,./j,_-s/

EPA 550_9-82-335

DEMONSTRATION OF NOISE CONTROL

FOR THE CUMMINS NTC-350 HEAVY DUTY

TRUCK DIESEL ENGINE

_,:,. ,_,_"_L PRO_-c_

-J JUNE]982

FINAL REPORTi;j,

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Document is available to the public in limitedquantities from the National Technical InformationService, Springfield, VA 22151.

....,i

EPA 550/9-82-335

DEMONSTRATION OF NOISE CONTROL FOR TIlE

CUMMINS NTC-350 H'IS_VY DUTY TRUCK DIESEL ENGINE

JUNE 1982

!f_" FINAL REPORT

Natan E° Parsons

Richard G. DeJong

Jerome E. Manning

_ Preparedfor

office of Noise Abatement

U.S. Environmental Protection AgencyWashington, D.C. 20460

Contract No. 68-01-4737

Cambridge Collaborative, Inc.

Kendall SquareCambridge, Massachusetts 02142

ii

Disclaimer Statement

This report has been reviewed by the Office of

Noise Abatement, EPA, and approved for publication.

Approval does not signify that the contents,/-_ necessarily reflect the views and policies of the

Environmental Protection Agency, nor does mentionof trade names or cormnercial products constituteendorsement or recommendation for use.

Cambridge Collaborative, Inc., the contractor,

is responsible for the facts, the accuracy ofthe data, the analysis, and the views presentedin this report. Neither the official views or

policies of the Environmental Protection Agency,the sponsor, nor those of the Cummins Engine

Company, who consulted in the work, are expressedin this report. With this disclaimer, PERMISSIONIS GRANTED TO REPRODUCE THIS MATERIAL WITI_3HT

FURTHER CLEARANCE.

iii

Preface

The work reported herein has been performed by

Cambridge Collaborative, Inc. under Contract No.

_f"h 68-01-4737 from the U.S, Environmental ProtectionAgency Office of Noise Abatement. The authorswish to acknowledge the Massachusetts Institute of

Technology which served as a subcontractor on the

project, and the Cu_nins Engine Company. Thecooperation and assistance of these organizations

have been an essential ingredient to this project.

iv

TABLE OF CONTENTS

.Pa_e

I, INTRODUCTION ............................................ 1

II, ENGINE DESIGN AND PERFORMANCE .......................... 3

A. Baseline Moise and Vibration Characteristics ....... 5

B. Engine Vibration Characteristies ................... 13

III. DEVELOPMENT OF A NOISE GENERATION MODEL ................ 18

A. Combustion Pressure ................................ 19

B. Piston Slap ........................................ 22

C. Injector Forces ..................................... 28

D. Vibration of Non-Load Bearing Covers ................ 32

E. Noise Radiation ................................... 32

IV. NOISE REDUCTION DESIGNS ............................... 40

A. Piston Slap Noise Reduction Designs ................. 42

B. Combustion Excitation Source Reduction .............. 47

r_ C. Injection Excitation Source Reduction ............... 49_ /

V. CONCLUSIONS ........................................... 51

VI. REFERENCES ............................................ 52

v

LIST OF ILLHSTRATIONS

___i_ure 25Lag__

1. Fuel System of NTC-350 Engine ......................... 4

2, Performance Curve for NTC-350 Engine .................. 6

3. Noise Measurement Instrumentation ..................... 7

4. Reverberant Room Facility ............................. 9

5. NTC-350 Engine Mounted in Test Cell ................... I0

6. A-Weighted Sound Power Levels of NTC-350 Engine

vs. Engine Speed at Three Load Conditions ............. ii

f_ 7. Frequency Spectra of Sound Power Levels ofNTC-350 Engine at 2100 RPM, '_qo Load Conditions ....... 12

8. Frequency Spectra of Sound Power Levels of

NTC-350 Engine for Three Speeds, Full Load ............ 14

9. Vibration Measurement Locations on External

Surfaces of NTC-350 Engine ............................ 15

i0. Lower Right and Left Block Surface Vibration

Levels on NTC-350 Engine at 2100 RFM, Full Load ....... 16

ii. Oil Pen Surface Vibration Levels on NTC-350

Engine at 2100 RPM, Full Load ......................... 17

12. Frequency Spectra of Measured Cylinder Pressure

on NTC-350 Engine at 2100 RPM, Full Load .............. 20

" 13. Instrumentation for Mobility Measurements ............. 21

14. Comparison of Vibration Transmission fromCombustion with Overall Vibration on NTC-350

Engine at 2100 RPM, Full Load ......................... 23

15a. Time Response of Liner Vibration, Major ThrustSide, on NTC-350 Engine at 2100 RPM, Full Load ........ 24

b. Frequency Spectra of Liner Vibration, Major _irust

Side, on NTC-350 Engine at 2100 RPM, Full Load ........ 24

16. Drivepoint Mobility of Cylinder Liner on

NTC-350 Engine ........................................ 25

17. Estimated Force of Piston Slap on NTC-350 Engineat 2100 RPM, Full Load ................................ 26

vi

LIST OF ILLlJSTBATIONS (coat.)

