Example of calculation of noise from ventilation systems. Acoustic calculation of the supply ventilation system

2008-04-14

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

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

The acoustic calculation of the SVKV is made for the purpose of optimal choice all necessary means of noise reduction and determination of the expected noise level at the design points of the room. Traditionally, active and reactive silencers have been the primary means of noise reduction in a system. Sound insulation and sound absorption of the system and the room is required to ensure that the norms of noise levels permissible for humans - important environmental standards - are met.

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 In the old SNiP II-12-77 "Noise Protection", the method of acoustic calculation of UHCW buildings is outdated and therefore has not been included in the new SNiP 23-03-2003 "Noise Protection" (instead of SNiP II-12-77), where it is still generally absent.

Thus, the old method is outdated, but the new one is not. It's time to create modern method acoustic calculation of UHCW in buildings, as is already the case with its own specificity in other, previously more advanced in acoustics, areas of technology, for example, on sea vessels. Consider three possible ways acoustic calculation in relation to the SVKV.

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

For pipe rectangular section side should be less than half the wavelength, and for round pipe- 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 side of the section 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 negotiating pipes, for example, for a long time on steamships. Typical scheme a long line of the ventilation system has two defining values: L wH is the sound power entering the discharge line from the fan at the beginning of the long line, and L wK is the sound power coming from the discharge line at the end of the long line and entering the ventilated room.

The long line contains the following characteristic elements. Let's 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 soundproofed R 5 and soundproofed outlet R 6. Sound insulation here means 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 tube 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 associated with 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 for the period of sound oscillations T in W:

where T = 1 / f is the period of sound vibrations, s; f - vibration 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 UHCS can be, of course, more than the above n = 6. Let's apply the theory of long lines to calculate the values of R i to the above characteristic elements of the air ventilation system.

Ventilation inlet and outlet 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 analogue of the interface between two media at 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, 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 above sections:

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 opened into a semi-infinite space can be considered from the point of view of sound insulating 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 silencer R 2. 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 flow area, m; l is the length of the muffler, m; S is the cross-sectional area of the muffler channel, m 2; α eq - 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

From the formula it follows that the sound insulation of the channel of the active muffler R 2 is the 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, 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 R 3. In ventilation systems, most often, the first pipe with a cross-sectional area S 3 then branches into two pipes with a cross-sectional area S 3.1 and S 3.2. Such a branch is called a tee: sound enters through the first branch, and passes through the other two. In general, the first and second tubes can be composed of a plurality of tubes. Then we have

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 tend 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 damper R 4. A chamber silencer is an acoustically narrow pipe with a cross section S 4, which passes into another acoustically narrow pipe of a large cross section S 4.1 of length l, called a chamber, and then again passes into an acoustically narrow pipe with a cross section S 4. We will use the long line theory here as well. Replacing the characteristic impedance in the well-known formula for sound insulation of a layer of arbitrary thickness at normal incidence of sound waves by the corresponding reciprocal values of the pipe area, we obtain the formula for sound insulation of a chamber silencer

where k is the wavenumber. The sound insulation of the chamber silencer reaches the highest 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 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 condition method. Effective chamber silencers sometimes require large dimensions. But their advantage is that they can be effective at any frequency, including low frequencies, where active mufflers are practically useless.

The zone of great soundproofing of chamber noise mufflers covers repeating rather 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 muffler is often lined with a sound absorber from the inside.

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

where c is the speed of sound in air. In the first method, the main question for us when developing a new method is to assess 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 supplied to the ventilated room - in this case, the values

We rewrite this expression in the following notation of the algebraic sum, namely

Note that the absolute maximum error of the approximate value is the maximum difference between its exact value y 0 and the 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 accepted, 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. We define this value 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 the 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

For 2Φ (t), = 0.9973, we have t = 3 = α and the statistical estimate for practically 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 “margin of ignorance”, which is suggested to be necessarily used in the acoustic calculation of the UHCS to ensure that the permissible noise standards in a ventilated room are met (this has not been done before).

But the probabilistic assessment of the errors of the result also indicates in this case that it is difficult to achieve high accuracy of the calculation results using the first method, even for very simple circuits and a low-speed ventilation system. For simple, complex, low- and high-speed SVKV schemes, 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... For a long time, ships have used a calculation method based in part on analytical dependences, 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 UHCW, R i is the sound insulation in the i-th element of the UHCW, dB (see the first method),

a value that takes into account the effect of a room on noise in it (in construction literature, sometimes B is used instead of Q). Here rj 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 radiation of the source, Φ is the factor directionality of the source, κ is the coefficient of disturbance of the diffuseness of the sound field).

