Gleason, Steve

Ma, Youngchan

Sjoholm, Lars I.

English

Purdue E-Pubs

Purdue UniversityPurdue e-PubsInternational Compressor Engineering Conference School of Mechanical Engineering2008Noise Reduction in Bus A/C Systems with ScrewCompressors Part IILars I. SjoholmThermo KingSteve GleasonThermo KingYoungchan MaThermo KingFollow this and additional works at: https://docs.lib.purdue.edu/icecThis document has been made available through Purdue e-Pubs, a service of the Purdue University Libraries. Please contact epubs@purdue.edu foradditional information.Complete proceedings may be acquired in print and on CD-ROM directly from the Ray W. Herrick Laboratories at https://engineering.purdue.edu/Herrick/Events/orderlit.htmlSjoholm, Lars I.; Gleason, Steve; and Ma, Youngchan, "Noise Reduction in Bus A/C Systems with Screw Compressors Part II" (2008).International Compressor Engineering Conference. Paper 1859.https://docs.lib.purdue.edu/icec/18591345, Page 1 International Compressor Engineering Conference at Purdue, July 14-17, 2008 Noise Reduction in Bus A/C Systems with Screw Compressors  Part II: Compressor With System Interaction Lars Sjoholm1, Steve Gleason2, YoungChan Ma31Ingersoll Rand Climate Control Technologies, Minneapolis, Minnesota, USA Lars_Sjoholm@irco.com, 952-887-3430 2Ingersoll Rand Climate Control Technologies, Minneapolis, Minnesota, USA Steve_Gleason@irco.com, 952-887-2485 3Ingersoll Rand Climate Control Technologies, Minneapolis, Minnesota, USA YoungChan_Ma@irco.com, 952-887-2526 ABSTRACT This paper evaluates screw compressors with different features and the response of bus air conditioning (A/C) systems to those features from a sound and vibration perspective.  Due to the relatively high discharge pulsation frequency of the screw compressor, the resonance situation in bus A/C systems can be challenging.  Evaluations are for complete bus A/C systems in which the compressor is belt-driven from the bus engine. Results are presented for both steady state operation and the transient conditions encountered with start/stop operations and gear changes during road tests. Data from a common coach bus with one rooftop unit and an articulated transit bus with two rooftop units are presented for systems with noise problems resolved using screw compressors with modified features.1. INTRODUCTION Whether we are taking the bus from the airport terminal to the car rental facility, riding the bus across town on our daily commute, or enjoying the scenery from a chartered luxury coach, we all appreciate the comfort that comes from an air conditioned bus.  It’s a great change from the days when an opened window offered the only respite from the heat of the day.  In most applications, a screw compressor quietly provides excellent capacity to cool a bus on the hottest of days, and few people give a second thought as to what mechanical system has relegated operable windows on buses to school kids.  While all buses of a similar type – i.e. coach, transit, articulated, etc. – tend to look very much alike, after close observation it quickly becomes apparent that they all differ in numerous ways.  The underlying bus structure tends to be unique, and there are numerous engine choices and mounting configurations used; these two factors affect where the air conditioning compressor is located and how it is driven, and then the style of the bus helps determine what type of A/C system is installed and where it is positioned on the bus.  Rooftop systems requiring long runs of refrigerant tubing from the back of the bus tend to be the applications where noise complaints associated with screw compressors have occasionally occurred.   It was becoming clear that one particular screw compressor application had a noise problem that was not going to be resolved by testing alternate designs in the laboratory.  Changes that led to noise improvements in isolated compressor tests typically had no effect in the bus.  A modified compressor bracket that showed notable improvement in the lab was not a solution in the field.  (Ma, et al., 2008)  Very capable field service personnel had already tried rubber isolation under the compressor to further eliminate structure-born noise as the likely culprit. The compressor was not audible over the engine at the back of the bus, which points away from an air-born noise 1345, Page 2 International Compressor Engineering Conference at Purdue, July 14-17, 2008 problem, so it was becoming apparent that it was a fluid-born noise problem.  The noise was transient and was affected by both compressor speed and the loading condition of the compressor.  This compressor has axial unloading that activates according to ambient conditions, and the noise was notably different at partial load compared to full load. Figure 1 depicts an overhead view of the rear seating area of the bus with the location of the compressor highlighted in green.  The noise was most intense in the seats closest to the compressor, so noise measurements were made in this area as shown by the red dots in Figure 1.  Sound pressure maps were generated to help compare the effects as changes were made in the compressor to mitigate the noise.  Figure 2 shows four of these sound maps at one particular speed for three different loading conditions as well as the engine-only case when the compressor was not operating.  These particular maps show the overall A-weighted sound level, and there is a notable change as the compressor transitions from the fully loaded case.  