The basis for designing sound attenuation of ventilation and air conditioning systems is acoustic calculation - a mandatory application to the ventilation project of any facility. The main tasks of such a calculation are: determination of the octave spectrum of airborne, structural ventilation noise at design points and its required reduction by comparing this spectrum with the permissible spectrum according to hygiene standards. After selecting 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 initial data for acoustic calculations are the noise characteristics of the equipment - sound power levels (SPL) in octave bands with geometric mean frequencies 63, 125, 250, 500, 1,000, 2,000, 4,000, 8,000 Hz. For indicative calculations, adjusted sound power levels of noise sources in dBA may be used.

Calculation points are located in human habitats, in particular, at the installation site of the fan (in the ventilation chamber); in rooms or areas adjacent to the fan installation site; in rooms served by a ventilation system; in rooms where air ducts pass through in transit; in the area of ​​the device for receiving or exhausting air, or only receiving air for recirculation.

The design 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 directional factor of noise emission, the number of noise sources, the location of the design point relative to the source and enclosing building structures, the size and acoustic qualities of the room.

The octave sound pressure levels created by the fan(s) at the installation location (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 of it, surrounding source and passing through the design point, m 2 ;

B is the acoustic constant of the room, m2.

Calculation points are located in the area adjacent to the building

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

If the distance from the fan to the design point is much greater than its size, the noise source can be considered a point source.

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

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

∆L Pneti - total reduction in sound power level along the sound propagation path in the air duct in the octave band under consideration, dB;

∆L ni - sound radiation directivity indicator, dB;

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

W is the spatial angle of sound radiation;

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

Engineering and Construction Journal, No. 5, 2010
Category: Technologies

Doctor of Technical Sciences, Professor I.I. Bogolepov

GOU St. Petersburg State Polytechnic University
and GOU St. Petersburg State Marine Technical University;
Master A.A. Gladkikh,
GOU St. Petersburg State Polytechnic University

Ventilation and air conditioning system (VAC) - critical system for modern buildings and structures. However, in addition to the necessary quality air, the system transports noise into the premises. It comes from the fan and other sources, spreads through the air duct and is emitted into the ventilated room. Noise is incompatible with normal sleep, educational process, creative work, highly productive work, proper rest, treatment, receiving quality information. IN building codes and the rules of Russia, such a situation has arisen. The method of acoustic calculation of HVAC buildings, used in the old SNiP II-12-77 “Noise Protection”, is outdated and therefore was not included in the new SNiP 03/23/2003 “Noise Protection”. So, old method is outdated, and there is no new generally accepted one yet. Below we propose a simple approximate method for the acoustic calculation of UHCR in modern buildings, developed using the best manufacturing experience, particularly on marine vessels.

The proposed acoustic calculation is based on the theory of long lines of sound propagation in an acoustically narrow pipe and on the theory of sound in rooms with a practically diffuse sound field. It is performed with the aim of assessing sound pressure levels (hereinafter referred to as SPL) and compliance of their values ​​with current permissible noise standards. It provides for the determination of ultrasonic sound from UHVV due to the operation of a fan (hereinafter referred to as the “machine”) for the following typical groups of premises:

1) in the room where the machine is located;

2) in rooms through which air ducts pass in transit;

3) in the premises served by the system.

Initial data and requirements

It is proposed to calculate, design and monitor the protection of people from noise for the most important octave frequency bands for human perception, namely: 125 Hz, 500 Hz and 2000 Hz. The octave frequency band of 500 Hz is the geometric mean value in the range of noise-standardized octave frequency bands of 31.5 Hz - 8000 Hz. For constant noise, the calculation involves determining the SPL in octave frequency bands based on the sound power levels (SPL) in the system. The values ​​of ultrasound and ultrasound are related by the general ratio = - 10, where - ultrasound relative to the threshold value of 2·10 N/m; - USM relative to the threshold value of 10 W; - area of ​​propagation of the front of sound waves, m.