18. Piston Slap Vibration Transmission from CylinderLiner to Block on NTC-350 Engine ...................... 27

19. Piston Slap Vibration Levels vs. Measured BlockVibration Levels on NTC-350 Engine at 2100 RPM,Full Load ............................................. 29

20a. Injector Rocker Velocity, Cylinder #3, on NTC-350

Engine at 2100 RPM, Full Load ........................ 30

b. Frequency Spectra of Injector Vibration Level on

MTC-350 Engine at 2100 RPM, Full Lend ................. 30

21. Drivepoint Mobility of Injector on MTC-350 Engine ..... 31

22. Vibration Transmission from Injector to Block onNTC-350 Engine ........................................ 33

23. Injector Vibration Levels to Lower Left Block vs.

Measured Vibration Levels on NTC-350 Engine at2100 RPM, Full Load ................................... 34

24. Sum of Piston Slap and Injector Vibration Levelsto Lower Left Block vs. Measured Vibration Levels

on NTC-350 Engine at 2100 RPM, Full Load .............. 35

i 25. Oil Pan Mobilities, Free, on NTC-350 Engine ........... 36

26, Oil Pan to Lower Block Vibration Ratio on NTC-350

Engine at 2100 RPM, Full Load ......................... 37

_ 27, Lower "In-Truck" Sound Power Levels vs. Total

Measured Sound Power Levels on NTC-350 Engine ......... 39

28. Contribution of Sources to Overall A-Weighted Truck

Sound Power Levels on NT0-350 Engine at 2100 RPMFull Load ............................................. 41

29. Comparison of Midway Suspended Liner with Standard

Liner Attachment on NTC-350 Engine .................... 44

30. Proposed Liner Attachment to }_ad on NTC-350 Engine... 46

31. Bolt Isolation Design for NTC-350 Engine .............. 46

32. Transfer Mobility of Piston Slap Force to BlockVibration with New Liner Attachments on NTC-350

Engine ................................................ 48

33. Injector Isolation Design on N_-350 Engine ........... 50

vii

I. iNTRODUCTION

q_lis report presents the resu].ts of an engineering study

to identify and design methods for reducing diesel engine block

vibration and radiated noise. The methods identified are such

that noise reductions could be achieved without degrading the

engine's performance, fuel economy, or emissions, qq_e goal of

the study was to demonstrate a 5 dB reduction in radiated noise

by retrofitting an existing, popula_ heavy duty truck diesel

engine, the Cu_nins NTC-350 Big Cam One engine. _f_e study fo-

cused on the noise radiated by the surfaces of the engine and did

not include exhaust, intake, or cooling fan noise since the noisef--t.

control for these sources has already been demonstrated [i].

_e work was carried out in five major phases:

i. Determining the engine noise characteristics.

2. Developing a model which relates internal forces to

external vibration or noise levels.

3. Designing noise controls with constraints provided

by the basic design and by the manufacturer of the

_ engine.

4. Developing a demonstration engine.

5. Testing and evaluating the demonstration engine.

The report is organized into sections discussing each of

these phases separately. Section II presents a discussion of the

engine's design and performance features as well as a detailed

description of the noise characteristics. First, the baseline

noise characteristics are determined which relate the total sound

radiation spectra of the engine to the speed and load of opera-

tion. Second, the magnitude of the sound intensity radiating

1

from each of the various engine surfaces is measured in order

to rank them according to their importance in generating nolse.

These measurements determine _le basic noise emission character-

istics of the engine in its standard configuration.

Section III presents the development of a noise generation

model for the engine. First, the levels of various internal

sources (such as combustion pressure, piston impacts, injectors,

bearings, gears, pumps, and air flow tbrough valves) are deter-

mined through direct or indirect measurements. Second, the

characteristics of the vibration transmission through the engine

structure, from each source location to the external surfaces,

are determined from vibration response measurements on the dis-

assembled engine. Third, the sound radiation characteristics

Of each vibrating surface are determined by relating the sound

intensity measurements to the vibration levels of the engine

measured during the operation of the engine in its standard con-

figuration. Combining _e results of these three procedures

gives a means of estimating quantitatively the contribution of

each internal source to the total radiated noise. The major

sources of noise are discussed in detail. _%ese include piston

slap, injectors, and combustion.

Section IV of _%e report discusses the development of engine

design modifications to reduce the A-weighted radiated noise in

_' a truck by 5 dB. First, the noise generation model is used to

develop analytical models which relate quantitative changes in

the engine design parameters to quantitative changes in the

engine noise. These models are used to identify values of engine

design parameters which will give the desired noise reduction.

Second, prototype designs are developed for the engine components

which will achieve the specified design parameters. These proto-

type designs are modified and improved based on the constraints

of other factors such as durability and performance.

II. ENGINE DESIGN AND PERFORMANCE

The Cummins N%_-350 Big Cam One is a heavy duty truck engine

designed to be used primarily in truck classes 6 and 7 for long

range hauling. _]e designation "Nq_-350 Big Cam One" refers to

the following characteristics of the engine:

N - New

T - _rboeharged

C - Custom-rated

350 - 350 horsepower

._ Big Cam One - The engine has a larger diameter cam

than the standard NTC-350 engine. In this version

the cam is increased from 2" to 2.5" in diameter

which results in shorter injection duration and

lower combustion temperature and yields lower

emissions and better fuel economy.

The NTC-350 Big Cam One's rated horsepower is 350 at 2100 rpm.