The resulting noise increase was very tonal near 1 kHz, so the level of disturbance is greater than might be assumed from the 2 – 4 dBA increases shown on the maps. Figure 1: The rear seating area in the bus with the compressor indicated in green and the microphone locations in red Figure 2: Baseline sound maps in the bus for different loading conditions at the compressor1345, Page 3 International Compressor Engineering Conference at Purdue, July 14-17, 2008 2. NOISE MITIGATION ATTEMPTS ON THE BUSPrevious experience in a somewhat similar application had shown there was a potential for the screw compressor’s internal check valve to affect the noise from the compressor, so one of the first modifications was to install a modified check valve. The particular modification that was tried had the unfortunate effect of exciting the discharge line not only in the rear corner of the bus but all the way to the front unit on the long, articulated bus as well.  While that was certainly not the desired result, it did show the nature of the problem was sensitive to changes we could make in the compressor.  The challenge was to sort out the most effective changes that could be made and quickly implemented to resolve the noise problem. Reducing the discharge pulse from the compressor would minimize the fluid-born noise problem.  Stabilizing the rotors was identified as one way to accomplish this within the design parameters.  A set of reversed-thrust bearings shown in Figure 3 was installed in the compressor after the baseline check-valve was reinstalled. Tests of that setup showed a much more favorable result as seen in Figure 4; the tonal noise from the screw compressor was no longer apparent at any steady-state condition that we could test.  The bus was taken for a test drive in this condition, which proved to be quiet until the bus reached high speeds on the expressway or the engine sped up quickly during gear changes.  It turned out that an electronic governor had limited the idle speed of the bus when it was in stationary, so the reversed-thrust bearings were only a solution up to a certain speed range. Figure 3: Arrows show the forces on the rotors from the reversed-thrust bearings1345, Page 4 International Compressor Engineering Conference at Purdue, July 14-17, 2008 Additional changes were necessary to account for the greater dynamic forces at high speeds and conditions with rapid speed transitions, and modifying the drive gears in the compressor proved to be a good way to do that. From our trials with reversed-thrust bearings, we know that a stabilizing force on the rotors had a positive effect. The blue arrows in Figure 5 show the axial gas force on the rotors. The red arrows show the axial forces from the original gears on the male rotor and the drive shaft.  The axial force from the gear on the male rotor is acting in the opposite direction of the gas force on the male rotor. This is typically done to balance the axial gas force on the male rotor and give the thrust bearings longer life. The green arrows show the axial forces from the reversed helix gears acting on the male rotor and the drive shaft. Notice that the blue and green arrows in Figure 5 are pointing in the same direction on the male rotor, making it very stable; however, this can reduce the life of the thrust bearings. These particular bearings were designed to handle this additional load. The reversed-thrust bearing on the male rotor was no longer needed because of the force from the reversed helix gears; however, the reversed-thrust bearing on the female rotor remained in the compressor.  Figure 6 shows a contour plot of noise measured in the rear seat of the bus before and after the reversed helix gears were implemented along with the reversed-thrust bearing.  The bus in this condition no longer had a compressor noise issue.  Figure 7 shows a contour plot of a speed sweep from low idle to high engine speeds and back to low idle both before and after these noise mitigation steps were implemented on a coach bus.  Note how the noise tends to intensify near 1 kHz and 2 kHz as the compressor went though a particular speed range in the baseline condition, but this noise was eliminated once the reversed-thrust bearing and reversed helix drive gears were implemented. Figure 5: Compressor with reversed helix gears. Blue arrows show axial gas forces on the rotors, green arrows show axial forces from the reversed helix gears and red arrows show axial forces from the original gears.  1345, Page 5 International Compressor Engineering Conference at Purdue, July 14-17, 2008 Figure 6: Steady-state noise contour plot for the baseline compressor in this application (top) and the same test after the compressor with the reversed-thrust bearing and reversed helix gears was installed. Frequency, Hz Time1345, Page 6 International Compressor Engineering Conference at Purdue, July 14-17, 2008 3. CONCLUSIONS With the modifications to the screw compressor to stabilize the rotors, noise from the compressor is no longer audible in the bus through the operating speed range.Compressor noise after the modifications is insensitive to the loading of the compressor.REFERENCES Ma, Y., Sjoholm, L., Gleason, S., 2008, Noise Reduction in Bus A/C Systems with Screw Compressors, Part I: Compressor Evaluation, International Compressor Engineering Conference, Purdue University, IN, USA ACKNOWLEDGEMENT The authors wish to thank the management of Ingersoll Rand Climate Control for permission to publish this paper.  We also acknowledge the assistance of Mr. Peter Lawrence and Mr. Jon DeTuncq, both of Ingersoll Rand. Figure 7: Noise contour plot of speed sweep from low idle to high rpm’s and back down again for the baseline compressor (top) compared to the same test with the modified compressor. Frequency, HzTime

