Compressor design should isolate the sources of noise and vibration so that these machines 'meet present and future legislative requirements
Noise control of air compressors Frederic M. Oran Executive Vice President IndustrialAcoustics CO.
Bronx, N. Y. 10462
-1
product of their primary function: supplying air at elevated pressures for operation of other equipment or devices. As a resuit of legislation promulgated at all levels of government, attention is now being paid to this noise problem. The Occupational Safety & Health Act (OSHA) of 1972 provided hearing damage criteria that related the exposure time in hours and minutes to the soundlpressure level in decibels dB(A). For example, a level of 90 dB(A) is established for 6-hr exposure, while 4-hr exposure at 95 dB(A) is permissible. MIL Std. 1474, "Noise Limits for Army Material," provides that for an 8-hr exposure where infrequent communication is required that "category octave band sound pressure levels be met at the observer. These levels reflect current hearing conservation guidelines and yield 90 dB(A) where dB(A) is a weighting network that approximates the frequency response and sensitivity of the human ear to sound pressure levels. Compressor noise is objectionable for reasons other than hearing conservation. The typical unsilenced 600 cfm compressor at 1 m generates about 115-120 dB(A). Reduction to 90 dB(A) still leaves levels that preclude community acceptance, reception of warning signals, adequate telephone communications, andlor tolerable speech communications. The results of on-going research on the effects of noise on man will lead to additional legislation that will require lower noise levels to safeguard hearing and satisfy other noise criteria such as communication and community disturbance. Municipalities such as New York City have passed or have under consideration legislation that will limit compressor noise to 70 dB(A) at a distance of 1 m. This goal may be unrealistically low in light of the existence of higher levels of traffic noise on the city's streets. Other municipalities have 90 dB(A) as their current requirement. At the time of passage of the OSHA regulations, little if any attention was paid by manufacturers to controlling the noise output of their compressors. Their attempts were limited to providing the lowest priced industrial muffler on the engine discharge. The level of attention paid to noise control is best highlighted by the fact that the lightweight sheet metai enclosure used for weather and vandalism protection must be open if the compressor is to operate without overheating.
Status of the problem -. . .. . . . . . . . m e passage ot OSHA provided tne impetus to manutacturers to give noise control new consideration. Recently, buyers of compressors have asked that the manufacturer guarantee that the noise levels will not exceed those permissible under OSHA andlor have asked for guidance in retrofitting the compressors. Industrial Acoustics Company (IAC) has developed a good library of test data on compressors of different make and size, and has applied or recommended solutions that range
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. _ r.--....= ,_,, =,,,,mu>, ,,#",,,=, ," enclosing the compressor in a soundproof housing or treating the existing housing and improving the ventilation system. Noise reduction of the order of 25 dB(A) has been obtained. Typical solutions include the noise control systems designed for the Continental TC-106 gas turbine driven compressor (Figure 1) and the Gardner-Denver 750 cfm diesel enginedriven compressor (Figure 2). Figure 3 shows sound pressure levels at 1 m before and after silencing of the Gardner-Denver unit. The noise control systems were designed and manufactured by IAC after an extensive evaluation program that yielded a rank order for each of the major noise sources and provided the OEM (original equipment manufacturer) with a list of recommendations which, if followed, on future equipment would result in lower noise levels. The study also indicated the degree and type of noise control necessary to reach each of several design goals. fl
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Figure 2.
