8, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). ESL-PA-8-6-8 Published in Vol., Part. For personal use only. Additional reproduction, distribution, or transmission SL-8- (RP-9) Qualification of Fan-Generated Duct Rumble Noise Part : Test Facility Joshua Kading J. Adin Mann, III, PhD Michael B. Pate, PhD Associate Member ASHRAE Member ASHRAE Member ASHRAE This paper is based on findings resulting from ASHRAE Research Project RP-9. ABSTRACT Duct rumble noise in HVAC air distribution systems is commonly attributed to the poor aerodynamic discharge conditions of fan outlets. To date, qualitative descriptors published in ASHRAE Handbooks have been used as design guidelines for engineers to limit the amount of duct rumble noise. The objective of this study was to design, build, and verify a test facility that could be used to accurately quantify the amount of duct rumble noise that is caused by the change in discharge conditions of a fan and duct system. The measured changes in the noise level needed to be due to the change of the discharge configuration and not outside influences on the system. To cover a large range of typical installation configurations the test facility was to be capable of changing between fourteen different discharge conditions, consisting of four different fan orientations and different distances between the fan outlet and the primary duct inlet. Further, the fan was operated at eight operating points on the fan curve that were chosen to cover the full operating range of a fan. The design and evaluation of the test facility is described. INTRODUCTION The discharge configuration of a fan affects the amount of low frequency duct rumble noise downstream of the fan. Rumble noise, defined for this study as noise in the 6 Hz to Hz one third octave bands, is considered by many practicing engineers as an indication of turbulence caused by poorly designed fan discharge conditions. The ASHRAE Applications Handbook () presents a qualitative diagram correlating the discharge conditions with a qualitative descriptor of the resulting noise. However, there is no sound level data that is connected with the qualitative descriptor. Specifically, even though the descriptor gives a rough guideline it does not give a means to predict the increased rumble for various designs. The predictions are needed to compare design alternatives such as weighing the cost of modifying the orientation and position of the fan compared to a noise control measure such as lagging or lining the duct. With noise level predictions, the various design alternatives can be effectively compared. ASHRAE therefore funded a project to provide information on the changes in rumble noise generated by various fan configurations. In addition to providing predictions of the rumble noise levels, an auxiliary goal of the work was to identify the mechanisms for the changing rumble noise. The specification on the design of the test system was that four different fan discharge conditions could be tested with four different lengths of duct from the fan to the first transition in the system. The measurements were to be performed at 8 operating points of the fan, to be accomplished by varying inlet restriction to the fan and the fan speed. The design and verification of the test system along with initial results are presented. TEST SYSTEM The system consists of three rooms shown in Figure : a fan room, a measurement room and an outlet room. The fan room was the inlet to the duct system, the measurement room is acoustically isolated from the rest of the system and the outlet room was open to the rest of the building, including the Joshua Kading is a mechanical engineer at Stanley Consultants, Inc., Muscatine, IA. J. Adin Mann, III, is an associate professor and Michael B. Pate is a professor in the Department of Mechanical Engineering, Iowa State University, Ames, IA. 8 ASHRAE 7 Published in Vol., Part.
8, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). ESL-PA-8-6-8 Published in Vol., Part. For personal use only. Additional reproduction, distribution, or transmission Figure Test system layout. fan room. The inlet to the fan room was controlled with a hole in the door with adjustable area, thus acting as the volume damper. The fan room was -7 x - and -8 tall. The measurement room was -7 x - and -8 tall. These were the desired room dimensions to limit the number of low frequency room modes in the measurement room. Both the fan room and the measurement room have structural insulation walls, as well as insulated floors. Further, the ceiling was not connected to the rest of the building. The volume damper in the fan room door had a maximum open area of 7 x. There were three boards, two inches and one inches tall, each constructed from ¾ thick medium density fiberboard (MDF) that could be added or removed from the bottom of the dampers damper. These served as a means to make large adjustments to the open area of the volume damper. In addition, there was a 8-/ tall sheet of MDF that slides over the top section of the hole. This sliding board allowed a finer analog control over the inlet area that was necessary to ensure that the system was operating at the desired operating points. A board was placed between the fan room inlet at the fan inlet. The board was placed at distance nearly the same at the distance from the fan inlet ant the wall opposite the fan room inlet. The board was used to provide the greatest possibility that the inlet to the fan was uniform from each side. 8 The duct work was made from gauge sheet metal without insulation lining. Each joint was constructed with easily modified duct joints. The fixed section of the duct work began at a transition piece that extended from the middle of the fan room to the fan room wall. This transitioned the outlet of the fan from 9. x 6. to the measurement section of the duct which was 8 x 8. The transition was 7 - long, creating a slope ratio of : from the fan ducting to the test duct section. ASHRAE Application () recommends a slope of :7 to :. The transition was enclosed in a 5 by 9 by tall box constructed from MDF and lined with duct board fiberglass. The box was designed to isolate the transition duct from break in noise from the fan room. The fiberglass insulation on the inside of the box was suspended against the MDF in order to limit any vibration damping that might occur if the insulation touched the duct. The duct work continued through the measurement room and out into the outlet room through a 7 long duct work and terminates into an outlet plenum designed to keep the noise in the outlet room from the measurement room. The test ducting consisted of 5 long sections with stiffening creases every and a -/ angle support bars every -6 running along the exterior of the 8 width of the cross section. In the measurement room there was a single section Published in Vol., Part.
8, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). ESL-PA-8-6-8 Published in Vol., Part. For personal use only. Additional reproduction, distribution, or transmission that was -7 long. This short length was in the measurement room because a 5 section was required to exit the measurement room to join into the outlet sound trap. Because of space limitations outside the measurement room, the location of the sound trap was fixed. On all the joints of the fixed section of the duct work a sealing adhesive was used along with bolts to join the duct together. The duct suspension system was standard and its vibration was not monitored. It was assumed that any excitation of the duct support would be part of a typical installation and should be included in the study. Further, while the duct walls could have resonance frequencies in the range of the study, they were again assumed to be part of the system being characterized, so were not identified or removed. The inlet to the duct system could be varied depending on the desired fan configuration. The fan outlet had a flexible rubber connector that consisted of a long fan flange, a length of rubber flex, and a length of duct flange. The duct section from the outlet of the flexible connector to the input of the transition section was variable. The fan could discharge horizontally, directly into the duct system, or vertically in an up blast position, into a length of duct and then into an elbow. Figure shows the fan exhausting in an up blast position, with a straight length of duct 8 long, and then the elbow at the inlet to the fixed section of ducting. When the elbow was not in place the farthest distance the fan can be extended from the transition is. The distance between the fan and the elbow could be up to 55 long. The duct work from the fan to the elbow was the same dimensions as the fan outlet, 9. x 6. The elbow had double wide turning vanes spaced evenly apart. The last piece of duct work that could be added to the system for the duct configuration was a short transition piece. The piece s inlet was 6 x 9. and the outlet is 9. x 6. The length is only 7. This allowed for the fan to be rotated 9 and have a discharge perpendicular to the elbow. On all of the joints in this section of the duct the ducting was bolted together. Clamps were used when possible and duct tape was applied to the exterior of the ducting joints. Fan Discharge Conditions The test system was flexible enough to achieve a total of fourteen discharge conditions. This was accomplished by manipulating the fan orientation and duct arrangement between the fan discharge and the down stream ducting. The four fan orientations are referred to as discharge configurations (shown in Figure ). Three of these configurations had four different duct lengths between the fan and the duct system inlet, and one had two different lengths between the fan and the duct inlet. Configuration : the fan was in front blast, blowing directly in to the transition piece. This configuration could only be moved away from the transition piece because of the room size. Configuration : the fan discharged air in a circular direction to the transition piece. The fan could be to 55 away from the elbow. Configuration : the fan discharged air against the circular direction of the duct. The fan could be to 55 away from the elbow. Configuration : the fan discharged air perpendicular to the duct. In this configuration there was a transition piece at the elbow that directed the air into the duct system. The transition piece was 7 long and therefore the fan was always at least 7 away from the elbow. This distance could also increase to 55. The distance between the fan and the elbow were set in terms of effective duct diameter, D h, which for a rectangular duct is D h =.