If there are m air distributors in the room of a modern building, the noise level from each of which at the design point is equal to L j, then the total noise from all of them must be lower than the noise levels permissible 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 UHCW, and the sound insulation R i, which takes place in all these elements, for each of them is preliminarily found experimentally. The fact is that over the past one and a half to two decades, the electronic technique of acoustic measurements, combined with a computer, has progressed.

As a result, enterprises producing UHCW elements must indicate in their 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 generation of noise not only in the fan (as in the first method), but also in all other elements of the HVAC, which can be of significant importance 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 means of standard measurements, which is now being done for the second method. As a result, the second method, in contrast to the first, covers almost all UHCW 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 permissible for a person in accordance with 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 acoustic power of the noise sources in watts, and the sound insulation of elements in decibels, is valid only, at least when there is no interference of sound waves in the system, according to the specified formula for the acoustic calculation of the UHCW. And when there is interference in the 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 entered the textbook and methodological 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 of ignorance” or, in critical cases, requires experimental refinement of the result to the required degree of accuracy and reliability.

For the choice of the "margin of ignorance", it is preferable, as shown above for the first method, a probabilistic assessment of errors, which is proposed to be necessarily applied in the acoustic calculation of UHCW buildings to ensure that the permissible noise standards in rooms 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 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 a thickness of δ 2, the sound propagation constant of which is γ 2 = β 2 + ik 2 and the 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 forward and backward directions with sound pressure

The input impedance of the entire 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 tube. And thus, with interference in narrow pipes, we have the formula for sound insulation in dB of a long line of the ventilation system:

The input impedances here can be obtained both, in simple cases, by calculation, and, in all cases, by measuring on a special installation with modern acoustic equipment. According to the third method, similar to the first method, we have the sound power emanating from the discharge duct at the end of the long line of the SVKV 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 get the following basic formula for the acoustic calculation of the ventilation and air conditioning system of buildings:

With the reliability of the calculation 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, directed to the future, is more accurate and reliable, but also more complicated - 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 of urgent questions of the development of a new method of acoustic calculation should take into account the best of the existing methods. A new method of acoustic calculation of UHCW of buildings is proposed, which has a minimum "margin of ignorance" BB, thanks to the allowance for errors by the methods of probability theory and mathematical statistics and the allowance for interference phenomena by the method of impedances.

The information on the new calculation method presented in the article does not contain some of the necessary details obtained by additional research and practice, and which constitute the author's "know-how". The ultimate goal of the new method is to ensure the selection of a complex of means for noise reduction of ventilation and air conditioning systems of buildings, which increases, in comparison with the existing one, efficiency, reducing the weight and cost of the UHCS.

There are still no technical regulations in the field of industrial and civil construction, therefore, developments in the field, in particular, of noise reduction of UHCW buildings are relevant and should be continued, at least until such regulations are adopted.

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Description:

The norms and rules in force in the country prescribe that the projects should 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 tech. sciences, head. laboratory of noise protection of ventilation and engineering-technological equipment (NIISF)

The basis for the design of noise suppression of ventilation and air conditioning systems is the acoustic calculation - a mandatory attachment to the ventilation project of any object. The main tasks of such a calculation are: determination of the octave spectrum of air, structural ventilation noise at design 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 calculated points is carried out, taking into account the effectiveness of these measures.

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

The initial data for the 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, 1,000, 2,000, 4,000, 8,000 Hz. For approximate calculations, the corrected sound power levels of noise sources in dBA are sometimes used.

Design 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 are in transit; in the area of the intake or exhaust device, or just intake air for recirculation.

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

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

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

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

S is the area of an imaginary sphere or its part 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 a room adjacent to the room where the fan is installed

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

(3)

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

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

S is the area of the enclosing structure, m 2;

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

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

The design point is located in a room served by the system

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

(4)

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 with i - the same as in formula (1) - formula (2).

Attenuation in the network (in the air channel) D L P network - the sum of attenuation in its elements, sequentially located along the course of sound waves. The energy theory of sound propagation through pipes assumes that these elements do not affect each other. In fact, the sequence of shaped elements and straight sections form a single wave system, in which the principle of independence of damping 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 cancel each other out, and therefore an energy approach that 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 ducts made of sheet material is due to losses due to deformation of the walls and radiation of sound outward. The decrease in the sound power level D L P 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. 1.

As you can see, in air ducts of rectangular cross-section, the attenuation (decrease in USM) decreases with increasing frequency of sound, and increases in circular cross-section. In the presence of thermal insulation on metal air ducts, shown in Fig. 1, the values should be approximately doubled.

