <nozzle case which i s dcir~!]nrC 10 oporatc usitrg :I spccilicd qua>rtity of water at stated head pressure

(n~~ i r imu i i i 0.7 kylcm' ~ I i f l i ~ ~ e n t ~ i t l ) , imd ;, !jvirii vacuum in the condenser. The

water jets are directed into the tail-piece at the lower end of the body, where they

unite to f o ~ m a single stream. Vapors entering the condenser come into direct

cont,ict ~ v i t h thc c o ~ i ~ e r ~ ~ i t i $ ~ water and ate condensed. The multi-jet and multi-jet

Figure 5.75. Typicd rnulr;+r spray direcr conracl

condenser.

spray type condensers are capable of flushing a certain amount of non-condcnsiblv

down the discharge pipe. The quantity is relatively small and varies with the opeia-

ting pressure of the condenser. Conversely, excess non-condensibles above the quan-

t i ty which the condenser is capable of handling wil l cause a rise in operating

pressure t o a point sufficient to flush the added non-condensibles through to dis-

charge. This increase in operating pressure above the design point i s undesirable for

several reasons including the possibility of overloading the supporting jet stages or

an objectionable pressure rise in the vessel being evacuated. The air handling cap,+ bilities of the multi.jet and multi-jet spray type condensers are shown in Figurr:

5.17. This graph denotes typical capacities over the air preienr in normal conli~irj

water (which i s illustratud in Figure 5.18). Both type units are capable of terminal differences of 6.1 to 7.2"C. Ther co:~.

densible vapor handling capabilities and maximum cooling water capacfties arc:

shown in Figure 5.19.

Air Pump

Fiyurrr 5. 16. T y p i c ~ l muiv-spray diircr cmrarcr condenser wrrh p r e ~ o o i e r andairpump.

The multi-jet splay unit differs from the multi-jet design in its flexibility of crpur;tt~on. This i s apparent from its design as seen in Figures 5.14 and 5.15. For full

vapoi load the rated water capaclty is l3assed through both the spray and jet nozrlcs. I f the vapor lo;id or water ternpeiature ducreascs, it is possible to throttle

tliu vv.ltc~ to the splay nozzles a ~ i d i~ l t~mate ly turn them off completely. In the I~t1i . l c.isc, the condenser I S operating in a (manner similar to the multi-jet type, but

wtI1 d mir>lmum of inicctlon water undur tlie given conditions.

The multi-spiay direct contact condenser was developed to handle applications

~nvolving l~mi ted water supply, high water temperatures in relation to vacuum

lequirement, or when removal of a large volume of non-condensibles is required.

This type condenser requires an alr pumD to draw the non-condensibles of f the

CAPACITY - PPH AIR PER 100 GPM WATER THRU JET NOZZLES

<' \

MULTI-JET AND MULTI.JET SPRAY (CORREC.TED FOR 0 L0S WATER VAPOR PRESSURE)

ATM PRES = 30" HG ABS

(CAPACITY FOR LOW-LEVEL MULTI-JET CONDENSERS EQUALS X VALUES SHOWN ON CURVES 1 I I I I l l I I l l l l l l l I

EFFECTIVE AIR PRESSURE = INS HG (CHANGING PRES - VAPOR PRES AT H,O DISCH il Mi'

Figure 5.17. Nonsondenrible /aid handling capxity of multi-jet and muiii-jer sproy dliccr c u : ~

t s t condensers. /Note: C~pscirier .shown art. in addition io air prcsrrrr ir? c u o i i ~ w.~rr,rl

Figure 5.18. Allowance ra be made for a r p r e r m r in condensmy water

0 10 i 5 20 25 30 35 40 45

LB PER HOUR VAPOR5 PER I GPM

: I I 1 I I . AS C ~ I I L x si!t!~l ir, Figure 5.16 the multi-spray co~liic~nst!! Ii;ls a ci lcrl la~ spray nozzle aclangemerit at the vapor inlet. I n this respect it 1s vaiy similar to a multi-jet spray type unit, however, i t does not have a jet spray seC1i011. Vapov elitcrs the condenser at the top and mixes thoroughly with the

. ~~ i iec t i on water which is delivered through spray nozzles. The downward action of these co~iverging sprays tends to create a suction in addition to their condensing action. The condensed vapors are taken to the hotwell through the barometric leg. The noti-condensibles are drawn through an air suction chamber to a small direct-

contact counter-current pre-cooler attached to the multi-spray condenser, The p je

cooler lowers the temperature of the air-vapor mixture and condenses as much of the remaining vapor as the water temperature and operating pressure allows. Sinci. the pre-cooler is using its own fresh water and only has to hanclle a relatively small amount of condensibles, i t can reduce the water vapor carry~over to the air pump to a point much lower than is possible with the larger condenser. The spray noirlus arc,

designed for a minimum 0.7 kg/cm2 throttling pressure with terminal differences of 1.7 to 2.8'C obtainable i n a well designed "tight" system. I t is also possible to

operate a multi-spray direct contact condenser without an air pump for short periods of time at slightly reduced vacua. Performance and maximum water capacities for the multi-spray condenser are shown i n Figure 5.20.

Figure 5.20. Performance curve for wprcal multi-spray dnect contact condenser wiih rnrrx~ni,,,,

water capaciry chart.

Maximum Maxarnurn Water Water

Size Capacity Size Capacity No. w r n No. w m 30 625 36 2200 3 1 750 37 3200 32 950 38 3800 33 1100 39 5000 34 1300 40 6000 35 1700 4 1 a000

42 9000

LOW-LEVEL EDUCTOR AND MULTI-JET EDUCTOR TYPE CONDENSERS

Tht, low-level eductoi and multi-jet condensers are des~gned to produce medium

tu l i g l ~ v.icu,i (735 Torr or higher) depend~ily upon water temperature, for use with

I I tw l j i~ws. Thcy B I ~ P I O ~ the energy of water issuing through jet

li,li,l,,r I&, c . ~ I . . ~ t ~ , YLII:LIIIIII. COII~~:IISL! stt:am, al,d handle small amounts of air lap-

in,,x~o).t~,,ly 20:: 0 1 i l ia , l i t lhatiiliing C~p:ll>illly <lenoif!d in Figure 5.171. No auxil- ,,it y v , i i : t , ~ ~ i n I I U I ~ , ~ 1s I , C Y : C S > ~ I ~ ;III<I tl,~, w:itci~ <lisi:I~a~q~:s dircclly i l l to a hotw(!lI with

n i l i ~ ~ l ~ i i ~ ~ : ~ v t , ~ l t 0 1 a L ) J I U ~ B I I C kg.

Tllr lc,w Itwe1 idoctor coi,ilr~iser (F~gc~ru 5.21 1 is capaljle of hancllirq capacities L I P 10 311 10 Ihqlh~ o l stodn, 31 G58 Tort vxu i i l n arlrl 21°C injectioir tt!mpel-ature.

Thi* lbocly o f the condenser is a closed cylindr~cill chamber with a water nozzle in

\ t i , ! p tmU whlch ext~nds downward into a combining tube, and discharges t i ~ ~ r i ~ i g h urntort tall piece. A short pipe ( two feet in length1 carries the discharge

,,>to i h~ : IIOIWPII. Exl)8ust steam enters the condenser through the side inlet in the

l i ~ l y , p~ssos into the combirling tube through holes in the tube, comes in contact wltl, thii w<<ter jet and is co~idensed. Air aod non-condensibles are entrained and

ilsi:lio~gcd with the condensed steam directly to the hotwell. Injection water must

bc dt'I~v~!l.ed dt a co,tstant 0.7 kg/cm' pressure and a water check valve should be

i ~ ~ s t : i l l ~ ~ r l i l l the steam line to pevent water back flow. A typical performance curve

.iil,i w.lter ccipacity chart are shown in F i g ~ ~ r e 5.22. The, low-ltw:l mdt i - jet co~idense~ (Figure 5.23) is very similar in operation to the

I,,w I,,vul ~!i lkjcio~ onit. I t is capable of handling capacities up to 69,000 kg/hr o f

stt,.iln ;it G58 Torr vacuum and 21°C ~niection temperature. The multi-jet condenser

I s d dusign features that are untque. I t has two steam inlets, one at the top.

,izlrl <ma dt the side, either one of which can be used. It is provided with an internal

i l i~. i t i>pu;~ied wcuum breaker to adn i t air to the vacuum space when there is a

L ~ c k !low o f water from the hotwell caused by any interruption or shut.off of

~ n j t x r o i ~ water. Larger unlts, 1132.511sec and higher), are normally equipped with

two v;icoum breakels. This unit's performance curve and water capacity chart are

show,, m Figure 5.24. A tyilical arrangement of a low-level multi-jet condenser under a turbine i s

r h o w ~ i 1 3 Fgore 5.25. In opetaton, the t u r b i ~ ~ e i s sta~ted with atmospheric exhaust.

A r s,,i,r, .ir the injection watel porno is p ~ ~ m e d anti running at normal speed, the v.iIvi3 $ 7 ihc pwnp discharge line is opened, ndmilting watel to the condenser. As the

, I 1 I j , q i , s t i ] coi~<lensc, stt:.iln. ll7o .~t~nosplleric rc:lief valve closes auto-

o ~ , i i ~ c d / y . , id V~CLIIIIII Lw11ds i l l the c o ~ l d ~ ' ~ i s e ~

I ~ i l ~ ! c t i ~ , r ~ w.itui ptesstrrv shot,ld b r 0.7 kg/crn2 or over ro create vacuum. After

111, v,ii:oiun 1 5 (OIII~D(~, thi! willt'l ilol7Ics a l ? tii:sigl~~!d ti) p i m the required f low at

r p ~ w l e i i v,,cutwn w ~ t l i 0.35 kg lcn~ ' piesswc?. Vaci~um in the condenser will not l ~ l d 10 th;ii ~ e q u r d for efficient operatlor, if thu injection pump is not properly

i d , f f ~ l c l i o n losscs dnri l i f t ~ n q l,cxJ ihiivc 1101 Own cillculaled properly, or i f

iittl,ioi, r lu l l tng boxcs or rxhaiist str:am con~~mt ions are not reasonably tiyht. Ex.

Figure 5.27. Twyplcal l o w level educroi

condenser

cessive air leakage i s indicated by turbulence in the hotweli. The water nozr1i.i

should also be protected from foreign matrer by using a screen or strai17er i r l t l l i '

water line between the pump and the condenser.

SINGLE STAGE JET EJECTOR

The general principles of operation for the single srage jet-ejector were diicusit.rJ

in an earlier segment o f this chapter. The jet-ejector, it may be i-ecalliiii, ,, ., thermodynamic device which utilizes the transfer of energy produced by a m o t \ , ,

gas to create vacua and compress the suctron load to an intermediate pressuri:, i h , .

enthalpv.entropv Path of the motive gas (steam in this case) passing throuqn ,, single stage unit is illustrated i r ~ Figure 5.26 by polnts 0 .1 - l ' ~2 . The amount of wr l ik

or energy available for compressing the suction fluid is rhr difference hetwi.i:,~ l t h

motive fluid-velocity energy h, --h, and the eneryy requlred to compress lhi: ,riijlI,,.

EFFICIENCY %

Flyure 5.30. Typ8cal sffrcr of steam rates on stage performance.

The corresponding jet efficiency curves, Figure 5.31,for thevarious nozzle pos- tion curves shown in Figure 5.29, illustrate that maximum efficiency i s ach~eved a: only one point, nearthe knee in the curve. In addition, i t also reveals that the jet efficiency is best at some nozzle position midway between the extremities of the operation. The information in Figures 5.29, 5.30 and 5.31 point out that although the air-capacity, suction, and discharge-pressure character~stics can be obtained by numerous combinations of nozzle position and steam rare, there will be only onc

combination that wil l produce the desired characteristic for the minimum amowit of steam.

JET E F F I C I E N C Y

Figure 5.31. Typ;caleffectofnorrleporrtion upon jet efficiency of onesrage.

From this unit selection i t i s now possible to establish a set of performairci: curves based on various motive pressures similar to that illustrated in Figure 5.27. Figure 5.32 shows jet ejector performance for varying motive pressures between 6.3 kg/cm2 and 14.0 kg/cm2. Suction load capacities increasewith increasing motlvu pressure in a curve which is typical for all jet-ejectors. Seiection of a particuiar u n r to meet specific operating criteria is made between points 0 and 1 in Figure 5.32. Movement of this flat portion of curve to either the right or left is possible through variations in expansion in both the nozzle and inlet taper to the diffuser. T h i movement corresponds to increased suction capacity handling capabilities at a glvtirt

suction pressure. Figure 5.33 illustrates a single stage jet-ejector with wqh!

Figure 5.34. Typical performance for S-3 unit from Fig. 7.7 wwirh varying rream

motive presrurer.

expansion changesin the nozzle and diffuser. As i s shown, this allows sulecrfon r,wr

a wide range of both suction Pressure and suction capacitie,. Each on,? of the eight units would then have an individual performance curve similar to that illustrated iri Figure 5.32. Figure 5.34 is one such curve for a single stage jet-ejector with steam motive.

The jet-ejector with air motive has the same operating and performance charac~ teristics as the steam operated jet-ejector. In some instances, particularly where heating and diluting are undesirable, or steam and/or water availability I S mlrlw mized, the air motive jet-ejector offers definite advantages. It must be rememberi:il, however, that air motive increases the non-condensible load under which the je t -

ejector is operating. In addition, i t may also increase size requirements on successlw?

jet-ejectors and condensers. In general the air motive jet-ejector requires a highii, rate of motive flow than a steam operated unit to handle the same load. Flgure 5.35 shows the performance o f an air operated jet-ejector using 3.16 kylcm' motive r l ~ .

as compared to the performance of a steam operated jet-ejector using 3.16 kgicm' motive. Note that the required motive rate for air is 261 kylhr versus 190 kglhr f m

steam. The percentage motive rate difference between air and steam motiviis 1 4 )

creases as motive pressure increases in single stage units arid rcquiri: a coni~ll i !~;rI,b.

increase ( in the range of 200-300%1 for multwstaye units.

MULTI-STAGE JET EJECTORS

Non-Condensing Type Multi-Stage Jet Ejectors

Wl?r~ l con,bi~?ing stages the supporting units can be designed so that the system 8s i:o~n,~ll;tlily s ta~~ ie owr the iwltire pr i~na iy stage performance curve as illustrated

by inus B aild 0 in Figure 5.37. The suction pressure-capacity characteristics, D, of the supporting stage are rarely identical to the discharge pressure characteristic 6' of

t l ~ . prim;luy stqe, 13 As a iesult, at the design point a supporting stage for a

co~nplrtely supported 2nd stdble system will us~~a l l y provide considerably more

capacity t h m is necessary, as indicated by area between lines 6' and D in Figure 5.37. Whri, i t is desired to obtain minimbm steam consumption and lower equip- rne.ll1 cost. iI,c slt:Jn, consumptlo~, iwd sire of the supporting stages can be reduced

1," i o p p i > ~ ~ : , g fllc ~ I ~ I T I J I ~ stwe in the region of the design point only, as illustrated

q I ,I t i . Tllir p ~ o d u c ~ s .> syrtun wl~ ich 1s s13bI~! at the design point, and It.), .i r r ~ ~ , t ~ I y . u r C ruplodilcjblc? sucrjoo prc,srulo at con<litions other than the design

I , ~ ) I ~ ~ I t F ~ l l tiit,! t -d i ,c t~~)os I,, sys tw~ S ~ L ' , I I ~ ams~ i inp t io~ i w e solnrtimes possible by

t i i l l ! ) sr,ig~'s wlucli I I~VP wisti~hli: shut-of1 p ~ i i l t s ilrc p~~iv iously illustrated by Curve

ij ,I, Ftgwi . 5.29. S~lch systems wil l be stable at the design point, but unstable and

ililsteiidy at the shut-off point.

Assumlng ,no rlgorous limitations on design variables -steam consumption, unit s i z e , number of stages required, equipment cost - systems can he designed to meet

special requirements. It could be advantageous, for example, ro design a i vs tc~ r~ with (a) two design i ~ a d points at different suctlon pressures, o r ( t i ) vacuulli

capacity at 150 percent or greater than the design polnt; or ic) the capabI\ty oi

evacuating vessels to a given pressure within a specified time limitation. All of t h ~

aforementioned systems can be optimum designed systerns with the ndividua 1'1

ejectors sized and arranged to operate at or near the maximum efficiency po~tlts. The general trend in this country i s to use compression ratios up to apploxl~

mately 10 to 1 for the primary and supporting non-condenslny stages. As ~nillcalwd

in Figure 5.28 higher stage efficiencies and consequenrly lower system stzam co l '~

rumption can be obtained by using stage compr6:ssion ratlos in the rarlgr of 4 : l t i )

6:l. I n sections of the world where water is scarce, users ofteh (itid i t rno!!:

economical t o use the lower compression ratios. This requires usng one or two

more stages than is general practice in this country.

Figurer 5.38 through 5.40 illustrate typical vacuum producing capacities and

1 . I ... L _ - - i . M 75 100 125 1%

DESIGN L O A O %

Fiyrirr. 5.37. Tvpic.rl combmalion o f rwo noncondensing stager.

5 -

", z* -

1 VI V)

L 0 * C'

T A W E 0 - - 0 10 20 30 40 50 60 70 80

AIR HANDLING CAWCITY l n R Y A I R A T 7 O F l

Figtin, 5.38 Pwformance curve o f two stage non-condensing multi-stage jet

t o I h tvpe T2 high virciiiim unit and S-3low vacuum unit wi th90 : '5 ,. ,? >5"P?tC ,,,,> ( j , , , ,

20 40 60 80 100 120 140 160 180 21 AIR HANDLING CAWCITY

( D R Y A I P A T 7 0 F l

Figure 5 3 9 . Performance curve for two stage iion-condenring rysnvn s i as F!g. i i 1.2

except type TC2 high vacuum unir uied.

corresponding steam consumptions of various multi-stage non:condensng slc!an,

jet-ejector systems. The curves shown are for systems that are stable in thi? qi:rwral

region of the design point and steady down to the shut-off point. Bi!cauw ( r f

differences in the shapes and character~stics o f the individual stage-pi:rforjn;$nii: curves, and the manufacturing necessity to construct units t o standard s ~ i c ' s , thi:~,!

are some sets o f conditions o f capacity and suction pressure w h ~ c h c a ~ t ~ i o t lh

accomplished at the rates shown b y the generalized curves. A t the same time tht :c

are also sets o f conditions where steam~consumption rates bettter than thosi! show11

are possible.

