1. Paper_3_Noise reducing methods.doc

Noise Reduction of an I.C. Engine by Enclosure method with 25mm thick Glass wool as Noise Absorbing medium

S. R. Patil1 , S. S. Shinde2

1 Assistant Professor A.I.S.S.M.S.C.O.E., Mechanical Engg. Department, Pune.2 Assistant Professor, A.I.S.S.M.S.C.O.E., Mechanical Engg. Department, Pune.

Abstract: — Major noise sources in a noisy portable I. C. engine driven generator set have been identified. The separation of engine combustion and mechanical noise indicated that their contribution was almost equal. The main noise sources were considered to be mechanical in nature. For noise reduction muffler, silencer, vibration control, barriers and enclosures can be used and should be considered at the design stage. A of method of noise reduction is depends upon the application of an I. C. Engine. Enclosures are specifically designed structures that reduce the amount of transmitted sound power impinging on the space of interest. In either barriers or enclosures, passive control is achieved by the use of an acoustic absorbing material or sound reflection. An absorbing material is used in this context to describe the material’s ability to attenuate or “absorb” the sound’s incident pressure wave, thus reducing the amplitudes of the transmitted and reflected pressure waves. The resulting sound pressure level on the opposite side of the absorbing material will therefore be reduced. The absorbing material is something as simple as a metal or plastic panel, or as elaborate as a multilayered partition which consists of specially engineered acoustic foam or fiberglass, different thicknesses of panel layers, as well as different panel materials. In this paper we have considered I.C. engine as a prototype of mobile generator and enclosure design for that. First SPL is measured without acoustic enclosure and then with acoustic enclosure of 1mm thick M.S. material with glass wool of 25mm thickness. Keywords— I.C. engine, Noise reduction, Acoustic enclosure, Acoustic material, Glass wool, thickness, Octave frequency band.

I. INTRODUCTION

Noise is defined as a sound, generally of a random nature, the spectrum of which does not exhibit distinct frequency component. Enclosure is a commonly used form of noise control at the workplace. In an enclosure, any pipes from the machine should not be rigidly attached to the enclosure, but should be supported by vibration isolating hangers. Air gaps need to be minimized. Ducts for fresh air and exhaust gases should be adequately silenced. The insides of the panels of enclosures are lined with absorbent materials to avoid build-up of noise due to reflections. The wall of an enclosure may consist of several elements, each of which may be characterized by a different transmission loss [2]. The noise sources in the engine were identified and the passive control measures taken, including a partial enclosure, helped to achieve a maximum noise reduction of about 8.5dBA. Passive control may not be very effective or could result in significant increase in weight. This has led to interest in other means of noise control, particularly active control, which is more easily implemented at low frequencies [5]. While designing the enclosure, two types of enclosure resonances are taken into account. The first is the mechanical resonance of the enclosure panels and the

other is the acoustic resonance of the air-space between the enclosed machine and the enclosure walls [6]. The light weight, close fitting enclosure will have some attenuation of the radiated sound due to the mass of the structure. Its effectiveness will also depend on the source to panel distance, the stiffness of the enclosure material, the enclosure panel vibration and the panel thickness [10]. Such element must be considered in turn in the design of an enclosure wall, and the transmission loss of the wall determined as an overall area weighted average of all of the elements.For this calculation the following equation is used:

TL =10log10

Where, Si is the surface area (one side only), andTLi is the transmission loss of the ith element.The first few panel resonance frequencies of the enclosure should not be in the frequency range in which sound attenuation is desired, acoustic resonances occur at the standing wave frequency given by (Ubhe 1996) [2].

F = c / 2d,Where c is the speed of sound and d is the distance between source and panel.Among various kinds of sound absorption materials, porous materials such as glass wool quilting and polyurethane foams are the most common and significant technique which are widely used for room acoustics and various electric devices [7]. For noise absorption fibrous sound-absorbing materials, 15 types of glass wool (thickness 8–54mm and bulk density 20–102 kg/m3) are available commercially [8].

II. DESIGN OF EXPERIMENT

Calculation of sound pressure level is done by two methods rectangular parallelepiped and hemisphere parallelepiped. In the present work, rectangular parallelepiped method is used. In this method, the first step is to make a grid according to the dimensions of engine. Length breadth and height of engine are approximately 375mm, 175mm and 100mm. The grid is made by placing an engine at centre position and with the help of wire at required positions mark the different points. There are minimum 10 Grid points formed as shown in figure 1. Initial noise levels of the generator were measured in order to have reference

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values for comparison after applying control measures. The sound pressure levels obtained at a distance of 1m from generator [1].

Figure 1. Layout of microphone positionsSound pressure level can be measured for every grid point shown in figure 2 for different throttle positions to change engine rpm in increasing manner by adjusting the throttle position at N1, N2 and N3 of engine respectively such that, speed at N1< speed at N2< speed at N3.

