B - room constant in the considered octave band in m 2, taken in accordance with paragraphs. 3.4 or 3.5;

AZ-i - an amendment that takes into account the distribution of sound power levels of ceiling lamps and gratings in octave bands, taken according to Table. 6, in dB;

D - correction for the location of the noise source; when the source is located in the working area (not higher than 2 m from the floor), A = 3 dB; if the source is above this zone, A *■ 0;

0.7 - safety factor;

F, B - the designations are the same as in paragraph 2.9, formula (5).

Note. The determination of the allowable air speed is carried out only for one frequency, which is equal to VNIIGS 250 Shch for ceiling lamps, 500 Hz for disk ceiling lamps, and 2000 Hz for anemostats and gratings.

2.14. In order to reduce the sound power level of the noise generated by bends and tees of air ducts, areas of a sharp change in the cross-sectional area, etc., it is necessary to limit the speed of air movement in the main air ducts of public buildings and auxiliary buildings of industrial enterprises to 5-6 m / s, and on branches up to 2-4 m/sec. For industrial buildings, these speeds can be respectively doubled, if this is permissible according to technological and other requirements.

3. CALCULATION OF OCTAVE SOUND PRESSURE LEVELS AT CALCULATED POINTS

3.1. Octave levels of sound pressure at permanent workplaces or in rooms (at design points) should not exceed the established norms.

(Notes: 1. If the regulatory requirements for sound pressure levels are different during the day, then the acoustic calculation of the installations should be made for the lowest permissible sound pressure levels.

2. Sound pressure levels at permanent workplaces or in rooms (at design points) depend on the sound power and the location of noise sources and the sound-absorbing qualities of the room in question.

3.2. When determining the octave levels of sound pressure, the calculation should be made for permanent workplaces or settlement points in rooms closest to noise sources (heating and ventilation units, air distribution or air intake devices, air or air curtains, etc.). In the adjacent territory, the design points should be taken as the points closest to noise sources (fans located openly on the territory, exhaust or air intake shafts, exhaust devices of ventilation installations, etc.), for which sound pressure levels are normalized.

a - noise sources (autonomous air conditioner and ceiling) and the calculated point are in the same room; b - noise sources (fan and installation elements) and the calculated point are located in different rooms; c - source of noise - the fan is located in the room, the calculated point is on the arrival side of the territory; 1 - autonomous air conditioner; 2 - calculated point; 3 - noise-generating ceiling; 4 - vibration-isolated fan; 5 - flexible insert; in - the central muffler; 7 - sudden narrowing of the duct section; 8 - branching of the duct; 9 - rectangular turn with guide vanes; 10 - smooth turn of the air duct; 11 - rectangular turn of the duct; 12 - lattice; /

3.3. Octave/Sound pressure levels at design points should be determined as follows.

Case 1. The noise source (noise-generating grille, ceiling lamp, autonomous air conditioner, etc.) is located in the considered room (Fig. 3). Octave sound pressure levels generated at the calculated point by one noise source should be determined by the formula

L-L, + I0! g (-£-+--i-l (8)

Oct \ 4 I g g W t )

N o t e. For ordinary rooms that do not have special requirements for acoustics, according to the formula

L \u003d Lp - 10 lg H w -4- D - (- 6, (9)

where Lp okt is the octave sound power level of the noise source (determined according to Section 2) in dB\

B w - room constant with a noise source in the considered octave band (determined according to paragraphs 3.4 or 3.5) in g 2;

D - correction for the location of the noise source If the noise source is located in the working area, then for all frequencies D \u003d 3 dB; if above the working area, - D=0;

Ф - radiation directivity factor of the noise source (determined from the curves in Fig. 4), dimensionless; d - distance from the geometric center of the noise source to the calculated point in g.

The graphical solution of equation (8) is shown in fig. 5.

Case 2. The calculated points are located in a room isolated from noise. Noise from a fan or unit element propagates through the air ducts and is radiated into the room through the air distribution or air inlet device (grille). Octave sound pressure levels generated at design points should be determined by the formula

L \u003d L P -DL p + 101g (-% + -V (10)

Note. For ordinary rooms, for which there are no special requirements for acoustics, - according to the formula

L - L p -A Lp -10 lgiJ H ~b A -f- 6, (11)

where L p in is the octave level of the sound power of the fan or installation element radiated into the duct in the considered octave band in dB (determined in accordance with paragraphs 2.5 or 2.10);

AL r in - the total reduction in the level (loss) of the sound power of the noise of the fan or electric

installation time in the octave band under consideration along the sound propagation path in dB (determined in accordance with clause 4.1); D - correction for the location of the noise source; if the air distribution or air intake device is located in the working area, A \u003d 3 dB, if it is higher, - D \u003d 0; Ф and - directivity factor of the installation element (hole, grate, etc.) emitting noise into the isolated room, dimensionless (determined from the graphs in Fig. 4); rn is the distance from the installation element emitting noise into the isolated room to the calculated point in m

B and - the constant of the room isolated from noise in the considered octave band in m 2 (determined according to paragraphs 3.4 or 3.5).

Case 3. The calculated points are located on the territory adjacent to the building. Fan noise propagates through the duct and is radiated to the atmosphere through the grate or shaft (Fig. 6). Octave levels of sound pressure generated at design points should be determined by the formula

I = L p -AL p -201gr a -i^- + A-8, (12)

where r a is the distance from the installation element (grid, hole) emitting noise into the atmosphere to the design point in m \ p a - sound attenuation in the atmosphere, taken according to Table. 7 in dB/km

A is the correction in dB, taking into account the location of the calculated point relative to the axis of the installation element emitting noise (for all frequencies, it is taken according to Fig. 6).

1 - ventilation shaft; 2 - louvre

The remaining quantities are the same as in formulas (10)

3.4. The room constant B should be determined from the graphs in fig. 7 or according to table. 9, using the table. 8 to determine the characteristics of the room.

3.5. For rooms with special requirements for acoustics (unique

halls, etc.), the constant of the room should be determined in accordance with the instructions for acoustic calculation for these rooms.

3.6. If the design point receives noise from several noise sources (for example, supply and recirculation grilles, an autonomous air conditioner, etc.), then for the considered design point, according to the corresponding formulas in clause 3.2, the octave sound pressure levels generated by each of the noise sources separately should be determined , and the total level in

These "Instructions on the acoustic calculation of ventilation units" were developed by the Research Institute of Building Physics of the USSR State Construction Committee together with the institutes Santekhproekt of the USSR State Construction Committee and Giproniiaviaprom of Minaviaprom.

The instructions were developed in development of the requirements of the chapter SNiP I-G.7-62 “Heating, ventilation and air conditioning. Design Standards” and “Sanitary Design Standards for Industrial Enterprises” (SN 245-63), which establish the need to reduce the noise of ventilation, air conditioning and air heating installations for buildings and structures for various purposes when it exceeds the sound pressure levels allowed by the standards.