Direct Contact Condensing Type Multi-Stage Ejector Systems

Bccause there are many various methods o f arranging multi-stayr ~ j u c t ~ , ~ s , I ~ I .

t icularly when used w i th inter-stage condensers, variations in descriptive i~orni:rtr:I;l

ture may also be encountered. Some users and manufacturers count stages f l o m th,: process side t o atmosphere, others from the atmospheric side t o the process. S n r x

nomenclature is critical i n discussing jet-ejector systems, the Heat Exchariyer l n i l n

tute developed a standardized indentification system which is illustrated in Figurii

5.41. The stage discharging t o atmosphere is considered the "2" stage w t h p rw

ceeding stages denoted alphabetically backwards t o the proct?ss. Intercondi!l~s~:ri 21,:

[ I ~ e c t contact type condensers used in jet-ejector systems are of the counter-

ci111~:11t, 01 caunti!r-flow type. Injection water enters the condenser through a water , io ) i l r 21 ti>,, top of the unit. A distribution tray, weir, or disk and donut arrange- inlent L O tho shell p,ovidr a "water curtain" through which the vapor must pass. Most 0 1 thr v;rl,or r~ l ter ing through a side inlet is condensed in the lower part of the s l l l , i~id 111~1 I~OI~-CO~I~L: I~S~I) /~S arc then req~lired to travel upward through the watei c o \ l ~ i < i , . A baffle arlilngement is normally provided at the air suction connec~

tion to ,educe the carry-over of water that may be entrained as the non-condi:nr~blos pass through the condenser.

The d~ruct contact, counter-current condenser operates with a terminal difference of 1.7 to 2 . 8 ' ~ between tail water and vapor dew point temperature. Large condensers will cool the non-condensibles to within 2.8'C of the water tempera- tun! I n cdsrs where large percentages o f non-condeosibles are present, however, t,30ni~;i l <iillwcnces ;is high as 16 to 28°C may be expected since the condenser is ~c tua l ly parforinng as a gas cooler. This is particularly tl.ue in multi-stage jet-ejector

condensing systenis. F I ~ L I I C 5 4 2 illustiiites a counter-current condenser with a distribution tray and

F151111,: 5.43 SIIOWS o co~~dcr i se with a disk and donut arrangement.

The performance characteristics of a two stage condensing system utilizing the same jet-ejectors shown in Figure 5.38 and 5.40 are illustrated in Figures 5.44 through 5.46. As indicated in the curves, a two stage non-condensing system handling 29 kg/hr dry air (at 2l0C1 and 89 Torr requires a 2" by 3" jet system using 228 kgihr steam at 6.3 kgicmZ or 199 kglhr steam at 10.55 kg/cm2. A t the same operating condition, and using a two stage ejector system with an intercondenser, i t

can be noted that only a 2" by 2" jet system is required. Also note that the motive steam required is lowered to 125 kglhr at 6.3 kglcm2 or 111 kglhr at 10.55 kg/cm2. The intercondenser requires .88 Ilsec at 2 4 " ~ .

Figure 5.44. Performance curve of fwo srage condensing multi-sfage jet ejector system wirh T2 high vacuum unir and S3 low vscuum unir with direct conracf inrercondenser.

It then becomes obvious that the choice of system to be utilized is dependent on many variables. Duration of operation, auailability of motive steam, availability of

water, water temperature, initial cost, cost of steam and water, etc., are all requlred to properly define the system best suited for any application.

MULTI-STAGE SYSTEMS WITH SURFACE CONDENSERS

Under some conditions, the use of surface condensers provides certain, d c f r ~ r , : advantages over the use of direct contact condensers.

2 B ! I" t , 1 101 101 ,

~ - 119 i l l , I . . ' ,,,,

; 4 1 "I I "" I" ; "*i I .~~~ l i b 7LI ? d l ; ,a:,

F y i i i , S d 5 . I'lvliirrn.riice cvrn . lor rwo rr.rqe condr~ruiw rvsc~r,? (Note: Svrrenr i s ido?ricill r c i > y r r w i i i r Fvi, 8 2 . 3 ~ r c w r Tvp,: TCZ high wcooin jm clccror used.

1 1 , tho c,lx 01 ilbjcctionable conrainln.mts, where the condensate effluent must

l x tc.;xt~% lx lu l t , ICIIS~! or i i i s ~ h a ~ g ~ d to the sawcr, the vdume of effluent i s greatly ~t i t loc~i i l by tlw LISC of si~rfacc co~ldensets since the cooling water and condensate I

st,c.ims .jut! lhtild repiil.alu as compared to mixing o f the two as in direct contact

cuiidenrerr.

h i the case of the recovery of process chemicals, such as solvents, the surface

co~>denser yields less total volume of condensate to separate or refine.

Separation of the cooling and condensate streams by surface condensers provides

not only clean condensate but ancillary conservation opportunities. In power plants,

tor exa~nple, i t i s common to use the turbine condensate as cooling water in the air

ejrctol cwlde~lsrls before i t is returned to the feed water system. This conserves both l I ~ ~ ~ h i ! . ~ l fh<,ln thostr;mr used i l l the i3ir ejector as well as pure water f i t for boiler use. > 1 1 , o, ; i i 1 1 , . syslwnr wl~;r:h u s e SE;~ water for coolirlq. surface condensers keep the con-

I , I thc se;] watcf 411d ,lllow the reuse of the condensate, which

\voulil ,lot be pussibe i f cllrect cont,~ct cmndense~s were ilsed.

Slit l;iw type corldcnsrrs used in conjunction with multi-stage jet-ejector systems

a l e identiiicd in the same manner as direct contact type units. That is, their loca-

t lo~ ' 1 1 the system will identify them as the "XY." "YZ," etc. condenser. The Heat

Exchange Institute and Tubular Exchanger Manufacturers Association have both

Figure 5.46. Steam morive requirements and ware7 requiremenrr for two rroge condensing system shown in Fig. 8.2.4. (Nore: Number preceding TC-2 and S-3 is size of rhe ejector ,,i

inches.

further described this type condenser with alphabetical head and shell designations.

That designation plus its shell diameter, tube diameter, and length wil l compr!ti:y

define the condenser.

Surface type condenser size i s based on requ~red surface area to acheve p~ok,t:l heat transfer. First the heat load is determined by the amount of desupe~hcal~ng,

condensing, and condensate subcooling required. This i s most accurately calculari:ri

by determining the total enthalpy o f the gaseshapors at thc ~nlliit of thr: cond<:~tii,~ and subtracting the sum of the enthalpy of both vent garcs and the conrlrrisst,. drains.

S u 1 . i ~ ~ : 1y1x co?do~lsi:!s art: In reality, and are therefore correctly named, partial

co~idu~is~! is for two reasons. First, some amount of non-condensibles are always

l ) l i s ~ v ~ f 181 v , i w t l r n wmk. Secondly, a portion of the condensibles escape condensa- tiun .inif a r i l t~shed throuyh the unit; the quantity of escaping condensibles de-

p w ~ i l i III~ ll l i! ~ ~ n o ~ ~ r i t o f non-condensibles present, the vent temperature, and the ~ i p ! ~ ; i l ~ ~ o 1)11,ss i )1~, c i l lhc condi?ns(:~ at rlischarge. An additional influence on opera-

I<,I ch,il-,i~crt:mt~cs IS that o l temperature. Condensing temperature is the dew point

of t l ~ u gaslvapor mixture at any given point within the unit. Condensation i s not an

isolI1~:rmal process in a paltial condenser as would be the case in "pure" condensa-

t ix i , , i r : ~ ~ ~ ~ ~ l ~ i i , s , i t ~ ~ , ~ ~ UI <i "p~~rc:" vapo, I. Ralhcr. condensing ternperaturc decreases as

I s o mx tu l t ! llows down alol~g the t ~ ~ b i ! bundles. Concurrently, as the

I:~,II~~~~IISLIILY .iic r:or,du~~sed out, the pclcentage of non-condensibles increases. This

I S I I?~, ~ i , i~ t i , i I /IILCSSIIIII III the o o ~ ~ c o ~ ~ ~ I ~ : ~ ~ s i b l c ! s and decreases the partial ~>~,.sru,, &,I the, i:o~~du~wl,lus. Bt!cai~se the dew pol111 of the condensibles is dependent

i i : l l l ~ t . p.ii 1 . 1 ~ ! I ISSI I IL ! 01 the i : o ~ i d i : ~ ~ ~ l ~ l e s , the dew point (condensing temperature)

,viI db>,.i~~;iic ,IS lhc coi,<lt,~~sblcs , j le colldensed out of the mixture.

I,, '';?Iw,,'' C : W ~ ~ ~ J I I S ~ : I S , wllcire ~ I ~ C C O ~ ~ ~ ~ U I I S I I , ~ tempe~aturr is isothermal through-

O L ~ I 111,. ~ : o i ~ ~ f ~ v , s i ! ~ 311~1 i!,? c o ~ l i ~ ~ g watt% irrmpr,8ture is ~easonably corxtant, i t i s

I X ~ , I X I t i ' usu .i log rnc~iu, ti'mlx:r;\tuit? d i f i i i ~e~ lce ILMTD) (the derivation of which , . . it\ I,,, I . i t l11~1 I,, 1nios1 IIOO~S 0 1 1 Ihcat tr i$nsfm/. But i l l the case of partial condensers

,vl,,,l&, II?,, ~.~,rnf,~i~sti,g twrnpc:~;ttu~v vat 1 1 5 . tI1c LMTD is not the plopel- temperature

i l i f i t ~ ~ i ~ ~ c c r 10 ~ i s u - J we~ghtci i mean temperature difference is required.

As 1 1 , ~ c<mr!mnsil~lc.s i:ondmsr, thi: !mass (flow1 velocity decreases and a f i lm of

n~~!l-ci,,dviis~blus b ~ l ~ l d s ~ l p 011 the su~face of the tubes. The condensing vapors must

dlffusr through this film. Each one of these phenomena causes the rate of heat/

m,iss t~:msii:i to decrease, not as a straight line function, but as an irregular curve. Thi. ~ i i q l ~ u d su~face calculat~ons are, therefore, done in small sections and the total

ru$f;ice ~eqoired becomes the sum of that required by the smaller, individual

si'ctlons. Pressure drop calculations are performed in the same manner. In partial

culidensers for vacuum service, the pressure drop calculations are just as important

.is thr r h c ~ n a l siring calculations since the miscalculation of either can cause failure

u l tilt! lutiil ~11ciium system.

TIic i:lass\cdl equation for heat transfer, Q = UAAt is also applied to partial

~ : c , r > < l ~ ~ , s t , , s cvrn rhough the condensing i s really a heatlmass transfer process. The

i v , , ~ h <,I COI~LIIII iinil HDII~CII ( 6 ) has ful-nished equations for converting the mass

~ ~ . u i r t ~ ~ . I I I~ < i l i u s i o i ~ l o st1 lctly heal tiansfel units. The ove~all heat transfer coeffi- I s ~ ~ ~ c i p ~ o c . , l 01 the sum of all the thermal resistances between the bulk of

111,. il,,r, v . l l~ i ,% 1:011111~1111119 n i i x t w t ~ w d t h ~ L ~ l l k 111 the cooIin(l water. The total of all r l ,~~m. i l i~wstdl,ces would i i~clude the rt?sistai>cr of thi! water film, water side foul-

II,!~, ! I 1 AKI vlil)or si<ie fouliilq (ill1 of whictl ale constant or relatively

cm~sr.ii,i) .is wull as no~l-condcnsible f i lm and the sensible heat transfer (which do

vmy cons~dmably from the i ~ d e t to the discharge of the condenser). Accurate

determination of these latter resistances necessitates the division of heat load into

sections for calculation o f total thermal load.

The values of the overall U and A t are deter~nined at beginning and end of each

section and each i s averaged for that section. The surface area required for each

section then is:

section section A (section) == Q Isection)/ U averayeX A t average

The total surface i s then determined by summing the surfaces reqilired for thri

individual sections.

The single most importarit parameter in ~:valuatirig tht: ~i:~forrnarmx ol con densers is the temoerature of the cooling water. Water temperature must be low

enough to effect condensation at the pressure involved with a direct proportion between the two: the lower the pressure, the lower the water temperature required.

Conversely, for a given water temperature. the dew point of the gaslvapor mixture

must be high enough (above the water temperature) for condensation to occur at

all. For a given gaslvapor mix, if the dew point is not high enough to bring about a

sufficient amount o f condensing to make the condenser effective, a steam jet

booster can be used to raise the pressure (and the dew point) high enough to allow

condensing to take place. On the other hand, i f the dew point is high enough

andlor water temperature is low enough, and the suction load contains a reasonable amount o f condensibles, a "precondenser" can be used. The "precondenser" con-

denses as much of the suction load as possible and thereby reduces the size of the following steam jet and condenser, therefore the utilities required.

When the vacuum load consists of a great quantity of condensibles at a pressure

(and dew point) too low to condense with Plant cooling water, but high enough to

be condensed with "chilled water." a refrigeration system should be considered,

and the economics weighed.

As should be noted once more, judicious design choices require care: the evaluation of various types of multi-stage jet-ejector condensing systems, the loca-

tion, type, and number o f both jet-ejectors and condensers required, and the overall economics of each, must all be fully analyzed to make a proper selection.

INSTALLATION AND TROUBLE SHOOTING JET VACUUM EQUIPMENT

The installation of ejectors is normally quite simple due to the inherent advaw

tages of equipment with no moving parts, and relatively light weight. I t is necessary,

however, to adhere to certain requirements in order to obtain satisfactory opera^

tion. A n ejector is essentially a fixed-flow device with nozzle and throat dimensons

configured at manufacture for operation at a fixed set of conditions. Thus, it is

essential to adhere as closely as possible to the design conditions for oplimum

operation of the unit. For high velocity units, such as steam jet-ejectors, the morvc

fluid pressure can be quite critical. To obtain optimum performancr! in these casi:$.

t I S .ilw.lv, cssrilt~al thdt the steam pressure be not less than the des~gn pressure.

w ~ t h CIS httle ovwpinssure as possible. I t i s also essential that the proper cooling wale, bc supplied to any condenser involved in multi.stage ejectors and, to achieve

low mainten:ince requirements, the motive fluid should be of as high quality as

possible. A strainer and separator should always be installed -directly at the inlet

10 th,! <!j,!ctol

The ~ n s t ; ~ l l ~ t ~ o n positton of ejectors is not critical. Some water jet-ejectors and

high v a c u ~ ~ m uiiits (such as jet condensers) should be installed with the motive f luid

flowing vertically downward.

The majority o f steam jet.ejectors are installed at barometric height, especially

whe~l dilccl-conttict condensers are used. I t is essential that a full barometric leg be Ljsed to prowde water removal from the condenser; allowance must be made for the

Inlxtures o f a l r and water discharged from jet condensers in calculating the required lhi~~c~l,l. It s iilso impo~tant that the barometric leg (tail pipe) be as straight as

pnss i~ l r .i~,il 111111 l ro in ally Ihorizontal runs where air entrained in the water can

sepalLite to cause discontinuities in the flow. I n general. any ejector installation should ;illow f o ~ free f low o i the f luid without odd arrangements which will cause

tn~bul~:nce and excess pressute drop in the piping. The supply pipes to liquid jets

must br such as not to cause any tendency toward vortexing in the entrance ch.imbe,

Tht! l i ~ l l o w ~ ~ ~ g scvrn stcps present an ordelly approach to diagnosing system 1 . 1 1 l ~ i ~ ~ ~ w h d l will S;WP apprcciablc timi: mid c l i n i ~ ~ a t c 1:11111s.

S r q , I I ) , , ~ , , I ~ I I ~ I ~ wl~t,tlt,,~ thr 1;10I1 lies wi l l ) ll lu v,i,:uum k,nmp or the system.

Thh I > , t , :~ :<~~>~~) l~s l~ , !~ l by b l w ~ h ~ c ~ ~ ] <>ll thc s<~: t (c~o to thc elccto~ a11d operating the tiit: it: ~11111. If ~IIC no-103d 0 1 shut-off ~oncl t t ioo apploaches the minimum suction

I S I thi? ejector, then the fault her n the system. This shut-off pressure

wotlld I,? .lppioximately 50 Torr for a s~ngle stage unit, 6.3 to 12.7 Torr for a two

i u g a . 1 to 3 Tori for a three stage, and 100 to 200 r ic rons for a four stage. If the test shows the ejector operatng satisfactorily at shut-off, i t can usually be expected

to operate satsiactorily at the load.

S t 2: Naxl chcck the system for leaks and correct. To determine the amount

of d i ~ Icak, approximate the total volume of the system, operate the ejector to

s ~ c i l ~ o d PcrssLlte somewhat less than 381 Torr. and then isolate the ejector from thr system. Measure the time rcqwred for a rise of 50 Torr in pressure in the vessel.

It I S usst!ntta th;it the absolutt! plessurc does not rise above 381 Torr during this tma,. T h l f o l l ~ ~ w ~ ~ g fornula will then give the li:akag<::

A p = Pressure me, inches Hg

t = Time, seconds

I f the volume of the system i s not known, the leakage can still be determined, but

two tests will be required. First, the test described above must be run. Then a

known air leak must be introduced to the system. This can be done by means of a

calibrated air orifice. A second test is then made to obtain a new pressure rise and r time. The unknown leak is then given by:

w' (n P') ( t )

i 1 where I W' = Known leak, kilograms per hour

Ap' = Second pressure rise, inches Hg

t' = Second time, seconds

I f the air leakage is greater than the load for which the jet was designed, then tho

alternatives are to correct the leaks or to use a larger ejector. The ejector must be:

large enough to handle not only the leaks but the normal load from the process.

Srep 3: Assuming that the test indicates the ejector is not operatlrlg satisfact<~r ily, the blank should be left in and t h i ejector utilities checked. First install c d

brated steam gages (with pig tail sypholi) on thi! steam chest or im~ncdiately arli.8

cent to the inlet to each stage o i the ejector. The operatiny pressure is normally

stamped on the nameplate or parts of the ejector, and the pressure determined by

the gage must not be lower than this rated steam pressure. I f steam pressure checks satisfactorily, the steam should be checked for excess moisture. The drainage lines

from steam separators should be checked to make certain that they are dralnillq properly through a steam trap or bleeding through a valve. Open the discharge from

the separator by means of a valve and let i t bleed to see whether this makes any change in the operation. A bucket trap discharging in cycles may cause f luc t~ !a t~nq

steam pressures. Step 4: Assuming that the steam pressures have been found satisfactory, all o l

the ejectors except the atmospheric stage should be shut off and the resultinq

vacuum read. I f this pulls down to approximately 50 Torr. then this stage is opera

ting satisfactorily. I f i t does not pull down to this vacuum, then either the atmos~

pheric stage is at fault or there is a leak in the ejector unit. The atmospheric staye

should be isolated by blanking off its suction and checking shut-off pressure as

before. I f this stage does not operate i t should be checked ior : (alclogged steam nozzle, or steam strainer reducing flow through the norrli!. I f

nozzle inlet i s red or black, look for a scale deposit which can be removed l i y

H t l ' s "St~ l~dar t is lo! D ~ c c t Contact arid Low Level Condel~sers" glves nomenclature, tiJrm>nology, performance, constructlon. and installation of both "baro- ~mcttic" ,111d "low li?w:I ini its."