SPL should be determined at the design points of premises rated for noise using the formula = + , where - SPL of the noise source. The value taking into account the influence of the room on the noise in it is calculated using the formula:

where is a coefficient taking into account the influence of the near field; - spatial angle of radiation from the noise source, rad.; - radiation directivity coefficient, taken from experimental data (to a first approximation, equal to unity); - distance from the center of the noise emitter to the calculated point in m; = - acoustic constant of the room, m; - average sound absorption coefficient of the internal surfaces of the room; - total area of ​​these surfaces, m; - coefficient taking into account the disruption of the diffuse sound field in the room.

The specified values, design points and permissible noise standards are regulated for the premises of various buildings by SNiP 03/23/2003 “Noise Protection”. If the calculated SPL values ​​exceed the permissible noise standard in at least one of the three frequency bands indicated, then it is necessary to design measures and means to reduce noise.

The initial data for acoustic calculations and design of UHCR are:

- layout diagrams used in the construction of the structure; dimensions of machines, air ducts, control fittings, elbows, tees and air distributors;

- speed of air movement in mains and branches - according to technical specifications and aerodynamic calculations;

- drawings of the general location of the premises served by the SVKV - according to the construction design of the structure;

- noise characteristics of machines, control valves and HVAC air distributors - according to data technical documentation for these products.

The noise characteristics of the machine are the following levels of noise level of airborne noise in octave frequency bands in dB: - level of noise propagating from the machine into the suction air duct; - ultrasonic noise propagation from the machine into the discharge duct; - Ultrasound noise emitted by the machine body into the surrounding space. All noise characteristics of a machine are currently determined on the basis of acoustic measurements according to the relevant national or international standards and other regulatory documents.

The noise characteristics of mufflers, air ducts, adjustable fittings and air distributors are presented by UZM airborne noise in octave frequency bands in dB:

- ultrasonic noise generated by system elements when air flow passes through them (noise generation); - USM of noise dissipated or absorbed in the elements of the system when a flow of sound energy passes through them (noise reduction).

The efficiency of noise generation and reduction by UHCR elements is determined based on acoustic measurements. We emphasize that the values ​​of and must be indicated in the relevant technical documentation.

Due attention is paid to the accuracy and reliability of the acoustic calculation, which is included in the error of the result in terms of and .

Calculation for the premises where the machine is installed

Let there be a fan in room 1, where the machine is installed, the sound power level of which, emitted into the suction, discharge pipeline and through the machine body, is in dB, and. Let the fan have a noise muffler with a muffling efficiency in dB () installed on the side of the discharge pipeline. Workplace is located at a distance from the car. The wall separating room 1 and room 2 is located at a distance from the machine. Sound absorption constant of room 1: = .

For room 1, the calculation involves solving three problems.

1st task. Compliance with permissible noise standards.

If the suction and discharge pipes are removed from the machine room, then the ultrasonic sound level in the room where it is located is calculated using the following formulas.

Octave SPL at the design point of the room is determined in dB using the formula:

where is the noise level emitted by the machine body, taking into account accuracy and reliability using . The value indicated above is determined by the formula:

If the room contains n noise sources, the SPL from each of which at the design point is equal to , then the total SPL from all of them is determined by the formula:

As a result of the acoustic calculation and design of the HVAC for room 1, where the machine is installed, it must be ensured that the permissible noise standards are met at the design points.

2nd task. Calculation of the value of the UZM in the discharge duct from room 1 to room 2 (the room through which the air duct passes in transit), namely the value in dB, is made according to the formula

3rd task. Calculation of the value of ultrasonic noise emitted by a wall area with sound insulation of room 1 into room 2, namely the value in dB, is performed according to the formula

Thus, the result of the calculation in room 1 is the fulfillment of noise standards in this room and the receipt of initial data for the calculation in room 2.

Calculation for premises through which the air duct passes in transit

For room 2 (for rooms through which the air duct passes in transit), the calculation involves solving the following five problems.