Figure 2. Experimental set up

III. EXPERIMENTATION

FFT analyser of model Bruel & Kjaer Photon+ 986A0186, Frequency limit1Hz to 20 KHz, Operating temperature range of 100C to 500C with microphone Bruel & Kjaer 4188 is used.First SPL is measured for condition of without acoustic enclosure (EC1) at throttle position N1, N2, N3 in increasing manner. Then sane procedure is applied for the condition of acoustic enclosure with 1mm thick M.S. wall and glass of 25mm thickness (EC2) as shown in figure 3.

Figure 3. Measurement of SPL at different microphone positions

IV. EXPERIMENTAL DATA

Table 1. For EC1 and N1

Location

No.

Overall SPL dBA

Sound Pressure Level at Octave Band Centre Frequencies in (dBA)

63Hz

125

Hz

250 Hz

500

Hz

1k Hz

2k Hz

4k Hz

8kHz

1 83.3 48 64 70. 72 72 70 66 612 81.8 46 61 69. 76 70 66 64 633 80.2 45 62 67. 73 67 63 63 624 80.7 45 61 64. 73 70 62 65 605 82.4 47 63 63. 77 72 68 65 596 85.3 50 67 68. 77 76 73 68 627 94.0 58 75 80. 88 81 72 77 678 87.3 51 68 69. 83 76 69 67 609 83.1 47 61 69. 73 71 67 64 6110 81.1 45 59 67. 71 68 64 64 5911 82.5 48 63 68. 74 70 68 64 5912 90.1 55 71 76. 83 76 72 72 6113 82.0 46 63 68. 75 68 64 63 58Av 84.14 48 64 69. 77 72 68 66 61


Location No.

Overall SPL dBA


63Hz

125

Hz

250

Hz

500

Hz

1k Hz

2k Hz

4k Hz

8kHz

1 87.1 42 64 71 82 72 66 67 662 84.9 41 68 68 79 72 66 67 643 82.4 43 62 69 73 67 66 66 614 83.9 43 64 64 78 68 69 65 595 86.1 43 65 66 79 71 72 74 616 90.7 47 70 74 83 76 76 77 627 96.9 54 76 82 91 81 79 78 698 88.7 45 67 70 82 75 72 69 639 85.2 40 62 70 80 70 67 67 6410 83.2 40 63 67 77 70 69 66 6211 85.1 41 65 66 79 71 70 69 6012 92.0 52 73 77 86 79 70 77 6413 84.4 40 63 68 82 72 66 67 66Av 86.97 44 66 70 80 72 70 70 63

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Location

No.

Overall SPL dBA


63Hz

125

Hz

250

Hz

500

Hz

1k Hz

2k Hz

4k Hz

8kHz

1 89.2 55 65 70 82 74 7 74 682 88.0 50 63 68 82 74 7 71 673 87.2 50 63 68 80 71 7 70 664 87.7 52 60 62 77 74 7 72 645 89.5 51 62 65 80 74 7 75 646 93.0 52 68 67 85 80 7 78 667 100.1 61 74 79 92 86 8 79 718 93.0 54 66 69 85 76 7 75 669 89.6 51 62 68 80 78 7 72 7110 87.6 48 60 66 79 75 7 72 6711 89.6 51 60 64 79 76 7 73 6612 96.5 58 72 75 88 82 7 82 6913 88.8 47 61 65 80 76 7 74 66Av 90.75 52 64 68 82 77 7 74 67


Location No.

Overall SPL dBA


63Hz

125

Hz

250

Hz

500

Hz

1k Hz

2k Hz

4k Hz

8kHz

1 79.5 44 62 66 73 68 61 59 492 74.0 41 56 60 67 63 55 49 443 72.7 40 61 62 66 60 52 47 414 75.9 49 62 60 71 65 57 50 425 80.4 48 63 67 75 68 65 59 496 84.4 54 68 71 75 75 72 62 547 92.9 59 75 75 88 82 76 65 608 84.9 55 70 71 79 74 67 64 529 81.4 51 64 63 77 70 63 56 4910 75.9 48 59 61 71 63 54 48 4111 80.3 47 59 66 75 67 63 54 4612 88.8 55 70 72 84 75 71 62 5613 80.1 46 60 63 75 65 60 51 43Av 80.8 4 6 6 7 6 6 5 4


Location No.

Overall SPL dBA


63Hz

125

Hz

250

Hz

500

Hz

1k Hz

2k Hz

4k Hz

8kHz

1 81.6 41 61 64 74 69 64 63 502 75.3 44 57 56 69 63 57 53 433 76.2 45 58 63 69 62 59 52 434 79.3 46 58 64 74 64 63 56 445 84.5 47 61 70 79 69 69 60 516 87.6 52 68 72 81 73 75 64 577 93.7 57 73 76 89 81 75 63 618 85.5 51 66 71 80 73 68 62 529 80.6 50 61 63 74 70 63 55 4810 76.7 49 61 61 72 62 55 51 4311 83.1 52 65 68 77 68 64 58 4812 88.8 56 71 72 84 75 71 63 5813 78.3 44 58 63 72 66 55 50 42Av 82.4 4 6 6 7 6 6 5 5


Location No.