It is a six-cylinder, in-line, four-stroke engine which is turbo-

charged and intereooled. The NTC-350 Big Cam One utilizes a PT

fuel system (p - pressure and T - time). This fuel system incor-

._-_ porates a low pressure metering fuel pump that controls the fuel• delivery rate to the injectors located in the head at the center

of the combustion chamber. The injectors are activated through a

rocker arm push rod assembly by the camshaft. These injectors

perform the same task as unit injectors in that they receive fuel

at a low pressure and raise the pressure significantly through

the use Of a plunger activated by the camshaft to achieve proper

fuel spraying into the combustion chamber. The injectors are

different, however, from the so-called unit injectors because

the metering of fuel is done by the PT pump rather than by the

injector (see Figure 1 for a schematic of the fuel system).

The NTC-350 Big Cam One engine comprises a single block

with three separate heads, each head serving two cylinders. It

utilizes a cast aluminum oil pan and rocker boxes and stamped

3

INJECTORS

p.

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PT PUMP

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Figure i. Fuel System of NTC-350 Engine

steel valve covers and is equipped with an air compressor driven

by the same shaft as the PT fuel pump. 'lq]eengine uses a 24V

electric starter attached to the SAE #2 flywheel housing.

The specifications of the NTC-350 Big Cam One are as follows

(see Figure 2 for engine performance):

power rating 350 bhp

governed rpm 2100

peak torque 1065 ib-ft at 1400 rpm

nominal torque rise 22%

no. of cylinders 6

bore and stroke 5.5 x 6 in.

:/ _ piston displacement 855 in 3

lube system oil cap 11.5 gallons (U.S.)

net weight with standard

accessories, dry 2580 ibs.

The NTC-350 Big Cam One incorporates the following elements:

single piece aluminum pistons, cast iron liners attached to the

block at the top flange, a single piece head gasket per head,

cast iron block, cast iron heads, and a positive displacement oil

sump pump.

!'f'i A. Baseline Noise and Vibration Characteristics

The noise radiated from the NTC-350 Big Cam One engine was

measured in one-third octave bands in the frequency range from

I00 to lO,000 HZ. A 200 m 2 reverberant test cell which was cali-

brated for sound power level measurements of noise sources with

puretones according to American National Standard ANSl SI.21 and

International Standard ISO 3742 was used for the measurements.

The instrumentation required for the measurements is shown sche-

matically in Figure 3. Using the rotating microphone boom, the

sound pressure in the room was measured and averaged over time

and space. The sound power level is computed from the average

sound pressure level by comparing it to the measured sound

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Figure 3. Noise Measurement Instrumentation

pressure level of a standard noise source with a constant sound

power output (s centrifugal fan mLlde by ZLG Manufacturers,

C_icago was used for this project). _le standard noise source

was measured after each engine measurement sequence ho account

for changes in the temperature and humidity of the air.

The vibration levels of the engine structure were measured

at various points wit]] aeeelerometers attached to the engine sur-

face. The noise and vibration signals were analyzed in one-third

octave bands using a Nieolet 444 Spectrum Analyzer interfaced

with beth analog and digital outputs. Selected test measurements

were also tape recorded on a 2-channel Nagra IV-SJ Tape Recorder

for storage and more detailed analysis.

A load was applied to the engine with a Go-Power DT-2000J

Waterbrake Dynamometer attached to the clutch housing. The dyna-

mometer indicates both speed and torque during the operation of

the engine. Other engine parameters such as temperatures and

pressures were also monitored during the Zests to assure that the

engine was operating properly.

A photograph of the reverberant room facility is shown in

Figure 4, A photograph Of the engine mounted in the facility is

shown in Figure _. In order to reduce the intake and exhaust

noise, the intake and exhaust pipes were wrapped with fiberglass

and lead during the noise test. The exhaust was cooled with a

,, water spray, The dynamometer and stand were also wrapped to

reduce their radiated noise. The engine was run without a

e0oling fan to eliminate that noise source. The total sound

power radiated by the engine was measured as a function of speed

and lead.

Figure 6 shows a summary of the noise _nissions of the die-

sel engine. The A-weighted sound power level is plotted versus

speed for 50% and 100% of the full load at each speed. The noise

emissions show a dependence on both speed and load. The sound

power level is approximately proportional te 25 log g where N is

the engine speed in RPM. A more detailed representation of the

noise emissions is given in the next two figures. Figure 7 shows

the frequency spectra of the sound power levels at 2100 RPM for

Figure 4. Reverberant Room Facility

Figure 5. NTC-350 Engine Mounted in Test Cell

10

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Figure 6. A-weighted Sound Power Levels of NTC-350vs. Engine Speed at Three Load Conditions

ii

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Figure 7. Frequency Spectra of Sound Power Levels ofNTC-350 Engine at 2100 RPM, Two Load Conditions

12

the two load conditions. At higher frequencies the increase in

sound level with load is small but fairly uniform with frequency.

Figure 8 shows the frequency spectra of _e sound power levels at

full load for three speeds. Except at low frequency the increase

in sound level with speed is fairly smooth with frequency.

B. Engine Vibration Characteristics

After the conclusion of overall engine noise measurements,

the Nq_-350 Big Cam One was instrumented with accelerometer moun-

ting pads at several points on the external engine surfaces. The

locations of the measurement points can be seen in Figure 9.

The engine was run at 2100 RPM, 100% load and an accelerometer

was moved from location to location. Accelerometer output for

each point was analyzed with an FFT analyzer, averaged over time,

and then plotted using an X-Y plotter.