------------------------------ ( H W).65 ( H + W).5, where H is the height and W is the width of the duct. The effective duct diameter for the fan outlet used in the study (9. x 6 ) was. For the measurement duct (8 x 8 ) the effective duct diameter was. The distance away from the elbow was measured in increments of this duct diam- () Figure Four fan outlet configurations with distance D as the variable length ( to in. in Configuration, to 55 in. in Configurations and, and 7 to 55 in. in Configuration ). Published in Vol., Part. 9
8, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). ESL-PA-8-6-8 Published in Vol., Part. For personal use only. Additional reproduction, distribution, or transmission eter. There were distances away from the elbow that were achieved by the system:, ( ), (8 ) and. (55 ) duct diameters. It was hypothesized that the distance between the fan outlet and the first flow disturbance would have a great impact on the aerodynamics of the system. In order to measure all of these configurations, the fan needed to be capable of rotating on two axes. The fan also needed to be mounted in such a way that it could be lifted up and down depending on the distance D. To do this, the fan was always mounted to a base. The mounting base was a double steel strut that held the fan and motor assemble together. The mounting base had six vibration isolation springs that it supported by. The four corner springs were 8 lb, isolation springs, the middle two are lb, isolation springs. For Configuration the motor needed to be raised to reach the keyway of the fan shaft. This was done by mounting two additional cross beams directly under the motor. The mounting base was made with -5/8 gauge x combination framing strut. The mounting base was supported by the support base, also made of -5/8 gauge framing strut, and kept up by six pillar supports. The top and bottom cross bars helped to support the structure. The middle bars could be shifted up or down depending on the configuration or duct diameter that was desired. The isolation springs from the mounting base rested directly on the middle bar of the support structure. Additional diagonal beams were added on the sides and across the center to give extra rigidity and support. Because of the changes in the structure, the diagonal beams were not in the same location between configurations. The fan and motor were mounted to a single structural base with an 8 direct drive coupler shaft according to ASHRAE Standards (999) arrangement 7. The coupler was chosen so the assembly did not need to be rigid like a concrete pad. This was necessary because of the large number of configurations required for the test plan. The 5 hp motor was controlled with an inverter control variable speed drive that could reach a peak speed of 9 rpm. The fan was a Size 8 Class, double wide/double inlet, forward curve, 8-inch diameter wheel, scroll fan. Outlet Plenum The air flow exhausted into an outlet plenum that was constructed as a sound trap. The sound trap was designed to eliminate the noise path from the outlet to the rest of the system (Figure ). The outlet plenum had an internal path that forms an arc from the duct at the top of the plenum to the outlet at the bottom of the plenum. The exterior was made from medium density fiber board (MDF) shell that was mounted on a x structure. The interior consisted of three layers. The first was fiberglass board x x, these boards were cut to be stacked and fill in between the MDF and the duct lining insulation. The second layer was duct board fiberglass that consisted of x 8 x sections. This lining was the section that the air was directly against and was designed to withstand high velocities. The last layer was perforated aluminum sheet metal, oval pattern, with a width of., gauge, that was used to create a rigid lining to the insulation. The perforated aluminum sheet metal was held in place by wire hangers and stapled to the exterior walls (where possible) at the inlet and exit. The geometry of the sound trap resulted in a center airflow radius of. The inlet to the plenum was the same as the duct 8 x 8 ; the air flow was sloped to an exit of width of 5 -. This gave a ¾ inlet to exit ratio in order to control the air velocity and the turbulence that was generated at the exit. Selection of Operating Conditions The required operating points for the fan were stated in the research project RFP. These populate the fan curve to give a representation of the effects on sound level at different typical operating points. This fan curve was used as a basis for setting all of the operating points. The eight required operating condition matrix were as follows: One fan speed at. inch of water pressure drop that represents the 9% wide open flow load curve. Two fan speeds at. and.5 inches of water pressure drop that represents the 8% wide open flow load curve. Three fan speeds at.,.5 and. inches of water pressure drop that represents the 6% wide open flow load curve. Two fan speeds at.5 and. inches of water pressure drop that represents the 5% wide open flow load curve. Since there was some discussion in the research monitoring committee on the procedures for setting these conditions, the procedures are presented in detail. The values in the first row in Table are taken from the fan curve provided by the fan manufacturer. These values are all at a fan speed of, rpm. For the next rows, the flow rate is calculated as a percentage of % WOF and then from the fan curves, the fan pressure drop is found. These points define the load curve which is referred to by the percent wide open flow (%WOF). Figure Outlet plenum diagram the inlet is screwed to the outlet of the system duct. Published in Vol., Part.
8, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). ESL-PA-8-6-8 Published in Vol., Part. For personal use only. Additional reproduction, distribution, or transmission Fan laws (ASHRAE Systems and Equipment ()) were then used to determine the values for operating conditions for each required pressure drop and flow rate (%WOF). where N = the desired fan speed (rpm) N =, rpm p = the desired pressure drop (inch wg) p = the pressure drop of the %WOF (inch wg) Q = the desired volume flow rate (CFM) Q = the desired volume flow rate of the desired %WOF (CFM) Using these equations and the fan curve supplied by the fan manufacturer, the conditions for each of the specified operating points were calculated, Table. During testing, the first step was to set the system to a desired load curve. This was done by setting the fan at the speed of one of the operating points on the load curve and then Table. Fan Curve Values Fan Flow Rate % of Wide Open Flow (%WOF) -- p N = N ---- p -- p Q = Q ---- p Volume Flow Rate (CFM) Fan Pressure Drop (inches water) %, 9% 9,98.7 8% 7,76. 6%, 5. 5% 7,77.9 () () varying the inlet flow area until the required flow rate was obtained. The other operating points were obtained by changing the fan speed to the required value in Table. The flow rate was measured, but the inlet restriction was not varied to compensate for any difference between the measured and desired flow rate. Thus, setting the operating points requires determining the setting of the inlet restriction for each of the load curves for each of the fan/duct configuration. Acoustic Isolation Because the sound measurements were taken in the measurement room, it was critical that the measurement room be acoustically isolated from the rest of the system. There were many potential sound paths (ASHRAE Applications ().) Since there were no return or supply air vents in the room the major concerns were the airborne noise transmitted directly through the wall, the structural noise that is transmitted through the floor and walls (also known as flanking paths), and the breakout noise that is generated from the duct itself. Of these three, the only one that needed to be measured was the breakout noise of the duct. The other two paths had to be minimized so the measurements could reflect the change in noise of the duct. The ASHRAE research RFP required a transmission loss between the measurement room and fan room of at least db at one-third octave bands between 6 Hz and 5 Hz. Noise reduction measurements were performed in the rooms with the duct and outlet plenums installed. The duct work passes through holes in the walls that were ½ to ¾ larger than the duct. The duct was framed around the wall openings with MDF (medium density fiberboard) leaving a small gap so the MDF was not touching the duct work. Fiberglass insulation was added in the large gaps and the remaining gaps are sealed with caulking. The interior of the measurement room separating wall was lined with MDF for additional mass. After preliminary tests the door inlet to the measurement room was covered by a sheet of MDF board. Table. Calculated Values for Operating Points in Terms of %WOF and Pressure Drop Flow Rate (%WOF) Fan Speed (rpm) Volume Flow Rate (CFM) Fan Pressure Drop (inches water) 9% 986 5,7. 8% 7 9,7.,6 5,88.5 58 5,99. 6% 96 9,9.5,59,878. 5% 99 5,55.5,75 7,. Published in Vol., Part.
8, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). ESL-PA-8-6-8 Published in Vol., Part. For personal use only. Additional reproduction, distribution, or transmission Sound tests were performed by placing a broad-band white-noise source in the fan room and measuring the sound pressure level at 8 locations in the fan room and 9 locations in the measurement room, Figure. The duct work entrance in the fan room was covered by a sheet of MDF that covered the inlet. The requirements were a minimum noise reduction of db at all one-third octaves, which was achieved as indicated in Figure. Airflow Pressure Traverse The airflow pressure traverse station was used to measure pressure drop and airflow rate. The airflow station was 6. duct diameters (6-6 ) downstream from the transition exit and.9 (5 ) duct diameters upstream from the outlet plenum. The airflow station consisted of six pitot tubes connected to a single pressure transducer. It was assumed in using these pitot tubes that the airflow was parallel to the duct axis. The pressure traverse was set at five positions in the duct height in order to obtain air velocity measurements at the points specified in the ASHRAE Handbook Fundamentals (ASHRAE Handbook Fundamentals 5). The five heights were -/8, 5-/ 8, 9, -7/8, and 6-5/8 above the duct bottom and the six pitot tubes were spread across the bottom of the duct at, - /,, 7, 6-/ and 5 from the duct side. The measured velocity pressure was averaged over all pitot tube locations and then used to calculate a air flow rate through the duct in cubic feet per minute. Q C p = w ----- ( L W) ρ V = velocity (ft/min) p w = averaged measured velocity pressure (lbf/in ) C = correction factor 96.5 ρ = air density (lbm/ft ) H = duct height (ft) W = duct width (ft) The pressure drop across the fan was estimated rather than directly measured. ASHRAE Standard 5 specifies that to measure the pressure rise across a fan the taps must be a distance at least an 8.5 duct diameters away from the fan. There were a large number of discharge configurations some that included the fan outlet to be directly against the elbow, making it impossible to consistently measure the down stream pressure 8.5 duct diameters down stream. Therefore the static pressure of the duct system was measured at the airflow station. This measured pressure drop, Δp meas, includes the pressure drop across the fan, Δp fan, and the down stream ducting system, Δp duct. () 6 of duct length and any duct diameters between the fan and elbow. Using tabulated values for parameters to model each of these elements, the value for Δp duct was estimated as Δp duct = ----------- fl + C ρ V ----------- D h 97, (6) where ΣC = the summation of the local loss coefficients f = friction factor L = duct length D h = equivalent duct diameter V = air velocity. A separate friction factor was calculated for each section of straight duct work f. ε 68 = ------ + -----.5. (7) D h Re If f =.8: f = f If f <.8: f =.85f +.8 where ε = material absolute roughness factor, ft Re = Reynolds number. The loss coefficients, C, in Equation 6 were assumed to be constant for the duct geometry. Future work could be done using factors that take into account the flow profiles that are measured. Setup for Sound Testing The sound pressure level in the room and in the duct were measured in both discrete frequency spectra from to 8 Hz and one-third octave bands from.5 to, Hz. The primary output of the study was the one-third octave levels between 6 and 5 Hz. Δp meas = Δp fan Δp duct (5) In order to use equation 5 to solve for the Δp fan the pressure drop across the down stream ducting, Δp duct needed to be estimated. The duct system includes the elbow, transition, 6 - Figure Measured noise reduction between the fan and measurement room. Published in Vol., Part.
8, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). ESL-PA-8-6-8 Published in Vol., Part. For personal use only. Additional reproduction, distribution, or transmission Sound levels were measured at an array of 9 locations in the measurement room and 7 in the duct. The 9 microphone room locations were on a x array with the row spaced apart and each column spaced apart, Figure 5. The array was spaced from the building north wall and 6 from the building west wall. These measurements were made at different heights to avoid any room modes. The heights were 6, 56 and 8, denoted as L, L and L in Figure 5. The 7 locations on the duct were distributed along joints. The main purpose of these measurements was checking for duct modes. The microphones were mounted near joints in the duct to minimize the stiffness changes created by the microphone mounts. Joint A, at the entrance of the duct to the measurement room, consisted of a single microphone at the bottom corner of the duct. The second cross section, 5 down stream from the building west wall, joint B in Figure 5, contained six microphones. Five of the microphones were along the bottom evenly spaced apart; the sixth was along the side of the duct 9 high. Each duct microphone mount was designed to hold the microphone tightly, with minimal flow induced noise and a tight seal so that sound was not transmitted from the duct into the room through the microphone mount. The mount consisted of a plastic tube assembly that held the microphone with an o- ring around it. The microphone was secured with a rubber capped screw so as not to damage the microphone. The top of the tube contained a thick piece of fiberglass between the microphone face and the hole in the duct. A tube assembly at each of the seven measurement locations was positioned flush to the duct with a ½ inch diameter hole in the duct. When a microphone was not inserted into a holder, a ½ inch diameter wood dowel was inserted to minimize sound transmission through the assembly. bearings were mounted to and in the case of the motor, the accelerometer was mounted on one of the motor mount points. As an example of the results, Figure 7 shows the velocity peak values for the 8 operating points with Configuration and no duct diameters to the transition inlet. The values are all much higher than the.5 in/s requirement that are specified in the RFP. Table shows the fan speed in rpm and the resulting frequency in Hz. Table also contains a listing of the frequencies for the peaks in the frequency spectra for each operating condition. It is important to note:. The fan has 6 blades, so the blade passage frequency was far above the peaks that are causing concern. Thus the source of the peaks in the vibration response was not directly related to an interaction of each blade to an aerodynamic loading such as passing the cutoff in the fan shroud.. The fan and motor shaft were running at the same speed. The data in Table shows that the fan vibration was dominated by the shaft rpm and harmonics while the motor vibration was typically at multiples of three times the shaft rpm. These results indicate that the motor and fan were effectively decoupled with the shaft coupler. In the case of the fan, the results in Vibration and Balancing A maximum peak vibration velocity of.5 in/sec was specified in the research project RFP. In order to minimize vibration, each aspect of the system was balanced individually. For example, the tolerance of fan/motor shaft coupler was between.5 and.55. This measurement was made and verified at the beginning of each configuration. However, because of the large variations of the configurations for the tests, the support structure was more flexible than a typical field application. Therefore, it was important to monitor the fan vibration for every test. The fan vibration was measured with a tri-axial accelerometer since the direction of greatest vibration was not known. The accelerometer was attached with the orientation shown in Figure 6. The accelerometer signal was time integrated to produce a velocity time signal. The average of discrete frequency spectrum and 8 seconds of each signal was recorded for every test condition. The vibration measurements were taken at two locations, one on the fan and one on the motor. In the case of the fan, the accelerometer was mounted on the ¼ steel plate that the shaft Figure 5 Figure 6 Floor microphone locations. Accelerometer axis (the x-axis runs in plane with the duct work, the y-axis funs along the coupler shaft, and the z-axis runs perpendicular to the floor). Published in Vol., Part.
8, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). ESL-PA-8-6-8 Published in Vol., Part. For personal use only. Additional reproduction, distribution, or transmission Figure 7 Velocity peaks of the operating points for Configuration, duct diameters. Table. Operating Points, Fan Speed in Terms of RPM and Frequency and Peak Frequencies Operating Points Fan Speed Peaks in Frequency Spectra % WOF ΔP (inch) rpm 5% 6% 8% Hz (fundamental + two harmonics) Fan Motor.5 9 5.5,., 6.5 5.6,.6, 5 6.9,., 6.9.,7 9.6, 9., 58.8 9.,.6, 9. 58.8, 8., 76.9. 59 9.9, 8.8, 9.7.,.,.., 6.,..5 9 5.,., 5.6 5,.6, 5.6 5.6,.6, 6..,7 9.5, 9., 58.5 9.,.6, 8.8 58.8, 8., 76.9. 7.9,.8, 5.7.9,.8,.6 6., 7.9,.8.5,6 8.8, 7.6, 56. 8.8, 7.5, 5.9 56.9,.8, 7. 9%. 988 6.5,., 9.5 6.,., 9. 5.,., 5. Table also show that the vibration was dominated by a once per revolution imbalance in the fan and fan shaft assembly. The peaks in the discrete frequency sound spectra have been compared to the harmonics of the fan and motor assembly vibration. The peaks in the discrete frequency spectra generally corresponded to the vibration harmonics. The th and 8 th harmonics have the largest peaks. These peaks reached up to db greater than the rest of the frequency spectra. In Figure 8 the discrete frequency spectra of 9% at. inch pressure drop is shown. The major peaks occur at the st, th, 8 th, and 9 th harmonics. These are only slightly greater than the rest of the frequency peaks. Figure 9 shows autospectra of the vibrations. All of the operating points have been analyzed in the same way. The largest peaks consistently occur at the th and the 8 th harmonic from the noise data. However, as seen in these examples, the peak vibration response was not at the th and 8 th harmonic. At the peaks in the vibration spectra there was little or no effect on the sound data. Therefore, the spectra support a conclusion that the peaks in the measured sound spectra are not caused by the fan or motor vibration and the vibration levels of the fan and motor are acceptable. To identify the source of the th and 8 th harmonic in the measured sound spectra, it should be noted that the fan blades are supported by four rods which extend over three quarters of the inlet. This is strong evidence that the th and 8 th harmonics that are seen in the noise spectra are results of aerodynamic disturbance of the fan rather than a vibration problem. This is further evidence that the fan and motor vibration levels, while exceeding the RFP initial requirement, are not contaminating the sound measurements. Repeatability An uncertainty analysis has been performed with the system. Three tests were taken at all of the operating points and at a single configuration (Configuration, duct diameters). Each test used a different testing sequences. The first was test was done with a constant operating point and rotating the microphones among the locations in the room. The second test was Published in Vol., Part.
8, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). ESL-PA-8-6-8 Published in Vol., Part. For personal use only. Additional reproduction, distribution, or transmission Figure 8 9% WOF in. pressure drop, average discrete frequency spectra with fan rpm harmonics shown (fan frequency 6.5 Hz). Figure 9 Frequency spectrum vibration of the fan. The x-axis is the horizontal; the y-axis is the axial, and the z-axis is the vertical. 9 load curve at. in. pressure drop operating point. done with constant microphone locations and rotating among the operating points. The last test was done randomly changing both the operating conditions and the microphone locations. Once the three test were taken the results were averaged and the variance in db was analyzed. Figure shows the uncertainty of three trials at the operating point 5% of the wide open flow,.5 inch pressure drop for the one third octave bands between 6 and 5 Hz. These results are consistent with all the operating points collected. Table is the maximum uncertainty for each operating point in the test that was described above. These values are only for Configuration and duct diameters away from the elbow. The uncertainty pertains to one measurement without changing the configuration or duct diameter. These results indicate that the uncertainty in the measured data is db, well within the range required by the RFP. Published in Vol., Part. 5
8, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). ESL-PA-8-6-8 Published in Vol., Part. For personal use only. Additional reproduction, distribution, or transmission Table. Uncertainty of Sound Measurements (db) 9%,. 8%,. 6%,. 8%,.5 6%,.5 5%,.5 6%,. 5%,. ±.8 db ±.6 db ±. db ±. db ±. db ±.9 db ±.8 db ±.7dB Figure Uncertainty of the 5% WOF at.5 in. pressure drop. System Modes The presence of modes in the system was analyzed. The modes that have been identified can be seen in the measured sound spectra. Further, the variation in the sound levels at the frequencies of the identified modes also has a spatial pattern typical of the modes in the duct and room. However none of the modes identified have a great impact on the sound results. Room Modes The predicted modes, shown in Table 5, are based on the dimensions of the room. The X, Y, and Z are the axis of the room and directions of the modes. The estimations are taken from equation, where L x, L y and L z are the dimensions of the room. f n = c -- n ---- x L x ---- n y L y ---- n z + + L z c = ( m s) The X axis follows the duct along the room, the Y axis is perpendicular to the duct, and the Z axis is along the vertical of the room. These modes were then compared to the narrow band sound pressure levels that were recorded in the room (Figure ). At 68 Hz there is the indication that there is a (,,) room mode when looking at the sound pressure level distribution in the room. However, looking at the narrow band it does not have a large impact on the sound level. It was therefore concluded that room modes did not impact the measured sound levels. (8) Duct Modes The duct modes were estimated in the same way as the room modes and it was concluded that they did not had a significant impact on the sound pressure levels. The narrow band sound pressure levels are shown in Figure. These are the sound averages of 5 locations along the duct and not all 7. To identify the duct modes the sound pressure levels are observed along the cross section of the duct. By looking at the narrow band frequency levels and the distribution of the sound pressure level in the duct there are no distinguishable duct modes. There is possibly one at Hz that peaks 5 db higher than the broad band levels; however, this is not significant enough to isolate. Looking at both the room and duct narrow band frequency, sharp peaks can be seen. These peaks are not consistent between each operating point, and were therefore concluded not to be room or duct modes. These are likely caused by aerodynamic disturbances. CONCLUSIONS The purpose of this facility was to measure the effects of aerodynamic disturbances on the sound levels. The system that was developed was capable of fourteen different fan outlet conditions: four different discharge configurations and four different duct distances to the duct inlet. Three of the four configurations could use all four of the duct distances, one of the configurations could only use two. The system could regulate operating conditions by changing fan speed and an inlet restriction. The operating conditions could be kept constant and verified by using the fan laws and a fan curve that were provided by the fan manufacturer. The measurement room was acoustically isolated from the rest of the system by added mass and an outlet plenum, and other potential sources of problems such as fan vibration, motor vibration, room modes, and duct modes, were studied and concluded to not impact the measured sound levels. The output from the system is the one third octave band and discrete frequency sound levels transmitted from the duct to the measurement room. The one-third octave band values were repeatable to within db. With the control available for the fan speed, fan orientation, and system flow rate, the system was shows capable to study the impact of fan orientation in a duct system on rumble noise transmitted through a duct system. REFERENCES ASHRAE Fundamentals (5). ASHRAE Handbook-5 ASHRAE Fundamentals, American Society of Heating, Refrigeration and Air-conditioning Engineers Inc. 6 Published in Vol., Part.
8, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). ESL-PA-8-6-8 Published in Vol., Part. For personal use only. Additional reproduction, distribution, or transmission Table 5. Predicted Modes in the Measurement Room X Y Z Freq X Y Z Freq 5 5 68 56 69 68 7 7 77 78 77 8 8 88 88 99 96 7 5 9 5 5 5 6 6 7 6 6 7 6 Figure Average discrete frequency spectra in the measurement room, Configuration, duct diameters, ΔP =. in. Figure Average discrete frequency spectra in the duct, Configuration, duct diameters, ΔP =. in. ASHRAE Systems and Equipment (). ASHRAE Handbook- Systems and Equipment, American Society of Heating, Refrigeration and Air-conditioning Engineers Inc. ASHRAE Applications, (). ASHRAE Handbook- Applications, American Society of Heating, Refrigeration and Air-conditioning Engineers Inc. ASHRAE Standards, (999). ASHRAE Standard 5-999Laboratory Methods of Testing Fans for Aerodynamic Performance Rating, American Society of Heating, Refrigeration and Air-conditioning Engineers Inc. Bies and Hansen () Engineering Noise Control Theory and Practice, New York, NY: Spon Press. p. 7. Published in Vol., Part. 7
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