Noise control system for the 750 cfm diesel enginedriven air compressor
Approach to a solution The ultimate objective of a compressor noise program is the economic retrofit of existing equipment, and the provision of guidelines for the future design of quieter compressors. The first step is to obtain noise profiles of the existing compressor at specified locations and under specified conditions of air discharge and engine speed for essentially free field conditions. The results of these tests will broadly indicate the magnitude of the problem. The noise profiles could be used to designate safe disfance with reference to hearing conservation or satisfactory distance with reference to hearing regarding elimination of disturbances of personnel. The determination of noise profiles is the easiest part of the solution. The real difficulty lies in segregating the contribution of each of the noise sources at each position under each test condition. Based on previous experience in compressor studies, IAC believes that the following noise sources, at a minimum, must be isolated: engine exhaust noise (major source) radiator cooling fan noise (major source) engine casing radiated noise (major source) compressor noise (secondary source) other sources (generally negligible sources) Engine exhaust noise Selection of a method for isolating exhaust noise depends on understanding how exhaust noise changes when physical parameters of the system are altered. One logical approach used in previous studies was to locate the compressor inside a building and pipe the engine exhaust outside the building. This approach is satisfactory provided additional pressure drop does not change engine power output. Engine exhaust noise is directly proportional to the power output. The length and diameter of the exhaust pipe extension will affect the amount of sound energy radiated and measured. The engine exhaust noise of a 600 cfm compressor will dominate the low frequency part of the overall noise profile and peak unsilenced at about 0.13 times the engine rpm with an overall sound pressure level at 1 m of about 115 dB. The spectrum will decrease monotonically with a slope of about 7 dB per octave band. If the predictions are assumed to be correct, then the projected dynamic insertion loss (DIL) requirement for the exhaust muffler will be as shown in Figure 4. A typical design requirement less 5 dB has been plotted and compared to the estimated unsilenced engine exhaust level. It is assumed that the three major noise sources will be of equal value when silenced and, hence, will combine to result in a sound pressure level (SPL) 5 dB higher than each of the individual sources. A review of the applicable engine specification will indicate that the allowable back pressure of the exhaust system including 1032
Environmental Science 8 Technology
tions do not permit this approach, the fan propeller could be removed and tested outside the compressor unit. If the latter approach is used, the propeller will be mounted flush in the wall of a reverberant room and the test done in accordance with AMCA (Air Moving and Conditioning Association) Standard 300-67, Test Code for Sound Rating Test Setup No. 4. Based on the heat rejection rate as a percentage of the engine power, the cfm requirement will probably increase by about 20% for a total requirement of approximately 30,000 cfm when operating in a 125OF environment. Based on a fan operating efficiency of about 65 YO and total pressure of 2 in. of H20 we have predicted the unsilenced level at 1 m to be as shown in Table 1. It is unlikely that this DIL (see Table 1) could be obtained by fan modification and so we would expect that a short dissipative muffler at the intake and discharge of the cooling air into the enclosure will be needed.
muffler may be about 20 in. H20. The minimum normal thickness of the piping or mufflers used is about 0.06 in. To obtain the high DIL in the exhaust system, the muffler would either have to be inside the enclosure, which could cause excessive heat buildup, or lagged to reduce radiated noise. Typically a 4 in. or 5 in. diameter exhaust pipe is commonly used. The type of muffler suitable for the engine exhaust is one that must depend on reactive performance in contrast to dissipative or absorptive performance. The absorptive-type silencer normally requires the use of acoustic materials that could pick up unused fuel in the engine discharge and become a fire hazard. A reactive muffler, sometimes referred to as a chamber unit, can be of all-metallic design. It depends on the relative areas of the exhaust pipe and muffler and the volume of the muffler. To obtain the required DIL, an area ratio of about 16 is needed, or a low profile elliptic shape of comparable cross-sectional area should be used. The overall length (not necessarily in a straight line) will have to be about 60 in. Radiator cooling fan noise The primary source of noise in the air-cooled radiator system is the relatively inefficient fan used to propel the cooling air over the radiator. The fan noise problem is somewhat complicated in that a quieter compressor dictates the use of a housing-either the existing sheet metal housing with added dampening materials or a new enclosure. In any event, additional cooling may be required. For belt-driven fans some additional cooling could be obtained by increasing the speed. This approach, however, would mean greater horsepower consumption. As the additional horsepower may not be available, other approaches must be considered. One alternative is improvement in the operating efficiency of the fan. The typical fan used is of the propeller type and generally operates at efficiencies of about 50%. The addition of a shroud around the propeller will essentially convert the propeller fan to the tube axial type that results in increased efficiency of up to 6 0 % . This permits airflow increases of up to 20 % without additional horsepower. Another approach is to convert the fan to a vane axial type utilizing a more sophisticated wheel design with airfoil section blades. This approach may yield efficiencies of up to 75% and an associated increase in airflow of about 50%. A byproduct of the improved efficiency fan will probably be reduced noise and/or a shift in the noise spectrum. In consideration of the above, it would appear that the fan noise could be uniquely deterrrlined by driving the existing propeller fan with a quiet auxiliary power source. The power source could be a totally enclosed and silenced electric motor coupled to the fan by a belt drive in situ in the existing compressor unit with the engine not operating. If physical limita-