1 8 1 IIIC,II "S I ;~ I I~~ ;~ I~~S l o Stcam Sul.fitcc Co~~iensers," HE1 provides nomenclature . i r < l tu~n~~,r , l ,qy 1 0 1 swfiict: condi7ns~!~s, a di!tiilc!d svction on performance which i n : I ~ i c h 111'111111;115 '+I,</ 111<!11~lal s ~ ~ i r l i j ~ a l c ~ ~ l a l ~ o n s , jerlui~.ed venting capacities for 1 1 1 i ~ n d nucIc;u powel plants, and data for the sizing of hoyging ejectors and 0 1 ',t,,,,,\,,l,~!l,<: ,f!l,<?f "A",!S.

r h ~ : ~ , , m.iv , i p p ~ r to bti ailuplicati)n of effort III pubicatioris issued by the A I I H E . 111 ~c::iliIy, lhv ASME is colicerni!d with the physical strength [,:ilt.ryi 01 ,111 typo.; o l plessurr vessels and thus established standards for materials,

drs~yn, dnd welding. The HE1 has furnished standards for the manufacture of speci- fi<. tvpcs of pressi~l-e vessels, all operating under vacuutn - the steam jet ejector, the clit;ct c m t ~ i c t condense<, and the steam surface condensers. the latter especially for lage steam engines and turb~nes.

The Tubitlar Exchanger Manufacturers Association (TEMA), has produced a rmmual specific to shell and tube heat exchangers for all pressures and materials ent~t led "Standards of Tubular Exchanger Manufacturer Association." The i t , i ~ ~ ~ / . ~ ~ ~ I s a t ! complimentary to ASME VIII, Division 1, and invoke the Code Stamp u>,l,,sr ,, t l~~,~wlst. specified by the purchaser. The standards furnish nomenclature for tl,,~ ".$I loos types o f fixed heads, return heads and shell side flows and are divided A O I O tlirtrv classes: "13" fol refineries, "C" for commercial and "0" for chemical p~,,~:,.ss, tl,i: last mually i~icludes alloyed materials. Minimums are specified for wall th~cknt:ss and corrosion allowances for vents, drains, temperature and pressure con- Iwctjons. .%nd fo~mu la for calculation o f tube sheet thickness are provided. Typical Ioolinq liictms, pliys~cal properties of fluids a t i d some useful general engineering ~ n f < ~ ~ m a t l a n a l e also included.

EXAMPLE PROBLEMS

Problem No. 1 - Dry Air Equivalent IDAEI Conversion

Cmnv~.it 50 PPH of < i ~ y mi satorated with water vapor at 8" Hg Abs and 140°F I<, DAE ;it 1 0 " ~ .

C,ilcol.itil~os: At 8" Hg A h a~,il 140°F. o n e proimd of ~ I I holds 1.8 PPH of water , I . i t ~ . ~ I o I . ~ I ~ ~ ~ o , \o t l ~ t 50 PPH ,,I w o ~ ~ l t l I i d d 50 X 1.8 - 90 PPH 01 water V . I ~ > . H . Thi* r<,t,il lo;d t l i w I S 50 PPH a r plus 90 PPH ol watet vapor.

I.,) < , ~ > I I V ~ , , t I,) DAE, IC,I<>[ ($7 F i q ~ ~ t c 5.5. All 50 PPHil 50 DAE 63 140°F W V 90 PPHl.81 111.1 DAE @ 140UF --

161.1 DAEO14OUF To collvrj t DAE O 140°F to DAE @ 70°F. refer to Figure 5.5.

Air 50 DAEl.982 = 50.9 DAE @ 70°F Warer Vapor 11 I DAEl.975 = 115.0 DAE @ 70°F -

164.9 DAE @ 70°F Answer. 164.9 DAE @ 70' F

(Note: Unless otherwise spectfied, the capacity curves for steam le t exhausters (ejectors) use PPH of DAE O 70°F as the suction load units. SCFM of air at 70'"F. if more appropriate is sometimes used instead of DAE).

The DAE of the above suction load can also be determined from Figure 5.6 after the quantity of water vapor has been calculated. 50/(50190) 35.7% alr in m l x ~ ture. From 3.2 the entrainment ratio i s 0.85. Dividing 140 by .85 yives 164.7 DAE @ 7 0 ' ~ .

Problem No. 2 -Steam Jet Syphon

Calculate the size and steam consumption of a steam jet syphon to evacuate 200 CU. ft. o f air from atmospheric pressure to 20" Hg vacuum (10" Hg Abs) in ftve minutes or less using the #60 nozzle and 30 PSlG steam.

Refer t o Figure 5.10. Note that the sizing chart has a capacity factor of 1 .OO for the 1%'' syphon, which means that the accompanying curves are for the 1'/." size unit. A t 20" Hg vacuum, the #60 nozzle gives a rate of 2.4 seconds per cu. it. of volume to be evacuated.

200 cu. ft . would require 200 X 2.4 seconds or 480 seconds. But the time limit i s 5 minutes or 300 seconds, so that a unit 4801300 or 1.6 times the capacity o f tlrl! 1%'' unit i s required.

Referring again to the sizing chart, the smallest unit that has at least 1.6 tlmf!s the capacity of the 1%'' size is the 2" size. The steam consumption from thc s a r n c

chart and S O nozzle is 523 PPH. The above calculations can be condensed to :

volume X time (seconds) per cu. it. -. - capacity factor time allowed in seconds

If 60 PSlG steam is available, the time rate is 1.7 secondslcu ft. and

200X 1.37 capacity factor required 5 X 60

From the sizing chart under r i l l 5 norzle (60 PSlG steam) the smallest s ize w ~ t h

at least 1.17 capacity factor would be the 2" syphon with a capacity factor o f 1.38 The corresponding steam consumption would be 375 PPH @ 60 PSIG.

c:?:eluI scriiplng and subsequent polishing. I f nozzles are graphite or plastic,

nii!,iswi, the oilflee to check for any shritlkage or wear.

(h i Fkcc,sb wrssule at dschalge o f ejector. I f the low stage is discharging directly

t i ) .itmospl!e~t!. check the rlischa~ge linu to make certain that i t i s cleal. Check I .illy I : w l ! ? c o s t ! ,may i !c~umi~Ia tc to cause excessive back I I s t i I : s s . I f tlw ciischarq~! lioe is connected to a

hulwull o r to .In .~ltt!r-condunsu~, disconrirct thc well 01 after-condenser from

the cjector and allow the ejector to discharge to atmobphere.

I f the dbove po~i i ts do not provide satisfactory operation, then the ejector must

Iw disassemblt!d The norzlc and t h ~ o a t dimi!nsions should be checked against the

original n i a ~ i ~ ~ f a c t u r r d dimensions. The diffuser should be examined for corrosion

01 we,ri. Pock-like irregdar indentations indicate corrosion. Longitudinal wear

i~idlhs i l l t h ~ steam orifice and d~ffusel are normally evidence of wet steam if

: I S t i t > l pfesr!il. A iwghening o f the irilet tapels through the diffuser,

!v~ l l !c~ i~ t . l i ly nl.~l~,ri.!l cll;!l lg~~ $ 1 1 iIi,l~!iotul inlay r:alise ma l f~ l~ l c t i on i~ lg o f the ejector.

1 lh,, h t l u s i ! ~ sl,ailld be smoi~tlied out or icpl,icr!d.

Step 5: Assuinng that the atmospheric stage ejector has been found to operate

s.irsfactol~ly, i t will be necessaly to look for difficulties elsewhere. Inter-condensers

oprrari~ig under v~icuuln should be carefully checked for satisfactory operation. The

inlet w.jtrr temperature must be equal to or less than the design temperature. The

I~III pipe temperature should be approximately 11°C above inlet temperature for

two stage umts, and 5.6"C for three stage. Note that these figures are approximate,

mcl will vary depending upon equipment design and the amount o f non-condensi-

I~lcs prrst!nt. The ejector inanufacturer can furnish design temperatures. Malfunc-

tioiiing o f thi: condenser can occur with either too little or too much water. Too

littlc water will be indicated by a high tail pipe temperature, and a high temperature in ll>e heid of the condenser before the suction o f the atmospheric stage. Too much

water $nay cause flooding of the condenser and would be indicated by pulsations in

the atmosphetic stage and discharge of water from that stage. I f the atmospheric stacye i s iemoved from the condenser, this carryover of water would be indicated by

rc.ilc <it*posils it> the si~ction co~lnectioi i - which will reduce the area and cause

mcush pirssul u drop. Waw cairy-ovev can also be caused by a leaking t-ril pipe. Air i ~ s i r g it, the pipe will prevent f low from the condenser and cause surging of con-

drns~n(j wale,. Look for leaks at the condensel outlet flange and at the surface of

the W ~ ~ L ' I 111 thv hotwell.

Assuming that thc trmperatu!r ;IIKI quantity of the ~t t i l i t y supply to the con-

tfmisa~ .I,,! cm~i!ct thc c o m l ~ ~ ~ ~ ~ ~ r i tsdl sho~l ld be chucl<etl 101 i n a l f u n c t i o ~ ~ i ~ ~ g . In the

usu.11 cyl~~id~ii: ; , l condwnsui. watai-stt!am c o n t x t i s obt;iincd by a series o f trays

I I i s o f I . At points whciri tlicse curr;~uns strike rho wall of the : I tI1cle s h ~ u l d III! ;i ~ $ n i f ~ ~ ~ n low t cmpcr i~ tu r~ without any brreaks. At

doint: ,wIic:ic IIX' WSIIOI is flee ffom the wall of the condenser thcle should be a i l i ,~t i ,~~n lligh icmperaturii. T1,o steam tcrnperatiire should decrease toward the top

of the condenser, and the water temperature increase toward the bottom of the

condenser. The head of the condenser should be relatively cool, aoproaching the inlet water temperature by 2.8 or 5.6 degrees. so that the load to the armospher~c

stage is as low as possible. High temperatures at this point, with adequate water

flow, indicates by-passirrg o f the curtains. This may be caused by clogg~ng of the inlet distribution piece, either nozzle or trays, or failure or clogging of the subse-

quent trays. This would then indicate that the condenser should be disassembled

and examined.

Step 6: Assuming that the condenser check indicates sat~sfactory performancc.

proceed to the stage immediately ahead of the condenser and examine it n thi.

same fashion as specified for the atmospheric stage.

Step 7: The same procedure should be followed step by step from the atrnas~

pheric end of the m~ilti-stage untt up to the highest stage in order to eliminate the

difficulties.

CODE STANDARDS (ASMEIHEIITEMA)

A steam jet-ejector, a direct contact condenser or a surface condenser can each

be considered a pressure vessel, and as such would come under the rules for such

vessels as set up by the American Society of Mechanical Engineers (ASME). In 191 1 the ASME created a committee (now known as the Boiler and Pressure Vessel

Committee) for the purpose of establishing rules of safety in the design, fabrication.

and inspection of boilers and pressure vessels, and for the interpretation of these

rules in case o f question.

Three publications of the Society are pertinent. Section VIII, Division 1. fur-

nishes formulas for calculating wall thickness for shells, heads, and nozzles, gradi:s of materials and allowable stresses for each. In addition, i t delineates criteria for thti

fabrication of ferrous, nomferrous, alloy, and cast iron unlts, forgings and the heat

treatment after fabrication, flange, gasket, and bolt loading calculation, hydrostatic and pnuematic testing and magnetic particle and liquid penetrant examination.

Section II, "Materials," specifies chemical and physical requirements - Ferrous Materials in Volume A and in Volume B, Non-Ferrous Code Materials. Section IX

gives the Weiding Qualifications for Code Welders.

The Heat Exchange Institute (HEII, i s an association of manufacturers of such

products. Its objective is to promote and further (in every lawful manner] hi: interests of the manufacturers of heat exchange and steam jet vacuum app;iratus,

and the interests of the public in manfuacturing, engineering, rafcty, transportatlo:,, and other problems of the industry.

HEl's publication, "Standards for Steam Jet Ejectors," cmtalrn nummclatu~ii, operating principle types, desiyn specifications, materials u f constructloll. HE1 flanges for vacuum service. pressure protection and hydrostatic tustng mmhods. A

large part is apportioned to "Test Standards" and the last part to "Vacuum E n . gineering Data."

Problem No. 3 -Water Jet Exhaustel Problem No. 5 - Multispray Condenser

What r , , : wale, jct exhauster would be required to handle 15 SCFM of air at 10" H!]. Abs whcn using 30 PSlG )motive watei at 80"F? What would be the water 1 :01 is l l l l l l 1 t111 ,1~

Ri,Ic,i to I F i g 5.12. Er,tcm t lw CII.III at (1) 011 Ic l t hand margin at the 80°F w,itm I , ! I I I ~ C I ; ~ I ~ ~ ~ , . Maw ho~irontal ly to l lw light l o (2) at 10" Hg Abs suctio~i i>ri,ssm<,. Plocwd vartically upwaid l o tlie 30 PSlG water pressure 131, and then ho>i io~, ts l ly to the ugh1 to tlx! 15 CFM line (4) . Read the size corresponding I I s I I I ( 5 T I 7 l c t d I S NO.^ with a citpacity factor o f 7.5 ( l ~ w r l "C;~,~.~,:~ty T :~ l~ lc" l .

Tliu iluanrrty of watt required fur a 2" u ~ i i t [wi th a capacity factor of 1.01 using 30 PSIG motive pressure can be determined from Figure 5.13 which is 67 GPM. Since .I Nci. G size exhauster has a capacity factor o f 7.5 times that of a No. 2, the water I C C I L I I ~ I ? ~ would also be 7.5 times that required for a No. 2 or 7.5 X 67 GPM or 503 GPM of motive water.

The- ;inswur then woold be a S (size) water jet exhauster with 503 GPM of

Inotlvv watt3( at 30 PSlG and 80°F.

P~ohlcm No. 4 - Multi jct and Multijet Spray Condenser

C,III:LII.II~, thc S I ~ B , co(>li,lg watt!r requred a ~ i d ail handling capacity o f a multijet .ii,(l ;i mi~lrilc!t sl,~sy condmso~ ti) hand,! 20.000 PPH stcam at 4" Hg Abs using 80°F r:<,i,l,x~ watc:~

Rt:1<:1 to Figwe 5.19. The curves g~ven for 4" Hg Abs and 80°F cooling water i j l v t i s ; ~ c;lpacity of 16 PPH steam per GPM of cooling water. The total water required

wmlcl Lx 20.000 PPH steam116 PPH per GPM - 1.250 GPM water. From the sizing chaft ti, tlii! l ight of tht: curve. 1.250 GPM woold requirc a 34" condenser that has a maxlmLim capacity of 1,300 GPM. Since the 33" condenser has a maximum capacity of 1,100 GPM, the 34" size would be requred.

To rh,ti.<mine the air hzindling capacity, it is necessary to determine the water disch;inqe tumpcratu$e horn thc condense,. Refel to Figure 5.17 A relatively close approxi~nat io~i of the condenser heat load can be had by using the latent heat of steam at the operating pressure of 4" Hg Abs. Thus the total heat load would be 20,000 % 1,021 o 20,420,000 BTU1hr. The cooling water temperature rise would IIC 20,.120,00011 1,250 X 500) - 3 2 . G " ~ rlsr.

Th? iv.ltt,i i l e t I I 80" + 32.flt'F 01 11 2.Ci1'F and the IV~~II<>>IXIIII!~ V , ~ ~ > C I I pit!sso~t! is 2.8" IHg Ails. T h , "t,!lm:tivc, ,111 p~t!sscw<!," tllc <qx! r i l t i~~$ j ~xcsswc rnli11~1s Ill<! v a p o ~ pr<!ssLlrl! at

riw <lisd,.i~gt, ten1pwi~tuc~i W O ~ < / bc: 4.0" Hq Abs minus 2.8" Hg Abs 01 1.2" Hg. Usnlg the 5 PSlG waicl pressu~e cuwc of Ftgure 517, for 1.2" Hg gives the value

of 2.8 PPH ail per 100 GPM. Using this value, 1,200 GPM would handle 1.250 X 2.81100 - 35 PPH ail.

Calculate the size, cooling water required and air handling capacty of a mu l t i~ spray condenser for the same load and conditions given in Problem No. 4.

From Figure 5.20 the curve for 8 0 ' ~ water and 4" Hg Abs gves a capaclty of 20 PPH steam per GPM of cooling water. The total coolirig water required would be 20,000/20 or 1,000 GPM. Referring to the siring chart to the right (Figure 5.20). the minimum size condenser that can handle 1.000 GPM i s 33" with a maxlmllm flow of 1,100 GPM. Therefore, a 33" condei~scr using 1.000 GPM cooling watt!( would be the answer.

The maximum ;III handhg cilpiicity of mult~sp~;iy cor~d~insiits 15 oot d<!!~rut~!.

since the precooler (direct contact, countercurrent condenser) would be s lLe i I lo suit the air load and it in turn would be supported by a suitably sized air pump. Should the above 20.000 PPH steam contain 200 PPH air. the precooler size would be a #3 Fig. 597 and the water for it required would be 75 GPM. The vent temperature of the precooler would be 8 2 ' ~ so that the air pump load would be

200 PPH air saturated at 3.9" Hg Abs and 82°F.

Problem No. 6 - Low Level Multijet Eductor Condenser

Calculate the size, cooling water required, and air handling capacity of a low- level multijet eductor condenser for the same load and conditions given in Problem No. 4. (Note, after start-up, the cooling water pressure i s reduced from 25 PSlG to 9 PSIG).

Using Figure 5.24, the 8 0 O ~ water curve at 4" Hg Abs gves a capacity of 14.6 PPH of steam per GPM. The cooling water required would be 20,000114.6 or 1.370 . GPM. From the sizing chart to the right of the curve, the smallcst condenser that will handle 1,370 GPM would be a 36" size unit.

For the air handling capacity, refer to Figure 5.17. Note the statement that for these low-level units that the air capacity is only 115 the curve value. The heat load would be the same as Problem No. 4: 20,000 X 1,021 BTU == 20,420.000 BTU pel hour. The water temperature rise would equal 20,420.0001(1.370 X 500) = 2 9 . 8 F . The water discharge temperature would be 8 0 ' ~ + 2 9 . 8 " ~ = 1 0 9 . 8 ' ~ and the corresponding vapor pressure 2.58". The effective air pressure= 4.0" Hg - 2.58": 1.42" Hg. From the 9 PSlG curve 1.42" Hg corresponds to 4 PPH of air per 100 GPM. The total air would be 11.370 X 411100' 54.8 PPH.