Editors: A. No. 1. Koshkin (Gosstroy of the USSR), Doctor of Engineering. sciences, prof. E. Ya. Yudin and candidates of tech. Sciences E. A. Leskov and G. L. Osipov (Research Institute of Building Physics), Ph.D. tech. Sciences I. D. Rassadi

The Guidelines set out the general principles of acoustic calculations for mechanically driven ventilation, air conditioning and air heating installations. Methods for reducing sound pressure levels at permanent workplaces and in rooms (at design points) to the values ​​established by the norms are considered.

at (Giproniiaviaprom) and eng. | g. A. Katsnelson / (GPI Santekhproekt)

1. General Provisions............ - . . , 3

2. Noise sources of installations and their noise characteristics 5

3. Calculation of octave levels of sound pressure in the calculated

points................. 13

4. Reducing the levels (losses) of the sound power of noise in

various elements of air ducts ........ 23

5. Determining the required reduction in sound pressure levels. . . *. ............... 28

6. Measures to reduce sound pressure levels. 31

Appendix. Examples of acoustic calculation of ventilation, air conditioning and air heating installations with mechanical stimulation...... 39

Plan I quarter. 1970, No. 3

3.7. If the calculated points are located in a room through which a “noisy” duct passes, and noise enters the room through the walls of the duct, then the octave sound pressure levels should be determined by the formula

L - L p -AL p + 101g --R B - 101gB „-J-3, (13)

where Lp 9 is the octave level of the sound power of the noise source radiated into the duct, in dB (determined in accordance with paragraphs 2 5 and 2.10);

ALp b is the total reduction in sound power levels (losses) along the sound propagation path from the noise source (fan, throttle, etc.) to the beginning of the considered section of the duct that emits noise into the room, in dB (determined in accordance with Section 4);

State Committee of the Council of Ministers of the USSR for Construction Affairs (Gosstroy of the USSR)

1. GENERAL PROVISIONS

1.1. These Guidelines are developed in development of the requirements of the chapter SNiP I-G.7-62 “Heating, ventilation and air conditioning. Design Standards” and “Sanitary Design Standards for Industrial Enterprises” (SN 245-63), which established the need to reduce the noise of mechanically driven ventilation, air conditioning and air heating installations to sound pressure levels acceptable by the standards.

1.2. The requirements of these Guidelines apply to acoustic calculations of airborne (aerodynamic) noise generated during the operation of the installations listed in clause 1.1.

Note. These Guidelines do not consider calculations of vibration isolation of fans and electric motors (isolation of shocks and sound vibrations transmitted to building structures), as well as calculations of sound insulation of enclosing structures of ventilation chambers.

1.3. The method for calculating airborne (aerodynamic) noise is based on determining the sound pressure levels of the noise generated during the operation of the installations specified in clause 1.1 at permanent workplaces or in rooms (at design points), determining the need to reduce these noise levels and measures to reduce sound levels pressure to the values ​​allowed by the standards.

Notes: 1. Acoustic calculation should be included in the design of mechanically driven ventilation, air conditioning and air heating installations for buildings and structures for various purposes.

Acoustic calculation should be done only for rooms with normalized noise levels.

2. Air (aerodynamic) fan noise and noise generated by air flow in air ducts have broadband spectra.

3. In these Guidelines, under the noise should be understood any kind of sounds that interfere with the perception of useful sounds or break the silence, as well as sounds that have a harmful or irritating effect on the human body.

1.4. When acoustically calculating a central ventilation, air conditioning and hot air heating installation, the shortest duct run should be considered. If the central unit serves several rooms, for which the normative noise requirements are different, then an additional calculation should be made for the duct branch serving the room with the lowest noise level.

Separate calculations should be made for autonomous heating and ventilation units, autonomous air conditioners, units of air or air curtains, local exhausts, units of air shower installations, which are closest to the calculated points or have the highest performance and sound power.

Separately, it is necessary to carry out an acoustic calculation of the branches of the air ducts that exit into the atmosphere (suction and exhaust of air by installations).

If there are throttling devices (diaphragms, throttle valves, dampers), air distribution and air intake devices (grilles, shades, anemostats, etc.) between the fan and the serviced room, sudden changes in the cross section of air ducts, turns and tees, acoustic calculation of these devices should be made and plant elements.

1.5. Acoustic calculation should be made for each of the eight octave bands of the auditory range (for which noise levels are normalized) with the geometric mean frequencies of the octave bands 63, 125, 250, 500, 1000, 2000, 4000 and 8000 Hz.

Notes: 1. For central air heating, ventilation and air conditioning systems in the presence of an extensive network of air ducts, it is allowed to calculate only for frequencies of 125 and 250 Hz.

2. All intermediate acoustic calculations are performed with an accuracy of 0.5 dB. The final result is rounded to the nearest whole number of decibels.

1.6. Required measures to reduce noise generated by ventilation, air conditioning and air heating installations, if necessary, should be determined for each source separately.

2. SOURCES OF NOISE IN INSTALLATIONS AND THEIR NOISE CHARACTERISTICS

2.1. Acoustic calculations to determine the sound pressure level of air (aerodynamic) noise should be made taking into account the noise generated by:

a) a fan

b) when the air flow moves in the elements of the installations (diaphragms, chokes, dampers, turns of air ducts, tees, grilles, shades, etc.).

In addition, the noise transmitted through the ventilation ducts from one room to another should be taken into account.

2.2. Noise characteristics (octave sound power levels) of noise sources (fans, heating units, room air conditioners, throttling, air distribution and air intake devices, etc.) should be taken from the passports for this equipment or from catalog data

In the absence of noise characteristics, they should be determined experimentally on the instructions of the customer or by calculation, guided by the data given in these Guidelines.

2.3. The total sound power level of the fan noise should be determined by the formula

L p =Z+251g#+l01gQ-K (1)

where 1^P is the total sound power level of vein noise

tilator in dB re 10“ 12 W;

L-noise criterion, depending on the type and design of the fan, in dB; should be taken according to the table. one;

I is the total pressure created by the fan, in kg / m 2;

Q - fan performance in m^/sec;

5 - correction for the fan operation mode in dB.

Table 1

Notes: 1. The value of 6 when the deviation of the fan operation mode is not more than 20% of the maximum efficiency mode should be taken equal to 2 dB. In the fan operation mode with maximum efficiency 6=0.

2. To facilitate the calculations in fig. 1 shows a graph for determining the value of 251gtf+101gQ.

3. The value obtained by formula (1) characterizes the sound power radiated by an open inlet or outlet pipe of the fan in one direction into the free atmosphere or into the room in the presence of a smooth air supply to the inlet pipe.

4. When the air supply to the inlet pipe is not smooth or the throttle is installed in the inlet pipe to the values ​​specified in

tab. 1, should be added for axial fans 8 dB, for centrifugal fans 4 dB

2.4. The octave sound power levels of fan noise emitted by an open inlet or outlet of the fan L p a, into the free atmosphere or into the room, should be determined by the formula

(2)

where is the total sound power level of the fan in dB;

ALi - correction that takes into account the distribution of the sound power of the fan in octave bands in dB, taken depending on the type of fan and the number of revolutions according to table. 2.

table 2

Amendments ALu taking into account the distribution of the sound power of the fan in octave bands, in dB

Notes: 1. Given in Table. 2 data without brackets are valid when the fan speed is in the range of 700-1400 rpm.

2. At a fan speed of 1410-2800 rpm, the entire spectrum should be shifted an octave down, and at a speed of 350-690 rpm, an octave up, taking the values ​​\u200b\u200bfor the extreme octaves indicated in brackets for frequencies of 32 and 16000 Hz.

3. When the fan speed is more than 2800 rpm, the entire spectrum should be shifted two octaves down.

2.5. Octave sound power levels of fan noise radiated into the ventilation network should be determined by the formula

Lp - L p ■- A L-± -|~ L i-2,

where AL 2 is the correction that takes into account the effect of connecting the fan to the duct network in dB, determined from the table. 3.

2.6. The total sound power level of the noise radiated by the fan through the walls of the casing (housing) into the ventilation chamber room should be determined by formula (1), provided that the value of the noise criterion L is taken from Table. 1 as its average value for the suction and discharge sides.

The octave levels of the sound power of the noise emitted by the fan into the room of the ventilation chamber should be determined by the formula (2) and Table. 2.

2.7. If several fans operate simultaneously in the ventilation chamber, then for each octave band it is necessary to determine the total level

sound power of the noise emitted by all fans.

The total noise sound power level L cyu during operation of n identical fans should be determined by the formula

£sum = Z.J + 10 Ign, (4)

where Li is the sound power level of the noise of one fan in dB-, n is the number of identical fans.

Table 4.

2.8. Octave sound power levels of noise radiated into the room by autonomous air conditioners, heating and ventilation units, air shower units (without air duct networks) with axial fans should be determined by formula (2) and Table. 2 with a 3dB up-correction.

