The Design and Perform ance of an
A xial-Flow Fan
AER-56-13
LIONEL S. MARKS1 a n d JOHN R. W ESKE,2 CAMBRIDGE, MASS.
T h is paper d eals w ith th e d esig n a n d p erfo rm a n ce o f a n
axial-flow fa n for co m p arativ ely h ig h p ressu res. T h e
d esig n is based largely o n ex ten siv e in v estig a tio n s o f th e air
flow th rou g h a fa n o f w ell-k n o w n d esig n w h ich y ield ed cer­
ta in co n sta n ts. T h e p roced ure in d esig n is sk etch ed very
briefly— it in volves th e u se o f b o th a irfoil th eo ry an d
circu latio n th eory. F u ll d eta ils are g iv en o f th e co m p leted
fan . T h e p erform an ce o f th is fa n is sh o w n in a series o f
graphs an d is com p ared w ith th a t o f th e fa n u sed for th e
p relim inary in v estig a tio n s. T h e in flu en ce o f th e n u m b er
and lo ca tio n o f g u id e v an es w as in v estig a ted . D iffu ser
a ctio n is a lso d iscu ssed . T h e n o ise e m issio n w as m ea su red
and w as fo u n d to be co n sid era b ly low er th a n th a t fro m th e
origin al fa n . A rela tio n b etw e en n o ise a n d fa n p erform ­
an ce is p o in ted o u t.
T
HIS PAPER deals with the design and performance of an
axial-flow fan for comparatively high pressures. It was
hoped that some improvement in efficiency over values
previously recorded might be obtained by giving careful
sideration to the aerodynamic principles involved, including
both airfoil theory and circulation theory.
To analyze fully the operation of an axial-flow fan, it is neces­
sary to have knowledge of the pressure, direction, and velocity
of the air in every part of the fan while in operation. To obtain
these data a fan was built, following closely a design which has
given good performance, and an extensive investigation was
made of the air flow through this fan. The details of this in­
vestigation and its results would require too much space to be
1 Professor of M echanical E ngineering, H a rv a rd U niversity, C am ­
bridge, M ass. M em . A .S .M .E . Professor M arks w as b o rn in B ir­
m ingham , E ngland. H e received th e degree of B.Sc. from th e U ni­
versity of L ondon in 1892 and M .M .E . from C ornell U n iv ersity in
1894. H e w as w ith th e Am es Iro n W orks, Oswego, N . Y ., in 1894 and
th en w ent to H a rv a rd U n iversity as in stru cto r in m echanical engineer­
ing. In 1900 he w as m ade a ssistan t professor and in 1909 w as ad­
vanced to his present position. Professor M ark s is a u th o r of “ S team
T ables and D iagram s,” “ G as and Oil E n gines,” “ M echanical E ngi­
neers’ H an d book,” “ T he A irplane E n gine,” and has co n trib u ted
num erous articles to th e technical press.
2 B ethlehem Shipbuilding C orporation, Q uincy, M ass. M r. W eske
entered th e H an o v er In stitu te of T echnology in 1920 and in 1923
was g raduated w ith th e degree of D iplom Ingenieur in m echanical
engineering. F rom 1924 u n til 1930 he w as engaged in m echanical
engineering and design w ork w ith several in d u stria l concerns. T hese
included D eutsche Schiffs u n d M aschinenbau A. G ., B rem en, G er­
m any; several firm s in D e tro it and San F rancisco; a n d th e tu rb in eengineering d ep artm en t, G eneral E lectric C om pany, Schenectady,
N. Y. Since 1930 M r. W eske has been in te rm itte n tly w ith th e
B ethlehem Shipbuilding C orp., Q uincy, M ass. F ro m 1931 to 1934
he has been engaged in g rad u ate studies and research a t th e H a rv ard
E ngineering School, receiving in 1932 th e degree of M .S. and in 1934
th a t of S.D .
C on trib u ted b y th e A eronautics D ivision for p re se n ta tio n a t th e
A nnual M eeting, N ew Y ork, N . Y ., D ecem ber 3 to 7 , 1 9 3 4 , of T h e
A m e r ic a n S o c ie t y
of
M
e c h a n ic a l
E n g in e e r s .
D iscussion of this p ap er should be addressed to th e Secretary,
A .S .M .E ., 29 W est 3 9th S treet, N ew Y ork, N . Y ., a n d will be ac­
cepted u n til Ja n u a ry 10, 1935, for p u b lication in a la te r issue of
T ransactions.
