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Predicting the sound emission from air-conditioning and ventilating systems
PREDICTING THE S O U N D EMISSION FROM AIRC O N D I T I O N I N G AND VENTILATING SYSTEMS
D. J. CROOME
Department of Civil Engineering, University of Technology, Loughborough, Leics. (Great Britain)
SUMMAR Y
Air-conditioning noise is, o f course, only one part o f the background noise spectrum .['or a space, but it has become significant over the last decade as more high velocity systems are being installed, and as people have become more sensitive to internal noise sources in buildings well insulated f r o m the external environment. Research has been carried out to investigate the nature and level o f sound emitted in rooms served by airconditioning systems. An initial survey measured the sound spectra in 74 university lecture rooms. Fans, motors, vee-belt drives and an airflow system were then installed to serve one lecture room, It was found that the soundpressure level in the room could be predicted by the equation:
or:
where L is the sound pressure level in dB( A ), v is the air velocity in the main duct and N is the fan speed (rpm). Further field studies are being undertaken on man); s),stems in buildings to find out ira general Jormula can be applied in practice.
CASE STUDY 1 : NOISE C O N D I T I O N S IN UNIVERSITY LECTURE R O O M S
In order to ascertain the importance of the problem, a questionnaire was circulated to 49 universities and colleges throughout Great Britain. The questionnaire consisted of ten questions requesting details of any ventilation or air-conditioning 303 Applied Acoustics (9) (1976)-- 0 Applied Science Publishers Ltd, England, 1976 Printed in Great Britain
304
D.J. CROOME
services, the acoustic acceptability of the lecture rooms provided with ventilation, steps taken at the design stage to ensure satisfactory noise levels, specific cases worthy of investigation and details of any amplification systems used. Replies from 16 universities were received from architects and building and planning officers concerning some 120 lecture rooms. In about a third of these the acoustics were considered unsatisfactory from some point of view by lecturers, building officers or students. Although such rooms are built for the sole purpose of allowing communication between lecturer and student, the inclusion of background noise level limits in lecture room specifications appears to be rare. In only one instance was a noise level specified;'NC 25 was chosen. Nearly half the cases made mention in a vague manner that the noise level should be kept to a minimum.
O 70, "
+ One s p e e d s y s t e m
F ~
~
0 Two speed system (slow)
60
,Q s o
4 0 ~ o
.S 3 0 lb
b i
62"5
125
?
250
500
I000
2000
4000
8000
Mean frequency of b o n d ( H z )
Fig. l(a). it is difficult to describe and locate noise sources, but',people thought that the fans and motors were responsible for noise in over half the cases and the airflow in the ducts in about a third of the cases. Duct wall resonance, grille noise, structural transmission and traffic noise were contributory factors. Often, the ventilation systems were designed to provide a greater air-change rate when the lecture rooms were fully occupied. This meant that the fan speed increased in proportion to the extra air supplied. Acoustic treatment was not usually recommended at the design stage. Frequently the system was installed, noise complaintswere then received and further money had to be spent to provide acoustic
S O U N D EMISSION F R O M A I R - C O N D I T I O N I N G SYSTEMS
305
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o
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b
m
70
oT;~o-spe#d syste(lost) m
1=
¢
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.~.
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.
30
"----...-..----
Ib L.
~ ...---._.__._~
20 Q.
I0
o
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30
62-5
125
250
500
Meon t r , q u , . c y
I000
ot bond
2000
4000
8000
(Hz)
Fig. l(b).
