CHAPTER 1

F A N S AND BLOWERSK. S. PANESAR Houston. TXINTRODUCTION Pumps, compressors, blowers and fans all belong to the same family of niacl1: called TURBOMACHINERY or ROTATING EQUIPMENT. Pumps can handle I ) ! incompressiblefluidsviz. liquids while compressors, blowers and fans, on the oti 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 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 aositive dirolacement machines. The oerformance characteristics of both are sho in Figure 1.1.

Ij

J

i3

0

l0

I

I100

50 CAPACITY (%)

Figure I . 7. Pressure-volume diagram for rhe cenrrifugal and rhe positive displacement rypes o f machines. Capacity is urually specified in terms of rhe inlet C F M /at rhe inlet condirionr) and the head in fee1 or pressure in PSI for compressors and blowers. For fans, however, rhe pressure is usually specified 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. A l l these machine5 can handle fluid flows a ranging l r o ~ n few hundred cubic feet pel- minute t o several m i l l i o n 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 r o m a fraction 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 r o m 2 Psi t o about 1 5 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 t w o or three cases, each case having several stages. FUNDAMENTALS Most centrifugal {machines have a housing w i t h an inlet and an outlet. Inside the housing is a wheel or an impeller which rotates and imparts kinetic energy t o the fluid. The f l u i d comes i n at a pressure. P , . and leaves the housing at a higher Ipiessilre, P-. . These inacliines will, therefore, always generate a constant pressure differentiill whet1 operating at the same f l o w . Thaovetic.?lly, the pressrlte~volurneline is supposed t o be a straight line, when lhc! i l i s ~ : l i ; ~ i q !angle o n the I~ladesis 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 prxtici:, however, this is never the case. They have a curvature t o them as already shown i n Figure 1.1. The reason f o r this is that there are certain losses within the housing: - disk f r i c t i o n (wheel friction). blade inefficiency, circulation w i t l i i n the bl.ides. cfc. Tile k i n e t x energy is generated b y the rota1.y m o t i o n o f the impeller and is irnpa~teclt o the f l u i d moving through the machine. Each machine, in fact, each ~ m p l l e r designed t o produce a certain pressure rise or (sometimes called "Head") is at a given capacity w i t h a m i n i ~ n u i nloss or, in other words, w i t h a m a x i m u m efficiency, as shown in Figure 1.3. I n oldel to undeustand h o w the head is generated b y an impeller, let us look at Figure 1.4. F g o ~ e . 4 shows a typical sectional elevation view o f hydraulic path o f 1 ~ an ~ ~ n p e I i tw lii.e r ~ d s ? Figure 1 4 b shows the same impeller in plan view w i t h inlet and discharoe v d o c ~ t v i ;males. t 'dt,' L e i 0 s . K S L I ~ L ' tIi;il ill t ~ ~ n e there i s a Inass 'dm' o f very thin layer of f l u i d j;i!i, g.3~ or ltiliwcl) lc~iwfngthc impellev and a t the salne time 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 the f l u i d contained between the t w o on blades, Let T denote the m o m e n t o f external forces. Then T is given b y the following equation:

and when the term ( d m l d t ) is applied t o all the fluid in all the blades, i t becomes Qplg; where the f l o w 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 fluid at the (lischarge: the absolute velocity o f the f l u i d 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; Now, substituting f o r ( d m l d t ) , equation number 1 becomes:

i

CAPACITYFigure 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 t o calculate the theoretical or actual brake horsepower as shown: THP == w Hp/33000 (7) Just divide the THP by the polytropic efficiency t o get the BHP. In fans, however. the horsepower is calculated i n a little b i t different form, which is shown below: BHP = ,000157 X ACFM X S.P./S.E.

iwhere

or BHP= ACFM X S.P.16356 X S.E. I n fans the discharge temperature i s hardly ever a problem because they are a very low pressure ratio machine. I n 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. T o calculate the discharge temperature, the following equations are used:

wlip

-

THP and BHP are theoretical and brake horsepowers respectively. flow i n Ibs per minute, and polytrop~c efficiency i n decimal.