For autonomous units with centrifugal fans, the octave sound power levels of the noise emitted by the suction and discharge pipes of the fan should be determined by formula (2) and table. 2, and the total noise level - according to table. 4.

Note. When air is taken from outside by installations, it is not necessary to take a higher correction.

2.9. The total sound power level of noise generated by throttling, air distribution and air intake devices (throttle valves.

Ventilation calculation

Depending on the method of air movement, ventilation can be natural and forced.

The parameters of the air entering the intake openings and openings of local exhausts of technological and other devices located in the working area should be taken in accordance with GOST 12.1.005-76. With a room size of 3 by 5 meters and a height of 3 meters, its volume is 45 cubic meters. Therefore, ventilation should provide an air flow rate of 90 cubic meters per hour. In the summer, it is necessary to provide for the installation of an air conditioner in order to avoid exceeding the temperature in the room for the stable operation of the equipment. It is necessary to pay due attention to the amount of dust in the air, as this directly affects the reliability and service life of the computer.

The power (more precisely, the cooling power) of the air conditioner is its main characteristic, it depends on what volume of the room it is designed for. For approximate calculations, 1 kW per 10 m 2 is taken with a ceiling height of 2.8 - 3 m (in accordance with SNiP 2.04.05-86 "Heating, ventilation and air conditioning").

To calculate the heat inflows of this room, a simplified method was used:

where: Q - Heat inflows

S - Room area

h - Room height

q - Coefficient equal to 30-40 W / m 3 (in this case 35 W / m 3)

For a room of 15 m 2 and a height of 3 m, the heat inflows will be:

Q=15 3 35=1575 W

In addition, heat dissipation from office equipment and people should be taken into account, it is considered (in accordance with SNiP 2.04.05-86 "Heating, ventilation and air conditioning") that in a calm state a person emits 0.1 kW of heat, a computer or a copier 0.3 kW, By adding these values ​​to the total heat inputs, the required cooling capacity can be obtained.

Q add \u003d (H S opera) + (С S comp) + (P S print) (4.9)

where: Q add - The sum of additional heat gains

C - Computer heat dissipation

H - Heat dissipation of the operator

D - Printer Heat Dissipation

S comp - Number of workstations

S print - Number of printers

S operas - Number of operators

Additional heat inflows of the room will be:

Q add1 \u003d (0.1 2) + (0.3 2) + (0.3 1) \u003d 1.1 (kW)

The total sum of heat gains is equal to:

Q total1 \u003d 1575 + 1100 \u003d 2675 (W)

In accordance with these calculations, it is necessary to choose the appropriate power and number of air conditioners.

For the room for which the calculation is carried out, air conditioners with a rated power of 3.0 kW should be used.

Noise calculation

One of the unfavorable factors of the production environment in the computer center is the high level of noise generated by printing devices, air conditioning equipment, cooling fans in the computers themselves.

To address questions about the need and feasibility of noise reduction, it is necessary to know the noise levels at the operator's workplace.

The noise level arising from several incoherent sources operating simultaneously is calculated based on the principle of energy summation of radiation from individual sources:

L = 10 lg (Li n), (4.10)

where Li is the sound pressure level of the i-th noise source;

n is the number of noise sources.

The obtained calculation results are compared with the permissible value of the noise level for a given workplace. If the calculation results are above the permissible noise level, then special noise reduction measures are necessary. These include: lining the walls and ceiling of the hall with sound-absorbing materials, reducing noise at the source, proper equipment layout and rational organization of the operator's workplace.

The sound pressure levels of noise sources acting on the operator at his workplace are presented in Table. 4.6.

Table 4.6 - Sound pressure levels of various sources

Typically, the operator's workplace is equipped with the following equipment: hard drive in the system unit, fan(s) of PC cooling systems, monitor, keyboard, printer and scanner.

Substituting the values ​​of the sound pressure level for each type of equipment into formula (4.4), we get:

L=10 lg(104+104.5+101.7+101+104.5+104.2)=49.5 dB

The obtained value does not exceed the permissible noise level for the operator's workplace, equal to 65 dB (GOST 12.1.003-83). And if you consider that it is unlikely that such peripheral devices as a scanner and a printer will be used simultaneously, then this figure will be even lower. In addition, when the printer is working, the direct presence of the operator is not necessary, because. The printer is equipped with an automatic sheet feeder.

The ventilation and air conditioning system (VAC) is one of the main sources of noise in modern residential, public and industrial buildings, on ships, in sleeping cars of trains, in various salons and control cabins.

Noise in UHKV comes from the fan (the main source of noise with its own tasks) and other sources, propagates through the duct along with the air flow and is radiated into the ventilated room. Noise and its reduction are influenced by: air conditioners, heating units, air control and distribution devices, design, turns and branching of air ducts.

The acoustic calculation of the UHVAC is carried out in order to optimally select all the necessary means of noise reduction and determine the expected noise level at the design points of the room. Traditionally, active and reactive silencers have been the main means of reducing system noise. Soundproofing and sound absorption of the system and premises is required to ensure compliance with the norms of noise levels permissible for humans - important environmental standards.

Now, in the building codes and regulations of Russia (SNiP), which are mandatory for the design, construction and operation of buildings in order to protect people from noise, an emergency situation has developed. In the old SNiP II-12-77 "Noise Protection", the method of acoustic calculation of the SVKV of buildings is outdated and therefore was not included in the new SNiP 23-03-2003 "Noise Protection" (instead of SNiP II-12-77), where it is still at all is absent.

So the old method is deprecated and the new one is not. The time has come to create a modern method of acoustic calculation of SVKV in buildings, as is already the case with its own specifics in other, previously more advanced in acoustics, areas of technology, for example, on ships. Let's consider three possible methods of acoustic calculation, as applied to UHCS.

The first method of acoustic calculation. This method, which is established purely on analytical dependencies, uses the theory of long lines, known in electrical engineering and referred here to the propagation of sound in a gas filling a narrow pipe with rigid walls. The calculation is made under the condition that the pipe diameter is much less than the sound wave length.

For a rectangular pipe, the side must be less than half the wavelength, and for a round pipe, the radius. It is these pipes in acoustics that are called narrow. So, for air at a frequency of 100 Hz, a rectangular pipe will be considered narrow if the section side is less than 1.65 m. In a narrow curved pipe, sound propagation will remain the same as in a straight pipe.

This is known from the practice of using speech tubes, for example, for a long time on steamships. A typical diagram of a long line of a ventilation system has two defining quantities: L wH is the sound power coming into the discharge pipeline from the fan at the beginning of the long line, and L wK is the sound power coming from the discharge pipeline at the end of the long line and entering the ventilated room.

The long line contains the following characteristic elements. They are R1 soundproof inlet, R2 soundproof active muffler, R3 soundproof tee, R4 soundproof jet silencer, R5 soundproof damper and R6 soundproof outlet. Sound insulation here refers to the difference in dB between the sound power in the waves incident on a given element and the sound power radiated by this element after the waves have passed through it further.

If the sound insulation of each of these elements does not depend on all others, then the sound insulation of the entire system can be estimated by calculation as follows. The wave equation for a narrow pipe has the following form of the equation for plane sound waves in an unbounded medium:

where c is the speed of sound in air and p is the sound pressure in the pipe, related to the vibrational speed in the pipe according to Newton's second law by the relation

where ρ is the air density. The sound power for plane harmonic waves is equal to the integral over the cross-sectional area S of the duct over the period of sound vibrations T in W:

where T = 1/f is the period of sound vibrations, s; f is the oscillation frequency, Hz. Sound power in dB: L w \u003d 10lg (N / N 0), where N 0 \u003d 10 -12 W. Within the specified assumptions, the sound insulation of a long line of a ventilation system is calculated using the following formula:

The number of elements n for a specific SVKV can, of course, be greater than the above n = 6. Let us apply the theory of long lines to the above characteristic elements of the air ventilation system to calculate the values ​​of R i .