treatment. This usually took the form of sound absorption material placed in the ducts near the room concerned, or at bends. Several instances were quoted where the fans were positioned adjacent to the lecture room. Sometimes insufficient thought had been given to plant room construction, resulting in high noise levels in neighbouring work rooms. As a result of the questionnaire survey, it was decided to carry out a site appraisal of the acoustic conditions in 74 of the lecture rooms. The results of the measurements, superimposed on a set of N R curves, are shown in Figs. 1 (a) and (b) and acceptability ratings are given in Table 1. In 13 per cent of the cases for single speed systems the acoustic conditions were intolerable and in another 21 per cent, unsatisfactory. TABLE 1 ACOUSTIC ACCEPTABILITY OF LECTURE ROOMS (I~ERCENTAGE AVERAGED OVER THE FREQUENCY RANGE
125--4000 Hz)
Subjective acoustic rating degree of satisJaction Satisfactory ( < NR 28) Moderate (NR 28-NR 32) Unsatisfactory ( > NR 32) Intolerable ( > NR 37)
Single-speed systems
35 per cent 31 21 13
: Two speed systems (17 lecture rooms) Slow speed
Fast speed
81 per cent 13 1 5
15 per cent 17 19 49
306
D.J. C R O O M E
It is difficult to say which of the rooms classified as moderate would be unsatisfactory, but, according to the standard recommendations, those which exceed NR30 by more than 2 dB should be classed as unsatisfactory, the 2 dB being the accuracy of the measurements. This means that about 50 per cent of the rooms are acoustically inadequate. The field studies exposed two broad categories of noise source; those originating from the basic nature of the airflow system and those arising from bad planning or installation. Magnetic tape recordings were made in the lecture rooms, unoccupied, with the ventilation system on and off. These were analysed using 10 Hz constant bandwidth and 6 per cent bandwidth audiospectrometers. In order to identify the noise sources,
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~
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Fig. 2 (a), (b), (c) and (d).
2OO
2OOO Hz
Supply plant room spectra for room 2 (6 per cent bandwidth).
measurements were taken in some of the plant rooms. The spectra for the room and the plant room were then compared and the spectral components transmitted from the plant identified. The noise spectra for the plant room are shown in Figs. 2(a), (b), (c) and (d) for two fan speeds. It can be seen that at fast speed fewer peaks are present than for the lower speed. This suggests that, as the speed increases, the broadband aerodynamic noise in the fan gradually masks the discrete noise from mechanical sources and even the fundamental fan blade passage frequency and its harmonics. In this ventilation plant the major mechanical noise sources were the vee-belt and the fan bearings.
307
S O U N D EMISSION FROM A I R - C O N D I T I O N I N G SYSTEMS
Many factors contribute to the noise which can be emitted from vee-belt drive systems. First, the belt travels over pulleys and can vibrate along its passage due to the formation of standing waves. The fundamental frequencies for this installation are 10 Hz and 3 Hz on low speed and twice these values on high speed; the higher modes of these vibrations will contribute to the low frequency noise content. Pulleys not running true or belts which are slipping can lead to a variation in angular velocity--the so-called Drehfehler phenomenon; this means that oscillations will occur in the standing wave frequency. A common fault, very prevalent in this system, was due to belt joints which gave a characteristic belt squeak each time contact was made with the pulley.
Rototionol noise Belt foint noise ,
-.
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O
Frequency
/:'on speed
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l
to
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620 r.p.m. ~ 6
Frequency
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Fig. 2 (e). Sound spectra in lecture room served by the plant room shown in Fig. 2 (a), (b), (c) and (d). Referring to Fig. 2 (a) and (b), it can be seen that at 420-510 Hz discrete frequencies are emitted which correspond to the belt joint noise. The reason for the wide spread is that three belts are used which will not necessarily all sound at the same instant, and, further, each joint will not necessarily cause the same note to be emitted on contact with the small or large pulley. This sound has a high harmonic content which can be seen at around 900, 1350, 1800, 2250 and 2700Hz. This plant delivered air to the lecture r o o m whose ventilation system noise spectrum is shown in Fig. 2 (e). It can be seen that, at the lower fan speed, some discrete frequency components are still present, otherwise the spectrum has a broadband character. The geometrical form of the air distribution duct influences
308
D..t. C R O O M E
the transmission of sound since bends, expanders, contractions and branch offtakes cause some frequencies to be filtered out. This investigation was only concerned with air velocities below 5 m/s (1000 ft/min) where noise generation due to airflow in the distribution network has, until now, been considered unimportant. In some cases high noise levels are due to planning which necessitated that some lecture rooms had to be positioned next to a plant room, or a single glazed wall had to be adjacent to a main road. Lack of choice concerning available space may force certain environmental factors, in this case noise, to become critical. Commissioning and maintenance procedures must pay attention to noise problems which may arise from faulty installation.