The concept of attenuation (decrease) of the level of the sound energy flow cannot be equated with the concept of change in the level of sound pressure in the air duct. As a sound wave moves through a channel, the total amount of energy it carries decreases, but this is not necessarily associated with a decrease in the sound pressure level. In a narrowing channel, despite the attenuation of the total energy flux, 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 sound attenuation in a section with a variable cross-section is equal to:

(5)

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

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

Attenuation at bends (in bends, 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 channel cross-section increases due to the resulting non-uniformity of the velocity field. The square rotation acts like a low pass filter. Cornering noise reduction 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 turning the channel. The maximum noise reduction occurs when the turning depth is approximately half the wavelength. The value of the theoretical modulus of the coefficient of sound transmission through rectangular bends can be judged from Fig. 2.

In real structures, according to the data of 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 range of wavelengths close in magnitude to the doubled channel width. Then it smoothly increases again at high frequencies, reaching 8-13 dB. In fig. 3 shows the curves of noise attenuation 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 maximum noise reduction at a = l / 2 can be used to reduce noise with low-frequency discrete components by adjusting the channel sizes at bends to the frequency of interest.

Noise reduction on bends less than 90 ° is roughly proportional to the steering angle. For example, the noise reduction in a 45 ° corner is equal to half the noise reduction in a 90 ° corner. Noise reduction is not taken into account when cornering less than 45 °. For smooth turns 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 channel branchings, the transverse dimensions of which are less than half the wavelength of the sound wave, the physical causes of attenuation are similar to those of attenuation in the elbows and branches. This attenuation is determined as follows (Fig. 5).

Based on the equation of continuity of the medium:

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

and the decrease in the sound power level with a branch cross-sectional area

	(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 in the case of 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 channel cross-sectional area to the smaller one.

The decrease in sound power levels when the channel sizes are larger than the half-wavelength of non-planar waves with 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 the reflection of waves with a length less than the channel dimensions does not occur.

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

Directional characteristics of sound emission by air distribution grilles are shown in Fig. eight.

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

The attenuation of the noise level in the room is determined by the formula (2). The design 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 generated by several gratings, then the acoustic calculation is performed taking into account their total impact.

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

(16)

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

D L 'Pseti - attenuation in the network between the source and the considered transit section, dB;

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

S T is the surface area of the transit section that 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.

Design points are located in the area adjacent to the building

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

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

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

(17)

where L Pokti - octave sound power level of the noise source, dB;

D L Pnetsi is the total decrease in the sound power level along the path of sound propagation in the duct in the considered octave band, dB;

D L ni - directivity index of sound radiation, dB;

r is the distance from the noise source to the design point, m;

W is the spatial angle of sound radiation;

b a - attenuation of sound 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.

Structure-borne noise calculation

Structure-borne noise in rooms adjacent to ventilation chambers results from the transfer of dynamic forces from the fan to the ceiling. The octave sound pressure level in the adjacent insulated room is determined by the formula

For fans located in a technical room outside the overlap above the insulated 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 vibration isolator elements on which the refrigerating machine is installed, N s / m;

Z lane - the input impedance of the floor - the bearing slab, in the absence of a floor on an elastic foundation, the floor slab - if available, N s / m;

S is the conditional overlap area of the technical room above the insulated room, m 2;

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

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

S u - area of the insulated room, m 2;

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

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

Determining the required noise reduction

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

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

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

The total number of noise sources n in determining D L tri of the required octave sound pressure level reduction in the urban area 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 the noise of the ventilation system, 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 ventilation system under consideration, - 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 reduction of noise generated by fittings and air distribution devices of the considered branch, - the number of fittings 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 a sound pressure level at the design point that is 10 dB lower than the permissible one, with their number no more than 3 and 15 dB less than the permissible one with no more than 10 of them.

As you can see, acoustic calculation is not an easy task. The required accuracy of its solution is provided by acoustics 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 incorporated directly into the project. So, just by installing mufflers, the length of which is 300-500 mm longer than the required one, additional costs for medium and large objects can be 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, AVOK, no. 2004. No. 4.

4. Guidelines for the calculation and design of sound attenuation of ventilation units. Moscow: Stroyizdat, 1982.

5. Yudin E. Ya., Terekhin A.S. The fight against the noise of mine ventilation units. Moscow: Nedra, 1985.

6. Reducing noise 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. Fight against fan noise. M .: Energoizdat, 1981.