After all the data were collected they were divided into

the following six areas: front, right side, left side, oil pan,

valve covers, and the after cooler. The data were then used to

obtain the average surface velocity for each surface. Greater

emphasis was placed on the surface velocity calculations for the

lower right block, lower left block and oil pan surfaces since

f_ the major objective of the program was to evaluate the noise and,>

measured vibration patterns of these surfaces within a truck

engine compartment. Figure I0 represents the lower right and

left block surface velocities and Figure ii depicts the oil pan

surface velocity.

It can be seen in Figure i0 that the lower ].eft block has

greater vibrational energy levels than the lower right block in

the frequency range from i00 Hz to 2 kHz with a pronounced peak

around 2 kHz. Since the left side of the engine block is the

major thrust side with respect to piston motion, piston slap is

expected to produce greater vibration on the left block side.

In Figure ii it can be seen that the cast aluminum pan exhibits

fairly large vibration levels in the frequency range from 800 Hz

to 2.5 kHz.

13

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Figure 8. Frequency Spectra of Sound Power Levels ofNTC-350 Engine for Three Speeds, Full Load

14

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Figure 9. Vibration Measurement Locations on ExternalSurfaces of NTC-350 Engine

15

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Figure ll. Oil Pan Surface Vibration Levels onNTC-350 Engine at 2100 RPM, Full Load

17

III. DEVELOPMP:NT OF A NOISI_ Gf_NI':RAT]IO_ MOD[_[,

T11e previous section discussed identification of the

surfaces of the engine that contribute most to the overall noise

radiation. At this point it wouh_ be possible to design covers

and shields to reduce the radiation from these surfaces. It

would also be possible to design a damping treatment for these

surfaces in an attempt to reduce the vibration levels and sub-

sequent radiated noise. However, the addition of damping is

not generally effective because of the relatively high value

of damping already present in a built-up engine structure.

The addition of covers and shields can be effective in reducing

noise, but is not desirable because of the potential damage to

these items in a service environment, the interference with

maintenance, and the added weight.

To proceed with a practical demonstration of engine noise

control we have focused on changes to the internal structure of

the engine which reduce the vibration generated by internal

sources, such as combustion or piston slap, or reduce the vibra-

tion transmitted to the radiating surfaces of the engine. To

accomplish this objective we have continued the modeling work by

developing methods to identify the sources of vibration in the

-. engine and the different paths by which that vibration is trans-

_.-_ mitred to the engine surfaces.

q_e noise generating process in engines can be divided into

four steps:

i. dynamic pressures and forces are produced within the

engine by internal sources such as combustion, piston

slap, fuel injection, gear mesh, etc.;

2. the pressures and forces act on the internal engine

structure and cause a local vibration;

3. the engine structure transmits the vibration to external

surfaces of the engine;

18

4 . the vibrating e×terua] sl*rf_._ces of the engine pro(]uce

radiated noise.

Based on previous test resu].ts 6rein Vee and in-line 6-71

engines ns well as the DDA 6V-92TTA ungine [2], three internal

sources were chosen for study: combustion, piston slap, and

injectors. For each source the level of the excitation and the

vibration transmission to the engine surfaces have been measured.

The results ]lave been combined into a noise generation model

which predicts the noise radiated by ti_ese engine sources and can

be compared to the measured noise levels of the engine. The

following paragraphs outline the procedures for studying each of

these three sources and their associated vibration transmission

paths. A more detailed discussion of the procedure is presented

elsewhere [3].

A. Combustion Pressure

The source level for combustion noise was measured directly

during operation of the engine with a pressure transducer in the

combustion chamber of one of the cylinders. The transducer was

water-cooled and flush-mounted in the cylinder head. Figure 12

_, shows the measured frequency spectra of the pressure signal.

' _ This cylinder pressure exerts a force both on the piston crown

and on the cylinder head surface, indicating that there are two

paths of vibration transmission. The first path is through the

piston, connecting rod, crankshaft, main bearings, and into the

block. The second path is through the head and into the block.

The vibration transmission characteristics of the engine

structure are determined from measurements taken in a non-running

engine of the transfer mobilities of the various components along

each vibration transmission path. The transfer mobility is a

transfer function which measures the ratio of the vibration

velocity at one point on the structure to the force exerted at

another point. The mobility is measured with the instrumentation

shown in Figure 13.

19

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Fiqure 12. Frequencv Spectra of Measured CylinderPressure on NTC-350 Engine at 2100 RPM,Full Load

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Figure 13. Instrumentation for Mobility Measurements

21

_le two transfer mobilities for t:he paths by which coln-

bustion noise is transmitted, i.e. the head-block path and the

piston-connecting rod path, are measured mobilitiee of one-third

octave band averages of mobilities measured between eight points

on the block and three cylinders. A prediction of the block

vibration due to combustion when the engine is running is

obtained by multiplying _:e total mobilities for each path by

the combustion force (pressure times piston area) and adding the

contributions from each path. A comparison of this prediction

Of block vibration due to combustion with the measured vibration

during Operation of the engine is shown in Figure 14. The com-

parison shows that except for the frequency band at 3200 Hz,

combustion is not a major source of block vibration and _herefore

is not a major source of noise for this engine,

B. Piston Slap

The source level of piston slap cannot be measured directly

by any known methods. Therefore, an indirect measurement pro-

cedure has been used to determine the magnitude Of the force

generated. This was done by measuring the local vibration

response of the cylinder liner at a point on the liner where

piston slap is believed to occur. Figure 15 shows a typical time

_'_ _ response of the liner vibration during operation of the engine

and the corresponding velocity spectrum. A second experiment was

conducted to measure the drive-point mobility at the same point

On the liner where _he vibration response was measured. Since

the drive-point mobility is the ratio of the velocity to the

applied force at the point where the force is applied, the force

generated by piston slap can be determined by dividing the mea-

sured velocity level by the drive-point mobility. Results are

shown in Figures 16 and 17.