Noise Reduction in Bus A/C Systems with Screw Compressors Part II

Purdue UniversityPurdue e-PubsInternational Compressor Engineering Conference School of Mechanical Engineering2008Noise Reduction in Bus A/C Systems with ScrewCompressors Part I: Compressor EvaluationYoungchan MaThermo KingLars I. SjoholmThermo KingSteve GleasonThermo KingFollow this and additional works at: https://docs.lib.purdue.edu/icecThis document has been made available through Purdue e-Pubs, a service of the Purdue University Libraries. Please contact epubs@purdue.edu foradditional information.Complete proceedings may be acquired in print and on CD-ROM directly from the Ray W. Herrick Laboratories at https://engineering.purdue.edu/Herrick/Events/orderlit.htmlMa, Youngchan; Sjoholm, Lars I.; and Gleason, Steve, "Noise Reduction in Bus A/C Systems with Screw Compressors Part I:Compressor Evaluation" (2008). International Compressor Engineering Conference. Paper 1858.https://docs.lib.purdue.edu/icec/1858 1344, Page 1  International Compressor Engineering Conference at Purdue, July 14-17, 2008 Noise Reduction in Bus A/C Systems with Screw Compressors, Part I: Compressor Evaluation YoungChan Ma1*, Lars Sjoholm2 , Steve Gleason3  1Ingersoll Rand Climate Control Technologies, Minneapolis, MN, USA E-mail: youngchan_ma@irco.com,  Phone: 952-887-2526  2Ingersoll Rand Climate Control Technologies, Minneapolis, MN, USA Email: lars_sjoholm@irco.com  3Ingersoll Rand Climate Control Technologies, Minneapolis, MN, USA E-mail: steve_gleason@irco.com  *Indicate Corresponding Author    ABSTRACT Screw compressors for bus air conditioning (A/C) have many advantages: low noise, reduced vibration eliminating the need for isolation mounts, a large operating speed range, and quiet unloading. However, sometimes interesting resonance situations can occur. The opportunity for this to happen is largely increased for roof-mounted A/C systems, where the discharge tubing can become quite long, given that the compressor is situated in the engine compartment. This paper discusses different features, such as unloading, volume ratio, oil viscosity, oil flow and bracket design, etc. from sound and vibration perspectives.  1. INTRODUCTION This particular bus A/C screw compressor consists of a step-up gear (68/17 gear ratio) that drives a male rotor with 5 lobes, which drives a female rotor with 7 lobes. The input shaft has an electric clutch, and the pulley on the shaft is belt driven by the main bus engine. The speed of the compressor varies within a range of about 700 to 3,000 rpm. This compressor has an oil separator and oil sump located on the discharge side as shown in Figure 1. The rotors, the bearings, the gears and the shaft seal are lubricated with oil supplied with pressure differential between the discharge pressure and the lubrication points.  One dominant noise source of the screw compressor is the pressure pulsations coming from discharging the gas from the closed flutes under compression as they open into the discharge cavity. This discharge pulsation is described in the following equation:  p(t)= *c (Q(t)-Avg. Q)/S         (1)  Where: p(t) = pressure pulsation amplitude in the discharge cavity    = density of the gas,   c = the speed of sound in the gas,   S = the cross section area of the discharge cavity,   Q(t) = the volume flow in the discharge cavity,   Avg. Q = the average volume flow in the discharge cavity,   1344, Page 2  International Compressor Engineering Conference at Purdue, July 14-17, 2008 Besides the pressure pulsation amplitude, the corresponding frequency is of high interest when identifying noise sources.  Fp = nm * ZM / 60         (2)  Where: Fp = the basic frequency or first harmonics of the pulsations, Hz  nm = the male rotor speed, rpm,   ZM = the number of lobes on the male rotor,  When the compressor has a step-up gear, following equation applies:  nm = G * n          (3)  Where: nm = the male rotor speed, rpm,  G = the gear ratio,  n = the compressor input shaft speed, rpm,  Combining equation (2) and (3) gives:  Fp = G*n*ZM / 60         (4)  Example: The gear ratio G = 68/17 and the male rotor lobes ZM = 5 give a frequency Fp of 333 Hz when the input speed is 1000 rpm (Fp = (68/17)*1000*5 / 60 = 333 Hz).  