Problem No. 7 -Steam Jet Exhauster (Ejector1

What size sngle stage jet and what would he the steam consilmption r f ! q ~ ~ l ~ l ? r I t r ,

handle the loading stated in Problem No. 1 using 90 PSlG steam? 200 PSlG steam? The suction loading for 50 PPH air saturated with water vapor at 8" Hy Abz an0

1 4 0 ~ ~ was calcualted in Problem No. 1 as 164.9 DAE @ 70°F.

1. R e f e r to F iy~ l le 5.34. From the 9 0 2 steam curve the capacity of a 1%" jet at 8" Hy Abs s 128 PPH DAE (at 7 0 ' ~ ) . If the suction load is 164.9 DAE and the capoc~ty of thc l i4"C jet is 128 DAE, then teh req~lired capacity factor is 16491128 0 1 1.29 capacity factor.

Ftoin t h l sirirlg ch;trt, rhu 2 " ~ jet has a capacity factor of 1.24% and requires 420 I'PII t,I n~o tv , i st<i;lni ;I! 90 PSIG m,d the 2 ' ' ~ i l l l i t has a capacity factor of 1 5 i % w ~ t h ,i sto;,nl cr,nsi~nlption o f 538 PPH.

Th,, <liic~sion as to whch of the two al,ovi. sizer should bc made upon the accur~icy of the 50 PPH a r . If i t s just an estimate and it is believed that 48 PPH of alr would suffice then the smaller jet with less steam consumption could be chosen. I f thi! 50 PPH of air 1s an absolutr! (ninimum than the larger jet with the greater steam coi~sumption should be used.

2. Refcr to the 200ssream curve on the same Figure 5.34. The suction capacity f o ~ the l i / . . "~ jet at 8" Hy Abs is 149 DAE (70°F). The capacity factor required wodd be the same suction load divided by the 1%" jet suction capacity or 164.9 DAEi149 DAE or 1.197Fcapacity factor.

The 1 ?>"C jet with a capacity factol. o f 1.0y is far too small so the minimum size su~tablc IS the 2 " ~ w t h a capacity factor of 1-24:, - which requires 378 PPH motive steam at 200 PSlG.

Proble~n No. 8 - Two-Stage Condensing Steam Jet Vacuum Pump With Direct-Contact condensers

W h ~ t s i c two- st:^^! c o ~ ~ d e r w r q steam jet vacuum pump would be required to Ih;i~rdlc 20 PPH dry air saturated with water vapor at 2" Hg Abs and 90°F using 90°F cuplny water ;,nd 90 PSlG motive stcam? 150 PSlG steam?

1. Refer to Figure 5.44 for 90 PSlG steam. A t 2" Hg Abs the intersection of 90'F saturation temperature and the 9 0 ' ~ cooling water gives an air handling capacty of 13 PPH of Dry Air, under Table A.

(Note: This set of curves, as well as those shown on Figure 5.45 for 150 PSlG steam, are set up for dry air and air saturated with water vapor at various tempera- turvs, so that the suction capacity can be read d ~ r e c t y in units of PPH of Dry Air. Thus, i t i s not necessary to determine the quantity of water vapor involved and ,.ol,vc, t thi! tot,il ;i~r/w;itrr vapol lo;ld to DAE). -

TI,,, i:.i~,.rcly f;ictol i iwwred woulcl ba 20i13 ol l 5 4 X . .....

I I I 5 . 4 . l J ~ i c l c Tt~J>lt; A. 1 1 1 ~ t i i h ~ i l i ~ t ~ o r gives a rapacity factor of - 2 . I S 111,. In,o>(non, th.it <.ot1lii lh,,i,~il,, 1 1 1 ~ COICIII~II~!~~ ~ q t l i r e d capiiclty factor of 1.54X. - iloclr,i ' 'Sit!' ' ~ ~ ~ ~ ~ ~ i > i , d n g to tlvs c i l p , ~ ~ t y factof of 2, is listed a 3" TC-2 first- s I I :< S-3 sucol~il stogr. The l o t d str;ini cmlsumption for the two jets is 213 PPI i .it '30 PSIF. Frr i the ln tc co~~clmsc!~ co~resl~onding to this capacity would Oo ;I s r e r:l with 18 GPM at 90°F.

2. For 150 PSlG steam refer to Fiyure 5.45. At 2" Hy Abs under Table A, the ~t?tursectcin of the 9 0 ' ~ satrlration tr:mpararure wilt1 the 90°F cooling water line

gives a suction capacity of 15 PPH of Dry Air. The required capacity factor would be 20115 = 1.33%. Refer to Figure 5.46. Under Table A, the minimumcapacity factor greater tha~l

the required 1.33% is 2. The corresponding jets would be 3" TC-2 first stage and a

1%" 5-3 second stage with a total steam consumption o f 195 PPH O 150 PSIG. Thi- intercondenser would be a size #l with 18 GPM at 9oUF.

REFERENCES

1. Standards for Steam Jet Elecrorr, Third Edtrion. Hear Exchange lnsrtuce. N Y . , N Y

1956. 2. Power Test Code 2 4 ~ 1 9 7 6 "Eiecrorr" American Sociely o f Mecharllcsl Eng~neers. 3. "Thermodvnamic Propertier of Steam'' J. H. Keenan and F. G. Keyer. First Edirron. john

Wiley and Sons. Inc.. N.Y., N.Y. January 1967. page 76. 4. Standards tor Steam Surface Condensers. Sixth Editlo". Hear Exchanger Iniriture, N Y

N.Y. 1970. 5. Standard of Tubular Exchanger Manufacturers Association, Fifrh Editton, 1968 and 1970

Addenda. 6. Colburn. A. P. and 0. A. Hougen. Ind. Eng. Chem. Volume 26 119341. 7. "Characterirricr of The Steam Jet Vacuum Pump" by L. S. Harrlr and A. S. Frcher, ASME

Paper 63-WA-132. 8. "How to Get the Mort From Eiectorr." bv C. G. Blatchley Perraieum Refiner, 12-58 9. "Controlling Ejector Performance," bv C. G. Blamhlev. AMETEKISchurre 5 Koerl~ng

Division. 10. "Selection of Air Eiecrorr:' by C. G. Blarchlev. Chemjcal Engineering Progress. 10- 61 , ~ n d

11-61, 11. "Jet Ejectors and Processing." by H. J. Stiarron. Food Eng~neermg. 11-68. 12. "Steam-Jet Air Ejectors:' R. 8. Power. Oil and Gar Equipment. 10-65 rhrough 1 1 -66. 13. "Steam-Jet Air Ejectors: Specificatmn. Evaluation and Operation by R. B. Power, ASME

Paper 63.WA-143.

serve as an approximation. The correlation takes advantage of the observation that

gasses diverge from the ideal gas law in similar ways, which suggests that a common

scali~l(i t x t o r can be used to correlate values of Z for a wide variety of gasses. In

actual p jxtrce, this scaling factor is determined by the critical point o f the gas in

(11~estio11. Genrr,ilized compressibility charts 131 (4) use the reduced parameters P, and T,

to cot relate 2. where P, ::: PIP, and T, = T n C . G e n w ~ 1 1 ~ e d c l i ~ r t s ate not entirely unive!sal and should be used only when Z-P-T

data to! a specific gas are unavdilable.

Equatlons of State

When a programmable computer is avaliable, the engineer may make use o f one

of ;I niimi)e~ of rqilations o f state to describe the behawor of a gas. A t conditions

ran<liml from low pressure up to approximately twice the critical density, the

Benedict-Webb-Rubiti (or Kelloggl 161 equation will give satisfactory results. Above

this level, the Redlich-Kwong equation 171 may be used.

Gar Mixtures

P V T d.it,l lo, .I n o ~ n l x . ~ i l l O X inlxtwi:s c a i be foc~nd in the literatute. However.

vhm, ~ ~ ~ ~ ~ c ~ ~ ~ n ~ : i ~ t a l i y < l t , t < , ~ ~ ~ t i ~ l c i l i ~ , l um i i t im t I S # lot itviiilal>l~), il method o f approxi-

~ n : t l ~ ~ ~ g ths d,?trl tmust bc used. The pseudocritical point method is one of the simpler techniques. The method

produces a pseudo T, and a PC that may be used as parameters todetermine a

reduced T, and P,. These reduced indeces may be used on a generalized com-

pressibility chart to produce a value of Z.

The pseudocritical temperature equation is:

S~lnilat ly the pseudoclitical pressure may be deternined by:

All that is needed then to determine the components of k, is a value of C,. A (mole wi:ighted average (nay be i m d according i n the equation:

Horsepower Requirements

The work required to compress a stream of gas adiabatically is approximately by

the formula:

where P, a,, and 2 , ark! muasului at ilirake cond~t~oi ls. m d Z2 is det~.inw~i!d i i l

discharge conditions.

Although this relation does not depend on the actual method of cornpresslol?, the adiabatic formula approximates the performance of rec~procating compresruts

more accurately than dynamic units. The brake horsepower of a reciprocating compressor is expressed as the adiabatc

horsepower modified by empirical factors that account for fluid losses at the cy l r l -

der (CE1 and mechanical losses throughout the compressor (ME]. The relation is.

Centrifugal compressors are often rates in terms of adiabatic horsepower, but

with a correction factor known as adiabatic efficiency (qadi applied to the de-

nominator:

Brake horsepower for centrifugal units may be approximated by the additton of

a factor to account for bearing and seal losses. This factor l ~ i l l visually have a villuc

between 7 and 50 hp (101.

I n the case of air compression, the humidity of the gas often has a s~gni f~cmi: effect on the performance of the compressor. Charts are available for direct readnil

of the actual k value that should be used for any particular specific humldty 181 (91. This figure will vary from k - 1.4 to 1.32 at practical temperatures.

When adiabatic head is known, the horsepower requirement may be descrbal ibv

the equation:

C e ~ i I ~ l ~ i ! i ; ~ l cornpressols may also be rated in terms of polytropic horsepower. l h t ! .ipl)ii~dbIt? equations are:

G H P: hpp,/ljOt (Eq. 1-17)

B H P - G H P I ~nechan~cal loss~s (Eq. 1-18)

Thi. p d v r ~ o p i c clficiency used i l l these formulas is related to the ratio o f specific

heats by the relar~on:

INLET VOLUME - THOUSANDS OF CFM

Figure 6-2.

Compressor Pressure Performance

Adiabatic head (Ha,) and polytropic head (H,,) are concepts that are parricw larly useful t o designers and users of dynamic type compressors. The head referred to i s analogous to the head developed by a centrifugal pump. In the case of the pump. the head developed at a particular flow and at a particular impeller speed I S

the same regardless of the density of the fluid being handled. In the same way, thi, polytropic head developed by a specified impeller handling a particular volurni! rare of gas at a particular speed is constant regardless of gas density.

Polytropic head may be expressed in terms of enthalpy IBTU/lb) or unr t i 0 1

ft-lbsllb. It is determined by the formula:

Adiabatic head is determined by the substitution of k for n in the above re la to^;

Adiabatic head may also be determined through the use of a Mollif!r dagts~!i Following a constant entropy line from inlet conditions to the required disch;i~w pressure establishes two points on the chart from which an ir~it ial and i3 f1#!;11 enthalpy may be read. This enthalpy change is related to the adiabatic head by ,i

simple unitPconversion factor:

RECIPROCATING COMPRESSORS

The basic components of a reciprocating compressor are the piston, cylinder, IIIX

valves. Reciprocators are the most commonly used compressors in process swvic~~s.

They tend to be the most energy efficient of all alternative designs. This i s partico larly true of slower speed and large cylinder-bore units. The principle or corn-' pression i s adaptable to a great number of services including those with variabo suction or discharge pressures, variable volume load, low vacuum inlet, or high discharge pressures. Units range in horsepower from fractional to 12,000 HP.

A reciprocating compressor should be strongly considered for applications in^

volving: - High discharge pressure.

Plunger-piston type units can compress gasses to over 25,000 psig the hlyhes! of any of the common mechanical compressor styles;

- Variable volume duty.

Reciprocating compressors can be unloaded in discreet steps eech c h o ~ ~ q n i l

the inlet volume with a nearly proportional saving in power input; Simultaneous multiple services. Larger reciprocating compressor Mmes can accommodate up to 10 crank throws some o f which may drive one to three cylinders in tandem. The i ~ c i m l n , ~ I inilepw,rlsnt sowci!s possiblc with a single driver i s obviously l.lrgl!.

Cylinder Arrangements

Srnallcr compressor units (less than 10 HP) are frequently single-acting in design. Thesc onits compress gas using o i ~ l y one side of a pistol!, the other end being open to the crankcase. The advantages of this design incl~ide a minimum of wearing parts, 3s well as the simplic~ty of splash lubrication of the cylinder in many cases. Single ~ c t w ~ g units are rnost oftcn arranged with cylinders in a vertical or "Y" arrangement (F~gu~c: 6-31.

Whw, pocoss gasstrs must be contained within the cylinder, the double acting cylindei, with a piston rod sealed at the stuffing box, is preferred

Vertical cylinder units range in power capability from fractional HP to about 250 HP. Vrrtical cylinders arc preferred whcn non-lube scrvice i s required (see

Figure 6-4). Larger units, with cylinders measuring 5" or greater in d~ameter are most frequently single cylinder in design. Vertical cylinders will not discharge liquids readily and are not recommended for wet or potentially wet services.

Figure6-4. Mulnplecylinder arrangement. (Courtesy lngeriol Rand Corp.1.

Single-crank horizontal compressors provide flexibility in design for services ranging from 20 t o approximately 200 HP. Cylinders are most commonly double^ acting. Multiple cylinders may be driven in tandem from a singlecrank (F~gure 6~51

"V" or "Y" type compressors with two double-acting cylinders supported f r o ~ n a single frame range from 50 to approximately 500 HP. These units are among thi. most compact designs for their capacity when floor space is a consideration. A th i r l cylinder mounted vertically may supplement the basic design for increased capacty (Figure 6-13),

Angle type compressors are often used in two stage servlce. The larger ftrst-stiqi, cylinder is placed vertically to minimize ring wear and to m~nimize space r i ! q i u , ,

InentS. A smaller second stage cylinder is mounted horizontally from th~: w r r m

frame. These compressors arc available for servces ranglrig from 200 ro 700 HI'

(Figure 6-7).

mticijiated allowable down-time: type of motive power available.

Other factors such as chemical corrosiveness, chemical toxicity, available tounriation space, or special oil-contamination limitations may also take on varying degrees of inlpoltance.

A p~el~minary calculation of the horsepower can be arrived at using equation 1-12 which makes use of the flrst six parameters listed above. Allowances must be modr fo, the mechanical efficiency and the compression efficiency as noted on page 282. At this point, the design cnglneer must rely on model availability data from manufacturers. The anticipated maximum horsepower will place the service in a iange of frame styles. However, consideration of the seventh parameter, allowable doivnttmc, w ~ l l ilarlow the range considerably.

H ~ g l i r ~ speed units, although less expensive, wil l tend to have shorter run times l ,~ twt~, r ) W W C . ~ , p c ~ o c i r thm1 s l m w spt:c!d tillits. Again, the ;lnticipated life of the pluccsr tm.iy br l i ~n~ ted , 0 1 capital sensitive suggesting the application of high speed >eclp!ocatois (900 RPM and abofei.

The selection of a frame style from any individual manufacturer wil l f ix the ialige of compatible cylinders that are available. Each cylinder has a specific inter. nal d~arneter and a maximum working pressure. The total piston displacement (PO) of the cylinder being considered is determined by the internal area of the cylinder, the size of the piston rod, the stroke (piston travel distance allowed by the crank-

shaft), atid the rotative speed of the crankshaft.

For single-acting compressors:

CA ( inZ) X stroke (in) X S (rpm) Po (CFM) = (Eq. Il.li

1728

For double-acting compressors the piston displacement is twice the value predicted by the single-acting equation, less rhe displacement of the piston rod (and tail rod i f applicable).

Consideration o f piston displacements, which must begreater than the ACFM ol the intended service, together with the cylinder pressure ratings will narrow t l ~ c range of cylinders for the application.

Consideration must also be given to the "rod load" or "pin load" rating of thc compressor frame. This must be compared to the maximum force imposed by the differential pressure on the cylinders, since this force is transmitted down the piston rod and/or connecting rod and onto the running gear and bearings withrn the frame.

For a typical double-acting cylinder the rod loading is determined by:

R L (lbs.) = PA ( in2) X Pd (psia) - (PA ( in2 ) - rod area (in") x P5 jpsia)

( E q . 11~21

When the range of cylinders has k e n narrowed to th~s point, a closur appsoxl~ mation of the gas handling capabilities o f il cylindcr may be obtained by all eva lu ,~ .

tion of the volumetric efficiency. Volumetric efficiency, piston d~splacement. and cylindcr capacity, are rc:lateil 2s

follows:

ACFM = PD X VE (Eq. 11-31

Evaluation of the volumetric efficiency requires a knowledge of the cylinde, clearance volume, which is commonly expressed as a percentage of the p~stoi l displacement. I n most cases, the clearance designed into a cyllnder will be betwr,i.rl 3 and 16%.

Precise evaluation of the volumetric efficiency also requires information cow cerning the pressure losses at the inlet valves, as well as empirical data concelruog the rate at which a particular gas slips past the piston rings, rod packing, and cylinder valves. Reliance on the manufacturer's method of calculation is, of course, necessary for guaranteed performance. However, for preliminary evaluations, thc following approximation may be used:

1

Note that i l l most cases Z d,sc, /Z ,,,,, wil l be nearly equal t o un i ty

Horsepower Determinations

A nwn l x :~ o f f~ ic to rs causl: thc compression of a gas wi th in a cylinder t o require I I I I V i ~ i p ~ t t 0111 ~11111 V ~ L I I ~ ~ l h m llii! th~ore t i ca l iscntropic equation will prudict. C h e f among the teasons (or this extra power requirement is the pressure loss through the valves. Other factors are turbulence, valve leakage. heating of rhe incoming gas, and slippage past the packing rings. These losses collectively ~nf luence the compression efficiency o f a cylinder design. Compression efficiency is exptcsscd as the theoretical compression horsepower compared t o the actual cy!inder i n < l i c ~ t e d liorsepower. Present day compressors range f r om 8 5 t o 93% i n C.E.

111 add i t io~ i , there i s 2 certain amount o f horsepower expended i n simply ' turning the 1113cIi111u OYI:~.' These losses include the work lost t o f r ic t ion i n the bearings, p.~chi<rg. wrd pssto\r 1tngs. These losses influence the mechanical efficiency of an rnt i re compfessor. Mechantcal efficiency is expressed as the ratio o f cylinder indicated horsepower to the total brake horsepower of the unit. Preesent day compressors range f rom 8 8 t o 93% i n M.E.