The vibration transmission path from piston slap to the

engine surfaces is primarily through the liner and block side

walls for the NTC-350. The transfer mobility between the

cylinder liner and the block side walls is shown in Figure 18.

22

Figure 14. Comparison of Vibration Transmission fromCombustion with Overall Vibration onNTC-350 Engine at 2100 RPM, Full Load

23

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Figure 15a. Time Response of Liner Vibration, Major ThrustSide, on NTC-350 Engine at 2100 RPM, Full Load

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Figure 15b. Frequency Spectra of L£ner Vibration, Major ThrustSide, on NTC-350 Nngine at 2100 RPM, Full Load

24

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FREOUENC¥ (Hz)

_igure 16. Drivepoint Mobility of Cylinder Liner onNTC-350 Engine

25

8Z

z

_ 2

_ 2

_ 40

0 = -

0 30: i

20

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10 _125 250 500 IK 2K IlK 8K 16K

:REQUENCY (Hz)

Figure 17. Estimated Force of Piston Slap onNTC-350 Engine at 2100 RPM, Full Load

26

................--_ _ _,_._._ __ _ ...............- ........................................ _.,

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125 250 500 1K 2K LIK 8K 16K

FREQUENCY(Hz)

Figure 18. Piston Slap Vibration Transmission from

Cylinder Liner to Block on NTC-350 Engine

27

Combining the force exc[tatior* ]eve] of the piston slap with

the ineasu_ed trassf<_r mobility gives a predictiorl of the engine

block vibration clue to piston slap. Figur_ 19 compares this

pre4_iction with the measured vibration levels for the left block

side, which is the major thrust side and the major radiating

surface. This comparison shows that piston slap is a major

source of block vibration, and therefore radiated noise, in the

frequency range from 500 to 5000 _ with a peak at 2000 Hz,

C. In_ector Forces

'_le NTC-350 8ig Cam One is equipped wit)] PT injectors, each

/'_h of which injects fuel into a cylinder by means of a cam-driven

plunger. A direct measurement of the dynamic force generated

during injection is difficult. Therefore, an indirect measure-

laent of the injector source level was obtained in a procedure

similar to that don_ for piston slop. The velocity of the

injector rocker was measured during the operation of the engine

and is shown in Figure 20. The designed velocity level of the

injector caln follower is also shown by the dashed line. It can

be seen that the plunger exhibits a large vibration near the end

of injection with a largo 1600 _ component ia the spectrum and

a sharp drop beyond that point in the frequency spectra.

f .._' The force exerted on the injector plunger was obtained by

dividing the velocity spectrum of the plunger by the measured

mobility of the plunger in its fully assembled condition (see

Figure 21).

Forces generated by the injectors are transmitted to the

engine block by two paths; the first through the injector rocker

box to the head and the engine block, and the second through the

cam push-rode, camshaft, hearings, e_n gears to the block. The

vibration transmission through each path was determined by

physically disconnecting the injector plunger from the c_n d_ive

train and exciting the structure with a shaker attached to one

or the other of the disconnected members. _e resulting measured

28

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, i!!!9O

URE"" 80

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so

125 250 500 IK 2K /IK 8K igK

FREQUENCY (llz)

Figure 19. Piston Slap Vibration Levels vs. Measured

Block Vibration Levels on NTC-350 Engine

at 2100 RPM, Full Load

29

INJECTOR ACCELERATIONMEASURED

/

CAM PROFILE

_'-_ i__--, i I210 3JO 4=0 510 60 70 80 CRANK DEG.

Figure 2Oa. Injector Rocker Velocity, Cylinder #3, onNTC-350 Engine at 2100 RPM, Full Load

T-4O

-o 30

.J

Z

o_ loI-<

0 !:0<

-11

125 250 500 IK 2K tfK 8K 16K

FREOUENCY (Hz)

Figure 20b. Frequency Spectra of Injector Vibration Level onNTC-350 Engine at 2100 RPM, Full Load

3O

-- ! '+Z

0 -50:LUr_

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125 250 500 1K 2K qK 8+< 16K

FREOUENCY(Hz)

Figure 21, Drivepoint Mobility of Injector on NTC-350 Engine

31

,x++_.++ ..,..+_ ++..... _++ +. ++_+. :;.._+....,_,.+ ........................ + .................................

transfer mobilities to the block are shown it] Figure 22. The

path through the camshaft dominates the vibration transmission

above 2000 {_.

Combining the force excitation level of the injector and the

measured transfer functions to the block gives a prediction of

the engine block vibration due to _]e injectors. Figure 23

compares this prediction with the measured block vibration levels

during operation of the engine. The comparison shows that He

injectors are a major source of vibration at L600 Hz.