Inlet and outlet openings of the ventilation system with R 1 and R 6 . The junction of two narrow pipes with different cross-sectional areas S 1 and S 2 according to the theory of long lines is an analog of the interface between two media with normal incidence of sound waves on the interface. The boundary conditions at the junction of two pipes are determined by the equality of sound pressures and vibrational velocities on both sides of the connection boundary, multiplied by the cross-sectional area of ​​the pipes.

Solving the equations obtained in this way, we obtain the energy transmission coefficient and the sound insulation of the junction of two pipes with the above sections:

An analysis of this formula shows that at S 2 >> S 1 the properties of the second tube approach those of the free boundary. For example, a narrow pipe open into a semi-infinite space can be considered, from the point of view of the soundproofing effect, as bordering on a vacuum. For S 1<< S 2 свойства второй трубы приближаются к свойствам жесткой границы. В обоих случаях звукоизоляция максимальна. При равенстве площадей сечений первой и второй трубы отражение от границы отсутствует и звукоизоляция равна нулю независимо от вида сечения границы.

Active noise suppressor R2. Sound insulation in this case can be approximately and quickly estimated in dB, for example, according to the well-known formula of engineer A.I. Belova:

where P is the perimeter of the passage section, m; l is the silencer length, m; S - cross-sectional area of ​​the silencer channel, m 2; α eq is the equivalent sound absorption coefficient of the lining, depending on the actual absorption coefficient α, for example, as follows:

α 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1.0

α eq 0.1 0.2 0.4 0.5 0.6 0.9 1.2 1.6 2.0 4.0

It follows from the formula that the sound insulation of the channel of the active silencer R 2 is the greater, the greater the absorption capacity of the walls α eq, the length of the silencer l and the ratio of the channel perimeter to its cross-sectional area П/S. For the best sound-absorbing materials, for example, the PPU-ET, BZM and ATM-1 brands, as well as other widely used sound absorbers, the actual sound absorption coefficient α is presented in.

Tee R3. In ventilation systems, most often the first pipe with a cross-sectional area S 3 then branches into two pipes with cross-sectional areas S 3.1 and S 3.2. Such a branch is called a tee: through the first branch, sound enters, through the other two it passes further. In general, the first and second pipes may be comprised of a plurality of pipes. Then we have

The sound insulation of a tee from section S 3 to section S 3.i is determined by the formula

Note that due to aerohydrodynamic considerations in tees, they strive to ensure that the cross-sectional area of ​​the first pipe is equal to the sum of the cross-sectional area in the branches.

Reactive (chamber) noise suppressor R4. The chamber silencer is an acoustically narrow pipe with a cross section S 4 , which passes into another acoustically narrow pipe of large cross section S 4.1 with a length l, called a chamber, and then again passes into an acoustically narrow pipe with a cross section S 4 . Let us use the theory of the long line here as well. Replacing the characteristic impedance in the well-known formula for the sound insulation of a layer of arbitrary thickness at normal incidence of sound waves by the corresponding reciprocals of the pipe area, we obtain the formula for the sound insulation of a chamber silencer

where k is the wave number. The sound insulation of a chamber silencer reaches its greatest value at sin(kl)= 1, i.e. at

where n = 1, 2, 3, … Frequency of maximum sound insulation

where c is the speed of sound in air. If several chambers are used in such a silencer, then the sound reduction formula must be applied sequentially from chamber to chamber, and the total effect is calculated by applying, for example, the boundary conditions method. Efficient chamber silencers sometimes require large overall dimensions. But their advantage is that they can be effective at any frequency, including low frequencies, where active jammers are practically useless.

The zone of large sound insulation of chamber silencers covers repeating fairly wide frequency bands, but they also have periodic sound transmission zones that are very narrow in frequency. To improve efficiency and equalize the frequency response, a chamber silencer is often lined on the inside with a sound absorber.

damper R 5 . The damper is structurally a thin plate with an area S 5 and a thickness δ 5, clamped between the flanges of the pipeline, the hole in which the area S 5.1 is less than the inner diameter of the pipe (or other characteristic size). Soundproofing such a throttle valve

where c is the speed of sound in air. In the first method, the main issue for us when developing a new method is the assessment of the accuracy and reliability of the result of the acoustic calculation of the system. Let us determine the accuracy and reliability of the result of calculating the sound power entering the ventilated room - in this case, the values

Let us rewrite this expression in the following notation for the algebraic sum, namely

Note that the absolute maximum error of an approximate value is the maximum difference between its exact value y 0 and approximate y, that is, ± ε= y 0 - y. The absolute maximum error of the algebraic sum of several approximate values ​​y i is equal to the sum of the absolute values ​​of the absolute errors of the terms:

Here the least favorable case is adopted, when the absolute errors of all terms have the same sign. In reality, partial errors can have different signs and be distributed according to different laws. Most often in practice, the errors of the algebraic sum are distributed according to the normal law (Gaussian distribution). Let us consider these errors and compare them with the corresponding value of the absolute maximum error. Let us define this quantity under the assumption that each algebraic term y 0i of the sum is distributed according to the normal law with the center M(y 0i) and the standard

Then the sum also follows the normal distribution law with mathematical expectation

The error of the algebraic sum is defined as:

Then it can be argued that with a reliability equal to the probability 2Φ(t), the error of the sum will not exceed the value

At 2Φ(t), = 0.9973, we have t = 3 = α and the statistical estimate at almost maximum reliability is the error of the sum (formula) The absolute maximum error in this case

Thus ε 2Φ(t)<< ε. Проиллюстрируем это на примере результатов расчета по первому способу. Если для всех элементов имеем ε i = ε= ±3 дБ (удовлетворительная точность исходных данных) и n = 7, то получим ε= ε n = ±21 дБ, а (формула). Результат имеет совершенно неудовлетворительную точность, он неприемлем. Если для всех характерных элементов системы вентиляции воздуха имеем ε i = ε= ±1 дБ (очень высокая точность расчета каждого из элементов n) и тоже n = 7, то получим ε= ε n = ±7 дБ, а (формула).

Here, the result in the probabilistic estimation of errors in the first approximation can be more or less acceptable. So, the probabilistic estimation of errors is preferable, and it should be used to select the “ignorance margin”, which is proposed to be used in the acoustic calculation of the SVKV to ensure that the permissible noise standards are met in a ventilated room (this has not been done before).

But the probabilistic estimation of the result errors also indicates in this case that it is difficult to achieve high accuracy of the calculation results by the first method even for very simple circuits and a low-velocity ventilation system. For simple, complex, low- and high-speed UTCS circuits, satisfactory accuracy and reliability of such a calculation can be achieved in many cases only by the second method.

The second method of acoustic calculation. On ships, a calculation method has long been used, based partly on analytical dependencies, but decisively on experimental data. We use the experience of such calculations on ships for modern buildings. Then, in a ventilated room served by one j-th air distributor, the noise levels L j , dB, at the design point should be determined by the following formula:

where L wi is the sound power, dB, generated in the i-th element of the UHCS, R i is the sound insulation in the i-th element of the UHCS, dB (see the first method),

a value that takes into account the influence of the room on the noise in it (in the construction literature, sometimes B is used instead of Q). Here r j is the distance from the jth air distributor to the design point of the room, Q is the sound absorption constant of the room, and the values ​​χ, Φ, Ω, κ are empirical coefficients (χ is the near field influence coefficient, Ω is the spatial radiation angle of the source, Φ is the factor directivity of the source, κ is the coefficient of violation of the diffuseness of the sound field).

If m air diffusers are placed in the room of a modern building, the noise level from each of them at the calculated point is equal to L j , then the total noise from all of them must be below the noise levels acceptable for a person, namely:

where L H is the sanitary noise standard. According to the second method of acoustic calculation, the sound power L wi generated in all elements of the UHCS, and the sound insulation R i that takes place in all these elements, for each of them is preliminarily determined experimentally. The fact is that over the past one and a half to two decades, the electronic technology of acoustic measurements, combined with a computer, has greatly progressed.