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

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

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

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

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

Design points are located in the area adjacent to the building

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

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

where L Pokti - octave sound power level of the noise source, dB;

∆L Pnetworki is the total decrease in the sound power level along the path of sound propagation in the duct in the considered octave band, dB;

∆L ni - directivity index of sound radiation, dB;

r is the distance from the noise source to the design point, m;

W is the spatial angle of sound radiation;

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

Ventilation systems make noise and vibrations. The intensity and area of sound propagation depends on the location of the main units, the length of the air ducts, overall performance, as well as the type of building and its functional purpose. The calculation of ventilation noise is designed to select the mechanisms of work and the materials used, in which it will not go beyond the standard values, and is included in the ventilation system project, 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 is noisy due to the sharp pressure drop from one side to the other. Engine - due to breakage or improper installation. Chillers make noise for the same reasons, and compressor malfunction is added.
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 the elements of the building. The solution is carried out by installing special supports or vibration damping gaskets. A vivid 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 gaskets, then its operation can cause acoustic discomfort for the owners of the installation or their neighbors.

Transfer methods

There are three paths for sound propagation, 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. It is distributed both inside and outside the building. The main source of stress for people. For example, a large store with air conditioners and refrigeration units located at the back of the building. Sound waves travel in all directions to nearby houses.
Hydraulic: noise source - pipes with liquid. Sound waves are transmitted over long distances throughout the building. It is caused by a change in the size of the pipe section and a malfunction of the compressor.
Vibrating: source - building structures. Caused by improper installation of fans or other parts of the system. Transmitted throughout the building and beyond.

Some experts use scientific research from other countries in their calculations. For example, there is a formula published in a German journal: with its help, the generation of sound by the walls of the duct is calculated, depending on the speed of the air flow.

Measurement 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 "sound level meter": it determines the strength of the propagation of sound waves. Measurement is carried out using three filters that allow you to cut off unnecessary 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 multiple points. For example, in the immediate vicinity of a noise source, then at a certain distance, then at the most distant point.

Code of practice

The rules for calculating noise from ventilation and algorithms for performing calculations are specified in SNiP 23-03-2003 "Protection against noise"; GOST 12.1.023-80 “Occupational safety standards system (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 guideline values are given for intermittent mechanical ventilation and open windows. If closed windows and a forced air exchange system capable of providing the design frequency are taken into account, then other parameters are used as norms. The maximum noise level around the building is increased to a limit that allows maintaining the normative parameters inside the building.

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

A - the best conditions.
B - a comfortable environment.
B is the noise level at the limit.

Acoustic calculation

It is used by designers to determine noise absorption. The main task of the acoustic calculation is to calculate the active spectrum of sound loads at all points determined in advance, and the resulting value is compared with the normative, maximum permissible. 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.

Calculation points:

direct place of equipment installation;
adjacent premises;
all rooms where the ventilation system works, including basements;
rooms for the transit application of air ducts;
air inlet or exhaust outlet.

The acoustic calculation is carried out according to two basic formulas, the choice of which depends on the location of the point.

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

where L Pi is the sound power in each octave;
∆L for i - a 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 i:

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 point of calculation, m 2;
B - constant value of the acoustic constant in the room, m 2.

The point of calculation is taken outside the building in a nearby area. The sound from the work spreads through the walls of the ventilation shafts, grilles and the fan housing. It is conventionally assumed that the noise source is a point source (the distance from the fan to the calculated position is an order of magnitude larger than the size of the apparatus). Then the octave noise pressure level is calculated using Equation 3:

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

If several noise sources act on one point, for example, a fan and an air conditioner, then the calculation methodology changes slightly. You can't just take and add all the sources, so experienced designers take a different path, removing all unnecessary data. The difference between the largest and the smallest source in terms of intensity is calculated, and the resulting value is compared with the standard parameter and added to the level of the largest.

Reducing the sound load from the fan

There is a set of measures to neutralize the noise factors from the fan operation, which 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 the standard microclimate parameters, but, at the same time, without a large power reserve. There is a wide range of fans with mufflers on the market, they are well protected from unpleasant sounds and vibrations.
Choice of installation site. Powerful ventilation equipment is installed only outside the served 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 tenants on the top floor will immediately feel discomfort. Therefore, in such cases only roof fans are used.
Selection of the speed of air movement through the channels. The designers are guided by an acoustic design. For example, for a classic 300 × 900 mm air duct, it is not more than 10 m / s.
Vibration isolation, soundproofing and shielding. Vibration isolation involves the installation of special supports that dampen vibrations. Soundproofing is carried out by pasting the enclosures with a special material. Shielding involves cutting off a sound source from a building or room using a shield.

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 challenging task that requires skills and experience in this area.

The company "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 neighboring regions. Technical support service will answer all questions by phone numbers indicated on the page. Remote collaboration is possible. Contact us!