There is another way of calculating the theoretical or actual horsepower when you do not know or d o not have t o know the polytropic head. The calculation of head equation is very useful i n centrifugal and axial machinery because from this you call deterniine the number of stages or wheels required t o get the desired pressure o i head. For reciprocating cornpressots, o n the other hand, you do not need to r.;ili:oliilu the lhcxi. There co~npressio~i achieved in a cylinder, and if the is III,:l,,!s O t "";,tt31 gage. c. Class Ill -This is a heavier frame and the pressures are limited up t o about 12 to 13 inches of water gage.Figure 1.15. AMCA Standard 2407-66, showing rhe different motor 00s;tions with respect to the fan.

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Fans Operation Once you understand the fan fundamentals and have some knowledge of the system i n 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 a close to its maximum s or best efficiency point a possible. I f a fan i s selected t o the right of its best s 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 its 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 t o operate in a system shown in Figure 1.19a. The fan selection for systems shown i n 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-

3. Motor PositionsThe AMCA also specifies the motor position with respect t o fan a shown below. s 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 w ~ r l the f o l l u w i ~ ~ g l arrangements (see Figure 1.161.

I I ~

5. Inlet Box Positions

Thr AMCA Standard No. 2405.66. has the following positions standardized for tht3 ~ i ~ l box ( S L W F i w r e 1.17). ct6. 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.

I'rocess I:ijuip~ne,ilSeries Vulumr

3

Right t o p angular i n t a k e

Horizontal r i g h t intake

Right bottom angular i n t a k e

0ot t b m intake

L e f t bottoin angular i n t a k e

liorizonral l e f t inrake

L e f t top angular intaue

TOP

i rnta-v

Figure 1.17. Showing the different positions for the inlet box. are at 45-degree angle.A R l i 1 S W S i u ,llluct i i r U ill,l ,,r" ,I,,,,,,> o,, D,,"'. ,n ,,,, N, b , . . , # ~ ! , , > > I,,,, P r # m e "C, n , " . , , u*ru i . i i , i , l i l ? * "8 imcgr.my

The angularposirionr show,

,.,,..,

,r.c,

I\Rn I S W S Ii",1,1.,.1,0,,

Foi L l U I d r l w 0 6,.A

,>,",".>>,:~",,>4,,,,u,,~~,",> :n::ia!,17r.?es13-lt

. . .

1 . 5011-7, 5 0 0

. /.- I . .'1-.,.1 . - ,! t 1

~ . o o o - ~ ~ . ~ ~ ot .C00-1E.Ooo~,,cw>-j~j.ooo 70.11W0-t,0. 0 0 0 h0.U00-100.000

to 0 . 5 2t o 0.5.

0.57-torJ.5.j

0.51

39- I:,.jt

L C 0.56

.!..'

0 . 5 4 to 0 . 5 5

,L [., . , :

,

Figure 2.5. Prelimii>ery selection data for cenriiluyai comirierrois.

:

H w d pcr stage (actual)

.g

u;~11111 I TB

(281

f l o w c u e f f c ~ e nmay be determined at the inlet or exit o f each impeller. A t t uthsl thdil the d r s g n rated condition, similar valttes o f and qI can be found. ~ c s u l t ~ n g the display o f a given stage or an overall machine performance curve i n in several manners (Figure 2 . 7 ) , showing f l o w versus head or f l o w versus pressure, both f o r one single speed, also f l o w coefficient versus head coefficient (dimensionlc:ssi ~ v l i ! c h 1s LISUIUI in generating performance curves at varying speeds or o n ;iItt!~ n:iw g.is c o n d i t i o ~ i sin the same compressor. These curves also show efficiency, : I I I s s i t l t 1111~1 exit flow angles f o , off-design i117dtiiiii, ~ i ~ t i f n u i b l, i l ~ !~ ~ i c i d ~ i n I ~!S S C S na ct higher, resulting n v,iy11!1ilvu1.1l1 ! l f ~ c ~ u n cthrouglioor thi! stable apelatin(] riinge. c y F i , b.~chwti~il-lu:iz,~n