As a result, enterprises producing SVKV elements must indicate in passports and catalogs the characteristics L wi and R i measured in accordance with national and international standards. Thus, the second method takes into account the noise generation not only in the fan (as in the first method), but also in all other elements of the UHCS, which can be significant for medium- and high-speed systems.

In addition, since it is impossible to calculate the sound insulation R i of such system elements as air conditioners, heating units, control and air distribution devices, therefore, they are not in the first method. But it can be determined with the required accuracy by standard measurements, which is now done for the second method. As a result, the second method, unlike the first one, covers almost all SVKV schemes.

And, finally, the second method takes into account the influence of the properties of the room on the noise in it, as well as the values ​​\u200b\u200bof noise acceptable to a person according to the current building codes and regulations in this case. The main disadvantage of the second method is that it does not take into account the acoustic interaction between the elements of the system - interference phenomena in pipelines.

The summation of the sound power of noise sources in watts, and the sound insulation of elements in decibels, according to the indicated formula for the acoustic calculation of UHCS, is valid only, at least, when there is no interference of sound waves in the system. And when there is interference in pipelines, then it can be a source of powerful sound, on which, for example, the sound of some wind musical instruments is based.

The second method has already been included in the textbook and guidelines for building acoustics course projects for senior students of St. Petersburg State Polytechnic University. Failure to take into account interference phenomena in pipelines increases the "margin for ignorance" or requires, in critical cases, experimental refinement of the result to the required degree of accuracy and reliability.

For the choice of "margin of ignorance", as shown above for the first method, the probabilistic error estimate is preferable, which is proposed to be used in the acoustic calculation of the SVKV of buildings to ensure that the permissible noise standards in the premises are met when designing modern buildings.

The third method of acoustic calculation. This method takes into account interference processes in a narrow pipeline of a long line. Such accounting can dramatically improve the accuracy and reliability of the result. For this purpose, it is proposed to apply for narrow pipes the "impedance method" of Academician of the Academy of Sciences of the USSR and the Russian Academy of Sciences Brekhovskikh L.M., which he used when calculating the sound insulation of an arbitrary number of plane-parallel layers.

So, let us first determine the input impedance of a plane-parallel layer with thickness δ 2 , whose sound propagation constant γ 2 = β 2 + ik 2 and acoustic impedance Z 2 = ρ 2 c 2 . Let us denote the acoustic resistance in the medium in front of the layer from where the waves fall, Z 1 = ρ 1 c 1 , and in the medium behind the layer we have Z 3 = ρ 3 c 3 . Then the sound field in the layer, with the omission of the factor i ωt, will be a superposition of waves traveling in the forward and reverse directions, with sound pressure

The input impedance of the whole system of layers (formula) can be obtained by a simple (n - 1)-fold application of the previous formula, then we have

Let us now apply, as in the first method, the theory of long lines to a cylindrical pipe. And thus, with interference in narrow pipes, we have the formula for sound insulation in dB of a long line of a ventilation system:

The input impedances here can be obtained both, in simple cases, by calculation, and, in all cases, by measurement on a special installation with modern acoustic equipment. According to the third method, similarly to the first method, we have the sound power coming from the discharge air duct at the end of a long UHVAC line and entering the ventilated room according to the scheme:

Next comes the evaluation of the result, as in the first method with a "margin of ignorance", and the sound pressure level of the room L, as in the second method. Finally, we obtain the following basic formula for the acoustic calculation of the ventilation and air conditioning system of buildings:

With the calculation reliability 2Φ(t)=0.9973 (practically the highest degree of reliability), we have t = 3 and the error values ​​are 3σ Li and 3σ Ri . With reliability 2Φ(t)= 0.95 (high degree of reliability) we have t = 1.96 and the error values ​​are approximately 2σ Li and 2σ Ri . With reliability 2Φ(t)= 0.6827 (engineering reliability assessment) we have t = 1.0 and the error values ​​are equal to σ Li and σ Ri The third method, directed to the future, is more accurate and reliable, but also more complex - it requires high qualifications in the fields of building acoustics, probability theory and mathematical statistics, and modern measuring technology.

It is convenient to use it in engineering calculations using computer technology. It, according to the author, can be proposed as a new method of acoustic calculation of the ventilation and air conditioning systems of buildings.

Summing up

The solution of urgent issues of developing a new method of acoustic calculation should take into account the best of the existing methods. A new method of acoustic calculation of the UTCS of buildings is proposed, which has a minimum "margin of ignorance" BB, due to the inclusion of errors by the methods of probability theory and mathematical statistics and the consideration of interference phenomena by the impedance method.

The information about the new calculation method presented in the article does not contain some of the necessary details obtained by additional research and work practice, and which constitute the author's "know-how". The ultimate goal of the new method is to provide a choice of a set of means to reduce the noise of the ventilation and air conditioning system of buildings, which increases, in comparison with the existing one, the efficiency, reducing the weight and cost of HVAC.

Technical regulations in the field of industrial and civil construction are not yet available, therefore, developments in the field, in particular, noise reduction in UHV buildings are relevant and should be continued at least until such regulations are adopted.

1. Brekhovskikh L.M. Waves in layered media // M.: Publishing House of the Academy of Sciences of the USSR. 1957.
2. Isakovich M.A. General acoustics // M .: Publishing house "Nauka", 1973.
3. Handbook of ship acoustics. Edited by I.I. Klyukin and I.I. Bogolepov. - Leningrad, "Shipbuilding", 1978.
4. Khoroshev G.A., Petrov Yu.I., Egorov N.F. Fighting fan noise // M .: Energoizdat, 1981.
5. Kolesnikov A.E. Acoustic measurements. Approved by the Ministry of Higher and Secondary Specialized Education of the USSR as a textbook for university students studying in the specialty "Electroacoustics and Ultrasonic Engineering" // Leningrad, "Shipbuilding", 1983.
6. Bogolepov I.I. Industrial soundproofing. Foreword by acad. I.A. Glebov. Theory, research, design, manufacture, control // Leningrad, Shipbuilding, 1986.
7. Aviation acoustics. Part 2. Ed. A.G. Munin. - M.: "Engineering", 1986.
8. Izak G.D., Gomzikov E.A. Noise on ships and methods of its reduction // M.: "Transport", 1987.
9. Noise reduction in buildings and residential areas. Ed. G.L. Osipova and E.Ya. Yudin. - M.: Stroyizdat, 1987.
10. Building regulations. Noise protection. SNiP II-12-77. Approved by the Decree of the State Committee of the Council of Ministers of the USSR for Construction of June 14, 1977 No. 72. - M.: Gosstroy of Russia, 1997.
11. Guidance for the calculation and design of noise attenuation of ventilation installations. Developed for SNiPu II-12–77 by organizations of the Research Institute of Building Physics, GPI Santekhpoekt, NIISK. - M.: Stroyizdat, 1982.
12. Catalog of noise characteristics of technological equipment (to SNiP II-12-77). Research Institute of Construction Physics of the Gosstroy of the USSR // M .: Stroyizdat, 1988.
13. Construction norms and rules of the Russian Federation. Noise protection. SNiP 23-03-2003. Adopted and put into effect by the resolution of the Gosstroy of Russia dated June 30, 2003 No. 136. Date of introduction 2004-04-01.
14. Soundproofing and sound absorption. A textbook for university students studying in the specialty "Industrial and civil engineering" and "Heat and gas supply and ventilation", ed. G.L. Osipov and V.N. Bobylev. - M.: AST-Astrel Publishing House, 2004.
15. Bogolepov I.I. Acoustic calculation and design of ventilation and air conditioning systems. Methodical instructions for course projects. St. Petersburg State Polytechnic University // St. Petersburg. SPbODZPP Publishing House, 2004.
16. Bogolepov I.I. Building acoustics. Foreword by acad. Yu.S. Vasilyeva // St. Petersburg. Polytechnic University Press, 2006.
17. Sotnikov A.G. Processes, devices and systems of air conditioning and ventilation. Theory, technology and design at the turn of the century // St. Petersburg, AT-Publishing, 2007.
18. www.integral.ru Firm "Integral". Calculation of the external noise level of ventilation systems according to: SNiP II-12-77 (part II) - "Guidelines for the calculation and design of noise attenuation of ventilation installations." St. Petersburg, 2007.
19. www.iso.org is an Internet site that contains complete information about the International Organization for Standardization ISO, a catalog and an online standards store through which you can purchase any currently valid ISO standard in electronic or printed form.
20. www.iec.ch is an Internet site that contains complete information about the International Electrotechnical Commission IEC, a catalog and an Internet store of its standards, through which it is possible to purchase the current IEC standard in electronic or printed form.
21. www.nitskd.ru.tc358 - a website on the Internet that contains complete information about the work of the technical committee TK 358 "Acoustics" of the Federal Agency for Technical Regulation, a catalog and an online store of national standards through which you can purchase the current required Russian standard in electronic or printed form.
22. Federal Law of December 27, 2002 No. 184-FZ "On Technical Regulation" (as amended on May 9, 2005). Adopted by the State Duma on December 15, 2002. Approved by the Federation Council on December 18, 2002. For the implementation of this Federal Law, see Order No. 54 of the Gosgortekhnadzor of the Russian Federation dated March 27, 2003.
23. Federal Law of May 1, 2007 No. 65-FZ “On Amendments to the Federal Law “On Technical Regulation”.