N o t e : S tatem en ts a n d opinions ad v an ced in p ap ers are to be
u nderstood as individual expressions of th eir au th o rs, and n o t those
of th e Society.
included in this paper and it is expected that they will be pub­
lished elsewhere.
Certain constants obtained from this preliminary investiga­
tion have been used in the new design and the fan developed
on the basis of these researches has high efficiency. In addition,
consideration was given throughout the design to the question
of minimizing noise, and tests appear to show that in this respect
also the fan performance shows improvement over previous
designs.
D e s ig n
A consideration of the two-dimensional flow around a blade
element in a fluid of infinite extent, with corrections for mutual
blade interference and for finite blade length, leads to certain
basic conclusions which after having been verified by test were
applied to the design.
The velocity diagram, Fig. 1, shows the conditions of flow at
inlet and discharge in the usual manner and with the standard
symbols for velocities and their components. It is drawn for an
airfoil operating with a constant axial-velocity component. It is
con­
apparent from Fig. 1 that an increase in the angle of attack, a,
accompanies a diminution of the axial velocity and an increase
F ig . 1
V e l o c it y D ia g r a m
for
I d eal F an
in the discharge circumferential velocity. At the same time,
it increases the lift and drag until the stalling angle is reached.
The decrease in relative velocity of the air with respect to the
blade corresponds to a static pressure difference across the
wheel.
The endeavor to obtain constant axial velocity leads, in the
first approximation, to constant pitch or a pitch angle inversely
proportional to the radius.
The fan can be designed so that the same amount of work is
done on each particle of air, at the flow conditions corresponding
to the point of maximum efficiency. To accomplish this, varia­
tion in the angle of attack was utilized, but this procedure can­
not be effective through a large range of maximum to minimum
radius of blade. Consequently a large hub diameter is necessary.
Increase in hub size will increase the necessary diameter of the
fan for a desired capacity. A compromise had to be reached and,
for the present design, the hub diameter was made one-half the
fan diameter.
Choice was made of an airfoil profile with a straight front,
for which the center of pressure difference across the blade is
well toward the leading edge. With this profile, eddy formation
at the trailing edge is reduced—a condition favorable to the
minimizing of noise.
A study of the lift and drag characteristics of such airfoils
807
TRANSACTIONS OF THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS
808
shows that thin profiles have their best lift-drag ratio at small the blade at an angle to the radius and of the leading edge of the
angles of attack, while thicker profiles produce an optimum ratio guide vanes in the opposite direction had also for its purpose the
at larger angles of attack and have also a large stalling angle. reduction of noise.
The lift produced by these sections can be increased by increasing
The guide vanes were designed on the basis of measurements
the angle of attack above the value giving the best lift-drag ratio. of the air stream at discharge, and the varying direction and
This increase in the angle of attack above the optimum was varied velocity of the air stream approaching the guide vanes was con­
from a negative quantity at the periphery to a maximum at the sidered when selecting a suitable profile. While the angle of
hub and resulted in an increase of pitch along the radius from tip incidence was selected for best results at maximum efficiency, the
to hub. Some decrease in the angle of attack is necessary as a thick profiles were intended to obtain good flow under other
result of mutual blade interference. The angle of attack was operating conditions. The change of section of the guide vane
selected so as to give constant pressure on each blade element.
with radius is shown in Fig. 3.
Diffuser action was obtained in a cylindrical casing by the
tapering of the hub. Diffusers for moderate deceleration are
F ig . 3
F ig . 2
S t r e a m l in e s
a n d V e l o c it y
ose o f H ub
N
C y l in d r ic a l S e c t io n s
of
G u id e V a n e s
D is t r ib u t io n A r o u n d
The theory of the individual blade element does not take into
account the mutual interference of neighboring blades, or induc­
tion phenomena near the tip and the hub, or the effect of rota­
tional flow in the interval between two blades. The last is a
rotation of the air relative to the blades and in a direction op­
posite to the direction of rotation of the fan and is due to the
fact that the air enters without rotational motion.
A more precise estimate of the pressure difference was obtained
through application of the circulation theory. According to
this theory, the circulation at the discharge side is equal to the
sum of the circulations around the individual blades, provided
that the circulation at the suction side is zero. Through Joukowsky’s theory, relating the circulation to the lift of an airfoil,
the connection between airfoil theory and circulation theory is
established. To investigate this relation, measurements were
made of the rotational velocity at discharge from the fan, in the
preliminary investigations. The results of these investigations
were summarized in the computation of a mean-value coefficient,
which is the ratio of the arithmetic average of measured circum­
ferential velocities and the circumferential velocity of the ideal
fan. This coefficient has the value of 0.5 to 0.6 for a fan in which
the chord of the blade section is approximately equal to the
normal pitch at that section. This factor is not greatly altered
for deviations from this ratio of chord to normal pitch.