Engine casing radiated noise The noise generated by diesel engines is generally considered to be a consequence of the sudden pressure rise in the cylinders during compression and to the mechanical impact resulting from piston slap, the closing of valves, and the meshing of gears. These actions impart structure-borne vibration that causes air-borne noise. It would appear that the best method of isolating and measuring the engine/compressor casing radiated noise would be to pipe out the engine discharge, lag the air receiver, pipe out the compressor, and operate with the radiator fan disconnected (see Table 2). If the engine manufacturer's operating instructions prohibit operation even for a short time without cooling, an auxiliary silenced air source should be used to force cool air across the radiator. Data obtained in this manner can be correlated with information from design of diesel engine test cells for several leading manufacturers. When possible, data should be obtained by direct measurement on diesel engines at the manufacturer's plant. The ideal means of obtaining this type of reduction is to modify pressure rise rates in the cylinders or eliminate mechanical impact. Both of these approaches are beyond the scope of a retrofit program. Therefore, as the cause cannot be eliminated, the effect must be treated. Based on some semblance of correlation between accelerometer readings on the casing surface and measured sound pressure level (SPL),
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it would appear feasible that areas with high vibrational fields could be modified to decrease their radiation efficiency. Modifications could take the form of added reinforcement or the addition of viscoelastic materials to attain damping. The use of close fining or applied materials is normally limited to areas requiring little maintenance or access. In addition, most damping-type materials are temperaturesensitive with peak damping capacity at 70'F or 80°F and substantial decreases below 0°F or above 125'F. Both t e r n peratures are likely operating points if weather extremes and engine surface temperatures are considered. The typical existing sheet-metal housing (minimum of 0.075 in. thickness) would, at first glance, be a means of minimizing the radiation of engine casing noise. However, the typical sheet-metal housing basically lacks the design features necessary to ensure satisfactory reduction of noise. These features are: adequate acoustic seals and closure at all openings damping of all large surfaces to reduce drumming adequate acoustic absorption to minimize buildup of energy inside the enclosure. Theoretically, an enclosure without absorption will provide 0 dB noise reduction. A proven economical concept has been to discard the existing enclosure antj provide a pre!fabricated enclosure with known acoustical PI operties and pnwen performance. Compressor and other noise sources Noise level and spectrum content vary with the type of compressor be it reciprocating, axial, rotary, or screw. For purposes of analysis, it is assumed that the rotary-type air compressor is.commonly used. In any event, the compressor casing-radiated noise, independent of the type of compressor, is normally overshadowed by the casing-radiated noise of the driving engine. The rotary compressor essentially consists of a rotor containing eight blades that slide in and out of slots relative to the housing and that rotate counterclockwise. Centrifugal force moves the blade outward trapping incoming air in a triangular pocket that becomes smaller as the rotor turns. As the pocket diminishes in size, the air is compressed and finally moves through a labyrinth to the air discharge. Oil is injected or forced by air pressure into the housing and acts as a lubricant and cooling agent for the heat of the compressor, and also as a sealant. An air-oil separator and thermal bypass separates the air from the oil and returns the oil to the cooler prior to reuse. Some oil is carried out of the primary loop and separated from the air by impingement on the fiberglass element of the separator. An understanding of compressor design and operation is necessary to be able to rank the compressor noise in terms ^I
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estimate of the noise levels can be made by analysis if it is not possible to isolate the compressor noise during the test. The primary noise source is the high-speed rotor. For example, an Eblade rotor moving at 1800 rpm will generate a peak frequency of about 240 Hz. Based on the mass flow through the compressor, the number of blades, and the pressure rise, it is estimated that a sound pressure level of 115 dB at the peak frequency inside the CBsing will exist. Assuming a minimum transmission loss for ttl e compressor casing of about 20 dB, it is concluded that the compressor radiated _. ?........a> I I U I S ~ wuuw alrec:~ LW rneasure.~ levels. lt is conservatively assumei that the internal noise of the compressor at 115 dB radiatesi to the intake. The presence of an air intake filter assembly and air intake control system nor__,I._ .Lr ~ r ~ yo e a r~~ r ~ a UBC t ~ U U ~ X noise C needs no additional silencing The noise radiated to the discharge of the compressor is similarly attenuated by the air-oil separator, filter medium, and receiver. The other major noise sources associated with the compressor operation are a consequence of the release of high.II._IIL.