Ventilation systems are noisy and vibrate. The intensity and area of ​​sound propagation depends on the location of the main units, the length of the air ducts, the overall performance, as well as the type of building and its functional purpose. The calculation of noise from ventilation is designed to select the mechanisms of operation and the materials used, in which it will not go beyond the normative values, and is included in the design of ventilation systems as one of the points.

Ventilation systems consist of separate elements, each of which is a source of unpleasant sounds:

• For a fan, this can be a blade or a motor. The blade makes noise due to a sharp pressure drop on one side and the other. Engine - due to breakdown or improper installation. Refrigeration units make noise for the same reasons, plus improper compressor operation.
• Air ducts. There are two reasons: the first is vortex formations from the air hitting the walls. We talked about this in more detail in the article. The second is a hum in places where the cross section of the duct changes. Problems are solved by reducing the speed of gas movement.
• Building construction. Side noise from vibrations of fans and other installations transmitted to building elements. The solution is carried out by installing special supports or gaskets to dampen vibrations. A good example is an air conditioner in an apartment: if the outdoor unit is not fixed at all points, or the installers forgot to put protective pads, then its operation can cause acoustic discomfort to the owners of the installation or their neighbors.

Transfer Methods

There are three sound propagation paths, and in order to calculate the sound load, you need to know exactly how it is transmitted in all three ways:

• Airborne: noise from operating installations. Distributed both inside and outside the building. The main source of stress for people. For example, a large store, which has air conditioners and refrigeration units located at the back of the building. Sound waves propagate in all directions to nearby houses.
• Hydraulic: Noise source - liquid pipes. Sound waves are transmitted over long distances throughout a building. It is caused by a change in the size of the pipeline section and a malfunction of the compressor.
• Vibrating: source - building structures. Caused by improper installation of fans or other parts of the system. It is transmitted throughout the building and beyond.

Some specialists use scientific research from other countries in their calculations. For example, there is a formula published in a German magazine: it calculates the sound generation by the walls of an air duct, depending on the speed of the air flow.

Measuring method

It is often required to measure the permissible noise level or vibration intensity in already installed, operating ventilation systems. The classical method of measurement involves the use of a special device called a "sound level meter": it determines the strength of the propagation of sound waves. The measurement is carried out using three filters that allow you to cut off unwanted sounds outside the studied area. The first filter - measures the sound, the intensity of which does not exceed 50 dB. The second is from 50 to 85 dB. The third is over 80 dB.

Vibrations are measured in Hertz (Hz) for several points. For example, in the immediate vicinity of the noise source, then at a certain distance, then at the most distant point.

Norms and rules

The rules for calculating noise from ventilation operation and the algorithms for performing calculations are specified in SNiP 23-03-2003 "Protection from noise"; GOST 12.1.023-80 “System of labor safety standards (SSBT). Noise. Methods for establishing the values ​​of noise characteristics of stationary machines.

When determining the sound load near buildings, it must be remembered that the standard values ​​are given for intermittent mechanical ventilation and open windows. If closed windows and a forced air exchange system capable of providing the design multiplicity are taken into account, then other parameters are used as norms. The maximum noise level around the building is increased to the limit, which allows maintaining the normative parameters inside the building.

Sound load requirements for residential and public buildings depend on their category:

1. A is the best condition.
2. B - comfortable environment.
3. B is the noise level at the limit limit.

Acoustic calculation

It is used by designers to determine noise reduction. The main task of acoustic calculation is to calculate the active spectrum of sound loads at all points determined in advance, and compare the obtained value with the normative, maximum allowable. If necessary, reduce to established standards.

The calculation is carried out according to the noise characteristics of the ventilation equipment, they must be indicated in the technical documentation.

Settlement points:

• direct installation site of the equipment;
• adjoining premises;
• all rooms where the ventilation system operates, including basements;
• rooms for transit applications of air channels;
• places of inlet supply or exhaust outlet.

Acoustic calculation is performed according to two main formulas, the choice of which depends on the location of the point.

1. The calculation point is taken inside the building, in the immediate vicinity of the fan. Sound pressure depends on the power and number of fans, wave directionality and other parameters. Formula 1 for determining octave sound pressure levels from one or more fans looks like this:

where L Pi is the sound power in each octave;
∆L pomi - decrease in the intensity of the noise load associated with the multidirectional movement of sound waves and power losses from propagation in the air;

According to formula 2, ∆L is determined by mi:

where Фi is the dimensionless factor of the wave propagation vector;
S is the area of ​​a sphere or hemisphere that captures the fan and the calculation point, m 2;
B is the constant value of the acoustic constant in the room, m 2 .

1. The settlement point is taken outside the building in the surrounding area. The sound from operation propagates through the walls of the ventilation shafts, grilles and the fan housing. It is conditionally assumed that the noise source is a point one (the distance from the fan to the calculated position is an order of magnitude greater than the size of the apparatus). Then the octave noise pressure level is calculated by formula 3:

where L Pocti - octave power of the noise source, dB;
∆L Pneti - loss of sound power during its propagation through the duct, dB;
∆L ni - sound radiation directivity index, dB;
r - length of the segment from the fan to the calculation point, m;
W is the angle of sound radiation in space;
b a - reduction of noise intensity in the atmosphere, dB/km.

If several sources of noise act on one point, for example, a fan and an air conditioner, then the calculation method changes slightly. You can’t just take and add up all the sources, so experienced designers go the other way, removing all unnecessary data. The difference between the largest and the least intense source is calculated, and the resulting value is compared with the standard parameter and added to the level of the largest.

Reduced sound load from fan operation

There is a set of measures that allow leveling the noise factors from the operation of the fan that are unpleasant to the human ear:

• Choice of equipment. A professional designer, unlike an amateur, always pays attention to the noise from the system and selects fans that provide standard microclimate parameters, but at the same time without a large power margin. There is a wide range of fans with silencers on the market, they protect well from unpleasant sounds and vibrations.
• Choice of installation location. Powerful ventilation equipment is mounted only outside the serviced premises: it can be a roof or a special chamber. For example, if you put a fan in the attic in a panel house, then the residents on the top floor will immediately feel discomfort. Therefore, only roof fans are used in such cases.
• Selection of the speed of air movement through the channels. Designers proceed from acoustic calculation. For example, for a classic air duct 300×900 mm, it is no more than 10 m/s.
• Vibration isolation, sound isolation and shielding. Vibration isolation involves the installation of special supports that dampen vibrations. Soundproofing is carried out by pasting the cases with a special material. Shielding involves cutting off a sound source from a building or room using a shield.