The principles of streamlining were observed throughout the
design of the fan and adjoining air passages. The nose of the
hub is shown in Fig. 2, which also gives the calculated flow lines
for the case in which the air approaches the hub in an axial direc­
tion. It will be seen that this shape has the advantage of giving
somewhat greater axial velocities in the vicinity of the hub.
The number of guide vanes was chosen so as to avoid simul­
taneous encounters of the trailing edge of the wheel and the lead­
ing edge of a guide. This should tend to keep down the in­
tensity of sound emission. The tilting of the trailing edge of
F ig . 5
A rrangem ent
T h r e e -B laded P r o pel l e r
C a s in g
of
in
AERONAUTICAL ENGINEERING
AER-56-13
809
fairly efficient—up to 85 per cent. For area ratios in excess of
1.22 with a conical diffuser and with laminar flow, back flow sets
in and the efficiency drops. With a cylindrical casing and a
tapering hub the inequality of decrease in velocity is counteracted
and a good efficiency is possible.
T e st in g A e r a n g e m e n t s
The details of the fan design are given in the pattern drawing,
Fig. 4; its arrangement in the casing, with its stationary stream­
line continuation, is shown in Fig. 5. The well-rounded entry
to the casing is omitted from Fig. 5 but is indicated in Fig. 6,
which shows the test set-up. The stationary streamline con­
tinuation of the fan hub was centered in the casing by a spider
of five radial blades and it contained the ball bearings which
support the shaft at the fan end. The guide vanes are located
between the fan and the radial blades.
As shown in the set-up for performance tests, Fig. 6, the fan dis­
charges directly into the atmosphere. The air enters through a
calibrated nozzle and the resistance is controlled by screens and
slats located 22 ft past the nozzle. The resistances are suc­
ceeded by a 12Va-ft length of square duct of about 50-in. side.
The air enters the fan casing through a well-rounded bell mouth.
The fan is driven by a long shaft which permitted the use of
diffusers of any desired length and located the dynamometer at
such distance as to offer no disturbance to the air flow. It had
the disadvantage, however, of limiting the permissible speed of
operation to about 3000 rpm.
The volume of air passing through the fan was measured by an
impact tube arranged on the center line of the nozzle, one-half of
its diameter distant from its outlet.
The total pressure is the difference between the impact pres­
sures at the inlet to the bell mouth and at discharge from the
fan casing or diffuser. The velocity pressure at the bell mouth is
too small to be measurable at any operating condition and conse­
quently a static-pressure measurement was substituted at the
location indicated. As the velocity over the discharge area is
variable, the discharge impact pressure was calculated from the
static pressure (which is atmospheric) and the mean-velocity
pressure computed from the air flow. This gives a smaller value
than the actual pressure.
Tests were made at 1800, 2400, 2700, and 3000 rpm and a
stroboscopic device, consisting of a neon lamp and a disk with
radial markings, rotating with the shaft, served to adjust the
speed to within one rpm of the desired speed.
The power input shown in the performance curves is the net
input, i.e., the difference between the measured and the frictional
horsepower. The latter was determined by tests in which a
plain cylindrical hub was substituted for the fan.
The test set-up differs in many ways from that of the Standard
Test Code for Propeller Fans of the American Society of Heating
and Ventilating Engineers. After the completion of the tests at
Harvard University the fan was tested in another laboratory,
following the methods of the Standard Test Code. The pres­
L o n <h
F ig . 6
A rrangem ent
of
t u d /m a l
F ig . 7
P e rfo rm a n c e C h a r a c te r is tic s o f T h re e -B la d e d
P ro p e lle b F a n
(Tests with ten guide vanes and diffuser; clearance, d = 1 in.)
sures, volumes and efficiencies as determined by the Standard
Test Code are greater than those recorded here.
T est R esu lts
The curves of Fig. 7 give the results of tests in which the fan
was provided with 10 guide vanes and discharged through a
conical diffuser, which increased in diameter from 20 in. to 38 in.
in a length of 80 in. All the remaining tests described in this
paper are with a cylindrical casing and no conical diffuser.