D. Vibration Transmission of Non-Load Bearing Covers

By summing _e contributions to block vibration from the

major sources, piston slap and injectors (see Figure 24), a

nearly complete model of the engine vibration can be obtained

for this specific engine. In order to complete the model it is

necessary to relate the vibration of non-load bearing covers,

su_1 as the oil pan, valve covers and front cover to the block

vibration since these covers can be major radiating surfaces.

Following the mobility approach the vibration transmission from

the block to covers has been determined by combining measured

mobilities on a non-running engine with measured block vibration

levels. Examples of the mobilities for the oil pan are shown in

'_.+ Figure 25. A prediction of the ratio of velocities of the oil

pan and the block is shown in Figure 26 along with a ratio deter-

mined from measured data. The comparison shows that the oil pan

has higher vibration levels over most of the frequency range Of

interest.

E. Noise Radiation

To complete the modeling, the noise radiated by each engine

surface must be related to the predicted velocity. The approach

chosen was determining a radiation efficiency by simultaneously

measuring the surface vibration and the sound power radiated

using the acoustic intensity technique. The measured values of

32

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Figure 23. Injector Vibration Levels to Lower Left Block

vs. Measured Vibration Levels on NTC-350 Engineat 2100 RPM, Full Load

34

E_D

100,;0

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Figure 24. Sum of Piston Slap and Injector VibrationLevels to Lower Left Block vs. Measured

Vibration Levels on NTC-350 Engine at2100 RPM, Full Load

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Figure 25. Oil Pan Mobilities, Free, on NTC-350 Engine

36

]'

radiation efficiency were combined with the prediction of block

and cover vibration to generate a prediction of the noise

radiated by the engino. Results are shown in Figure 27 and are

compared to the total measured sound power levels. There is good

agreement between prediction and measurement except in the 300

to 500 Hz range.

38

120 .: !

°

-=_ii0: ::

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t-tA

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o o MEASURED

X ._< MODEL

Figure 27. Lower "In-Truck" Sound Power Levels vs.Total Measured Sound Power Levels on

NTC-350 Engine

39

............... ...,,., .

IV. NOISI_ RI_DUCTION DESIGNS

_%e objective of the work presented is this report was to

demonstrate s meth_I of achieving a 5 dB reduction of in-truck

noise. Toward this end we have developed new and _nproved

modeling procedures that allow us to predict the major sources of

high frequency engine vibration and to understand the paths by

which that vibration is transmitted to the surfaces of the engine

and radiated as noise.

Based on the results _)tained in the modeling phase, it is

apparent that the major source of high frequency vibration for

the NTC-350 Big Cam One is piston slap, with lesser contributions

If-_ to overall A-weighted truck sound power levels from injectors and

combustion. As can be seen in Figure 28, piston slap dominates

the spectrum from 500 _ to 8 k l_ with equal contributions by

injectors at the 1600 Hz bandwidth.

The combustion source is important below 500 I_ and at

frequency bands of 3.15 to 4 kHz. The limited contribution of

the combustion process to overall A-weighted sound power levels

is common to turbo-charged engines and the peak in the frequency

spectrum of the combustion model can be correlated to the cavity

resonances of the combustion chamber.

\ Despite the overall domination of piston slap as the majorI,

noise contributor, it was felt that to achieve an overall 5 dB

reduction in truck noise it would be necessary to treat all

three sources, piston slap, injection, and combustion, since the

expected reduction in transmission of piston slap excitation

would result in a 4 dBA overall noise reduction. Our approach

to reducing the noise from these sources was to modify the engine

structure to reduce the vibration transmission from each source

to the radiating surfaces. Other methods for achieving the

desired noise reduction included: (a) modifying the sources to

reduce the impulsive forces generated during injection and piston

slap, (b) using engine covers and enclosures to reduce the noise

radiation from the vibrating engine surfaces, and/or (c) using

damping and vibration isolation treatments to reduce the

40

120- b _ --il:C -_ -L-

I10 :

.. /0 -- qZ ,-

70 _: :d _ -

50 _ , T125 250 500 IK 2K qK 8K 16K

FREQUENCY(Hz)

.... _ INJECTOR MODELc,-_-_ MEASUREDD D COMBUSTION MODEL_----_ PISTON SLAP MODEL

Figure 28. Contribution of Sources to Overall A-WeightedTruck Sound Power Levels on NTC-350 Engine at2100 RPM, Full Load

41

vibration and noise radiation from non-load bearing covers end

shields. These other techniques are valid noise control pro-

cedures and have received e great deal• of attention by others

working it] the area of engine noise reduction. In our work we

have not eliminated these techniques from consideration, but have

not spent any time developing them because of the work already

done by others and because of our belief that modification of the

engine structure is the most cost-effective means to achieve our

5 dB noise reduction goal.

A. Piston Slap Noise Reduction Designs .

In order to reduce the contribution of piston slap excita-

tion to overall noise generation, the three primary paths of

vibrational energy transmission from the excitation point to the

radiating surface were examined. It became apparent that the

major transmission path is from the top liner attachment point to

the upper block and from there to the lower radiating surfaces.

In seeking to alter this path two other possible arrangements

for attaching the liner to the engine structure were considered.

The first alternate approach involved suspending the liner

midway in the block right under the top water jacket. The second

_ approach called for attaching the liners to the head through some

fastening arrangement. To use either approaoh we found that it

would be necessary to increase the wall thickness of tbe liner to

achieve a higher local impedance resulting in a decrease of power

flow from the piston to the liner during piston impact. A more

detailed discussion of both approaches is given in the following

paragraphs.

i. Midway Suspended Liner

The primary reason for suspending the liner midway on the

block is the relative ease of implelaenting this design change.