The calculation of noise from ventilation systems involves finding such technical solutions when the operation of the equipment will not interfere with people. This is a complex task requiring skills and experience in this area.

Mega.ru has been dealing with ventilation and creating optimal microclimate conditions for a long time. Our experts solve problems of any complexity. We work in Moscow and the regions bordering on it. The technical support service will answer all questions by phone numbers listed on the page. Remote collaboration is possible. Contact us!

Description:

The norms and regulations in force in the country stipulate that the projects must provide for measures to protect against noise of equipment used for human life support. Such equipment includes ventilation and air conditioning systems.

Acoustic calculation as a basis for designing a low-noise ventilation (air conditioning) system

V. P. Gusev, doctor of tech. sciences, head. noise protection laboratory for ventilation and engineering equipment (NIISF)

The norms and regulations in force in the country stipulate that the projects must provide for measures to protect against noise of equipment used for human life support. Such equipment includes ventilation and air conditioning systems.

The basis for the design of sound attenuation of ventilation and air conditioning systems is acoustic calculation - a mandatory application to the ventilation project of any object. The main tasks of such a calculation are: determination of the octave spectrum of airborne, structural ventilation noise at the calculated points and its required reduction by comparing this spectrum with the permissible spectrum according to hygienic standards. After the selection of construction and acoustic measures to ensure the required noise reduction, a verification calculation of the expected sound pressure levels at the same design points is carried out, taking into account the effectiveness of these measures.

The materials given below do not claim to be complete in the presentation of the method of acoustic calculation of ventilation systems (installations). They contain information that clarifies, supplements or reveals in a new way various aspects of this technique using the example of the acoustic calculation of a fan as the main source of noise in a ventilation system. The materials will be used in the preparation of a set of rules for the calculation and design of noise attenuation of ventilation installations for the new SNiP.

The initial data for acoustic calculation are the noise characteristics of the equipment - sound power levels (SPL) in octave bands with geometric mean frequencies of 63, 125, 250, 500, 1000, 2000, 4000, 8000 Hz. For indicative calculations, corrected sound power levels of noise sources in dBA are sometimes used.

The calculated points are located in human habitats, in particular, at the place where the fan is installed (in the ventilation chamber); in rooms or in areas adjacent to the installation site of the fan; in rooms served by a ventilation system; in rooms where air ducts pass in transit; in the area of ​​​​the air intake or exhaust device, or only the air intake for recirculation.

The calculated point is in the room where the fan is installed

In general, sound pressure levels in a room depend on the sound power of the source and the directivity factor of noise emission, the number of noise sources, the location of the design point relative to the source and the enclosing building structures, and the size and acoustic qualities of the room.

The octave sound pressure levels generated by the fan (fans) at the installation site (in the ventilation chamber) are equal to:

where Фi is the directivity factor of the noise source (dimensionless);

S is the area of ​​an imaginary sphere or part thereof surrounding the source and passing through the calculated point, m 2 ;

B is the acoustic constant of the room, m 2 .

The calculated point is located in the room adjacent to the room where the fan is installed

The octave levels of airborne noise penetrating through the fence into the isolated room adjacent to the room where the fan is installed are determined by the soundproofing ability of the noisy room fences and the acoustic qualities of the protected room, which is expressed by the formula:

where L w - octave sound pressure level in the room with a noise source, dB;

R - isolation from airborne noise by the enclosing structure through which the noise penetrates, dB;

S - area of ​​the building envelope, m 2 ;

B u - acoustic constant of the insulated room, m 2 ;

k - coefficient that takes into account the violation of the diffuseness of the sound field in the room.

The calculated point is located in the room served by the system

The noise from the fan propagates through the air duct (air duct), partially attenuates in its elements and penetrates into the serviced room through the air distribution and air intake grilles. Octave levels of sound pressure in a room depend on the amount of noise reduction in the air duct and the acoustic qualities of this room:

where L Pi is the sound power level in the i-th octave radiated by the fan into the air duct;

D L networki - attenuation in the air channel (in the network) between the noise source and the room;

D L remember - the same as in formula (1) - formula (2).

Attenuation in the network (in the air channel) D L R network - the sum of the attenuation in its elements, sequentially located along the sound waves. The energy theory of sound propagation through pipes assumes that these elements do not influence each other. In fact, a sequence of shaped elements and straight sections form a single wave system, in which the principle of attenuation independence in the general case cannot be justified on pure sinusoidal tones. At the same time, in octave (wide) frequency bands, standing waves created by individual sinusoidal components compensate each other, and therefore the energy approach, which does not take into account the wave pattern in air ducts and considers the flow of sound energy, can be considered justified.

Attenuation in straight sections of air ducts made of sheet material is due to losses due to wall deformation and sound emission to the outside. The decrease in the sound power level D L R per 1 m of the length of straight sections of metal air ducts, depending on the frequency, can be judged from the data in Fig. one.

As can be seen, in rectangular ducts, the attenuation (lowering SAM) decreases with increasing sound frequency, while that of a circular duct increases. In the presence of thermal insulation on metal air ducts, shown in fig. 1 values ​​should be approximately doubled.

The concept of attenuation (reduction) of the sound energy flow level cannot be identified with the concept of a change in the sound pressure level in the air duct. As a sound wave travels through a channel, the total amount of energy it carries decreases, but this is not necessarily due to a decrease in the sound pressure level. In a narrowing channel, despite the attenuation of the total energy flow, the sound pressure level can increase due to an increase in the sound energy density. Conversely, in an expanding duct, the energy density (and sound pressure level) can decrease faster than the total sound power. The attenuation of sound in a section with a variable cross section is equal to:

(5)

where L 1 and L 2 are the average sound pressure levels in the initial and final sections of the channel section along the sound waves;

F 1 and F 2 - cross-sectional areas, respectively, at the beginning and end of the channel section.

Attenuation at bends (in elbows, bends) with smooth walls, the cross section of which is less than the wavelength, is determined by the reactance of the additional mass type and the appearance of higher order modes. The kinetic energy of the flow at the turn without changing the cross section of the channel increases due to the resulting non-uniformity of the velocity field. The square turn acts like a low pass filter. The amount of noise reduction at a turn in the plane wave range is given by an exact theoretical solution:

(6)

where K is the modulus of the sound transmission coefficient.

For a ≥ l /2, the value of K is equal to zero, and the incident plane sound wave is theoretically completely reflected by the channel rotation. The maximum noise reduction is observed when the turning depth is approximately half the wavelength. The value of the theoretical modulus of the coefficient of sound transmission through rectangular turns can be judged from Fig. 2.

In real designs, according to the data of the works, the maximum attenuation is 8-10 dB, when half the wavelength fits in the channel width. With increasing frequency, the attenuation decreases to 3-6 dB in the region of wavelengths close in magnitude to twice the channel width. Then it again smoothly increases at high frequencies, reaching 8-13 dB. On fig. Figure 3 shows the noise attenuation curves at channel turns for plane waves (curve 1) and for random, diffuse sound incidence (curve 2). These curves are obtained on the basis of theoretical and experimental data. The presence of a noise reduction maximum at a = l /2 can be used to reduce noise with low-frequency discrete components by adjusting the channel sizes at turns to the frequency of interest.

Noise reduction on turns less than 90° is approximately proportional to the angle of the turn. For example, the noise reduction on a 45° turn is equal to half the noise reduction on a 90° turn. On curves with an angle of less than 45°, noise reduction is not taken into account. For smooth bends and straight bends of air ducts with guide vanes, the noise reduction (sound power level) can be determined using the curves in Fig. 4.

In branching channels, the transverse dimensions of which are less than half the wavelength of the sound wave, the physical causes of attenuation are similar to the causes of attenuation in elbows and bends. This attenuation is determined as follows (Fig. 5).