The curves of static pressure, efficiency, and horsepower input
are typical of propeller fans. The static pressure rises steadily
with decreasing flow to a “no-flow” value which is 70 per cent
of the spouting pressure corresponding to tip velocity, or equal
to the velocity head of a particle rotating with the fan wheel 8
in. distant from the axis. At about 40 per cent of the flow giving
best efficiency, there is a disturbance, noticeable by a bend in the
curves and by a considerable increase in noise. Measurements of
flow within the fan show that this is caused by secondary currents
due to centrifugal effects.
Peak efficiencies vary with the velocity as indicated by the
measurements at different speeds, but at 2700 and 3000 rpm the
maximum total efficiency has a constant value of about 80 per cent.
s e c t / oa/
A x ia l - F l o w P r o p e l l e r F a n
V ie w "A'
fo r
P erform a n ce T ests
TRANSACTIONS OF THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS
810
The same phenomenon of variation of peak efficiency with rpm
has been observed in propeller and centrifugal pump operation.
As a considerable part of the loss incidental to the operation of
axial-flow fans is due to conditions of flow in the discharge, an
investigation of the interaction of guide vanes and fan blades
was undertaken. Two factors were investigated: (1) the
effect of the number of guide vanes upon fan performance
and (2) the effect of the axial clearance between the trailing
edge of the fan wheel and the leading edge of the guide vanes.
G u id e V a n e s
Efficiency tests were conducted with 10 and with 5 guide
vanes installed and also without guide vanes. For these tests
the axial clearance between the fan wheel and the guide vanes
was maintained constant at a value of l 3/s in., which other tests
had shown to be a favorable distance in respect to efficiency
and noise. The five radial blades supporting the bearing hous-
F ig . 8
S t a t ic P r e s s u r e , N e t H o r s e p o w e r I n p u t ,
T o t a l E f f ic ie n c y
and
(Stub outlet duct; ten guide vanes; clearance, d — l 3/s in.)
F ig . 10 S t a t ic P r e s s u r e , N e t H o r s e p o w e r I n p u t , a n d
T o t a l E f f ic ie n c y
(Stub outlet duct; no guide vanes; clearance, d = l 3/s in.)
F ig . 9
S ta tic P r e s s u re , N e t H o rs e p o w e r In p u t, a n d
T o ta l E ffic ie n c y
(Stub outlet duct; 5 guide vanes; clearance, d = l 3/s in.)
ing presumably functioned in part as guide vanes during the
operation which is designated as “without guide vanes.” The
tests which were made cover the range of normal operating condi­
tions, but a few additional points were included down to the
“no-flow” operation.
The results obtained are given in Figs. 8, 9, and 10 which give
static pressures, total efficiencies, and net horsepower inputs at
1800, 2400, 2700, and 3000 rpm. The following conclusions may
be drawn from them.
(1) The variation of number of guide vanes does not affect
the quantity of air flow through the fan.
(2) As the number of guide vanes is increased, the static
pressure rises. The rise is 4 per cent for five guides and 13 per
cent for ten guides, as compared with operation with no guides,
in the region of best efficiency.
(3) The gain from guide vanes is most clearly shown by the
AERONAUTICAL ENGINEERING
curves jf total efficiency. For no guides the total efficiency
has a maximum of 70 per cent at 3000 rpm but diminishes slowly
with change of volume flowing. With five guides the peak
efficiency is brought up to 79 per cent at 3000 rpm but the ef­
ficiency curve is steeper. At large flows, the efficiency obtained
with five guides is slightly less than with no guides, as the guides
are not designed for this condition. For all other operating
conditions the use of five guides yields higher efficiency than with
no guides. With ten guide vanes the maximum efficiency in­
creases to 81 per cent but the efficiency curve becomes steeper still
and a further moderate decrease of efficiency is indicated at largest
flows.
These tests seem to indicate that the optimum number of
guide vanes for a fan of this type is between five and ten.
E ffect
of
C learance B etw een F an
and
AER-56-13
811
pactness of the equipment is considered. The fan performance
with greater diffuser action was determined by substituting the
conical diffuser for the stub cylindrical duct. The dimensions of
this diffuser are given earlier in connection with the discussion of
Fig. 7.
A comparison of the performance with the two degrees of
diffuser action is presented in Fig. 15, which is for a speed of
G u id e s
Variation of the axial distance, d, between propeller and guides
has a considerable influence upon the noisiness of the fan. In-
F i g . 11 E f f i c i e n c y C u r v e s
W ith A x ia l C le a r a n c e
d = 3/ 8 I n .