Since the engine block is an in-line configuration, however, and

the lower attachment point is not necessarily stiffer than the

42

original attachment location (as is the ease for an engine block

of vee oo;]struction [1]) certain doubts ware raised based oe ou_"

fun(]amenta] design principles. We believe that to achieve a

major reduction in the contribution of an egeitatioll mechanism to

overall noise by making a structural change in the transmission

path, it is necessary to create a substantial impet]ance mismatch

along a particular path. The mismatch is needed to achieve the

reflection of vibrational energy back toward the excitation

mechanism and the hi,Jh damping values l_ossibly associated with

this occurrence.

In the case of the transmission path for the midway

suspended liner, the increase of the liner stiffness resulted io

I_--_ the need for a stiffer attachment point to the block so that an

impedance mismatch could be achieved. Achieving a softer attach-

ment point for the purpose Of an impedance mismatch presented the

problem of selecting an isolation material that would have the

proper stiffness and the proper load capabilities to sustain the

relatively high clamping loads applied to the liner top in order

to achieve combustion seal. These clamping loads a_e in the

magnitude of 40,000 to 60,000 PSI.

In determining the appropriateness of the midway attachment

of the liner to the block, several measurements of the drivepoint

impedance and the transfer impedance were conducted at the pro-

posed liner attachment point and from that location to six lower

block locations. The results shown in Figure 29 support our

argument that the transfer of vibrational er_ergy through this

attachment point is not reduced and in certain cases actually

increases. This design approach was rejected, therefore, in

favor of the attaching the liner to the !lead.

2. Liner Attachment to the Head

When the design having the liner attached to the !lead was

originally conceived, it appeared to }lave great potential for

reducing the contributions of piston slap and combustion to

overall noise reduction. Z_wever, it required major changes in

43

o ........!i_e 90li................. F

ii :"-_' _ -_ lo-

--I "-

o -12Gol

-180

-140@ 125 250 500 IK 2K 4K 8K 16K

FREQUENCY (Hz)

i_---.oMIDWAY LINER CON.

'P-"aSTANDARD LINER CON.

Fi@ure 29, Comparison of Midway Suspended Liner with StandardLiner Attachment on NTC-350 Engine

44

regard to engine maintenance and assembly procedures. The basic

design involves the attachment of the top portion of the liser to

a machined groove in the fire deck of the head (see Figure 30).

The fastener of the liner to the heed for combustion sealing and

alignment was designed to be a threaded connection using a low

taper 1O s thread pattern which enables a better load distribution

characteristic _lan standard 45 ° threads. The primary reason for

having liners attached to the head which can be separated for

maintenance as Opposed to casting tbe liners and head as an inte-

gral part, is that most heavy duty engines are designed with the

intentioa of replacing liners when they are worn down rather than

over-sizing the bores.

Another feature of this design approach is isolating _le

head from the block by using a relatively compliant material

between the fire deck of the engine head and the top deck of the

block as well as isolating the head bolts from the head (see

Figure 31 for bolt isolation design).

The attachment of the liner to the head eliminates the need

for applying a high preload on the head gasket by the head bolts

in order to achieve a compression seal. This allows installation

of a softer gasket material in between the head and the block so

that an effective isolation can be achieved. The isolation,

/._, coupled with the lower power flow into the liner resulting from

- the higher local impedance, should yield a substantial reduction

of piston slap contribution to overall noise.

The primary change _oncerning engine assembly caused by this

design approach is the necessity for installing _]e pistons in

the liner prior to installing the liner head assembly to the

engine block. In order to facilitate production of a demonstra-

tion engine within cost and time constraints, we found that the

following compromises had to be made:

a. the production of the prototype liners from

retrofitted standard liners ra_her than from new

liner castings, preventing us fr_, having sub-

stantially thicker liners as originally requested.

45

-- HEAD

fq

LINER

,"A

Figure 30. Proposed Liner Attachment to Head on NTC-350 Engine

HEAD BOLT

STEEL COLLAR

I$OLATtON MATERtAL _

HEAD HEAD I

. /' II _272777T., z///////////

/ISOLATION MA Ti_. RIAL

Figure 31. Bolt _solation Design for NTC-350 Engine

46

b. the attachment of the liners to the bead by

soldering rather than by threading, which created

substantial machining problems.

c. a small shift in the standard compression ratio by

about 2% as a result of using standard retrofitted

liners,

A mock-up assembly of the liner-head attachment was con-

structed to evaluate its dynamic effectiveness in a non-running

engine. Transfer path analysis was carried out to four lower

block points from piston slap excitation location and the results

are shown in Pigurs 32. It can be seen that in the frequency

range from 800 Hz to 5 kHz there is an average of 8.5 dB reduc-

tion in the vibrational energy transfer compared to a standard

liner configuration. An additional 3 dBA reduction could be

achieved by increasing the liner thickness as specified by our

original design.