Chapter 2. Richard F. Neerkenvariable, shutting o f f a reactor or other process item downstream o f a compressor can cause the pressure t o rise t o 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 will 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 t o 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, b u t are wasteful o f power. More economical c o ~ ~ t rmethods include the use of cylinder unloaders and clearance pockets. ol Cylinder unloaders are manually or automatically operated devices on one or both cnds o f a cylinder, designed t o unload or h o l d 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 ~rle d ~ c i r q , the net output o f that cylinder b y approximately onehalf when the ilcvices wele actuated Unloaders are usually supplied as a means of totally I ling ~ ~ ~ ~ l o . i;~compressor for startup, but prolonged operation with unloaders o n one elld o f a cylinder may cause problems o f frame load, valve life, o r pulsation damping. Clea~ancepockets are additional volumes of clearance built i n t o or bolted o n t o ; cylinder head or valve cap, either head end or crank end f o r both) t o increase the I cylinder clearance. Reference t o equation 32 w i 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 w i t h b o t h automat i c d l y operated unloaders and a manually operated valve-cap clearance pocket. Retlief valves must always be provided i n the piping immediately downstream o f any rec~urocatingcompressor. They should be sized for the full output capacity of the compressor, set t o 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 flow generated by the reciprocating Illston. Thrsr (nay be i n the form of volume bottles or special devices w i t h internals designed to absorb or cancel some of the pulsing flow. It i s common t o specify that the pulsations i n the piping leading t o or from a compressor shall not exceed 1% of the opelatiog pressure i n pressures up t o about 400 psig; lower values are expected and required i n pressures higher than this. These devices require a pressure drop as the gas passes through: careful calculations for sizing reciprocating compressors w i l l ~,cludr ;~lli>w:mcus such pressure losses. for

1

Process Equipment Series Volume 3

1

Figure 2.30. Variation in campressor capacity with increased clearance.

Var~ableSpeed 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 engineC O I ~ ~ ~ S S O I . wlitxc

Lobe-type 12-lobe or 3-lobe1 Vane-type Screw-type [wet-screw or dry-screw) Liquid-ring type

i h c 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.

Figure 2.33. Lobe-type rotary blower.figure

2.32. Gas e t w n e driven

recipra'ating comprerror

Steam turbines have been applied to reciprocating compressors by use o f single01 double-reduction gears t o reduce turbine speed. Although many successful insr,illations have been made, caution should be used when specifying this type drive 31~3ngZlnt?llt cons~derableadditional analysis of torsional vibrations, couplings, as ti l l y w l i ~ d zS i c c j ~ l ~ r e d . I

Most widely known is the lobe-type (Figure 2.33). Two figure- shaped I c r g equ:ltion: y

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

Thus the overall adiabatic efficiency relations hi^: '

in impeller is given b y the following

The individual losses can n o w 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 following i s the analysis o f each of these losses. 1. Shock in R o t o r 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 l o w shock which would cause the f l o w t o detach from the blade wall and the loss to bc much higher. 2 . Incidence Loss A t o f f design conditions, flow 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 w i t h incidence suffers an instantaneous change of velocity at the blade irilrit t o comply w i t h the blade inlet angle. Separation o f the blade also creates a loss associated w i t h this phenomenon. 3. Disk F r i c t i o n Loss - This is the loss due t o the frictional torque on the back surface o f the r o t o r 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 i n f l o w compressor, or a radial inflow turbine. I n 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! yiil, is o f the order o f magnitude o f the boundary layer, the effect of the gap p ? t ! 1,. negligible. A p o i n 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 - T h i s loss arises because of negative velocity gradvr:t\ i n the boundary layer. This deceleration of the flow 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.

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

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Thai x ~ a b a t i chead that i s actually avallable at the rotor discharge is equal t o the tliI, the blade loadings. (Aq,,, I the clearance (Aq,) between the r o t o r and the shroud. and the viscous losses @qSf) encountered in the flow passage.