CASE S T U D Y 2: R E S E A R C H ON A VARIABLE SPEED V E N T I L A T I O N SYSTEM S E R V I N G A LECTURE ROOM
The ventilation system serves a 74 seat lecture theatre in a university. Air is distributed from a unit comprising a 444mm diameter centrifugal fan with ten backward-curved blades. The drive shaft is mounted on ball-bearings and is driven by a variable speed motor with four fixed speeds via a double vee-belt and pulleys having a 1.6:1 diameter ratio. Table 2 gives the essential performance specification for the multi-speed centrifugal fan unit. TABLE 2 PERFORMANCE SPECIF|CATION OF MULTI-SPEEDCENTRIFUGALFAN UNIT
Speed I
Speed 2
Speed 3
Speed 4
4 1-2 760
6 1.3 1145
8 2.5 1530
12 5-5 2290
Motor No. of poles hp Speed (rpm)
Fan Speed (rpm) Volume flowrate (m3/s)
480 0.73
720 1.17
960 1.55
1440 2.33
Ductwork is of rectangular cross-section and constructed of 18 gauge galvanised sheet steel; the interior and exterior surfaces are unlined. The heater battery comprises a parallel bank of ten 9.5 mm finned hot water pipes arranged evenly across the duct, in the air stream. The outlet grilles have two rows of aerofoil vanes at right angles; the five of the inner row are fixed, whereas the six of the outer row are hinged and adjustable. The complete system is suspended from a void ceiling by 9.5 mm threaded rods and supported on brackets. The only contact with the void floor is at the grille outlets.
309
SOUND EMISSION FROM AIR-CONDITIONING SYSTEMS
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D.J. C R O O M E
Measurements of the air velocity across the duct cross-section revealed the airflow profiles shown in Fig. 3 for the two lower fan speeds only. Air turbulence is evident at the bend between test stations 5 and 6. Later measurements, not recorded on Fig. 3, verified turbulence at test station 18 after the bend following test station 17. The profiles are consistent for the two fan speeds. Now look at the in-duct sound profiles shown in Fig. 4. Attenuation of sound occurs at all frequencies but is offset almost entirely by generation of sound which is evident at each bend and at each fan speed, except at the first bend for fan speed 2. The generation happens for all frequencies but is most marked in the 500-1000 Hz frequency range. Sound is generated as the air travels through the attenuator at the three higher fan speeds. One significant conclusion is that sound generation is just as likely to occur in a low speed ventilation system (average air velocity in this example of about 4 m/s) as well as in a high speed system (average air velocity in this example of about 13 m/s); this contradicts the traditional belief that sound generation is insignificant at low air speeds. It is well established that ventilation system noise has a large low frequency content. This work reaches a similar conclusion. At the lowest fan speed, sound in the 31-5-250 Hz range has sound levels along the system of 63 to 88 dB (linear), whereas sound in the frequency range 500-1000 Hz has levels in the range 39-75 dB (linear). In either case the peak levels are due to sound generation. This pattern of behaviour is similar at the higher fan speed, although the frequencies in the range 500-1000 Hz gradually increase their share of the sound energy as the fan speed increases. The effect of sound generation begins at some distance before the bend; at low speeds this distance is a b o u t 1 m but extends to 2 m or more as the air velocity increases. The generation at the first bend (test stations 5 and 6) shows some interaction with that which has preceded it at the dampers (test stations 3 to 4), the breeches piece (test stations 2 and 3) and through the attenuator (test stations 1 to 2); such an interaction also happens to some extent between the generation effects at the bends themselves. Duct resonances do not appear to be significant in this system as no fundamental or harmonic frequencies are observed in the spectra. Notice that a given rise in sound level due to generation occurs over a much smaller distance than a similar decrease in sound level. This work confirms the point that airflow systems should be designed to have the most simple configurations wherever possible. G o o d aerodynamic design will assist acoustic design, although a precise relationship between turbulence and sound generation was not established for this system. A summary of results is given in Table 3. The results in Table 3 show that where there is a strong interaction between any elements of the ventilation system, any natural attenuation can be cancelled out by sound generation. The generation at the second .bend stands out more clearly than that of the first bend (see Fig. 4) because generation caused by elements preceding the first bend interact with the generation characteristic for the first bend. The varying attenuation r a t e s - - s o m e are even z e r o - - a r e also indicative of interaction effects.