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Environmental Science & Technology
pressure air to the atmosphere, either through the service valve or through a blowdown valve. Valve noise is normally jet-like in nature and depends on the mass flow, size of valve, and upstream pressure. In most cases, the velocity in the throat of the valve will be sonic. Some valve manufacturers have developed so-called "quiet trim" that permits more gradual expansion of the jet and hence less noise. If this approach is not satisfactory, an inline absorptive silencer can be added to the service valve. In the case of the blowina valve a diffuser tvDe element that reduces the discharge velocity and, hence; the noise can be Wed. Occasionally diesel engines will be turbocharged. If so, this noise source normally has to be treated. In addition, noise from pulleys, sheaves, and belt slap should be considered, but it is unlikely that such sources will be significant. Similarly, the engine intake, normally treated with a combined filter-siIlencer, should not be a problem. I
the above prescribed test program. A wellequipped acoustical laboratow with modern instrumentation can taoe or directly read the necessary data under controlled eniironmental conditions. Technicians familiar with compressor operation are necessary. In addition to having the necessary outdoor open space, indoor "free field' space is helpful. If persistent inclement weather occurs, all tests could be run indoors without reflections interfering with the measurements. Measurements and recording of data should be performed in accordance with applicable portions of CAGI-PNEUROP [Compressed Air and Gas Institute (U.S.) and the European Committee of Manufacturers of Compressed Air Eouiomentl Test Code for the Measurement of 'Sound from Pneumatic Equipment, IEEE No. 85, Test Procedure for Airborne Noise IMeasurements, ANSI S1.2 Physical Measurement of Sound, innd ANSI S1.13 Methods for Measurement of Sound Pressure I-evels. I General considerations
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The nnn.ramim+ir mnc+rsin+ctn ha n A a r i n n nr ".,mniirlara.4 i,,. y'-'y" noise control equipment are heavily biased by the need to minimize any effect on function. TO this end, the best guidelines can be obtained from the manufacturers and from government documents such as MIL-C-55G. Compressors, Air Reciprocating and Rotary, Diesel Engine Driven and MIL-E11276; noise control must facilitate installation and minimize cost. Most important is compatibility with the compressor unit, which may be designed to operate in extreme environments. These conditions are detailed, to some extent, in an Army regulation entitled "Operation of Material Under the Extreme Condition of Environment." The succe?is of a noise control program aimed at retrofitting existing enginedriven compressors will be heavily depen....>:_.I... >!-.!~-,!_-dent upon the cuuruma~~on DT many engineeringoiscipiines ~ n 1:luding acoustics, aerodynamics, and mechanical design. I ..I " . . . . . . I _ . . . ,
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Frederic M. Oran is executive vice president of Industrial Acoustics Company, hc. (New York, N.Y.). He has overall responsibifity for the company's
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projects from conception to installation. Mr. Oran has sales responsibility for special prcducts such as gas turbine silencing systems to jet engine test cells. He holds masters degrees in business administration and aeronautical engineering, and is a registeredprofessional