1st task. Calculation of the sound power emitted by the walls of the air duct into room 2, namely determining the value in dB using the formula:

In this formula: - see above the 2nd problem for room 1;

=1.12 - equivalent cross-sectional diameter of the air duct with cross-sectional area;

- length of the room 2.

The sound insulation of the walls of a cylindrical duct in dB is calculated by the formula:

where is the dynamic modulus of elasticity of the duct wall material, N/m;

- internal diameter of the air duct in m;

- thickness of the air duct wall in m;


The sound insulation of the walls of rectangular air ducts is calculated using the following formula in DB:

where = is the mass of a unit surface of the duct wall (the product of the material density in kg/m by the wall thickness in m);

- geometric mean frequency of octave bands in Hz.

2nd task. Calculation of the SPL at the design point of room 2, located at a distance from the first noise source (air duct), is performed according to the formula, dB:

3rd task. Calculation of the SPL at the design point of room 2 from the second noise source (SPL emitted by the wall of room 1 to room 2 - value in dB) is performed according to the formula, dB:

4th task. Compliance with permissible noise standards.

The calculation is carried out using the formula in dB:

As a result of the acoustic calculation and design of the HVAC for room 2, through which the air duct passes in transit, it must be ensured that the permissible noise standards are met at the design points. This is the first result.

5th task. Calculation of the value of the UZM in the discharge duct from room 2 to room 3 (room served by the system), namely the value in dB using the formula:

The amount of losses due to radiation of sound noise power by the walls of air ducts on straight sections of air ducts of unit length in dB/m is presented in Table 2. The second result of the calculation in room 2 is to obtain the initial data for the acoustic calculation of the ventilation system in room 3.

Calculation for premises served by the system

In rooms 3, served by SVKV (for which the system is ultimately intended), design points and permissible noise standards are adopted in accordance with SNiP 23-03-2003 “Noise Protection” and technical specifications.

For room 3, the calculation involves solving two problems.

1st task. The calculation of the sound power emitted by the air duct through the air outlet into room 3, namely the determination of the value in dB, is proposed to be performed as follows.

Particular problem 1 for low speed system with air speed v<< 10 м/с и = 0 и трех типовых помещений (см. ниже пример акустического расчета) решается с помощью формулы в дБ:

Here



() - losses in the noise muffler in room 3;

() - losses in the tee in room 3 (see formula below);

- losses due to reflection from the end of the duct (see table 1).

General task 1 consists of solving for many of the three typical rooms using the following formula in dB:



Here is the UZM of noise propagating from the machine into the discharge air duct in dB, taking into account the accuracy and reliability of the value (accepted according to the technical documentation for the machines);

- UZM of noise generated by the air flow in all elements of the system in dB (accepted according to the technical documentation for these elements);

- USM of noise absorbed and dissipated during the passage of a flow of sound energy through all elements of the system in dB (accepted according to the technical documentation for these elements);

- the value taking into account the reflection of sound energy from the end outlet of the air duct in dB is taken according to Table 1 (this value is zero if it already includes );

- a value equal to 5 dB for low-speed UAHV (air speed in highways is less than 15 m/s), equal to 10 dB for medium-speed UVAV (air speed in highways less than 20 m/s) and equal to 15 dB for high-speed UVAV (speed in highways less 25 m/s).

Table 1. Value in dB. Octave bands

2008-04-14

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

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

Acoustic calculation of UHVK is carried out with the aim of optimally selecting all the necessary means of noise reduction and determining the expected noise level at the design points of the room. Traditionally, the main means of reducing system noise are active and reactive noise suppressors. Sound insulation and sound absorption of the system and room 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 arisen. In the old SNiP II-12-77 “Noise Protection”, the method of acoustic calculation of HVAC buildings was outdated and therefore was not included in the new SNiP 03/23/2003 “Noise Protection” (instead of SNiP II-12-77), where it is not yet included absent.