SOUND EMISSION FROM AIR-CONDITIONING SYSTEMS
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19 20 21 22 21 22 23 241 i
D. J. C R O O M E
312
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SOUND EMISSION FROM AIR-CONDITIONING SYSTEMS
The attenuation rates are variable and very different to those used in practice (see IHVE Guide Book B, 1970). Sound generation magnitudes at bends are of a much higher order--from 4 dB at low frequencies to 18 dB at high frequencies--than attenuation magnitudes in straight metal ducts which range from 0-5 to 3 dB/m; no attenuation occurred at bends. These results dispute much traditional design data and further research is needed to resolve the conflicts. A satisfactory design criterion for the lecture room would be an Llo level of 36 dB(A) or NR 30. The lowest speed ventilation produces a sound level of 48 dB(A) in the room. The attenuator makes no effective contribution; in fact it worsens the situation at the three higher fan speeds.
80-
Lp m 70.
~ 60 ®x ® ~ 50
%
g 0
&
~ 20
(1976)} (19"75) E n v i r o n m e n t o / room t e s t s AIIbones (1976)] R Scott-Young .j" 023.Lecture room tests
® Croome o
(Co,'rek,tion,oe,; i, ntiO.9=S)
x AIIbones ® Bioletti
30
I0
x
26o
6&o 80o
(1976)
Questionnaire results for 8 buildings
,26o 48o Fort s p e e d
Fig. 5a.
(r. p. m.)
N
2, ,o 22oo 24oo
Room sound level related to fan speed (Croome3).
Figure 5 (a and b) shows the relationships between room sound pressure level and fan speed, and between room sound pressure level and main duct air velocity. It may be possible to predict room sound pressure level f r o m this type of relationship instead of the more complicated procedures given in the 1970 IHVE Guide mentioned above, but more research is needed in order to establish the influence of system configuration, cluct material and various makes of equipment on this function. This could be simply achieved by establishing the relationships derived from such measurements as those given in Fig. 5a for many installations. These preliminary studies show that the sound level in the room, L, can be predicted from the fan speed N (rpm) or the main duct air velocity v (m/s) by:
314
D.J. CROOME
L = 4 0 ( ~ 7 + 1)
in dB(A)
or:
=
+ 1
in dB(A)
I f N and v are fixed by other circumstances than acoustic ones the order of attenuation required can be estimated. Alternatively, if N can be chosen to match the selected criterion for L then this value of L can also be chosen to determine v, and hence the size of the main distribution duct can be calculated: Using Darcy's formula (see Chapter 5 of Croome and Roberts 2) the r o o m sound level may be related to pressure drop per unit length of main duct, Ap, by: L =
40F 1 f2gmAp~ 1'2 ] L ~ \ ~ / + 1
in dB(A)
for a duct having a mean hydraulic depth m, friction factor f, and where g = 9.81 m/s 2.
8O < 70 ,n "Io
J
"~ 60
J
~ so •~ 4 0
Approximote relotionsh/p from these meosurements :
b
x 30 Sound l e v e l i n r o o m
~
20
LdBA-40[~*i I]
~ ~o
o
2
4
Air
6 8 velocity in
ib
m
moin duct m/$
14
16
Fig. 5b. Room sound level related to air velocity in main duct (Croome3). A questionnaire has been circulated to 500 firms requesting them to measure the sound emission from one airflow system; these results will be entered on to Fig. 5, and a regression analysis carried out to seek the validity of this simple sound level prediction method. Some results of measurements made on another system by Bioletti 1 are already included on Fig. 5.
SOUND EMISSION FROM AIR-CONDITIONING SYSTEMS
315
REFERENCES 1. S. BIOLE'rTI,Environmental Engineering Finalist Research Project Thesis, Loughborough University of Technology, 1975. 2. D.J. CROOME and B. M. ROBERTS, Airconditioning and ventilation of buildings, Pergamon Press. 1975. 3. D.J. CROOME, Noise, buildings andpeople, Pergamon Press, 1976.