The discharge pulsations are affected by the contact between the two rotors. The rotor contact noise is affected by transferred torque between the rotors, rotor contact band defined by the rotor profile, rotor lead, and rotor divide.     It is also indirectly affected by oil flow, oil viscosity and proper oil drainage out of the discharge port shown in Figure 2.  The drive gears (see Figure 3) are another noise source.  Fg = n * ZG / 60          (5)  Where: Fg = the basic frequency or first harmonics of the gear meshing, Hz  n = the compressor input shaft speed, rpm,  ZG = the number of teeth on the drive gear,  Example: The number of teeth on the drive gear ZG = 68 gives a frequency Fg of 1133 Hz when the compressor input shaft speed is 1000 rpm (Fg = 1000*68/ 60 = 1133 Hz).  Noise coming from gas pulsations and gear meshing can be structural transmitted, in this case over the bearings into the housings and into a mounting bracket. From the mounting bracket it can continue into the frame of the bus. The noise can also be transmitted over the discharge fluid (refrigerant gas and oil) from the discharge port and into the oil separator. From there it can continue out into the discharge lines of the A/C system. Capacity control affecting the displacement can also give different sound characteristics. One reason for this is that usually the optimum volume ratio is not obtained during partial load. This particular compressor is arranged with axial capacity lift valves (Sjoholm, 1986).  For the noise evaluation of compressors, numerous research methods have been applied including Experimental Statistical Energy Analysis (Ma and Bolton, 2002).  To quickly resolve this noise issue, most of the testing was done in the lab with follow-through on the bus to confirm effectiveness of design changes.     1344, Page 3  International Compressor Engineering Conference at Purdue, July 14-17, 2008 2. TESTS OF THE COMPRESSOR IN THE LABORATORY The low-floor articulated bus equipped with a screw compressor is shown in Figure 4, and the sketch of the compressor and its installed view in the bus is also displayed in Figure 5.  In the beginning, we focused on the correlation that appeared in the spectrum data between the vibration frequency and the noise frequency coming from compressor body.  Figure 6 shows very close correlation of the vibration and the noise, in particular from the mounting bracket described in Figure 5. Therefore the first approach was to modify this bracket to avoid the resonance frequency around 1 and 2 kHz.   Figure 7 shows the noise improvement from this new modified design at 2000 rpm at a partial loading condition.   The compressor noise level was reduced about 8 dBA overall due to a 9 dBA reduction at the 1.25 kHz 1/3 octave band.  Other design parameters were also tested for their effects on noise level: oil mass flow rate, oil viscosity, internal check valve design and so on.  These various alternative designs are introduced in Figure 8, and the results are displayed in relative noise levels compared to the compressor that was returned from the field due to noise issues.    3. FEEDBACK FROM TESTING IN THE BUS  Based on those different approaches shown in Figure 8, noise tests were performed in a bus at the customer’s site. The results are plotted in Figure 9.  Although we proposed a couple of new alternative designs to reduce noise level and also confirmed their validity through lab testing, the field test result from the bus operation was not satisfactory.  During the comparison between the lab testing and the field test, some different circumstances were noted.  For the lab testing, the compressor was run at 1000, 2000, and 3000 rpm, and the sound pressure was also measured at one foot away from the surface of the compressor.  For the bus operation, however, the bus engine was run from 600 through 1200 rpm, which is equivalent to compressor speeds of 885 - 1770 rpm, and the sound was measured at the rear passenger seats.  We also recognized there was a lack of correlation between noise behaviors on the electrically- driven compressor in the lab and the engine-driven compressor in the bus.  Lab tests could not account for torsional fluctuations of the engine and its effects on the discharge pulsations.  Changes that showed an average of 4 dBA improvements at any speeds in the lab were not seen in the bus.  