Thus, the old method is outdated, but the new one is not. The time has come to create a modern method of acoustic calculation of UVA in buildings, as is already the case with its own specifics in other areas of technology that were previously more advanced in acoustics, for example, on sea vessels. Let's consider three possible methods of acoustic calculation in relation to UHCR.

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

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

This is known from the practice of using speaking pipes, for example, on ships for a long time. A typical design of a long line ventilation system has two defining quantities: L wH is the sound power entering the discharge pipe from the fan at the beginning of the long line, and L wK is the sound power coming from the discharge pipe at the end of the long line and entering the ventilated room.

The long line contains the following characteristic elements. We list them: inlet with sound insulation R 1, active silencer with sound insulation R 2, tee with sound insulation R 3, reactive silencer with sound insulation R 4, throttle valve with sound insulation R 5 and exhaust outlet with sound insulation R 6. 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 emitted by this element after the waves pass through it further.

If the sound insulation of each of these elements does not depend on all the 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 air duct over the period of sound vibrations T in W:

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

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

Inlet and outlet openings of the ventilation system with R 1 and R 6. According to the theory of long lines, the junction of two narrow pipes with different cross-sectional areas S 1 and S 2 is an analogue 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 junction boundary, multiplied by the cross-sectional area of ​​the pipes.

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

Analysis of this formula shows that at S 2 >> S 1 the properties of the second pipe approach the properties of the free boundary. For example, a narrow pipe open to a semi-infinite space can be considered, from the point of view of soundproofing effect, as bordering on a vacuum. When 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 silencer R2. Sound insulation in this case can be approximately and quickly estimated in dB, for example, using the well-known formula of engineer A.I. Belova:

where P is the perimeter of the flow section, m; l — muffler length, m; S is the cross-sectional area of ​​the muffler channel, m2; α eq is the equivalent sound absorption coefficient of the cladding, 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 active muffler channel R 2 is greater, the greater the absorption capacity of the walls α eq, the length of the muffler l and the ratio of the channel perimeter to its cross-sectional area P/S. For the best sound-absorbing materials, for example, 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 cross-sectional area S 3 then branches into two pipes with cross-sectional areas S 3.1 and S 3.2. This branching is called a tee: sound enters through the first branch, and passes further through the other two. In general, the first and second pipe may consist of a plurality of pipes. Then we have

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

Note that, due to aerohydrodynamic considerations, tees 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 noise suppressor is an acoustically narrow pipe with a cross-section S 4 , which turns into another acoustically narrow pipe with a large cross-section S 4.1 of length l, called a chamber, and then again turns into an acoustically narrow pipe with a cross-section S 4 . Let us also use the long line theory here. By replacing the characteristic impedance in the known formula for sound insulation of a layer of arbitrary thickness at normal incidence of sound waves with the corresponding reciprocal values ​​of the pipe area, we obtain the formula for sound insulation of a chamber noise muffler

where k is the wave number. The sound insulation of a chamber noise suppressor reaches its greatest value when 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 muffler, then the sound insulation formula must be applied sequentially from chamber to chamber, and the total effect is calculated using, for example, the boundary conditions method. Effective chamber silencers sometimes require large overall dimensions. But their advantage is that they can be effective at any frequency, including low ones, where active jammers are practically useless.

The zone of high sound insulation of chamber noise suppressors covers repeating fairly wide frequency bands, but they also have periodic zones of sound transmission, very narrow in frequency. To increase efficiency and equalize the frequency response, a chamber muffler is often lined on the inside with a sound absorber.

Damper R5. The valve 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 with an area S 5.1 is less than the internal diameter of the pipe (or other characteristic size). Soundproofing of 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 assessing 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 value

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

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

The least favorable case is adopted here, 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 an 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 determine this quantity under the assumption that each algebraic term y 0i of the sum is distributed according to the normal law with center M(y 0i) and standard

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

The error of the algebraic sum is determined as:

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

With 2Φ(t), = 0.9973 we have t = 3 = α and a statistical estimate with 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 of a probabilistic error estimate in a first approximation can be more or less acceptable. So, a probabilistic assessment of errors is preferable and it is this that should be used to select the “margin for ignorance”, which is proposed to be necessarily used in the acoustic calculation of UAHV to guarantee compliance with permissible noise standards in a ventilated room (this has not been done previously).