D. J. CROOME
Department of Civil Engineering, University of Technology, Loughborough, Leics. (Great Britain)
SUMMAR Y
Air-conditioning noise is, o f course, only one part o f the background noise spectrum .['or a space, but it has become significant over the last decade as more high velocity systems are being installed, and as people have become more sensitive to internal noise sources in buildings well insulated f r o m the external environment. Research has been carried out to investigate the nature and level o f sound emitted in rooms served by airconditioning systems. An initial survey measured the sound spectra in 74 university lecture rooms. Fans, motors, vee-belt drives and an airflow system were then installed to serve one lecture room, It was found that the soundpressure level in the room could be predicted by the equation:
or:
where L is the sound pressure level in dB( A ), v is the air velocity in the main duct and N is the fan speed (rpm). Further field studies are being undertaken on man); s),stems in buildings to find out ira general Jormula can be applied in practice.
CASE STUDY 1 : NOISE C O N D I T I O N S IN UNIVERSITY LECTURE R O O M S
In order to ascertain the importance of the problem, a questionnaire was circulated to 49 universities and colleges throughout Great Britain. The questionnaire consisted of ten questions requesting details of any ventilation or air-conditioning 303 Applied Acoustics (9) (1976)-- 0 Applied Science Publishers Ltd, England, 1976 Printed in Great Britain
304
D.J. CROOME
services, the acoustic acceptability of the lecture rooms provided with ventilation, steps taken at the design stage to ensure satisfactory noise levels, specific cases worthy of investigation and details of any amplification systems used. Replies from 16 universities were received from architects and building and planning officers concerning some 120 lecture rooms. In about a third of these the acoustics were considered unsatisfactory from some point of view by lecturers, building officers or students. Although such rooms are built for the sole purpose of allowing communication between lecturer and student, the inclusion of background noise level limits in lecture room specifications appears to be rare. In only one instance was a noise level specified;'NC 25 was chosen. Nearly half the cases made mention in a vague manner that the noise level should be kept to a minimum.
O 70, "
+ One s p e e d s y s t e m
F ~
~
0 Two speed system (slow)
60
,Q s o
4 0 ~ o
.S 3 0 lb
b i
62"5
125
?
250
500
I000
2000
4000
8000
Mean frequency of b o n d ( H z )
Fig. l(a). it is difficult to describe and locate noise sources, but',people thought that the fans and motors were responsible for noise in over half the cases and the airflow in the ducts in about a third of the cases. Duct wall resonance, grille noise, structural transmission and traffic noise were contributory factors. Often, the ventilation systems were designed to provide a greater air-change rate when the lecture rooms were fully occupied. This meant that the fan speed increased in proportion to the extra air supplied. Acoustic treatment was not usually recommended at the design stage. Frequently the system was installed, noise complaintswere then received and further money had to be spent to provide acoustic
S O U N D EMISSION F R O M A I R - C O N D I T I O N I N G SYSTEMS
305
rt =L
o
~
b
m
70
oT;~o-spe#d syste(lost) m
1=
¢
I=, ,,Q
60 50
0
.~.
40 NR35
.
30
"----...-..----
Ib L.
~ ...---._.__._~
20 Q.
I0
o
0
NR 25
30
62-5
125
250
500
Meon t r , q u , . c y
I000
ot bond
2000
4000
8000
(Hz)
Fig. l(b).