Recognizing this, other design factors had to be evaluated, especially those that are affected by torque variations into the compressor.  4. CONCLUSIONS  To resolve the sound issue of the bus in the field, a couple of new design parameters were studied: A very close correlation between the vibration and the noise from the compressor body was found. The improvements that were confirmed in the lab testing were not seen in the bus.  There was a lack of correlation between noise behavior on the electrically driven compressor in the lab and the engine-driven compressor in the bus.  REFERENCES Sjoholm, L., 1986, Variable Volume-Ratio and Capacity Control in Twin-Screw Compressors, Proceedings of International Compressor Engineering Conference, Purdue University, IN, USA, p. 494-508. Ma, Y., Bolton, J.S., 2002, Experimental Statistical Energy Analysis Applied to a Rolling Piston-Type Rotary Compressor, Proceedings of International Compressor Engineering Conference, Purdue University, IN, USA, C17-4.     1344, Page 4  International Compressor Engineering Conference at Purdue, July 14-17, 2008 ACKNOWLEDGEMENT  The authors wish to thank the management of Ingersoll Rand Climate Control for permission to publish this paper. The authors also acknowledge the assistance to this work by Mr. Peter Lawrence and Mr. Jon DeTuncq, both of Ingersoll Rand Climate Control.                                                   Figure 1: Screw compressor for bus A/C showing  (from left to right) clutch,  gears, rotors, oil separator  and oil sump Figure 2: The discharge port of a bus A/C screw  compressor,  the red arrow indicates important area for drainage of oil Figure 3: The drive gears are shown on a screw compressor for bus A/C  1344, Page 5  International Compressor Engineering Conference at Purdue, July 14-17, 2008                                                         Figure 4: Low-floor articulated Bus Figure 5: Screw compressor installed with a bracketFigure 6: Correlation of the vibration and the noise of returned screw compressor Acceleration of returned screw compressor housing and bracket at 3000RPM385919983595338990010000.02.04.06.08.010.012.014.00150300450600750900105012001350150016501800195021002250240025502700285030003150330034503600375039004050420043504500465048004950FrequencyGal [m/sec2]Noise leve of screw compressor at 3000RPM10003196 33963596199840455055606570758085900150300450600750900105012001350150016501800195021002250240025502700285030003150330034503600375039004050420043504500465048004950frequencydBA 1344, Page 6  International Compressor Engineering Conference at Purdue, July 14-17, 2008                                                       Figure 7: Noise improvement with a modified new bracket at 2000 rpm at a partial loading in the lab testing (delta Y: 5dBA) Compressor Noise Level at 2000RPM with partial loadingbetween current and modified new bracket 6570758085909510020 Hz25 Hz31.5 Hz40Hz50Hz63 Hz80 Hz100 Hz125 Hz160 Hz200 Hz250 Hz315 Hz400 Hz500 Hz630 Hz800 Hz1 kHz1.25 kHz1.6 kHz2 kHz2.5 kHz3.15 kHz4 kHz5 kHz6.3 kHz8 kHz10 kHz12.5 kHz16kHz20 kHzSUM(LIN)SUM(A)Frequency, 1/3 octaveSound Pressure [dBA]current bracket new bracketFigure 8: Alternative design changes tested in the lab for the noise reduction (delta Y: 4dBA)  Noise design options tested at 2000 RPM with partial loading84889296100Returned compressor w/original bracketModified new bracket Modified orifice w/original bracketThick oil w/ originalbracketModified internal checkvalve Noise design optionsSound Pressure [dBA]Figure 9: Noise level of various design changes tested in the Bus operation (delta Y: 2dBA)  Noise levels of alternative design configuratioins tested in the Bus operation707274767880Returned compressor Current bracket w/ thick oil Current bracket w/ currentoilModified bracket w/ currentoilModified bracket w/ thick oilAlternative design configurationsSound Pressure [dBA]1200 RPM - full load1200 RPM - partial load 600 RPM - full load 600 RPM - partial load

Noise Reduction in Bus A/C Systems with Screw Compressors Part I: Compressor Evaluation

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Noise Reduction in Bus A/C Systems with Screw Compressors Part I: Compressor Evaluation

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