The ovcrall eff ic iency o f a reciprocating compressor is the mathematical product o f the m<~ha i7 ica l and compression efficiencies. Most modern compressors range Ihwn 75 ti, 88% i n overall efficiency.

There IS unfortunately no exact and universal method o f calculating brake horsepower f r om service data. A manufacturer's guaranteed rating is the final word i n thfs matter shott o f startup and testing.

Staging

Tho srilection o f interstage pressures is ft.equently the responsibility o f the process engineer. This may be the result of constraints o n the process such as special i i~ t r rcoo l iny , oi tho inject ior or iemoval of gasses at specific interstage pressures.

Molt! f r t !qu~ l l t l y , the lleed for multistage operation in a reciprocating uni t is dc t r rm~ned by one or more o f the fol lowing condit ions:

1. Discharge temperature limitations. Most compressor manufacturers set a nominal l imi t on the anticipated discharge

tcrnperatLire o f a cylinder. The reasons include considerations of metal stress, rn; icl~int~ tale~ances, mid the! ma1 degradation o f lubricants. I n the case o f oxidizing y.rssi,j, co~,<,sion or l ub i can t fl;~iiim:ibility tn.ly hove to bi: considered. I n general, no i , y l i~n l< :~ s l l o i ~ l d cxcrvd 350°F in xiial,.ttic discll;vga tcinpetature. Adjust ing the 1.lli0 l i t i:c,m,,li'ssi<,n ;,I i: jc l l st;ig~: of conlplc!ssiall. nlunij wi th adequate interstage c m ~ o g . C I ~ I , i i x u ~ e tllis, ( T l i c : ~ ~ ilrc C I X C ~ : ~ ~ ~ O I I S 10 tliis rule, CIS I I ~ the case o f certain singlt3 sliigc air u ~ i i t s which opcrate I I the taiige o f 500°F i n adiabatic discharge trmpelature but which dissipate enough heat t o reduce the actual temperature t o bctwren 375 and 4 2 5 " ~ ) .

2. Power economy A rough estimation o f power savings can be made by assuming the f o l o w n g : a. No stage wil l exceed a ratio o f compression of four. b. The ratios o f compression per stage will be nearly equal and determined by

the formula

b r~ "5

- d r G overall stage +

where ns = the number o f stages anticipated. (This does rtot ent~re ly apply i

process conditions dictate interstage pressures). c. Pressure loss between stages wil l be 5% o f the upstream discharge pressurci. d. Where intercooling i s possible, the temperature approach of the heat r * ~

changer wi l l be 20°F. e. The discharge temperature of any stage wi l l be 3 2 5 ' ~ or less.

4. These assumed limitations wil l lead t o an opt imum number o f stages. Each stdgi, can then be treated as a separate compressor when making approximations of

t horsepower. The resulting total power requirement may be used as a basis of comparison

when determining whether increasing the number o f stages still further is l ikely t o be cost effective.

Fo r precise cost figures, a manufacturer's empirical knowledge o f compressul performance, as well as vendor's pricing information, i s needed.

3. Rod Loading.

C A high ratio o f compression can result i n a compressive or tensile load on thc

piston rod that exceeds safe limits. Increasing the number of stages reduces lhi: differential pressure across the piston of the intake cylinder whlch reduces th<i r w i

load. Frame bearings, as well as valves, and piston rings are all subject t o less wear ti,

the operating differential pressure per cylinder decreases. Lower maintenance ciists can be expected.

Compressor Components

Figures 6-9 and 6-10 illustrate the components of typical double acting compres. sors.

4 1. Frame 2. Cylinder 3. Piston and piston rings 4. Crankshaft 5. Valves 6. Connecting rod 7. Oil reservoir 8. Main bearings

9. C~osshrad 10. P~ston rod

11. Piston rod packing 12. Cylind<,f cnok!ng i;~ckats

13. Dist;lnco pic!cu 14. Oil sclapi!r rmigs

15. Oil dcflcction collar

16. Wr~st pin

Spec~al Considrrotions

Liners - A cy l fnde ltnt:~ is risetl whr:re chang~ng the diameter of the cylinder is

.ii~tlcpaterl at some time in the furute. The bore diameter may be changed either to

inert new capacity req~~iremrnts or to produce a new refinished cylinder surface

. i i t t , v tht. n~iiuirial surface has I)PW (lamaqed i)y weat

Liners may also be required when the material used for the bulk of thecylindi:~

will not provide proper wearing properties at the friction interface. This is the casv

in most steel cylinders.

Cylinder liners are commonly made of cast iron. although special wear resistant

or chemically resistant materials may be ordered.

There are two types of liners in general use. The most common is the dry typri,

which is essentially a shrunk-fit or pressed sleeve within the original bore. These

have the disadvantage of reducing the rate of heat rejection to the jacket water. The

alternative wet type liners. are designed for jacket water circulation immudiatcy

behind the liner. These have a sealed seam between the compression chamber mcl

the jacket which may leak under some extreme circumstances.

Piston Rods - Piston rods are subject to repeated compression-tension cycli,

loading. The surface finish of these rods should be as smooth as possible to avwil

fatigue cracking. Heat treating, nitrat~ng, or carbori~ing is sometimes used to h a r r h

the surface of this component.

Valves - Valves are the most frequently serviced of all compressor comporiitrtls Valve designs must strike a compromise between maximiany operating cycles a n d

minimizing valve losses. Maximum operating cycles are attained by valves w ~ r h

minimum lift, and few moving parts. Minimum valve losses ii.e. minimum pressulc

losses) are attained by valves with high lifts and/or several sealing components.

A : v l ~ t o f i i s s o r oflutirigs i h o ~ ~ l d include a comparison o f valve v : t 71is im, inrr t r t 1s lou i id by dividing the total piston displacement o f a

cvl ini I t2~ LIV tl?,, t ~ ~ l . i l l i l t ! xed o f all suction valves (1 11. I n general, this figure should 1 1 0 1 c , k c w i l 7,500 f t l l n in . how eve^, veiy low moleci~lar weight gasses, hydrogen i n I).I~ t1co1.11, !n;iv 1 1 0 1 I I I~LICI: 811 adequilte valve l i f t Lliiless velocities are relatively high.

SU~II<IO .iilil IIISL:JI;BI[~C! v i l l v ~ ! ~ 21,: 1101 ~II~CICI~;IIKJ<'~L)ID in function although they n i ~ i v somclirnus be iirturcharigeablc in por t seating dimensions. This is a particular

lhazairl io o1d?r units. Care must be exercised b y servicemen when replacing inlet and outlet valves. t o avoid interchanging the two. These valves may be polarized i f

desired, by machining o r inserting pins t o prevent interchanges. Packing - Piston rod packing i n heavy duty compressors i s almost invariably the

f ~ l l l f loating mechanical type. The most frequently used ring material is bronze. although micarta, phenolic resins, PTFE, and other materials o r combinations or

1 n ~ t e ~ i 3 l s may be used for corrosion resistance (See Figure 6-11].

I I S i l l 2 out attic long r u n n i x l periods w e n i n the cleanest o f st'~iwc~rs. D i l l y , w ~ i l , 01. high pressure services wil l require packing ring replacement 1:111(.11 11,111 i' O i I,',,

Distancs Piece - A dstdncc plc:ci? may be ostal led between the cylinder and , i~mic, l o r CII,,, ( 8 1 lhrce ~easo~is.

I Whim >I I S ~i~i i :css, i~y 10 plavmnl c;llry ovur o f ft.ame lubricants in to the

~ ~ l i i i i l c , ~ , t l r t i il~st;l~rc,: picci! assuior that no p o r t i o ~ i o f the rod can travel the cI~sI.~I,c,, I I O I ~ th r frdmc oi l w i p ~ ! ~ riligs to the C y l i d e r packing.

2. In i > t l > i ~ ~ c:ws, the distmce piece si?rves as a means o f venting the process gas 1l1.11 l<,.~hs past thi! r od vack~ng. Gassesmay be simply vented to the atmos- i4 i$ !1~ ! l l r ~ o u y h I;lr!)o ports, 01 they may be purged f r om an encloseddistance

vi?cc wi th 3 stream o f inert gas. The latter method is used when hazardous gasses are being plocessed.

3. The ~list,inci? piece also serves as a service access t o the piston rod packing and 011 ,.~,,ll,,~! , , , , g s .

Capacity Control

Reducing the volumetric capacty o f a reciprocating compressor can be accom~ plished by one o f the following means:

Start and Stop This means may be used when the i>rcssurizcd gas is being rleliwred t o ;i s to lay receiver. Caution must be exerc~sad when sizing a compressor for t h s type ol

service t o assure that the compressor motor is no t subjected l o too Inany starts wi th in a l imited t ime period. Starting ismost commonly done while the c o m ~ pressor i s unloaded in order t o l im i t the starting torque. Constant speed controls. These include cylinder inlet valve unloading mechanisms, special clearance p o c ~ kets, and external bypassing. Valve unloaders, the most commonly used capacity reduction device, hold the inlet valve! open during the compression stroke of the cylinder, thereby p t r - venting compression f rom occurring. In a double-acting cylinder, unloaders allow capacity reduction steps o f 50% and 0% of ful l ACFM. Mu l t i cylinder units allow stil l more combinations o f loaded and unloaded cylinders (Figure 6.121.

Fiyuic 6-72. Capscity COnriol using valve i idoadcrs. The cylinder in rhre lrrsr liyurc ir loaded uil

borh the head arid crank ends 10s shown oil the indicator caidsi. in thurecond figure ihc, hrad end is unloaded for 50% c ~ p a c i r y . I n ?he third figure borh ends are unloaded lor 0% cdp,,c,r8i.

/Courrr!iy lnyr i io i R3nd C o r p

Care must be exercised in the planning stage when d uni t will have valve iuv

loaders in a particular service. A check should be made t o guarantee that rod load reversals (compression followed by tension loading1 will occur whe17 a cylinder is partially unloaded. In addition, multi-cylinder horizontal opp6si:d units should be unloaded symetrically to avoid producing excessive unbalanci:ri forces on the crankshaft and frame.

Thii power saving o f t h ~ s rmethod o f capacity control i s not total since some

work musr be done to lnove the gas in and out through the valve Ports. GUI>~:IJI I~, t h ~ s work will be about 5% of the full load for the cylinder. The

mech;ri,lc~I frlction losses of the compressor also make a contribution to the

powcr consumptio~>. These factors result in a no-load horsepower demand of

~i,ugliIy 25"U 01 fill1 load CyImdc> u11 l0~di .1~ $nay haw! elr!ctrlwl, pncu~nalic, 01 manual operators.

~ / P J I . ; I I I C ~ ? poi;h~zs reduce the volumetric efficiency o f a cylinder by effectively ;ilc~c;is,nq the, cyl inde~ clearance by ;I fixccl percentage, or by a n incremental ,,,11l,ll,Il.

Athorigh cleaiance puckers may be used at the crank end and/or the head end o f ;I cylinder, they are (nore con7monly located at the head end.

C I e j ~ a x e pockets, ~ n d fixed volume cledl-ance botlles have an advantage over

trlet vdvc i~nloadrrs i ? that they do not litnit control to large increments of

c;ip,toiry ~ u d i ~ c t i o n . They (nay be machincd i l l order for small increments Such as

10% ~cduction. Mote than one pocket may be used on a cylinder for several

steps of ieduction. Variable volume pockets are available for still finer control ,,l<:I1~1,,1'1,l1.

TI,,! p i w , , ~ r , x l i ~ u o i ~ affected by clearanci' pockets i s approximately 85% of the

i:.tp.~c~tv ~ a d u c t ~ o n C/~I,II,IIC,, ~ 1 o ~ k r t s 1n1.1y lhiwr dt~ctrical, p~~eumatic, manual operators. Manual

.,I,! ti,,, ,>,Oil C < l l l l l l i O l l .

Whul, col,st,int speed controls are used on multi-stage compressors, all stages

shoold bo rod~icci! proportionately and s~multaneously to avoid large changes in

~ ~ ~ t t ~ ~ s t , ~ g c pressut!. t s r e i i u l !,y~msiiiy as a means of capacity control allows the use of conventional

s l o w i c g ~ ~ l a t n g valves. This method is not often used with reciprocating

compressors because of the lack of any saving in power, and because of the

te~?dency to build up heat in the recycle loop which requires the use of a

bypiiss.gas coo le~. External bypassing o f essentially all of a compressor's output

i s sornetllnes used as a means o f unloading a compressor during startup. How-

ever, suction valve unloaders, when available, are the preferred means of un- l < > ~ l < ~ l l ~ ~ l I n st:lltLlp.

Cylinilcr Coaling

higher horsepower demand and fewer Iblhr of gas delivered.

There are two chief cylinder designs used to dissipate this heat:

Air cooled machines have external fins which act to extend the surface area of cylinders in order to transfer heat to the surrounding atmosphere. In most cases,

the convention heat-dissipation rate is increased by incorporating an ~ntegial f a l l

into the design.

Air.cooled cylinder compressors are nearly always small urilts (less than 100 total HP). They are very commori in air package w i t s for l im~tud dcma~ld service.

Water cooled cyilnders a l e the commor choice for hcavy duty appiicat~o~i,; Cylinder castings include channels for circulating a cooling fluid to mantain u

uniform working temperature within the metal walls of the cylnder.

Although the heat dissipation to the cylinder cooling fluid does have a measui~

able effect on the cylinder discharge temperature as well as the compressioll

efficiency, most general methods of predicting this advantage are not preclsii.

The conservative approach of neglecting the cooling effect of the cylinder on the

gas is recommended for all bur the most critical process calculations. In those

cases, the manufacturer should be consulted for performance data.

In the case of the well-studied air compressor for uti l i ty service, the practice has been to assume that 15 to 20% of the accumulated heat that must be dissipated

between stages and by aftercoolers, is actually taken by the jacket coo l~ rq

system. In order to determine the quantity of water needed for cylinder jackets, an

approximate heat rejection rate o f 500 BTU/BHP.hr with a l ! i °F rise in water

temperature may be assumed for cast iron cylinders. The heat rejection rate will increase as the cylinder diameter i s reduced. The use of dry type cylinder l i ne~ i .

or applications involving gasseswith low k values (such as natural gas) will reduce

the heat rejection rate by some 50%. Cooling water should never be cold enough to cause condensation w~th in a cylinder. Severe wear or sudden damage to thecompressor can result, lncom~ng

water temperature should be 10 to 15'F abow the temperature of the mcamtny gas. An aftercooler or an intercooler may serve as a convenient supply of warp! water. Where only cool water is available, the rate of circulatiol~ through the j,icki!ri

should be controlled to maintain the water outlet temperdture at 15 to 20°F above the gas inlet temperature.

The discharge temperature of the cylinder jacket water should be less tliarl 130°F except where conde~isatiorl within rhc cylinder may result.

Some compressor applications will require no cooling. These includi: low

temperature services such as refrlgi!ration, and systems involvng lr,w ratios o f corn

pression (less than 3) with yasses hav~ng a low specific heat f a t o ai i r l thi: c a r i ,,I

light hydrocarbons. In these cases, the cylinder jackets may bi: filled w r h ii heat

c o ~ i d u c t ~ ~ g lluiil. such as all or an antifreeze solution, to distribute temperatures

i?vc.nly th~oilgl iour the cylinder casing.

A n rhmnosyphon system is often used to move the coolant slowly

~ l i ~ o u g l > the. j;icke~s. Tllis I I ~ V O I V ~ S the use of an external coolant reservoir tank and

I s n .imotii>t of piping to allow the coolant to circulate through a closed loop by

lhti;~tindoci:d conveiitlon.

Materials o f Construction

Lubrication

Hcci~>rocat~nq compressor lubrication is d ~ w d c r l intc four main categories:

F ~ a n r I u b ~ i c a t o ~ ~ ;

F ~ l l l cylinder lubrication;

Mi r~ imum cylinder itbrication;

Null-lube cylinder systems.

A i.ooiprrssols, ncluding noii-lube units, require frame lubrication

Frame Luhriciltion

1 1 , m;ilry sniallc,~ ttnlts, splash lubrication is provided by the agitating action of

tht, ~.~.i~,l,si~.ilt (ot . i t i r~g n r i l l o I teservor. In rhi! singlp-acting compressors, both the Ih;in,,i 111,.31 illgs .IIII~ c y l l ~ l d c m e lubricarod by tlia samr atotnized oil mist.

S : t i s lmh lubr~catcd compressors require frequent oil changes and I i ~ i l .tiiciitims .IS p w t of a ruydar m;,intcnanC<! program.

L,llgti pri~ccss cornptersors have force feed lubrication systems. The points of

lubl icdtor~ include the frame running gear, crossheads, the cylinders, and packing.

1 1 most c~siis, the rmning gear and crossheads will be pressure lubricated through a

single oil system, while the cylinders are lubricated throuyh an indepe~ide~ir 0 1 1

system. In most cases, the lubricating oil used in the cylinders w l l not be identcal

to the oil used in the running gear (Figure 6-13), This is particularly true n the case

of services involving unusual process temperatures, or gasses, requiring special u b r -

cants.

OIL COOLER

The frame lubrication system in laryer compressors commor.ly includes a dust

tight crankcase which selves as an oil reservoir, an oil straner, a small year-type 01

centrifugal pump which may be driven from a power take-off on the main drive and

an oil filter. Larger units may use an oil cooler.

Some moderate size units (less than approximately 200 HP) may use a floo0

type system for lubricating the running gear. Frame oil is carried up by a crankshaft

driven mechanism within the crankcase, and allowed to flow down to lubricate the

jornals.

Full Cylinder Lubrication

The cylinder lubrication system in larger compressors usually includes a muitw

point lubricator which is capable of individual and adjustable flow rates to r x h

point. Lubricants are not recycled (see Figure 6-14). The points of lubrication within a cylinder include the pston rlngs, the packtrlq,

and on occasion, the valves. Of these, the piston rings are the most critical and ,I,,:

invariably fed through ports in the cylinder bore.

1 , ) I,I,III~ c,z~~,s, tl>c: I O ~ I I C ~ I I lb~,i$~!j tcxl 10 thx, cyl1,1d1?1 Imre IS carricd ovcr onto

: n t , , )>1, )0 0 1 1 l l r l c l 0 1 1 1 0 1/11! ~ ) , ic l<~~l< j 1 1 , S L I ~ ~ I C I C I I ~ quantity to elimirmti! thu need for

si'pi.itc IeW to the pricking. V.ivcs i,suGilly do nor r ~ x i ~ l i r e d separate 1ub1ic:ant f w d . Oil is carried to the valve

t I 1 j s t . In so!ric i~~stallatians, oil rnay be injected into the gas

srle, i ln befnlc rhr compressor inlet t o guarantee lubrication of the inlet valves.