treatment. This usually took the form of sound absorption material placed in the ducts near the room concerned, or at bends. Several instances were quoted where the fans were positioned adjacent to the lecture room. Sometimes insufficient thought had been given to plant room construction, resulting in high noise levels in neighbouring work rooms. As a result of the questionnaire survey, it was decided to carry out a site appraisal of the acoustic conditions in 74 of the lecture rooms. The results of the measurements, superimposed on a set of N R curves, are shown in Figs. 1 (a) and (b) and acceptability ratings are given in Table 1. In 13 per cent of the cases for single speed systems the acoustic conditions were intolerable and in another 21 per cent, unsatisfactory. TABLE 1 ACOUSTIC ACCEPTABILITY OF LECTURE ROOMS (I~ERCENTAGE AVERAGED OVER THE FREQUENCY RANGE
125--4000 Hz)
Subjective acoustic rating degree of satisJaction Satisfactory ( < NR 28) Moderate (NR 28-NR 32) Unsatisfactory ( > NR 32) Intolerable ( > NR 37)
Single-speed systems
35 per cent 31 21 13
: Two speed systems (17 lecture rooms) Slow speed
Fast speed
81 per cent 13 1 5
15 per cent 17 19 49
306
D.J. C R O O M E
It is difficult to say which of the rooms classified as moderate would be unsatisfactory, but, according to the standard recommendations, those which exceed NR30 by more than 2 dB should be classed as unsatisfactory, the 2 dB being the accuracy of the measurements. This means that about 50 per cent of the rooms are acoustically inadequate. The field studies exposed two broad categories of noise source; those originating from the basic nature of the airflow system and those arising from bad planning or installation. Magnetic tape recordings were made in the lecture rooms, unoccupied, with the ventilation system on and off. These were analysed using 10 Hz constant bandwidth and 6 per cent bandwidth audiospectrometers. In order to identify the noise sources,
o
so.l~'-" 4O
O v e r o l l dBA l e v e l ~ - -
r"";
20
[
I
Fon speed
J
200
IL
2000
3 1 0 r.p.m.
Hz
L,
CO "o
20
2OO
°o i--ore.°.,0. lip
~L "b
20 GO ~
,eve,_41
I O v e r e l l 'dBA l e v e l
it2 0 0 ' I
0
~0
2OOO Hz
If
l
r.o
.o,.,.._i
2 0 0 0 Hz l
Fort s p e e d 6 2 0 r p.m -
~
40
Cd)
3O
20 20
Fig. 2 (a), (b), (c) and (d).
2OO
2OOO Hz
Supply plant room spectra for room 2 (6 per cent bandwidth).
measurements were taken in some of the plant rooms. The spectra for the room and the plant room were then compared and the spectral components transmitted from the plant identified. The noise spectra for the plant room are shown in Figs. 2(a), (b), (c) and (d) for two fan speeds. It can be seen that at fast speed fewer peaks are present than for the lower speed. This suggests that, as the speed increases, the broadband aerodynamic noise in the fan gradually masks the discrete noise from mechanical sources and even the fundamental fan blade passage frequency and its harmonics. In this ventilation plant the major mechanical noise sources were the vee-belt and the fan bearings.
307
S O U N D EMISSION FROM A I R - C O N D I T I O N I N G SYSTEMS
Many factors contribute to the noise which can be emitted from vee-belt drive systems. First, the belt travels over pulleys and can vibrate along its passage due to the formation of standing waves. The fundamental frequencies for this installation are 10 Hz and 3 Hz on low speed and twice these values on high speed; the higher modes of these vibrations will contribute to the low frequency noise content. Pulleys not running true or belts which are slipping can lead to a variation in angular velocity--the so-called Drehfehler phenomenon; this means that oscillations will occur in the standing wave frequency. A common fault, very prevalent in this system, was due to belt joints which gave a characteristic belt squeak each time contact was made with the pulley.
Rototionol noise Belt foint noise ,
-.
2
O
Frequency
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,
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620 r.p.m. ~ 6
Frequency
8
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~ 8
kHz
Fig. 2 (e). Sound spectra in lecture room served by the plant room shown in Fig. 2 (a), (b), (c) and (d). Referring to Fig. 2 (a) and (b), it can be seen that at 420-510 Hz discrete frequencies are emitted which correspond to the belt joint noise. The reason for the wide spread is that three belts are used which will not necessarily all sound at the same instant, and, further, each joint will not necessarily cause the same note to be emitted on contact with the small or large pulley. This sound has a high harmonic content which can be seen at around 900, 1350, 1800, 2250 and 2700Hz. This plant delivered air to the lecture r o o m whose ventilation system noise spectrum is shown in Fig. 2 (e). It can be seen that, at the lower fan speed, some discrete frequency components are still present, otherwise the spectrum has a broadband character. The geometrical form of the air distribution duct influences
308
D..t. C R O O M E
the transmission of sound since bends, expanders, contractions and branch offtakes cause some frequencies to be filtered out. This investigation was only concerned with air velocities below 5 m/s (1000 ft/min) where noise generation due to airflow in the distribution network has, until now, been considered unimportant. In some cases high noise levels are due to planning which necessitated that some lecture rooms had to be positioned next to a plant room, or a single glazed wall had to be adjacent to a main road. Lack of choice concerning available space may force certain environmental factors, in this case noise, to become critical. Commissioning and maintenance procedures must pay attention to noise problems which may arise from faulty installation.