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

The second method of acoustic calculation. On sea vessels, 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, 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 UAHV, R i is the sound insulation in the i-th element of the UHVAC, dB (see the first method),

a value that takes into account the influence of a room on the noise in it (in construction literature, B is sometimes used instead of Q). Here r j is the distance from the j-th 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 angle of the source radiation, Φ is the factor directivity of the source, κ—the coefficient of disturbance of the diffuseness of the sound field).

If m air distributors are located in the premises of a modern building, the noise level from each of them at the design point is equal to L j, then the total noise from all of them should be below the noise levels permissible for humans, 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 UHCR and the sound insulation Ri occurring in all these elements are determined experimentally for each of them in advance. The fact is that over the past one and a half to two decades, electronic technology for acoustic measurements, combined with a computer, has progressed greatly.

As a result, enterprises producing UHCR elements must indicate in their passports and catalogs the characteristics of L wi and Ri, measured in accordance with national and international standards. Thus, in the second method, noise generation is taken into account not only in the fan (as in the first method), but also in all other elements of the UHVAC, 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 included in the first method. But it can be determined with the necessary accuracy by standard measurements, which is now being done for the second method. As a result, the second method, unlike the first, covers almost all UVA 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 ​​of noise acceptable for humans 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 powers of noise sources in watts, and the sound insulation of elements in decibels, according to the specified formula for the acoustic calculation of UHFV, is valid only, at least, when there is no interference of sound waves in the system. And when there is interference in pipelines, it can be a source of powerful sound, which is what, for example, the sound of some wind musical instruments is based on.

The second method has already been included in the textbook and in the guidelines for course projects in building acoustics for senior students of the 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.

To select the “margin for ignorance”, it is preferable, as shown above for the first method, to use a probabilistic error assessment, which is proposed to be used in the acoustic calculation of UHVAC buildings to guarantee compliance with permissible noise standards in premises 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 radically increase 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 USSR Academy of Sciences and the Russian Academy of Sciences L.M. Brekhovskikh, 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, the sound propagation constant of which is γ 2 = β 2 + ik 2 and the acoustic resistance Z 2 = ρ 2 c 2. Let us denote the acoustic resistance in the medium in front of the layer from which 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 factor i ωt omitted, will be a superposition of waves traveling in forward and reverse directions with sound pressure

The input impedance of the entire layer system (formula) can be obtained by simply applying (n - 1) times 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:

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, similar to the first method, we have sound power emanating from the discharge duct at the end of a long UHVAC line and entering the ventilated room according to the following scheme:

Next comes the assessment of the result, as in the first method with a “margin for ignorance,” and the sound pressure level of the room L, as in the second method. We finally obtain the following basic formula for the acoustic calculation of the ventilation and air conditioning system of buildings:

With calculation reliability 2Φ(t) = 0.9973 (practically the highest degree of reliability), we have t = 3 and the error values ​​are equal to 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, aimed at 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 in engineering calculations using computer technology. According to the author, it can be proposed as a new method for acoustic calculation of ventilation and air conditioning systems in buildings.

Summing up

The solution to pressing issues of developing a new acoustic calculation method should take into account the best of the existing methods. A new method for acoustic calculation of UVA buildings is proposed, which has a minimum “margin for ignorance” BB, thanks to taking into account errors using the methods of probability theory and mathematical statistics and taking into account interference phenomena by the impedance method.

The information about the new calculation method presented in the article does not contain some necessary details obtained through additional research and work practice, and which constitute the author’s “know-how”. The ultimate goal of the new method is to provide the choice of a set of noise reduction means for the ventilation and air conditioning systems of buildings, which increases, compared to the existing one, efficiency, reducing the weight and cost of the HVAC.