Based on the medium continuity equation:

From the pressure continuity condition (r p + r 0 = r pr) and equation (7), the transmitted sound power can be represented by the expression

and the reduction in the sound power level at the cross-sectional area of ​​the branch

	(11)
	(12)
	(13)

With a sudden change in the cross section of a channel with transverse dimensions less than half-wavelengths (Fig. 6 a), the decrease in the sound power level can be determined in the same way as with branching.

The calculation formula for such a change in the channel cross section has the form

(14)

where m is the ratio of the larger cross-sectional area of ​​the channel to the smaller one.

The reduction in sound power levels when the channel sizes are larger than the non-planar half-wavelengths due to a sudden narrowing of the channel is

If the channel expands or gradually narrows (Fig. 6 b and 6 d), then the decrease in the sound power level is equal to zero, since there is no reflection of waves with a length shorter than the channel dimensions.

In simple elements of ventilation systems, the following reduction values ​​​​are taken at all frequencies: heaters and air coolers 1.5 dB, central air conditioners 10 dB, mesh filters 0 dB, the junction of the fan to the air duct network 2 dB.

Reflection of sound from the end of the duct occurs if the transverse dimension of the duct is less than the length of the sound wave (Fig. 7).

If a plane wave propagates, then there is no reflection in a large duct, and we can assume that there are no reflection losses. However, if an opening connects a large room and an open space, then only diffuse sound waves directed towards the opening, the energy of which is equal to a quarter of the energy of the diffuse field, enter the opening. Therefore, in this case, the sound intensity level is attenuated by 6 dB.

Characteristics of directivity of sound emission by air distribution grilles are shown in fig. eight.

When the noise source is located in space (for example, on a column in a large room) S = 4p r 2 (radiation in a full sphere); in the middle part of the wall, floors S = 2p r 2 (radiation into the hemisphere); in a dihedral angle (radiation in 1/4 sphere) S = p r 2 ; in the trihedral angle S = p r 2 /2.

The attenuation of the noise level in the room is determined by formula (2). The calculated point is selected at the place of permanent residence of people closest to the noise source, at a distance of 1.5 m from the floor. If the noise at the design point is created by several gratings, then the acoustic calculation is made taking into account their total impact.

When the source of noise is a section of a transit air duct passing through the room, the initial data for the calculation according to formula (1) are the octave sound power levels of the noise emitted by it, determined by the approximate formula:

where L pi is the sound power level of the source in the i-th octave frequency band, dB;

D L' Рneti - attenuation in the network between the source and the transit section under consideration, dB;

R Ti - sound insulation of the structure of the transit section of the air duct, dB;

S T - surface area of ​​the transit section, which goes into the room, m 2 ;

F T - cross-sectional area of ​​the duct section, m 2 .

Formula (16) does not take into account the increase in the density of sound energy in the duct due to reflections; the conditions for the incidence and passage of sound through the duct structure are significantly different from the passage of diffuse sound through the enclosures of the room.

Settlement points are located on the territory adjacent to the building

Fan noise propagates through the air duct and is radiated into the surrounding space through a grill or shaft, directly through the walls of the fan housing or an open pipe when the fan is installed outside the building.

When the distance from the fan to the calculated point is much larger than its dimensions, the noise source can be considered as a point source.

In this case, the octave sound pressure levels at the calculated points are determined by the formula

where L Pocti is the octave level of the sound power of the noise source, dB;

D L Pseti - total reduction of the sound power level along the path of sound propagation in the duct in the considered octave band, dB;

D L ni - sound radiation directivity indicator, dB;

r - distance from the noise source to the calculated point, m;

W - spatial angle of sound emission;

b a - sound attenuation in the atmosphere, dB/km.

If there is a row of several fans, grilles or other extended noise source of limited dimensions, then the third term in formula (17) is taken equal to 15 lgr .

Structural noise calculation

Structural noise in rooms adjacent to ventilation chambers occurs as a result of the transfer of dynamic forces from the fan to the ceiling. The octave sound pressure level in the adjacent isolated room is determined by the formula

For fans located in the technical room outside the ceiling above the isolated room:

(20)

where L Pi is the octave sound power level of airborne noise emitted by the fan into the ventilation chamber, dB;

Z c - total wave resistance of the elements of vibration isolators, on which the refrigeration machine is installed, N s / m;

Z lane - input impedance of the ceiling - the carrier plate, in the absence of a floor on an elastic base, the floor plate - if it is available, N s / m;

S - conditional floor area of ​​the technical room above the isolated room, m 2;

S = S 1 for S 1 > S u /4; S = S u /4; with S 1 ≤ S u /4, or if the technical room is not located above the isolated room, but has one common wall with it;

S 1 - the area of ​​​​the technical room above the isolated room, m 2;

S u - area of ​​the isolated room, m 2;

S in - the total area of ​​​​the technical room, m 2;

R - own insulation of airborne noise by overlapping, dB.

Determination of required noise reduction

The required reduction in octave sound pressure levels is calculated separately for each noise source (fan, fittings, fittings), but at the same time, the number of noise sources of the same type in terms of the sound power spectrum and the magnitude of the sound pressure levels created by each of them at the calculated point are taken into account. In general, the required noise reduction for each source should be such that the total levels in all octave frequency bands from all noise sources do not exceed the permissible sound pressure levels .

In the presence of one noise source, the required reduction in octave sound pressure levels is determined by the formula

where n is the total number of noise sources taken into account.

The total number of noise sources n when determining D L tr i the required reduction in octave sound pressure levels in urban areas should include all noise sources that create sound pressure levels at the design point that differ by less than 10 dB.

When determining D L tri for design points in a room protected from ventilation system noise, the total number of noise sources should include:

When calculating the required fan noise reduction - the number of systems serving the room; noise generated by air distribution devices and fittings is not taken into account;

When calculating the required noise reduction generated by the air distribution devices of the considered ventilation system, - the number of ventilation systems serving the room; the noise of the fan, air distribution devices and fittings is not taken into account;

When calculating the required noise reduction generated by shaped elements and air distribution devices of the considered branch, the number of shaped elements and chokes, the noise levels of which differ from one another by less than 10 dB; the noise of the fan and grilles is not taken into account.

At the same time, the total number of noise sources taken into account does not take into account noise sources that create at the design point the sound pressure level 10 dB lower than the permissible one, if their number is not more than 3 and 15 dB less than the permissible one, if their number is not more than 10.

As you can see, acoustic calculation is not an easy task. The necessary accuracy of its solution is provided by acoustic specialists. The efficiency of noise suppression and the cost of its implementation depend on the accuracy of the performed acoustic calculation. If the value of the calculated required noise reduction is underestimated, then the measures will not be effective enough. In this case, it will be necessary to eliminate the shortcomings at the operating facility, which is inevitably associated with significant material costs. If the required noise reduction is overestimated, unjustified costs are laid directly into the project. So, only due to the installation of silencers, the length of which is 300-500 mm longer than required, additional costs for medium and large objects can amount to 100-400 thousand rubles or more.

Literature

1. SNiP II-12-77. Noise protection. Moscow: Stroyizdat, 1978.

2. SNiP 23-03-2003. Noise protection. Gosstroy of Russia, 2004.

3. Gusev V.P. Acoustic requirements and design rules for low-noise ventilation systems // ABOK. 2004. No. 4.

4. Guidance for the calculation and design of noise attenuation of ventilation installations. Moscow: Stroyizdat, 1982.

5. Yudin E. Ya., Terekhin AS Fighting the noise of mine ventilation installations. Moscow: Nedra, 1985.

6. Noise reduction in buildings and residential areas. Ed. G. L. Osipova, E. Ya. Yudina. Moscow: Stroyizdat, 1987.

7. Khoroshev S. A., Petrov Yu. I., Egorov P. F. Control of fan noise. Moscow: Energoizdat, 1981.