CASE S T U D Y 2: R E S E A R C H ON A VARIABLE SPEED V E N T I L A T I O N SYSTEM S E R V I N G A LECTURE ROOM
The ventilation system serves a 74 seat lecture theatre in a university. Air is distributed from a unit comprising a 444mm diameter centrifugal fan with ten backward-curved blades. The drive shaft is mounted on ball-bearings and is driven by a variable speed motor with four fixed speeds via a double vee-belt and pulleys having a 1.6:1 diameter ratio. Table 2 gives the essential performance specification for the multi-speed centrifugal fan unit. TABLE 2 PERFORMANCE SPECIF|CATION OF MULTI-SPEEDCENTRIFUGALFAN UNIT
Speed I
Speed 2
Speed 3
Speed 4
4 1-2 760
6 1.3 1145
8 2.5 1530
12 5-5 2290
Motor No. of poles hp Speed (rpm)
Fan Speed (rpm) Volume flowrate (m3/s)
480 0.73
720 1.17
960 1.55
1440 2.33
Ductwork is of rectangular cross-section and constructed of 18 gauge galvanised sheet steel; the interior and exterior surfaces are unlined. The heater battery comprises a parallel bank of ten 9.5 mm finned hot water pipes arranged evenly across the duct, in the air stream. The outlet grilles have two rows of aerofoil vanes at right angles; the five of the inner row are fixed, whereas the six of the outer row are hinged and adjustable. The complete system is suspended from a void ceiling by 9.5 mm threaded rods and supported on brackets. The only contact with the void floor is at the grille outlets.
309
SOUND EMISSION FROM AIR-CONDITIONING SYSTEMS
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D.J. C R O O M E
Measurements of the air velocity across the duct cross-section revealed the airflow profiles shown in Fig. 3 for the two lower fan speeds only. Air turbulence is evident at the bend between test stations 5 and 6. Later measurements, not recorded on Fig. 3, verified turbulence at test station 18 after the bend following test station 17. The profiles are consistent for the two fan speeds. Now look at the in-duct sound profiles shown in Fig. 4. Attenuation of sound occurs at all frequencies but is offset almost entirely by generation of sound which is evident at each bend and at each fan speed, except at the first bend for fan speed 2. The generation happens for all frequencies but is most marked in the 500-1000 Hz frequency range. Sound is generated as the air travels through the attenuator at the three higher fan speeds. One significant conclusion is that sound generation is just as likely to occur in a low speed ventilation system (average air velocity in this example of about 4 m/s) as well as in a high speed system (average air velocity in this example of about 13 m/s); this contradicts the traditional belief that sound generation is insignificant at low air speeds. It is well established that ventilation system noise has a large low frequency content. This work reaches a similar conclusion. At the lowest fan speed, sound in the 31-5-250 Hz range has sound levels along the system of 63 to 88 dB (linear), whereas sound in the frequency range 500-1000 Hz has levels in the range 39-75 dB (linear). In either case the peak levels are due to sound generation. This pattern of behaviour is similar at the higher fan speed, although the frequencies in the range 500-1000 Hz gradually increase their share of the sound energy as the fan speed increases. The effect of sound generation begins at some distance before the bend; at low speeds this distance is a b o u t 1 m but extends to 2 m or more as the air velocity increases. The generation at the first bend (test stations 5 and 6) shows some interaction with that which has preceded it at the dampers (test stations 3 to 4), the breeches piece (test stations 2 and 3) and through the attenuator (test stations 1 to 2); such an interaction also happens to some extent between the generation effects at the bends themselves. Duct resonances do not appear to be significant in this system as no fundamental or harmonic frequencies are observed in the spectra. Notice that a given rise in sound level due to generation occurs over a much smaller distance than a similar decrease in sound level. This work confirms the point that airflow systems should be designed to have the most simple configurations wherever possible. G o o d aerodynamic design will assist acoustic design, although a precise relationship between turbulence and sound generation was not established for this system. A summary of results is given in Table 3. The results in Table 3 show that where there is a strong interaction between any elements of the ventilation system, any natural attenuation can be cancelled out by sound generation. The generation at the second .bend stands out more clearly than that of the first bend (see Fig. 4) because generation caused by elements preceding the first bend interact with the generation characteristic for the first bend. The varying attenuation r a t e s - - s o m e are even z e r o - - a r e also indicative of interaction effects.