There are no technical regulations in the field of industrial and civil construction yet, so developments in the field, in particular, of reducing the noise of UVA buildings are relevant and should be continued, at least until such regulations are adopted.

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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 Technology” // Leningrad, “Shipbuilding”, 1983.
6. Bogolepov I.I. Industrial sound insulation. Preface by academician I.A. Glebova. Theory, research, design, manufacturing, control // Leningrad, “Shipbuilding”, 1986.
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11. Guidelines for the calculation and design of noise attenuation of ventilation units. Developed for SNiP II-12–77 by organizations of the Research Institute of Building Physics, GPI Santekhpoekt, NIISK. - M.: Stroyizdat, 1982.
12. Catalog of noise characteristics of process equipment (to SNiP II-12–77). Research Institute of Construction Physics of the USSR State Construction Committee // M.: Stroyizdat, 1988.
13. Construction norms and rules of the Russian Federation. Sound protection. SNiP 23-03–2003. Adopted and put into effect by Decree of the State Construction Committee of Russia dated June 30, 2003 No. 136. Date of introduction 2004-04-01.
14. Sound insulation and sound absorption. Textbook for university students studying in the specialty “Industrial and Civil Engineering” and “Heat and Gas Supply and Ventilation”, ed. G.L. Osipova and V.N. Bobyleva. - M.: Publishing house AST-Astrel, 2004.
15. Bogolepov I.I. Acoustic calculation and design of ventilation and air conditioning systems. Guidelines for course projects. St. Petersburg State Polytechnic University // St. Petersburg. Publishing house SPbODZPP, 2004.
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18. www.integral.ru. Firm "Integral". Calculation of the external noise level of ventilation systems according to: SNiP II-12–77 (Part II) - “Guide to the calculation and design of noise attenuation of ventilation units.” 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.
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Ventilation calculation

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

The parameters of the air entering the intake openings and openings of local suction 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 of 90 cubic meters per hour. In summer, it is necessary to install an air conditioner in order to avoid exceeding the temperature in the room for 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 an air conditioner is its main characteristic; it determines the volume of the room it is designed for. For approximate calculations, take 1 kW per 10 m 2 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 a given room, a simplified method was used:

where:Q - Heat inflow

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 m2 and a height of 3 m, the heat gain will be:

Q=15·3·35=1575 W

In addition, the heat emission from office equipment and people should be taken into account; it is believed (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 copy machine 0.3 kW, By adding these values ​​to the total heat inflows, you can obtain the required cooling capacity.

Q additional =(H·S opera)+(С·S comp)+(P·S print) (4.9)

where:Q additional - Sum of additional heat inflows

C - Computer heat dissipation

H - Operator Heat Dissipation

D - Printer Heat Dissipation

S comp - Number of workstations

S print - Number of printers

S operators - Number of operators

Additional heat inflows in the room will be:

Q add1 =(0.1 2)+(0.3 2)+(0.3 1)=1.1(kW)

The total sum of heat inflows is equal to:

Q total1 =1575+1100=2675 (W)

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

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

Noise level calculation

One of the unfavorable factors of the production environment in the computer center is the high level of noise created by printing devices, air conditioning equipment, and fans of cooling systems 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 emissions 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 noise level for a given workplace. If the calculation results are higher than the permissible noise level, then special noise reduction measures are required. These include: covering the walls and ceiling of the hall with sound-absorbing materials, reducing noise at the source, proper layout of equipment and rational organization of the operator’s workplace.

The sound pressure levels of noise sources affecting 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: a hard drive in the system unit, fan(s) of PC cooling systems, a monitor, a keyboard, a printer and a scanner.

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

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 we take into account that it is unlikely that peripheral devices such as a scanner and printer will be used simultaneously, then this figure will be even lower. In addition, when the printer is operating, the direct presence of an operator is not necessary, because The printer is equipped with an automatic sheet feed mechanism.