Care must be exercised in the selection of cylinder lubricating oils. Lubricants

mtx Ihigl7ly i~,diwdualired froln process to process. Halocarbon-compressor oils can-

,not be interchanged with lubricants for ammonia compressors, uti l i ty air compres-

sol oils cailnot be interchanged with lubricants for high pressure cylinders, etc.

Thwnal decomposition of an incorrectly prescribed lubricating oil in an air

C O I ~ / ) I I ~ C S I I I c d i i i csdt $ 1 7 r l~c~lni i i lat ions of soot in the discharge piping. These

~11,1,1~111 ,311. c c l ~ ~ ~ b i ~ s t ~ l ~ l ~ , , , I I ~ I . ~ I ~ ~ ~ ~ . ~ ~ . I ~ I I , I c i i l s 3 ;)LIP to s p c ~ i l y rhc ~ p t i ~ n u m characteristics o f a

I I I .i ~ . I I I I C ~ J I ~ . I C ~ ~ I I I ~ ~ I I) d specif~!d SCIYIC<?. how eve^. thc final recom- ~ n ~ i ~ i c l . i : ~ t , o 1m lht. (IX;IC~ lubrci1111 to bt! ~ m t l should come from a reputable oil

s i l p p l ~ ~ ' ~ ivlic i, w i l i l lg to gildrantee the pet fo ima~~ce of his product. 1 1 , ymrral. the bleak-in peliod of a ni?w reciprocatingcompressor wil l require

i e a t v e y I81g.e amounts o f lhbr~cant. Most often, this oil will not be identical t o the lubricant that will be used during the process run.

The rate of cylmder feed is not completely predictable. Proper oil feed rates are

determined by observation at startup, and every six months or so thereafter.

Minimum cylinder lubrication is a compromise between full lubrication and

non-lube design. Full lubrication offers mechanical reliability to minimize down

time, but invariably results in a gas stream contaminated by the lubricant. N o n ~ lubricated cylinders wil l not contaminate the gas stream to any measurable degree, but are more subject to mechanical wear.

Minimum lube systems incorporate the materials of construction of i ionub i i

cylinders. The mechanical wear of these components is minimized by supplytny

them with a fine coating of a liquid lubricant. I n the most common design, the rate of lubricant additton i s so small thar lht:

packing is the only lubricated component. In this case, oil 1s carried into thu

cylinder by the rod and slowly migrates to the cylinder walls and valves. The rat i : i i :

oil feed to the packing itself may be minimized, or eliminated altogether if su f f :~

cient crankcase lubricant reaches the packing during normal operation (see Ftyure

6-15].

F,gure 6.75. /Courtesy lnyersol Rand Corp.1

Non-lubricated Cylinders

Non-lubricated cylinder compressors incorporate the f o l o w ~ n g construction

features: .

1 P x k ~ t ~ g rngs and compression rings made o f a self-lubricating solid such as

~ I~ ICI I I , I U I I ~ I O I C ~ ~ PTFE, combinations of carbon and PTFE, or molybdenum

~ l l s~~ l ,> l l l d ! ;

2. A il~rt;ini:i: i11rci: k t w c e ~ i the frame and cylinder to guarantee that frame oil

~ : . i ! ~ ~ l o l I,IIIC,I thii cyIin~lc?r through the packing. An oil-deflection collar may

IN. o s ~ ~ i 1 0 st011 thv migratio~i of frame oil along the metal surface of the rod;

3. Spcc~;i! valvc i~tscrts at the points of wear ~nade of materials similar to those

usi!C it, the piston rmtgs and pack~ng;

4. An extra ring or set of rings on the piston designed to support and guide the

piston. Thcse 'rider rings' are 1101 designed to seal against gas leakage, but

r;ithc.~ to rakii ,most o f the mechanical load and wear that would otherwise be ilnposed o ~ i the piston rings; (see Figure 6.16).

5. A l 1 1 1 ~ 4 y h m m l finish in thc cylinder bore and on the piston rod. Cylinder w~f;ic,,s ;IIC cwnrnoi~iy lhoncd with PTFE to f i l l castlng pores etc. This pro-

~ : i i s s ii.isi!s tlic wi3h o f ~ rngs dur~rlg bleak in.

14 -G

SINCE T H I S PISTON OPERATES VERTICALLY. ONLY ONE RIDER R l N G I S REQUIRED.

RlNG

In general, non-lube cylinders will have a lower compression efficiency than

corresponding lubricated cylinders. This added CE loss i s approximately 5%. This

reduction in capacity does not result in a noticeable reduction in horsepower sinci!

work is done on virtually all the gas within the cylinder even though some of i t sips

back to the suction side during compression. Special precautions must be observcd when a non-lube compresror is to bc used

1. The cylinders are particularly sensitive to dirt. Thorough prvstartup cln;ini~,g

of all process piping is necessary. A suction filter with a rating of 5 micron; 111

finer is preferred to a suction screen. The rate of wear on cylinder widli. piston rings, rod packing, and on the piston rod can be expected to i:ri:<it!rI

that of a lubricated unit. Careful monitoring of the rate of wear is rnact:is,ily.

2. Spare cylinder components will be used more frequently in a non-lube corn-

pressor than in a comparable lubricated unit;

3. Special prcautions must be taken to prevent the oxidation of metal surfactts

during periods of downtime. Ferrous cylinder bores and piston rods a r e not protected by a lubricant coating and are subject to attack by small amountr

of moisture and oxygen. The use o f warm circulating jacket water or an i r lc~t

gas purge may be necessary as part of a standard shut-dowti or equpmenr

storage procedure.

Compressor Gas Piping

Compressor cylinder ports are most olteri sized larger thar) gas l j lp i~ ig practws

would predict. This is because gas velocities or nominal prossurc losscs b;isi:d r,ll

steady state f low do not account for an all important criterion in reciprocatiicj

compressor line sizing, which is that pressure pulsations must be attenuated.

Gas piping to and from compressor cylinders should not run smaller in diameter

than the corresponding port connection size. Piping runs to the inlet and discharge side of compresror cylinders may have the

unplanned for effect of resonating with pulses from the compressor. If unchecked.

these oscillations can interfere with f low dynamics within the cylinders, causing a

loss o f capacity and an extra burden on the compressor driver. In hlgh pressure

installations, these pressure waves can do physical damage to the piping installation.

There are three main solutions to the problem of pressure pulsations, The suresr

of these is to rely on a firm specializing in the design and manufacture of pusatlori

dampeners. Commercial dampeners incorporate proprietary pulse attenuating dc:-

vices within a vessel that have the effect of reducing downstream pulses to I%, of

the lime pressure. They also may reduce upstream pressure pulses to a specified

limit.

Pulsation dampeners tend to be expensive n first cost, but offer the be;t prii-

tection to compressors in services involving high horsepower and high pressure. The next best alternative to the pulsation damperlor is thr? surgi: bottle. T h i : ~ 2 1 r .

enlarged chambers in the piping, located adjacent to the compressor port;, that ac t

ro db,ti!oe the p~piny system. Some rules-of-thumb can be applied for evaluation purposes: In services up to 700 prig, they are at least seven times the single-acting

displacement volume of the cylinders they serve. Higher pressures require larger

bottles. A t 2,000 pug, the suggested minimum is at least 15 times the swept

cyli~icler volume (131. Sulgr bottles should be located as close to the cylinder as possible. Pressure taps

'ilong the chamber may be useful in checking the performance of a volume bottle.

Whenever possible, bottles should be equipped with drains. The third method of averting pulsation resonance, is to design the inlet and

discharge piping systems to avoid straight piping lengths, and equivalent piping

li:nyths, that will oscillate in sonic resonance with the compressor. Compressor suppl~t:is can usually assist in this planning. For large installations, an analog study in,," Ih, n ~ ~ ~ ~ d t d

COIII~II~?SSOI pipinq must include an allowance for thermal expansion. It is an

WIWISC ~ ; I C I I C C 10 0011 a horizontal compressor cylinder down to a section of

dischalgc piping that i s rigidly braced against a floor. Other lines that may be

subject to plpe thermal expansion such as those with tracing and those handling

~cf r ig r lmts , should not be positioned to create a strain on a cylinder casting or risk disro~ting cyl i (>d<i l - f~ame alignment.

C s s discharge lines and manifolds must include a safety relief device

w r h ;I pli,ssillr ~r,+t~ng ade~uat? to protect the cylinder. as well as a capacity rating

I c l I I ~-OIIS~~~.I;IIIOO 01 1hv volume between the compressor and the nearest

/101111 "I 1S01;,11o11 ( 1 2 ) . 1 1 , ~ ~ s t ~ ~ 1 1 . 1 t ~ m ~ wI1c:i~ ~IIU 111Ii:l \) i l) i~lq will 1101 st i~nd lul l discharge pressure, i t is

h , s r ti, c m ~ s i d a ~ that the i:omprcssr>r valv<:s !nay at som! time leak high pressure gas

back to the suction side. I f the inlet piping can be isolated by an upstream valve, rhl.11 ;I ow-s~de ,<!lief valv,i is adviscd.

t3.xlc instlumentation should include inlet and outlet pressure gage taps and/or

~nstalled gages with snubbers, a mounted dial type thermometer to monitor dis-

I i p a t r , <and a means of measuring gas irilet temperatures. Where I ~ ~ ~ X ~ ~ I I I C ioC packng is ~ised. J r i iea~~s of m~asuring the packing temperature will

wove i~soful.

Alarms

TI?,! need l o rnol~itor the likely trouble spots in a compressor increases in propor-

11m1 1 0 thi! inipolt.lllce o f the cornplussol u ~ u t in process. I t i s often the responsi-

h l r y (11 rlir process engineel to select the appropriate alarm points. Piucess v;i~iablus that may endanger the operation of a compressor include:

Sii<:c,oi, piessiirs wriaNoi).

Reduction in inlet pressure will change the horsepower requirement of a compressor. I f capaclty contjol cannot accommodate this change, then the corn.

plrsso ilhver may encounrer a peak horsepower demand at some point. A well

designed system will be able to operate in this peak range wlthout overload~lig

the driver or compressor frame rating. However, economic considerations inay

warrant the use of a low inlet pressure alarm.

A more significant effect of low inlet pressure is the probable increase in the

ratio of compression. Piston rod loading may be exceeded, or discharge tempera-

tures may run excessively high causing a risk of lubrication failure, cyl~ndrr

damage, or valve damage. Discharge pressure variation

Operating a loaded compressor against a dead head can be disasterous. Under n o

circumstances should a compressed gas system be designed without regard to thi! maximum allowable working pressure o f the cylinders.

Variations in discharge pressure can also have horsepower, rod loading, an<! temperature effects similar to those caused by low inlet pressure. Excessive discharge pressure also resents the danger of condensation withill rl18'

gas being compressed. Lubrication failure, valve darnagc!, or severe! damaql? t i ) :h

cylinder head and drive train may result.

Inlet temperatures.

A t a constant ratio of compression, the discharge temperature of a compri:sm

cylinder is very nearly a multiple of the absolute tempeature of the n le t ijd,

Good process designs should not depend on the cooli~ig jackets of cylirides to

relieve high adiabatic discharge temperatures. Composition chanyes.

Failure to account for the addition of small amounts o f corrosive or churnxll,r active materials in the process stream can r~:sull in lrr lmc;i l ir~~i fulnro, a r : c d ~ ! ~ ; ~ l ~ : ~ I

corrosion of wearing parts, and the vodirig of manufacli~rer's warranty.

Composition variables should be controlled to preclurle condensation w t h ~ r , cylinders.

Features within the compressor car1 also endanger its operaton. Tliesc includl:

High discharye temperature.

No single symptom i s a more versatile diagnostic clue. High dscharge temlii!i.~

ture occurs when:

ratios of compression are running high; valves are broken, worn, or seating improperly.

cylinder cooling is inadequate; piston rings are worn or damaged, allowing excessive slip;

cylinder walls are out of round, scored, or worr~:

cylinder lubrication is inadequate. Low discharge presure.

LOSS of pressure may be attributed to valve failure, worn rlngs or worri packrig

However. this symptom is more likely to be the result of conditions exter~?al t r ,

the compressor, such as a downstream increase in demand for gas or a pIuq<ji.(J inlet filter.

LOb,, l,~,,!!t! <>,I pr'wore.

!'tab:iMi: rc3uscs nclude lube pump tal lu~e, auxiliary di-ive failure, contaminated

or!. ail oil li?.!k, a plugged oil filter, or loss of 01.

L,)," /r.,,,,t, <>,I /~"L,I.

J!>!,/l I,.,!,,<, <,,I I~,,~,,>~V',f,,,,,. Tlvs C.,>IRI~IIWI is I A d y 1 ~ 1 OCCLI, wli,!~i a bca~r,!, h:,s failed. I t ~n iyh t also bc: the

~<!s,,II <,I ,, !<,,I,,,<! I,, lh<! <>,I cool,!,

1 . 0 , ~ / , , I , >J<!<,,/ //<,w.

This is a co>ldit ion cdused by plugged lubr~cation ports, a plugged filter, a

~n<!cli;~~i!i:;~l f.lilurr of the pump, or 3 loss of oil. l . , > v / , ~ , , / i ! , d ~ r /,,!U!<:,,f,,,g <,,I /<?",!I.

C~/,/I</<V It/hx,,ror roraf iu!~ f.w/ur(?.

E \ , , v s s > ~ L ~ w l > r ~ t m , ~ .

h.; 111.!y 11,. thr, twul t of a bearlng failure, a valve failure, a valve unloader I,I~~IIII,, I i ~ l k i ( l I I :I cyliiidr,, a lubrication failure, a weakening in the compressor

luui,d.ition, 0 1 a la i lu~f ! in the hold down bolts.

H i y / ~ i;irr!rsr~ye rei~~perariire.

Multb,t,ige comPtb!ssors with integral watercooled intercoolers depend o n a con-

s l t L I couling water to avoid a high inlet temperature on the stage

r!owmt~t!am of the intercooler. A rise in the interstage pressure and an uncon-

t i d l ~ ~ r l s11ifri~r:l in tlic i ~ t i o s of compression wil l result from an intercooler failure.

M.lilv cmnilll,sssol m;ululactilrr~s are d l le to s~ipply standard annunciator panels CI)III.IIII~!I ,111 nucliblc tilkilrn. 1;1kImi warning lights (usu;llly arranged to identify the

111st i xx t i~ 01 i ;~i lurc). as wcil as timers and switch contacts to atuomatically shut

lht! coi~ii>xosso~ down whiin necessary. The alarm points are often selected by the c\istomc

Drivers

make the induction type motor the most popular electrical driver in current use. On

the other hand. they have the disadvantage o f a lagging power factor, as well as a

slightly higher power consumption to power output ratio when compared to the

synchronous motor. Induction motors may be direct couplrd, belted, or flange mounted onto a

reciprocating comprcssol Synchronous mofors opi:rati! 3 pr~!cis,! rotatlve speed. They tcnd to lbi! sl~ghtly

more power efficient than their induction motor counterparts. and lh;rvu the! ofterl

useful advantage of a leading power factor.

Synchronous motors arc most often applicd to larger compressors. usually with-

out speed reduction. Thr!y may be direct coupled, flanga mounted, ur mwrl t l i r l or1

the compressor shaft in an engine type arrangement.

Induction and synchronous motors may be required to operate in corrosive or

flammable atmospheres. Attention to the National Electric Code, the NFPA and

local ordinances is advised. Special venting, a special enclosure, or special materials of construction may be required.

The choice o f motor enclosure will also be influenced by the nature of the

operating atmosphere. Standard casing designs include open, drip proof, splash

proof, totally enclosed, and explosion proof, among others. A steam engine may be used to drive a compressor in applications where surplus

steam is available. Steam engine drive offers the advantage of relatively high effi-

ciency in comparison to the alternative steam turbine. I t also provides flexibility in

capacity control since its rotative speed is infinitely variable through a wide rangc. Capacity reduction is achieved with a very nearly proportional saving in power

consumption.

Units range from about 50 t o 4,000 HP. Compressors with integral steam engines are among the oldest and most reliable

compressor-unit designs available. Some less common designs use a compressor-

drive coupling arrangement.

Steam engines have three disadvantages:

The steam exhaust from most engines will contain small amounts of lubtlcating

oil which can affect tho performance of the boiler i f condensate I S t o be reused.

Non-lube adaptations are available, but oiily at a sacrifice of on-line availablity.

Steam engine drives rend t o be bulky in comparison to electric motor drives.

The efficiency advantage of the steam engine over the turbine tends to diminish

when high pressure (over 300 psig) and h ~ h temperature steam is to be LISI?~.

Steam turbines operate at relatively hlgh speeds and require a reducing gear to

drive standard reciprocating compressors. Steam tllrbinrr can match arid encr:c<l the ranqr of horsepowi:i cr,vl:ri!rl by stcam

engfnes. Thcy havc lh r addsd xlvantayc r i l 01 - l r r i i : dlscliargi:.

The chlef mechanical dlsadvantagi! of the steam turbiric 1s rhr: !ni!cessity o f thl: reducing gear.

Gas e,ryiires a,id gas turbiiws a r e the preferred type of driver where fuel gas i s p l e ~ ~ t i f u l . These sites (nclude gas transmission and well head stations, gas storage

f.1c~lit~1!s, iuol gas ptoccssng plants, and g;~solino refirling facilities, among others. Integral-cngcne compressors have been the preferred design in very large installa-

tlons , i l thoqh dilcct co~ipled engine drives are ,rot uncommon. Gas engines may be % t r , 4 ,II<,~,. cyclc iui < 1 ~ 3 1 ! 1 1 1 . w ~ t l ~ sr~pc~chi~c~l ing a,ldt!d to some units to improve

Ihtwm,il i ~ l l ~c : l~ !n i :~ . Ad j~s lmrn ts ill C O ~ ~ I I : ~ S I X capilclty can be )made eilrily by valy- ,o!j !ht' spucd of the (cnylnt: duver .

G;is ti~11,int:s lcquirt! a speed reducing gear to drive reciprocating compressors.

Somc jpplications have ncorporated exhaust gas heat exchangers to improve the

overall efficiency of plant processes.

Diesel oil fueled engines are used less frequently than other drivers. These may

be integral compressor-driver units or direct coupled compressor-driver assemblies.

These engines may be designed for dual fuel application to take advantage of

seasonal changes in fuel prices and fuel availability.

Installation

Ruc~p~ocatlny compressors vary in foundation requirements. Some compressors

Ihciva mulr i l~ lu pistons ;ind counterweights arranged to minimize net dynamic forces

t,.n,sm~rt,!ii rh,otigh thc iramt:. Those balanced units may be mounted on a struc- I I i d I 1 ' Y ' type units may need no molr? than a strong concrete

fl<,<,, ,,s dl, J,l,Y~,,:,t,, l"L,ll,i,ltlo~l.

t l , most ~ t . ~ : i p ~ u c ' i t i i ~ l compressor {iasiyns ~equi tc xi indi'pendent poured

,:,III<:,,,~C, foun~1atoi> 10 /i l l ld ~mi ic l~i , ,~! aliyilmi!~it, 3n(/ to assLlli: thiit vitxatioll forces 1 1 c m 1 1 , ~ c c ~ ~ n p ~ c s m a,! ! n c > l t r ~ ~ ~ s ! ~ i ~ l t ~ ? d to s ~ ~ r ~ ~ x ~ ~ l d ~ ~ l y s t c ~ ~ c t ~ ~ r ~ ? s .