F i g . 13 E f f i c i e n c y C u r v e s
W ith A x ia l C le a r a n c e
d — I 3/* I n .
F i g . 15
C o m p a riso n W ith a n d W ith o u t D if f u s e r
(3000 rpm ; 10 guide vanes; clearance, d = 1 in.)
F ig . 12 E f f i c i e n c y C u r v e s F i g . 14 P e a k V a l u e s o f T o t a l E f ­
fic ie n c ie s a t T h r e e S p eed s P l o tte d
W ith A x ia l C le a r a n c e
A g a in s t A x ia l C le a r a n c e , d
d = 3/ t I n .
crease of noise becomes noticeable as d is decreased below 1 in.
It becomes a maximum when d is made very small.
The effect of the axial distance d upon efficiency was in­
vestigated in the region of best efficiency. Tests were made at
3000, 2700, and 2400 rpm and in some cases at 1800 rpm. The
results are shown in Figs. 11, 12, 13, and 14. The axial clearance
was varied from 3/ s in. to l 3/ 4 in. The lower limit was deter­
mined by the increase of noise, while above l 3/ 4 in. changes in
clearance did not influence the test results appreciably.
In Fig. 14 peak efficiencies are plotted against the axial dis­
tance, d, for various speeds. It will be seen that highest ef­
ficiencies are obtained for a clearance of 1 in. A somewhat
larger clearance was found desirable in order to reduce noise
further and a compromise was made in the adoption of a clear­
ance of 13/ a in. At this clearance, the efficiency is only slightly
lower than at 1-in. clearance.
3000 rpm and an axial distance between blades and guides of 1 in.
The conical diffuser has too large an area ratio (1 to 3.6) for opti­
mum results. The static pressures are not increased except at
large volumes but the net horsepower input is diminished, the
computed diffuser efficiency is 70 per cent and the maximum total
efficiency is decreased by about 1.5 per cent.
C o m p a r a t iv e P e r f o r m a n c e
For purposes of comparison, performance tests were made on
the two-bladed propeller fan referred to in the second paragraph
of this paper. The fan was built, as stated, for the research
preliminary to undertaking the design of a fan. Its general
features are shown in Fig. 16. The ratio of hub diameter to tip
diameter is 0.3. During the tests the hub was provided with a
well-rounded nose and a streamlined after-body. The blade
surfaces are parallel over a cylindrical section, tapering off
to a slightly rounded edge on both sides, as compared with the
airfoil sections of the new design. In the cylindrical develop­
ment, the blade is a curved plate of constant thickness with the
front or driving side convex as against the straight front shown in
Fig. 4. The pitch is constant along a radius but increases in the
axial direction, from inlet to outlet, as shown in Fig. 17. The
ratio of pitch at the leading edge to that at the trailing edge is
0.68. This change in pitch is proportional to the axial depth
D if f u s e r
except in the region near the trailing edge where the pitch
The cylindrical discharge casing of the fan, Fig. 5, gives remains constant. This has the effect of reducing the pressure
probably as much diffuser action as is desirable when the com­ difference across the blade in this region. Both the leading
812
TRANSACTIONS OF THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS
edge and the trailing edge are radial lines in an axial plane. In
the new design the pitch is variable along a radius and is constant
in the axial direction.
Ten guide vanes of constant curvature along the radius and
with an angle of incidence of 28 deg were installed for the perform­
ance tests.
The test results with this fan are given in Fig. 18 and a com­
parison of the two fans is made in Table 1. The tabulation is
TABLE 1 COM PARISON OF A TW O-BLADED AXIAL-FLOW
FAN W ITH T H E NEW THREE-BLA D ED AXIAL-FLOW FAN
TwoThreebladed
bladed
fan
fan
Tip diam, in........................................................................
20
19.3
H ub diam, in......................................................................
6
10
Axial depth, tip, in...........................................................
6
3.5
Axial depth, hub, in.........................................................
6
5.5
R pm ....... ............................................................................. 3000
3000
Tip velocity, ft per sec................................................... 262
252.5
Air volume at maximum efficiency, cfm ................... 6500
4650
Mean axial velocity at fan discharge plane, fps. . . .
54.2
51.6
Static pressure at maximum static efficiency, in. of
w ater...............................................................................
2.40
Net horsepower input.....................................................
3.58
Maximum static efficiency, per cent...........................