B. Combustion Excitation Source Reduction

Based on the conclusions derived from the modeling effort,

it is apparent that treatment of piston slap e×eitation alone

6_ yields on the order of a 4 dBA noise reduction. If this reduc-

tion were not sufficient, it would be necessary to treat the

contribution of combustion excitation to overall noise. The

primary contribution of combustion excitation and its associated

transfer paths to the noise spectra lies in the frequency range

from 3,15 to 4 kHz. The two major transmission paths are the

piston-connecting rod-crankshaft-lower block and the head-upper

block-lower block. Since the path through the head can be

treated by noise reduction designs for piston slap, the focus

is shifted on treating the path through the piston.

To treat the transmission path through the piston, a design

was developed involving the application of a set of resilient

bearings. The resilient bearings are installed at the main

47

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bearings of the craekshaft and result is the reflection of vibra-

tional eae_'gy back toward the source. The high damping values

attributed to the piston motion also create a substantial bene-

fit. The treatment using resilient bearings could potentially

yield a reduction of vibrational energy transfer excited by

piston slap. This has already been demonstrated on other engines

where in particular frequency regions the transmission of piston

slap excitation is dominated by the path through the piston

rather than by through the liner.

C. Ie_ection Excitation Source Reduction

"-_' A final step in reaching a 5 dB Or more truck noise reduc-

tion involves reducing the excitation generated by the injection

process at the 2 kHz bandwidth. Since the main transmission

path of this excitation is through the rocker box, a design was

created to isolate the rocker arm shaft from the rocker box using

a frequency tuned resilient mounting bearing. This alleviates

the problem without affecting the injection performance require-

ments (see Figure 33 for the modification schematic).

49

ROCKER ARMRESILIENT BEARING

%_ ROCKER BOX

fINJECIOR t __.jPUSH ROD

ROCKER SHAFT

7///////////////////

THEAD

Figure 33. Injector Isolation Design on NTC-350 Engine

V. CONCLUSIONS

'_lis report has described a demonstration of the use of

transfer path analysis in the investigation of major noise

contributors and their transmission paths within a heavy duty

diesel engine, the Nq_-350 Big Cam One. T_e report also

describes suggested methods for reducing the contribution from

the various sources to overall noise emission which involve

alterations to the transmission paths of the sources.

'I¢,e following briefly summarizes our conclusions from

this study:

I. Transfer path analysis is a valuable tool for

gaining an understanding of the noise emission

mechanisms is diesel engines.

2. In the case of the NTC-350 Big Cam One diesel engine

piston slap is the major noise contributor and by

adequate treatment applied to this source a 4 dBA

overall noise reduction can be achieved.

3. The proposed method for reducing the piston slap noise

contribution by attaching the liners to the ]lead andisolating the head from the block appears to be a viable

technique having significant noise reduction potential

even though it has not yet been tested in a running

engine and is considered to be a somewhat unorthodox

practice in common engine design.

51

VII. RE_ EI{I_NCES

[iJ "A Baseline Study of the I_era_nste_s Affecting Diesel Engine

Intake an_ Exhaust Silencer Design, '_ Final Report to DOT,

prepared by the Donaldson Co., Inc., Apri_ 1973.

[2] DeJong, R.G._ N.E. Parsons, and J.E. Manning, "Demonstration

of Noise Control for the DDA 6V-92TTA Heavy Duty Truck

Diesel Engine," EPA 550/9-82-336, May 1982.

[3] DeJong, R.G. and J.E. Manning_ "Modeling of Vibration

Transmission in Engines to Achieve Noise Reduction, " SAE

Paper No. 790360, 1979.

52

_'l'l_d,_, q_=z,!/_l_;_l_it, _t r.F I/_L,zll, r_,, /_,./q,z,__Jrr,/,/=t_z_;

T*--_ _T_O 5 _N0_/D.82-3 _ 5 _"_ i"---'-_ _J--_'_t'_'_';_ _,............

Demonstration of Noise Control for the June 1982

Cummins NTC-350 Heavy Duty Truck Diesel Engine 'I"E'W{JHMINGOfI(_ANIZAJ_O'_CODE7,AUTt$O(4(,_) tlf'_flf(_f1&IdN(;(3_$GAt_IZAflON II_=pO_TN(

Natan E. Parsons, Richard G. DeJong, and-_n

Cambridge Collaborative, Inc.

P.O. Box 74, Kendall Square _C_C_q_A',_._CAN_Cambridge,MA 02142 68-01-4737

I2,_PONSO_IMa /,_EN{y NAM_ AN(} ADDRE_ 13 7 ¥PE OI _tEpOn T AND PERIOD COV[RE_

Office of Noise Abatement _]/]a_]

.,... U.S. Environmental Protection Agency 14_PUN_OHINGAGEN_Y(_OrH_Washington,D.C. 20460 EPA-ONA

_, SUPPLEMENTABY NOTES

_S+A_$TRACT

This report presents the results of an engineering study to designand demonstrate a method of reducing piston-slap noise for an in-line,heavy duty truck diesel engine. The Cummins NTC-350 engine was selectedfor the demonstration. The engine heads and cylinder liners were

structurally modified to reduce the vibration transmission to the engineblock and the radiated noise. The modifications were designed to achievea 5 dBA noise reduction without degrading engine performance, fueleconomy, or emissions.

DE._CFtI_TOR$ I_.I[3LN TIF IE_aS'OP[ N ENOF _ TE:_M_ C, CGSAT_ I iCh$1L;r_'up

NoiseVibration

EngineTruckDiesel

Release to Public .__U._ncIa_i_ied...........___59_,,=_.....Unclassified