Basc ar , i l ycne~al leculnmendatons on the type of foundation needed for the compressor n are provided by the manufacturer. However, the compressor

purchasel. assumes the responsibility of investigating the bearing capacity of the

s~~ppor tmg 8011.

ROTARY COMPRESSORS

Helical Comprussors

The basfc components of a rotary screw compressor are the main rotor, the

s~:cond,ity lotor, dnd i h c housing ur cyl~ntlcr. Comprcssior 1s accomplished without

The principle of compression involved may be envisioned by examimng e t h i , rotor separately. Each rotor, encased in a closely fitted cylinder housing, form,

hollow spiral-shaped cavities. Gas enters these hollow chambers as they rotate pas? an inlet port in the housing. The intermeshing between rotors effectively forms 2

moveable seal within each of the cavities. Compression occurs during rotation as th!.

cavity travels down the housing toward the discharge port. Cavity volunrc Is i , , ~

duced until the chamber encounters an opening discharge port. Like many other compressors the most frequent application of the rotary s c l w

machine is in compressing air. Helical compressors in this service are nearly always

supplied as packaged units.

Apart f rom air service, the rotary screw compressor has been used for handlinrl

HN,, coke oven gasses, fluorocarbon refrigerants, wet HIS, various organic vapors,

saturated steam, and a variety of other applications.

The oil-flooded design and the dry type are the most popular units in current use.

Dry Type Rotary Screw Compressors (See Figure 6-17)

The mating rotors of a dry type rotary screw compressor do not make contact

with each other within the compression chamber. Timing gears are used to maintar

a small but finite clearance between the mating surfaces as the rotors turn.

Typically, a dry type units operate at rotative speeds that are three times fasts! than oil flooded models. Speeds exceeding 10,000 rpm are possible depending on the size of the unit. Alternatively, some models are designed for drect couplnrg !o

3,600 to 1.800 rpm motors.

Dry type screw compressors may be considerrd fur applications ranyny up to 9,000 h.p. Capacities ranging to 26.000 CFM (of air1 are available.

The differential pressure across the compressor is limited by the allowable d e ~

flection of the rotors. In some standard units this differential l imit is approximately

80 psi, in others. i t extends to 170 psi.

The highest discharge pressure attainable is dependent on the strength of t h i casing. Special designs are available for single-stage services exceeding 450 psg.

Discharge temperatures are limited to those values that do not cause a significant

rotor growth through thermal expansion. Designs for temperatures up to approxl-

mately 450°F are available.

These compressors should be strongly cons~dered for services involwig:

1. Gasses with entrainments.

Non-contact operation of the rotors often permits the handling of droplets a ~ i c l

fine solid particles that would cause unacceptable wear in a reciprocating or ceritrl.

fugal unit of comparable size. Screw compressors can provide reliable srrvcr n

'dirty' applications as found in Pollution abatement systems, In compressing vertical

Pyrolysis furnace of f gases, or in wet applications such as the mechanical recom-

pression of vaporsjn evaporator systcms.

AltI l~w(l l1 .i y I s ! : p s s m;iy bc thl! 1x:st choico for a I I I I I I SU~VICI , , 501111' wiiill I I loss 01 x i : must be t ~ ~ ~ W < ~ l , , , l .

! 2. G.is,t3s w ~ t l i r ! n t~a inmi~~~ ts that contai~i fouling inawials. Sonir ( o t a y scl.ew cnrnpliissoi drs~gns will accept the continuous injection of a

IIOLIIC m lo t l i ~ ~:onip~esso~ C : I V I ~ ~ . This liqci!ri Jcts as both a 9.1s coolant and a rotor clr.in!n(, so l vc l , l .

Ex.im[~It,s ~,icludc h u inltct ion of water mto <i w l i t c~mpressing vapors from an evaooratol concuntrating ar ii>organic solut~on, where entrained droplets might o t h m w ~ s ~ ~csul t 1 1 , W~CI I IS~~I~I (? I I . In this caw, the iiijection of water p re~en t~so l i ds .~~ . ( -O I?~L I I~ I I IO~~ L ~ I I thl' c i l t ~ ~ s . Anothvi exiiniplc 1s [lie coniprcssiori of spray tower

~ I s i - l l . i ~ g c ' (j.isst,s ivl icl> 11r;iy I,.ivu ent~~,iinmt!nts wit11 i i i ss(~Ivd hard-water minerals or s ~ ' ~ ~ l o l l ~ ' ~ ~ l ~ ~ , ~ ~ ~ > ~ c ~ < ~ m p n ~ ~ ~ ~ < I s .

I!>,, ~b,i:g>\>)\,it.nd,iiii,n u l .i p;*iticul.i~ sulvm~t i s i~suii l ly the r t !spo~~~ib i l i ty of the g e g . Wate~ is the imost common tnedirim except where strongly reactive q:issrs a l e imcounte~ed. Alternatives include o~thodichlorobenzene, gas-oil, kero- srtlu, l i q~ i i c NH.; and others. The injection of a cuolant ur solvent may increase the coniprcssic~ri efficiency of a helical unit by partially sealing the rotor gap and

reducing the gas slip loss. On the other hand the injection of a liquid will induce some hydraulic power losses. I n addition, any vapors generated by an evaporative cooling effect must also be compressed with the main stream flow. The net horse-

Power advantage or disadvantage of coolant injection i s often negligible when compared to dry adiabatic performance.

3. Gasses that polymerize at elevated temperatures. Insomeapplications, thecoolingeffect of liquid injection can avert the formation

of sludges, films or resins that would otherwise form at higher temperatures a r id

pressures. Wet hydrogen cyanide and acrylic acid vapors are examples of materials that

have been successfully compressed in rotary screw machines in spite of their tendency to polymerize under conditions developed by adiabatic compression.

4. Gas streams of variable molecular weight. The screw compressor, being a positive displacement unit, will move gas a t a

fixed volume rate despite variances in gas density.

5. Gasses that dissociate or react at high temperatures. The injection of a liquid coolant can reduce the heat of compression in ydi

streams down to acceptable level Phosgene i s an example of a temperature sensitive compound that has been suc-

cessfully handled.

Construction

Casings are most often cast iron for services to 200 prig. Cast s t i ! e ciro I N

supplied for higher pressure services. Rotor shafts are commonly mad[: of stecl. The rotors thcmsr:lvr!s rnay l h I I I ~ , .

pendently machined components made of cast iron, or thr:y may be s t e i : c<>lr, ponents fabricated as one integral piece with the shaft for still yeater mechantcal

strength. Special materials of construction such as Ni resist. Hastalloy@ coatings, or stail).

less steels can be supplied for particularly corrosive services. Shaft seals are commonly carbon ring labyrinth des~yn, but may be modified by

standard options such as purge, vacuum vent, or liquid end fittings. A number of other more specialized designs, including mechanical seals, are also available.

Bearings may be splash lubricated in light duty compresssors. Force feed l u b i c a ~ tion of the bearings and timing gear is the preferred method in large units. Lube oil pumps may be shaft driven or external and motorized. Lubrication systems may be of a manufacturers standard design. to suit the installation site and operating cond- tions or built specially to meet particular industry or user specifications. Oil

pumping. cooling, filtering, and pressure control for single or multiple services arc common to all.

Operating Characteristics

The thcoletical ind~cator card for the rotary screw compressor resembles that of

;, ~~c i />~oc i i t i , , g comp!essor with zero cylinrlei clearance. However unlike the reci-

~ I O C ~ I I I I ~ c ~ m p r e s s ~ r , the openillg of the discharge port depends on axial port

~>I.ICUI,I'I~I .inlil 1101 011 d i f /~ : tm t i i i l ~ ~ O S S L I I C IIC~WWII the compression cavity and thc

~I,cl>,i~gt, 1 ~ o c A 1ott31y scjuw colnprassor will compluss gas within its lobes through

a sprci f~c volume reduction ratio despite the line pressure outside of the machine.

I f , when the compressor discharge port opens, the pressure within the chamber

i!xci?cidr thv pwssulu in lhi! lille, then tha gas cornpresscd within the machine will exp:i~,d ill10 I ~ C ~ I S C ~ J I < ] U line. On the other hand, i f the line pressure exceeds the

~ ~ U S S U I W witt i ir the comp~ession chamber when the discharge port opens, then some o f thu g . ~ f rom the line will flow back into the compression chamber. Conse-

~ ~ u u n t l y , m y m~smatch brtwt!en the design ratio o f compression and the actual

sc~v~cc! r;,to will resrilt i r a small loss of powel efficiency '

The optimum latio o f compression (p2/p , tor any rotary screw compressor is

determined by the thermodynamic properties of the gas and by the volume reduc-

t o n ratio built into the machine in question. Typical ratios of compression, based on atmospheric air inlet, are 1.5:l and 2 : l for units capable o f 35 psig discharge

plusrule, tilid 4.2: 1 for high pressure machines. Mtiltiplc-stag,: compression is achieved through the use o f multiple casing com-

p ~ r s s o ~ w>its with a common drive shaft. Thc overall ratio of compression within

two-stag? cornplussor arangements 1s typically 10.

External intercooling between stages is often more power efficient than direct

njcction of a coolant into the compression chamber. The ho~sepowet requirements of rotary screw compressors are fundamentally

dependent on features that vary from design to design. These features include lobe

clearances, rotor cross-section designs, and liquid injection capabilities. For this

leason, the ielative efficiencies o f these compressors varies widely between styles.

However, as a first approximation i t may be assumed that a dry rotary screw complt:ssm hschdrging above 20 ps~g will have a power efficiency comparable to

that ot ii celltlifugal compressor operating at its design point.

In general, as the ratio of volume flow rate to clearance area increases, the

c t f~cency of 3 rotary screw mach~ne rises. Thus these compressors tend to be more

r f t c e n t :it h ~ g h e ~ volumetuc flows.

Cooling Water

Exr,,pt 1 0 I,>w ~ ~ I I ~ W J I ~ ~ I C ,>I vwy Imv ,:>!to X J W W S , type compressors will

I I Y / L W ! ~ , coi)I~~,! j ~ ~ 1 1 0 1 to i ~ , a ! i t . i i ~ ~ W , I ! O I I ~ t c l~~~ l )~ :~a tu res 11~11xigho~il the casing foi

din?e~,sio~~;ll st jbihty. Thc bt:;11,t,g a n d ye,,! l uh~ca t ion systums may also have heat

cxchanguis that !eqoirc cooling water.

The effect of jacket watel cooling o r the theimodynamic performance of a

rotary compressor is usually negligible. However, the use of cool water in the

jackets, and consequential condensation within the compressor. IS less of a hararu

in these machines than i t is in reciprocating cylinders. The compressor manuiac

turer should be consulted for temperature limitations.

In general, cooling water f low rate requirements rangc from abproximately 1 l o

9 gpm. A typical cooling water discharye temperature is 1601'F.

Performance Quotations

Since the volumetric efficicricics of rotary screw cornprcssors art! not caidy calculated by the process engineer, the relative volumetric displacemettt is of littl i.

value as a basis of comparison between competitive units. Vendor performancii

quotations are essential in the early stages o f design.

Performance guarantees for domestic designs are most oftun limited to i 40:, I , !

capacity and horsepower.

Noise

Rotary screw compressors discharge gas in pulses. The number of pulses pel

second and the energy dissipated to the surroundings with each pulse often pro.

duces objectionable noise within the audible range. Most machines can exceed 103

dB in normal use. Some generate much less sound, and some installatio~x have required l i t t le noise control equipment.

If noise is seen as a problem but the gas being compressed contains materials that

can foul the complex internal surfaces of commercial silencers, then altrrnativr

control methods must be used. These alternatives may include the use of an acous tic enclosure surrounding the installation, or personnel access restrictions.

The proper selection of compressor silencers and noise abatement enclosures i i

part of the skill of most application engineers who are concerned wtth rotary screw

compressors as a product line. I t is usually wise to allow the vendor to spec~ty arid

supply the necessary silencing equipment.

Special Precautions

Rotary screw compressors can operate as expanders. A discharge check valvf. ! ,

recommended to prevent reverse rotation after shutdown, If two 01 more corn

pressors are installed in a parallel arrangement. a means of isolatng the discharge o !

each machine is necessary to preverlt possible damage to any compressor. 'ind I:.

prevent the back f low o f pressurizeti process gas.

Unresrrined reverse rotation posi!s thi? danger of rxcwdinq the rotatvc t p .;l,t:s:!i

l imit of the driver, or reversing ttlc loadincj diructlon 0 1 7 thi! thr~jsr bi!arngs n! i l ,dv ing the compressor shafts.

When simple check valves cannot be used because of toulng problems, thr:,~

valves with powered operators and suitable controls may substitute .

Cdpmty Control

Dry rotdly sclew compressors are best applied as base load units. The common collstatir.si)ocd methods of capacity conrrol tend to be relatively inefficient f rom the srandpwnr o f power sawngs.

V,~~i.~l>lt. spw<i c o n t ~ n l is possible with adjustable speed drivers such as engines am1 t o i b ~ ~ ~ ~ ' ~ . how eve^, the rate of gas slip becomes more significant as the inlet

capacity decreases. The lower l imi t of this method of control i s reached when the

volume ol gas slipping back to the inlet is equal to the forward rate, or when the dischat ge t<:mperature l imi t is reached. In general, these compressors rarely operate 31 lg:ss t l ~ ~ i u ~ 50°K o f des!gn speed. Nevertheless, va~iable speed control i s easily the

rnost cf l ic~t!nt ,means of capacity reduction. T l i ~ i ~ t t i ~ t i g t h ~ itilt:t of a dry type compressor usually does not result in a sub-

sl,iiil~,iI ~~ow.vii) s:iving Mo~iiovilr, thrott l ing thu inlet may result in excessively high ii1sc11.11qc l<!lTlpvl3tUies or ilndcceptable differential-pressure loading. This method of capacliv cont\ol is less preferable than gas bypass or speed reduction.

Alrhou!lI? rt,toliilng part o f the dischat.De stlearn back to the compressor inlet olli:is 110 POWLII sawng i i s a ,method of capacity control. the practice does permit I ~ r t t c ! ~ c < m t ~ o 01 Iicat I x i l d ~ i p . I f cneccssary, a bypass-gas heat exchanger may be 111sI.1ll~'~l lii II,IIIOY~, tI11, lht!;il (11 i : ~ ~ m p r t : s s i ~ ~ ~ .

011 Floodcd Rotiwy Screw Compressors

I v I : I S : I s I ~~ l j uc l i og 8 lubricatinq oil i l ~ i ~ c t l y intc th,! comprossio~> chamber. Thjs is done for three principal reasons: (1) Thr clll .~ISLI ,IC~S a s a coding agent ~ l l o w i n t l the compressol t o operate with low

~11scli.11gti tivnpm~itu,~:s. 121 In many units, the presf~ice of a lubricant also allows I v t n v loto! I n d l v v the secodary rotot directly within the compression I s (31 Thu oil x i s to seal the c1t:arance space between rotors and casing tlimi-by reducing gas slip loss. These particula~. designs do not require timing gears tc h c ~ y lhi, ~ o l o ! s f ~ o m tnakitq contact wfth each other.

C;iPdcit~cs o f oil flooded designs vaiy from 7% to 700 horsepower, with

discharge pressures up to 150 psig for air applications, and approximately 300 psig in some process units. The largest standard compressors are capable of handling more thdn 3,000 CFM o f air.

Rord11~1' sveeds vary but are always slower than comparably sized dry type

~ o t ~ ~ y SCII'!Y designs. In general, larger units operate at 1,800 RPM and smaller dw!jiis i t in . i t 3.600 RPM. Approximarcly 5,000 RPM i s a typical upper limit.

0pt1in;il tip rw!eds arc! in the range of 65 to 115 ftlsec. Below thisspeedrange, tht, i . 1 1 ~ of g.is slip p x t the rotors, which tcnds to bo constant, becomes significant t i1 cui11p;it tsw> 10 the volume o l gas be~rrg compressed. A t speeds above the optional range, dy~iarnic losses adversely affect the operating efficiency. These include h y d ~ a u l ~ c IOSSI?S introduced as a result of oi l injection.

The majority of oi l injected rotary screw compressors are supplied as part of ati package units. However, these compressors may also be considered for various other services where oil-gas contact i s not objectionable. These include the compression of ammonia and fluorocarbons in refrigeration systems, and the compression of inert gasses, such as nitrogen, for process or uti l i ty service.

Oil flooded compressors tend to be most competitive in the horsepowel ranges

between the very small (less than 10 hp) units where air cooled reciprocating umts

< dominate, and the low ranges o f the centrifugal compressors (approximatcly 500 hp) which dominate the large.capacity field.


The oil injected into the compression chamber absorbs much of the heat o f Compression. This resultsin a compression cycle that i s closer to isothermal than adiabatic. The heat transfer process i s so complete that for all praaical purposes oil and gas are discharged at a single temperature. Usually, this temperature is'well below 2 0 0 ' ~ .

Oil i s cooled to approximately 140" F before i t is reinjected into the compiei. sor. Lower temperatures would present a risk of moisture condensation anrl coiisi,. quenr oil contamination. Therefore, automatic oil temperature controls arc most often supplied as standard package equipment.

The relatively low discharge temperatures generated by thew compressors piit

mils them to apcrato at high ratios o f complesslojl. Most st;inda~d rljl l loo<lirl

machines are capable of discharging to 125 psig from atmospher~c air IIII:~. a single stage. Some are capable of still higher ratios.

Variances in design prevent any precise and universal correlations for predlctiri(1 horsepower from gas service data. However, single stage oil flooded rotary s c r w

1 compressors for air service require approximately 20 hp for each 100 SCFM o f a i l

compressed to 110 psig. In general, an oil flooded screw compressor will rcq~iiri!

slightly less than 110% of the power supplied to a two-stage water cooled recipro~ cating compressor in the same service.

Two-stage oil flooded compressors may offer some improved power efficiency over their single-stage counterparts. However, they are applied much less often than

the simpler and less costly single stage machines.

( Oi l flooded rotary screw comprssor units incorporate oil separation devices

which reduce the carry-over of oil within the discharge stream down to 8-10 ppm

! by weight, in most cases. Dual demisters are able to reduce this concentration to 2-5 ppm. and specialized units claim removal down t o less than 1 ppm.