Overall SPL dBA


63Hz

125

Hz

250

Hz

500

Hz

1k Hz

2k Hz

4k Hz

8kHz

1 87.1 42 65 71 82 72 66 67 662 84.9 41 62 68 79 72 66 67 643 82.4 43 62 69 72 66 67 66 604 83.9 43 64 64 78 68 69 65 595 86.2 43 65 66 79 71 72 74 616 90.7 45 70 74 83 76 76 77 627 96.9 54 76 82 91 81 79 78 698 88.7 45 67 70 82 75 72 69 639 85.3 40 62 70 80 70 67 67 6410 83.2 40 63 67 77 70 69 66 6211 85.1 41 65 66 79 71 70 69 6012 92.1 52 73 77 86 79 70 77 6413 84.5 40 63 68 78 71 67 68 61Av 87.00 44 66 70 80 72 70 70 63

From these observations following graphs are plotted.

Figure 4. Overall SPL level in dBA to Throttle position

Figure 5. Overall SPL level in dBA to Enclosure conditions

Figure. 4 and Figure. 5 shows the nature of variation of overall SPL at all three throttle positions for both enclosure conditions. Which shows that maximum SPL is for Throttle Position 3 and enclosure condition 1.

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Figure 6. Avg. overall SPL to Octave band central frequency

Figure 6. shows variation in Avg. overall SPL at different frequencies in octave frequency band for both enclosure conditions and at all three throttle position.

V. RESULTS

From this experiment following results can be calculated,

1. The overall SPL is maximum for more throttle opening, and for enclosure condition 1 i.e. without enclosure up to 90.75dBA.

2. For throttle position N1 overall SPL is reduced by 3.28dBA, for N2 by 4.57dBA and for N3 3.75dBA.

3. By using acoustic enclosure of 1mm thick M.S. wall with 25mm thick glass wool sound pressure level can be reduced up to 4.57dBA. This is shown in figure 7.

4. Average overall SPL is maximum up to 92Hz at frequency 500 Hz for EC1.

Figure 7. Reduction in overall SPL for all throttle openings due to aplication of enclosure method

VI. CONCLUSIONS

1. As the throttle opening increases overall SPL increases irrespective of enclosure condition.

2. After application of acoustic enclosure SPL reduces for all throttle openings N1, N2 and N3.

3. Average overall SPL increases from centre band frequency 63Hz up to 500Hz and then again starts decreasing

4. Difference between average overall SPL is very less for lower frequencies and for higher frequencies this difference increases widely.

5. As overall SPL can be reduced by application of acoustic enclosure up to 4.57dBA, it can be used for less noise reduction applications such as portable generator of medium speed.

REFERENCES

1) N. Tandon, B. C. Nakra, D. R. Ubhe and N. K. Killa, “Noise Control of Engine Driven Portable Generator Set”,Applied Acoustics, Vol. 55, No. 4, pp. 307-328, 19982) N Tandon, “Noise-reducing designs of machines and structures”, Sadhana, Vol. 25, Part 3, June 2000, pp. 331-339.3) Indian Institute of Technology Roorkee, “Vehicle Noise and Vibration Control”. 4) Professor Colin H. Hansen, Dr Berenice I.F. Goelzer “Engineering Noise Control”, Department of Mechanical Engineering World Health Organization University of Adelaide South Australia 5005.5) C.R. Krishna, J. E. Wegrzyn, “Survey of Noise suppression systems for engine generator sets”, Brookhaven National Laboratory, Communications and Electronics Command, U.S. Army, October 1999.6) Joseph E. Blanks, Optimal Design of an Enclosure for a Portable Generator, Thesis submitted to the faculty of the Virginia Polytechnic Institute and State University in partial fulfillment of the requirements for the degree of Masters of Science in Mechanical Engineering February 7, 1997, Blacksburg, Virginia.7) Toshihiko Komatsuzaki and Yoshio Iwata, “Modeling of Sound Absorption by Porous Materials using Cellular Automata”, Graduate School of Natural Science and Technology, Kanazawa University, Kakuma-machi, Kanazawa, 920-1192, Japan.8) Takeshi Komatsu, “Improvement of the Delany-Bazley and Miki models for fibrous sound-absorbing materials”, 2008 The Acoustical Society of Japan, Industrial Research Institute of Shizuoka Prefecture, 2078 Makigaya, Aoi-ku, Shizuoka, 421–1298 Japan.

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1. Paper_3_Noise reducing methods.doc