Therefore, the adiabatic efficiency i n the impeller is:

I ,lit ot the over-all stage efficiency must also include the losses I,, r ~ , c i , u n l r ~ e d the dlffuset. Thus, the overall actual adiabatic head attained w o u l d be l h c actual j d ~ a b i i t i chead o i the impeller minus the head losses encountered i n tllc d f l u s u r duc t o wake caused b y 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

rulccs lossis

01,

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

5. Exit Loss

- The exit loss assumed that one half of the kinetic energy leaving

i r s u ~ l l ydemmined by considering the f l o w a an equivalent circular cross s

the vaned diffuser is lost. Losses are a complex phenomena and a discussed are a function of many s parameters such a inlet conditions, pressure ratios, blade angles, flow etc. Figure s 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

sectioti wiih a hydraulic diameter. The loss is then computed based on the well known 1)ipr flow 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 l o w angle at the exit of the impeller a seen in Figure 3.38. Part of the fluid is recirculated from s 11111 illu use^ to the tmpeller and its energy is returned to the impeller.

I,>,,

'8,)

Figurc 3.38. Recirci,/ating loss.

Flgure 3.39. Losses in a centrifugal compressor.

2. Wake Mixing Loss - Thls loss i s duc to the impeller blades, causing a wake inLxhuid the l o t o r . This loss is m ~ n i m i r e di n a diffuser which i s s y n i ~ i w i~ J l O u l l d the axis of rotation. t~ 3. Vaneless Diffuser Loss -This loss i s experienced in the vaneless diffuser due t o rlli f r i c I ~ w aid the absolute flow 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 u s1.11o Thcy dso take into x c o u n t the blade ~ncidenceangle and the skin1 I i ~ .v;lnelrsssi);ice

Performance Characteristics

-&

11 c t ~ o iduc I n the vanes. l

A Plot showing the variation of total pressure ratio across acompressoriis a function of the mass flow 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. m l The actual mass flow rates and speeds are corrected by factor ( h and i l l respectively, i n order t o take into account the variation in the inlet c o n r j ~ t o r t i i of temperature and pressure. The surye line is the line which joins the po81~tir r ~ different speed lines where the compressor's operation begins to be unstabli!. A

compressor i s said t o be in surge when the main flow 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 i n back pressure causing the main flow to reverse again. This unsteady process, i f allowed to persist, may result i n irreparabit? dxn.qc to the machine. Lines of constant adiabatic efficiency (sometimes called the efflctency islands) are also plotted o n the compressor map. A condition known as choke o i commonly known as "Stonewall'' is indicated on the map which shows the maximutn mass flow rate possible through the compresor at that operating speed.Surge

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 t o change with flow 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; v e t , therr have been cases i n 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 Cvcl!c forcer from propclle~, u! fan. S e l k x c i ~ r d e m n q lorct";. b Pneumatic hammer. (Conciudcdi

Gravity Magnetic field, stationary or rotating Axial forcer

Dercription

Application

netic field Imprerred cyclic ground-or foundation-motion Air blarr, explorion or earthquake. Nearly unblanced machinery. Blows, impact Present in all rotatrng machinery. Mot~on around c u r v e o l varying radtus. Soace ap~llicalionr. Rotarycoordinufed unalynes. P ~ ~ w e ro y r o w r material which lf 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 . C r n ~ s t ! u c t ~ o n m p l n g a r ~ r i n gr o m da f ~cllative o t i o n between shrunk m l i t t e d arremblier. Dry-fricrion bearing whirl. Vircour shear of bearings. Fluid entrainment i n turbomachinery. Windage. Bearing load capacity Bear~ng rtiffnerr and damping properlies. Rotors with differing rotor lateral rliffnerrer Slotted rotors, electrical machinery, Keyway Abrupt meed change condltionr Significant i n 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 w i t h indexing or positioning errors Drive gear forcer

Table 3.6. Characteristics of Forced and Self Excited Vibration Forced or Resonant Vibration FrequencyIRPM Relilfionshlp Ampl$tudc/RPM Relatronrh~p I n 1 o f Damping Self Excited or Instability Vibration

' N F NRpM 0 N o r rational fraction

-

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 ~ l ;# ir hrgt~c:r RPM. W$llno! rrm~c:ts,8il'~ affect amplitude Independen~lyo f symmetry s,n;iIl dellsct8n t o m a x ( w m e t r < c system. Amplitude w d sell prapagare. same 1. Operating RPM below onset.

System Geometry

Peak in narrow lbanrlr 01 RPM Add. dampin