68.5
Maximum total efficiency, per cent............................
69.5
Volume coefficient,
_ / axial velocity at discharge plane\
0.207
0.205
\
tip velocity
/
Pressure coefficient,
_____ static pressure \
0.157
0.216
velocity head at tip speed/
Characteristic speed based on 1 in. of w ater........... 1580
1150
°W ith stub-discharge duct; all other figures with conical diffuser.
N o is e
U sed
F i g . 17
tD e p t h
for
C o m p a r is o n W
it h
N e w T h r e e -B laded F a n
High-speed axial-flow fans are noisy in operation as com­
pared with centrifugal fans. The problem of noise reduction
was kept in mind throughout the design as indicated at several
places in this paper. The completed fan was tested for noise
and similar tests were made on the two-bladed fan and on* a
centrifugal fan. In the noise tests of the axial-flow fans, the
stub cylindrical discharge duct was used and the microphone was
placed 2 ft from the end of the duct, near the edge of the air
V a r ia t io n o f P it c h W it h A x ia l
T w o - B l a d e d A x ia l - F l o w F a n
in
for 3000 rpm which is considerably below
the optimum operating speed. The new
design has a diameter slightly smaller
than the two-bladed fan and conse­
quently has a lower tip speed. The
F ig . 18 P e r f o r m a n c e C h a r a c t e r i s t i c s o f T w o - B l a d e d 2 0 - I n . P r o p e l l e r F a n
smaller discharge volume of the new
(3000 rpm ; 10 guide vanes and diffuser.)
design results from the larger hub di­
ameter, the lower tip speed, the thicker blades, and the decrease flow, and was oriented at 45 deg to the axial direction. In the
in pitch of the blades. The tests were made with the conical case of the centrifugal fan, the microphone was placed 2 ft away
from the edge of the well-rounded inlet to the fan and was
diffuser previously described.
It will be noted that the pressure coefficient of the new fan is oriented at 45 deg. to the fan axis.
The results of these tests are presented in Figs. 19, 20, and 21.
37.5 per cent greater than that of the two-bladed fan. This is in
accordance with the original purpose of designing a fan for com­ For the axial-flow fans, observations were made at two speeds,
2400 and 3000 rpm, with one additional observation at 1800 rpm
paratively high pressures.
AERONAUTICAL ENGINEERING
on the new fan. »The axial-flow fans had
10 guide vanes located 1 in. past the fan
blades. For the centrifugal fan, obser­
vations were taken at one speed only.
In each case the observations covered
the usual operating range of capacity.
The results obtained show an inter­
esting relation between fan performance
and noise. It will be observed that,
with the axial-flow fans, minimum noise
coincides approximately with maximum
efficiency and that noise increases rapidly
from that point, with decrease in ca­
pacity, until a break-down point is reached
where the noise intensity drops suddenly.
This break-down point coincides with
the inflection point in the static-pressure
curve. With further decrease in capacity
the noise increases again.
A comparison of the two fans at 3000
rpm and at maximum efficiency shows a
noise intensity of 84.3 db for the new
fan and 92.3 db for the two-bladed fan.
This represents a decrease in sound
AER-56-13
813
F ig . 20 N o is e M e a s u r e m e n t s o f T w o -B l a d e d P r o p e l l e r F a n
(2400 and 3000 rpm.)
F ig . 21
N o ise M e a s u re m e n ts o f 38-In. TVID S tu r t e v a n t
C e n tb ifu g a l F a n
(720 rpm.)
energy, for the new fan, to less than one-sixth of its value for
the two-bladed fan. The same ratio holds approximately at a
speed of 2400 rpm. The variation in noise with rpm, as shown
by the broken curve in Fig. 19, is exceedingly rapid. It is
obvious that the noise problem in axial-flow fans is not yet solved.
The noise characteristics of the centrifugal fans are quite differ­
ent from those of the axial-flow fans. Minimum noise occurs at
very low capacities and noise intensity increases regularly with
capacity, except for a break-down point which again coincides
with an inversion in curvature of the static-pressure curve. The
actual sound intensities for a given volume and static pressure
are much lower than with the axial-flow fans.
A cknow ledgm ent
F ig . 19
N o is e M e a s u r e m e n t s
of
N e w T h r e e -B la d e d F a n
(2400 and 3000 rpm.)
The authors desire to acknowledge the valuable assistance
of Mr. Thomas Flint, graduate student in the Harvard Engi­
neering School, in carrying out tests on which this paper is based.