SOUND EMISSION FROM AIR-CONDITIONING SYSTEMS
311
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19 20 21 22 21 22 23 241 i
D. J. C R O O M E
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The attenuation rates are variable and very different to those used in practice (see IHVE Guide Book B, 1970). Sound generation magnitudes at bends are of a much higher order--from 4 dB at low frequencies to 18 dB at high frequencies--than attenuation magnitudes in straight metal ducts which range from 0-5 to 3 dB/m; no attenuation occurred at bends. These results dispute much traditional design data and further research is needed to resolve the conflicts. A satisfactory design criterion for the lecture room would be an Llo level of 36 dB(A) or NR 30. The lowest speed ventilation produces a sound level of 48 dB(A) in the room. The attenuator makes no effective contribution; in fact it worsens the situation at the three higher fan speeds.
80-
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(1976)} (19"75) E n v i r o n m e n t o / room t e s t s AIIbones (1976)] R Scott-Young .j" 023.Lecture room tests
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(1976)
Questionnaire results for 8 buildings
,26o 48o Fort s p e e d
Fig. 5a.
(r. p. m.)
N
2, ,o 22oo 24oo
Room sound level related to fan speed (Croome3).
Figure 5 (a and b) shows the relationships between room sound pressure level and fan speed, and between room sound pressure level and main duct air velocity. It may be possible to predict room sound pressure level f r o m this type of relationship instead of the more complicated procedures given in the 1970 IHVE Guide mentioned above, but more research is needed in order to establish the influence of system configuration, cluct material and various makes of equipment on this function. This could be simply achieved by establishing the relationships derived from such measurements as those given in Fig. 5a for many installations. These preliminary studies show that the sound level in the room, L, can be predicted from the fan speed N (rpm) or the main duct air velocity v (m/s) by:
314
D.J. CROOME
L = 4 0 ( ~ 7 + 1)
in dB(A)
or:
=
+ 1
in dB(A)
I f N and v are fixed by other circumstances than acoustic ones the order of attenuation required can be estimated. Alternatively, if N can be chosen to match the selected criterion for L then this value of L can also be chosen to determine v, and hence the size of the main distribution duct can be calculated: Using Darcy's formula (see Chapter 5 of Croome and Roberts 2) the r o o m sound level may be related to pressure drop per unit length of main duct, Ap, by: L =
40F 1 f2gmAp~ 1'2 ] L ~ \ ~ / + 1
in dB(A)
for a duct having a mean hydraulic depth m, friction factor f, and where g = 9.81 m/s 2.
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J
~ so •~ 4 0
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b
x 30 Sound l e v e l i n r o o m
~
20
LdBA-40[~*i I]
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o
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Air
6 8 velocity in
ib
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Fig. 5b. Room sound level related to air velocity in main duct (Croome3). A questionnaire has been circulated to 500 firms requesting them to measure the sound emission from one airflow system; these results will be entered on to Fig. 5, and a regression analysis carried out to seek the validity of this simple sound level prediction method. Some results of measurements made on another system by Bioletti 1 are already included on Fig. 5.
SOUND EMISSION FROM AIR-CONDITIONING SYSTEMS
315
REFERENCES 1. S. BIOLE'rTI,Environmental Engineering Finalist Research Project Thesis, Loughborough University of Technology, 1975. 2. D.J. CROOME and B. M. ROBERTS, Airconditioning and ventilation of buildings, Pergamon Press. 1975. 3. D.J. CROOME, Noise, buildings andpeople, Pergamon Press, 1976.