Special Considerations

Rotary screw compressors will act as expanders. In order to prevent back llrjw through these machines after shutdown, a check valve i s commonly nstalled n thi:

uiiolioy o i [ l i e yas. Both liquid and gas phases immerge from the discharge separator

d l a cornmoo temperature (See Figure 6-79). Th~s type of complessor should be considered for services involving:

1. Wet Gasses. Slhgs o i liquid wil l n~l i ther damagqe inor greatly upset the performance of these

machines.

2. Temperatule Limitations. GJS IemporatLire (luring compression may be closely controlled by maintaining

1111. tolnpwattlre of the l iquid r'ng medium. Continuous recirculation of the medium !h<ough nc external hsat exchanger i s often used.

3. G;;sstrs with Dust. Tht! ~pinciple o f operallon does not depend on large solid sealing surfaces with

c i i~ i : f t~ I toIel:rnces. Moreover, the liquid within the compressor acts as a scrubbing

mrdwm during r:ornpression. 4. V.ipol Reuove~ y .

T h w cumplcssors can serve the function of gas separators by condensing satura- t ~ i i components out of a gas stream while passing the inon-condensibles.

5. Cor~<~r iv i : Gossrs. The, ii(loi(i m ~ d i u r n coats most of the internal surfaces o f the compressor. For

I l l i m i i t ~ g .I compl-rssol b u ~ l t with standad iron construction with an .ilh,,hnc, hquol allows the mdclnine to handle gasses with acid vapors.

6 . Lirnnitud Allow;iblu Down Ton,:. Thc s m p l ~ c i l y of these condensers is relfected in minimal servicing and down

lime.


L i q ~ l ~ d l ing compressors require a constant supply of liquid. The liquid com-

piessant may be supplied i n a once-through piping arrangement, or it may be ~ec~rculated from the discharge separator, through a heat exchanger, and back to the itllct. 111 the recirculation arrangement, the compressor itself may provide the

pumpiiig actioil needed t o move the liquid. (See Figure 6-20). In virtually every case, a discharge liquid-gas separator i s necessary. This is often

plovidrd 35 a sta~xlnrd compressor accessory. Liquid l ing compressors rend to be less power efficient tharother designs. Their

I I)"! l~>irn.itncc ch;liaclr~istics v;ir.y with the conditions of service. Standard

rl<'t lt>w?,.i!,i:t. c i . ? ! . ~ i s ~IIIII IsI~L'~ lo! .iir scrvicc us~ng water as a liquid compressant. TI>,, colnptmsoi c.ipaclry a i d lhorscpowei ior most se~vices can be approximated hni 11-11. st;intIiliti t:~bl<,s by using an ~!qwvale!~,t CFM with a calculated allowance k t i ~ u v . i p o ~ i . ~ . ~ t i i i ~ ~ 01 rhr cornplwalit . More precise predictions of performance

To! yasscs other than ail or a liquid compressant other than water wil l usually teqiri~e considi:~ation of the following:

1 ) Lj r l t~ id c imp~msainl specific g~avity, viscosity, specific heat and vapor pres- s,,,,,.

i Fiyure6.20. Instrument air compressor packaye. /Courresy Nash Engineermy Co.).

( 2 ) Solubility of the gas in the liquid compressant. 3) Effect of condensibles in the gas stream on liquid compressant characterist~cs.

Manufacturers' emperical data describing the effect of these factors on'comprrsso,

performance i s useful when available. The method o f compression and the porting arrangmentr designed into thcsc

compressors permit compression with a near absence of pulsation. Multiple stage, single casing liquid ring compressors are available for incraasc?d

1 ratios of compression.

Construction

Cast iron is thu most common material used in the bodies o f liquid ring conr(>ri:r.

soil.. Bronze may be used for internal fittings, and in some units, for rotors aitci cones. All iron construction i s also available.

Special corrosion resistant alloy construciton may also be supplied. Two styles predominate:

Thc sccei,riic lobe provides one compression cycle per chamber per revolution. This design Imposes lateral forces on the rotor shaft as compression occurs. I t i s

theletore l imited to the lower pressure ranges (approximately 20 psig and under). The eccentric lobe design tends t o be more power efficient than the alternative double lobe design, and is therefore the preferred style i n the higher volume ranges.

Thc: (/oi,hlr lobe design has two cycles of compression per chamber per rev0lu- ~ l i > ~ i . F O I C P ~ on the shstt arc: balance(l, permitt ing higher ratios of compression than

the eccentric lobe design (See Figures 6-21 and 6.22).

BODY

Capacity Control

Dtschalge presssure control and capacity control is most of ten accomplished t l ~ i o i ~ g l i an external bypass line. An alternative method o f constant speed control

uscd wit17 some imi ts consists of automatically dlaining part of the l iqu id "com- / ) I C S W I I ' ' wll ich allows the units t o operate unloaded for a l imi ted period while L i l i l s r ' i V i l l g pow,!,

Convcntonal s ta l t~s top control may also be used.

Alternative Designs

Ssvetal orher compressor styles contr ibute their particular advantages t o the fleid o f rotary compressors.

The sl!iiii,g vane is among the oldest and most reliable of rotary models.

Figure 6-22. Cornpor~enri of duiible lohe derwi liquid r ing coirrpn:rror. iCuuilr.,y N.,,/,

Ei iy i i i rerb~g Co.J.

Auxiliary package components are sim~ldr t o those used i n rotary screw units Single-stage jacket water-cooled vane type compressors rallge in capacity i rmv

33 t o 3.250 CFM and t o 50 psig i n discharge pressure. (See Figure 6.231. Two-stage water.coolecl units are available for capactties ranging f rom 115 tc.

approximately 3,000 CFM and pressures t o 125 ps~g.

Special three stage units may be supplied for pressures t o 250 psig. A l l jacket water cooled designs are force feed lubricated. Smaller oi l f looded vane type compressors are availble in capacities ranging troln

60 t o 600 CFM wi th discharge pressures in the range o f 80 t o 150 psig. The o ~ l flooded design reduces the temperatures o f the discharge stream to under 200°F.

Straight lobe t ype compressors are available i n capacities ranging f rom 5 to 30,000 corn and pressures t o 12 psig in single-stage operation. I n some servces. two-stage units may provide up t o 20 psig. (See Figure 6-24).

These compressors are more properly termed positive displacement blowers be^

cause they move quantities o f gas f rom one level of pressure t o another without an internal volume reduction mechanism.

INLET \ -+@

DISCHARGE

ACFM

BHP

C C A CE

C~ C" GHP H

Had

HP? hp k ME MW N n ns P PA

P C

PP PN

p, PI Q R

'c SCFM T

Tc Tr v VE W

y A &

%d

NOMENCLATURE

Actual, or measured cubic feet per minute. f t 3 m

Brake horsepower, hp

Cylinder clearance, in"ini Cylinder cross-sectional area, in2 Compression efficiency, dimensionless Heat capacity at constant pressure. Btu/lb.mole Heat capacity at constant volume, Btullb-mole " k Gas horsepower, hp Enthalpy, Btul lb Adiabatic head, ft-lbsllb Polytropid head, ft-lbsllb theoretical power. horsepower ratio o f heat capacities. cp/c,. dimensionless Mechanical efficiency, dimensionless Molecular weight, dimensionless Number of moles, Ib-mole Polytropic exponent, dimensionless Number o f stages, dimensionless Absolute pressure, psia. lbs/in2 Piston area, in2 Critical pressure, lbs/in2 Piston displacement, ft7/min Discharge pressure, lbs/in2 Reduced pressure, PIP,, lbs/inz Inlet pressure, Ibs/in2 Cubic feet of gas per minute, f t v m i n Ideal gas constant. 1595 f t lb l lb mole O R or 1.986 Btullb mole ' R Ratio of compression, pfinalIpinlr ial, dimensionless Standard cubic feet per minute at 14.7 psia and 6OUF, ft7/min Temperature. O R

Critical temperature. "R Reduced temperature, O R

Volume, ft' Volumetric efficiency, dimensionless Mass rate. lblmin Mole fraction of component A in a gas mixture, dimensionless Compressibility factor, dimensionless Adiabatic efficiency, dimensionless Polytropic efficiency, diverisionless

REFERENCES

PROCESS EQUIPMENT S E R I E S

Volume 3 Index

1 : I : E r e I l l F r l ~ h Ed.. R. H. Perry orid C. H. Chllron. editors. McGraw

l l ~ l l , 1913. pp. 3-232 illxi 3 2 3 3 5 ~ ~ , u ~ ~ r ~ ~ ~ ~ ~ ox2 B O ~ . G ~ S Procersorr supp~ierr Arrociation, ~ i n t h ed.. 1977. pp. 16-11. i

.>i>il 16-14

ii M I~c.,I,:~IIcI. G. 8. Webb. ,lnd L C. RuLII~. J. Chcm Phys.. 8, 334 119401. 1 0 l i ed l x l i and J , N. S. Kwong. Chem. Rev.. 44. 233 119491.

8 Cm>urrsseri Al l and Giir Handbook. 4 th ed.. j. P. Rollinr. ed.. Compressed Air and Gar

I 1 Ampllc.~n Petlulr i im ln r f i fu re Siimdiird 618. Second Ed.. July 1974, Sec. 2.7. 12. Appllc,ible rt.mdardr may be found i n Section V l l l of the ASME Boiler and Pressure Verrel

Code. the Amerrcan Perloleum lnmru te Standard 618. and i n ANSI 31.8.

13. G.ir Procesrorr Suppl~err Arrn. Eng~neering Data Book, Ninth Ed.. Third rev. 1977, pp. 4 1 7

j i

Acid vapors 303 Adiabatic and polytropic efficiency 102

efficiency. 99, 126 Aero-thermodynamics of centrifugal

compressors 98 Affinity law 16. 66 Aftercooling 7 Air handling curve for jet syphone 225

present in condensing water 233 Alarms 296 Altitude on pump performance 188 (AMCA) Standards 25 Angle type compressors 277 API Standard 671-Couplings 89 Application of centrifugal compressors

93 char for centrifugal compressors

98 Arrangements 97 Asymetric flexibility 144 Axial fans 15

flow compressors 67 fan 16 performance curve 69

thrust 18

Backward curved blades 15, 119 Backwardly inclined blade fan 14 Balancing 157 Barometric pressure vs, altitude 190 Bearing 23. 61. 147

stiffness 139 Bid tabulation 35 Blade angles 107

attachment 137 discharge angles 3

Blower 1 and mechanical pumps 185

Boundary layer development 118, 119

Boyles and Charles Law 191

Calculation of vapor pressure effecl 193

Capacity 42 control 287, 306 loads 220

of liquid ring pump 186, 188 Carbon ring seal 64. 65 Casing and foundation forces 141 Centrifugal compressor 45. 88

calculation 54 classification 94

fans 12

fan thrust 18 Centrifugal force 19

mach~nes 2 Check valve 209 Chemical cmrrosivencss 280

properties 196 Chlorine production 205 Cleanup of lubricant sysrcrn l t ib Clearance loss 129

r~ockets 78 Cl~mate 89 Code standards 263 Cold weather applications 89 Composition changes 297 Compressant selection 193 Compressibility 271

factors 41 Compression ratios 249 Compressor 2. 39

adiabatic efficiency 74. 75 application 39, 289 capacity 79 components 105, 283 cnnfiq~irarinn 95

D~a~rletcr change 16 varies 17

Diffusers 121, 218 D i f f ~ ~ s i o n blading loss 127 Dimensional analysisof a centrifugal

compressor 101 Directcontact condensers 227, 229,

23O.23l.25l. 254 Disc coupling 163 Discharge positions 30

pressure 297 temperature 7, 297

Disk friction 2 loss 127

Double acting compressor 177 f low impeller 56 lobe pumps 312 volute design 21

Driver 70. 298 selection 97

Drives on compressors 97 D l y triction whirl 146 Dyna~nic surge detection 134

type compressors 67

Eductor condenser 236 Effect of nozzle location 244 Efficieni:ies 9 9 Electtic motor drivers 70 Energy transfer 98 Entrainment ratio curves 223 Entropy-enthalpy diagram 100 E,ivironmunt 89 i s f s t 272 Eolw w w k dislribntion 112 Evaci~ationcurve for jet syphon 226

/priming time 228 Exit velocity 120 Explosive gases 181

F Guide vanes 23. 68

for compressors 67 Fan 1

and motor arrangements 26 H construction 23 controls 22 dynamics 17 laws 16 selection 38 vibrations 17 i n parallel 29 operation 27

Filtration 202 Flexible connector 25

coupling 162

j gear couplings 160 Flow analysis 112

characteristics 117 coefficient 48, 103 map 114 patterns involute 125 regions of teh vaned diffuser 122

Forces and moments 25 Forward curved blade 15

blade fan 14 Frame load 76

lubrication 290 ratings - reciprocating compres-

sors 74 Full recirculation seal system 199

Gas analysis 35, 40 composition 186 engines 80 laws 185 mixtures 272 piping 295

Gear 166 couplings 25

Graphical plot o f alignment 152

Head-capacity curve 4 coefficient 48 flow rate characteristics 107

Heat Exchanger Institute Standard 253, 263

of compression 288 Heated cylinder surfaces 288 Helical compressors 300 Helmholtz law 120 High speed balancing 158 Horsepower 6

calculation 6 determinations 282

Horsepower requirements 273 Hysteretic whirl 146

ldeal and actual heads 47 ldeal Gas Law 191 Immiscible seal liquid 193 Impeller 48. 57. 105. 108. 11 1 , 1 11.

136. 138 blades 113 blading 106 channel flow 116 design 7 plane 113 types 5 velocity diagram 46

Incidence loss 127 Induced heat load 182 Inducer 114

centrifugal compressor 116 inlet velocity diagrams 109

Inlet box 29 positions 26

guldu vanes 109 velocity triangles 128

lnst:jllat~on 206, 259, 300 01 electors 259

Instlil!ileol , 1 1 4 culrlpll!ssol pdckaqe 311 I I ~ I ~ I , : ~ ~ ~ ~ ! ~ ~ I ~ L W S ~ U ~ ~ : ~ m ~ ~ x c s s u r GO, G l

L , I t ! I 154 Lr?:llM,k! I I 9 L l l : S S e r e 187

u ! ~ j l ypr t a r colnprrssors 308. 310 vacuum pump 83, 174 pr~riciplc 176

Lobe-type rotary blower 81 Low-level eductor 236, 237

pressure bloweis 9, 62 viscosity ot colnpressant 196

Lubrlcaiit cm,taliimatfon 164 SL'I~XIIOII 164

L1101 c . i l i ) ~ ~ 1 0 ?90 ivsrt,rlls 1G3

M~l i l len. l~ icr 215 Matcials lor falmcatii~g 138

of coostroction 290 Mechatiicul contx:t shilft seal 156

friction 288 packing rings 286

Methods of rating compressors 279 Misal~gnment 150, 151 M~scible seal l i q~ l i d 193 Mo~srure extr:xtion 202 Mdccular weight and ratio of speufic

heats 40 entrainment 222

Mollif! <liilgr;un 249 Motive pressure 246

rate difference 247 steam consumption 226

Motor driven lubrication 292 positions 26

Multi-jet condenser 239 )

direct contact condenser 230 eductor 236 1

Multiple cylinder arrangement 277 Multiple direct contact condenser 232

Multistage compressor 56. 60

horizontally split compres-

sor 59 jet ejectors 248 systems with surface con.

densers 255

Noise 17. 305. 308 Nozzle position upon jet efficiency

245

Oil-til leil inachine 83 floo<led compressors 307 1

I 306

ro1;lry screw compressors 306

whlp 20. 146 Once-thru-seal systcm 197

Open faced impeller 137 Operatingcharacteristics 304,307,310

pressure 186 speed 19

Operation of liquid ring pump 184 Outlet dampers 23

Parallel operation 29 Partial condensers 258

recirculation system 199 Performance calculations 43

characteristics of centrifugal compressors 126 a diffuser 122

curve 13, 14.50.51.234.235. 238,240, 250, 251, 255, 256 of liquid ring pump 179

guarantees 305 of a two-stage liquid ring

pump 185 Piping 296

arrangements 206 errors 208

Piston rods 285 Pollution 182 Polytropic efficiency 49. 101

head method 54 equation 5

Positive displacement compressors 71, 270

Power consumption 190 Pre-cooler 234 Pressure coefficients 103

differential 2 -enthalpy method 55 gages 167 range 2 ratios 91 volume curves 3

Prewhirl 110

angle 11 1 distribution pattern 11 1

Priming 204 Pulsation dampers 78. 295 Pump selection 186. 190

Radial blade fan 13

design 7 Reciprocating compressors71.72.73,

275 Recirculating loss 130 Recovery of process chemicals 256 Relation of altitude to barometric pres-

sure 89 Resonant vibration 143 Rod load 76

packings 286 Rotor losses 127 Rotary compressors 80, 300

calculations 83 lobe compressor 314

type blower calculation 84 Rotation and discharge orientation 26 Rotor bearing system 140, 142. 143

instabilities 141 dynamics 138 motion 141 response curve 139. 160

Roulte's law 191

Saturated gas 187 Schematic of a centrifugal compressor

104 Screw compressors 301

type rotary compressor 82 Scroll 123 Seal 62. 153

oil 63

Suction loads 220, 252 capacity 243

pressure 221, 246 Surge 132

bottles 296 control 168 detection 133

System curve 32 resistance 31

Temperature entrainment ratio curve 22 1

rise 193 Terminology for centrifugal fan com- i

ponents 36 Testing 158 1 Thermal efiiclencies 238

effii:iency of vacuum putrips 182 I ' l i ~ ! r m ~ ~ i l y ~ ~ . ~ n i i c s 98

d1.11actu~ istics 185 cy,:I,: 1 0 StU.,," ,I!I.~B~IIIIII~

pump 241 of gas comflession 270

3 Throttling 306 Thrust bearings 149. 150 Tip speeds 306

i j Trouble shooting 259 Tube-axial design 15

fans 12 i 1

Tubular Exchanger Manufacturers As- sociation 264

Turbine drivers 70 j I

Turbomachinery 1

Types of compressors 90

fans 12

Typical jet ejectors 217

i i !

U 1

Ut i l i ty consumption 216

Vacuum 218

pump 199 relationships 220 technology 174

Valves 285, 286 Vaneaxial design 15

fans 12 thickness 119

Vaned diffuser loss 130 Vaneless diffuser loss 130 Vanes 107 Vapor pressure 194 Variable speed control 306

drivers 80 Velocity distributions 113

profiles 115 triangle 108

Vibration 19, 133, 143 stilnuli 144

Volute 123 desiq,~ 19

Wake mixing loss 130 Water conservation 201

cooled cylinders 289 jet exhauster 223, 224 piping 208 requirements 257

Wet-screw 83 Wheel friction 2

Documents

- Process Eqpt Series Volume 3 by KS Panesar-1