A COMPARATIVE STUDY OF ENGINE

MOUNTING SYSTEM FOR NVH

IMPOROVEMENT – PART-I

Chandra Prakash Mishra, Tata Motors Ltd, Lucknow, India, +91-9793099332

c.mishra@tatamotors.com

Vinay Kumar Singh, Tata Technologies Ltd, Lucknow, India, +91-8960009100

vinay.singh@tatatechnologies.com

Abhishek Pyasi, Tata Motors Ltd, Lucknow, India, +91-9473871242

abhishek.pyasi@tatamotors.com

Rajiv Singh, Tata Motors Ltd, Lucknow, India, +91-93898509

rajiv.singh@tatamotors.com

Rajit Ram Singh, Vindhya Institute of Technology, Indore, India, +91-9229122133

singhrajitram@gmail.com

Kapil Jain, Devi Ahilya University, Indore

Kapil_jain1411@yahoo.com

ABSTRACT

The purpose of this study is to present the

effects of several design actions on engine

mounting system for improving the NVH

characteristic of the vehicle, with the aid of CAE

and experimental investigations. Power train

mounting system is one of the most important

vibration isolators in vehicles. In order to

achieve good cabin comfort, the noise and

vibration levels have to be as low as possible.

Engine mount transmits the power train

vibrations to the body, and the chassis

vibrations excited by road to the power train.

The design of engine mounting system is an

essential part in vehicle safety helps in

improving the vehicle noise, vibration and

harshness (NVH) performances. This paper is

based on a Rear wheel driven front engine and

rear engine buses with a six-cylinder engine

placed longitudinally along the vehicle, the

uncertain characteristics of engine mounting

system are studied by interval analysis method.

Considering the design parameters including

locations, orientations and stiffness of the

mounts as interval variables, the lower bounds,

upper bounds of natural frequencies and the

mode kinetic energy distributions of a engine

mounting system are estimated by means of

interval analysis.

Keyword: Dynamic design; engine mounts;

automotive system; structure joints; vibro-acoustic

modelling; optimisation; engine vibration; engine

bounce.

I. INTRODUCTION

The vibration characteristic of the vehicle is one of

the most significant factors in ride and comfort.

All piston type engines generate vibrations due to

the firing forces and reciprocating components.

Some of these vibrations are internal to the engine

structures and compensated or balanced by

opposing forces within the engine. The other

vibration cause whole engine rigid body motions

and vibratory forces that act on the engine

mounting system. The engine mounting system

must isolate these vibrations from the vehicle or

machine structure in most cases. Depending on the

engine design the engine vibration will cause

translation or rotation about three orthogonal axes.

The vehicle engine mounting system generally

consists of engine, clutch, transmission

supplementary system & several mount rubbers

connected to the vehicle structure. The engine

mounting systems have been successfully applied

to isolate the driver & passenger from noise,

vibration and harshness (NVH) generated from

power train system. Since the engine is the largest

concentrated mass in the vehicle. The power

system is excited by engine firing force,

rotating/reciprocating unbalances (unbalance forces

increases as the square of rotational speed), road

profile, and dynamic forces/motions transmitted by

accessories and other vehicle components. The

major excitations are an internal oscillating torque

on the crankshaft; internal forces in cylinder

direction and corresponding moments; and

excitations from road and wheels. The crankshaft

oscillating torque will always occur at the firing

frequency, the other direction of engine excitation

will depends on the engine types. Firing forces

acting on the cylinder head are cancelled out by

their reaction forces on the main bearing. The

frequency of the cylinder firing impulses is

determined by the number of cylinders, engine

design and the engine speed. A 4-stroke cycle

engine fires each cylinder once every two engine

evolutions, so an inline six-cylinder engine has

three cylinders firing per engine revolution. So the

engine firing frequency is three times the engine

rotational speed. Thus, for an engine idling at 700

rpm, 10 revolutions per sec. the engine firing

frequency is three times this, 30 firing/sec, causing

a 30Hz vibrational input to the mounting system.

The effectiveness of an engine mounting system in

isolating the vehicle structure from engine vibration

depends on the relationship between the frequency

of the vibration coming from the engine and the

natural frequency of the engine mounting system.

II. LITERATURE REVIEW

Optimization of Engine mounts is an interesting &

challenging research field among automobile

researchers. In the past lot of experiments and

studies have been carried out to understand the

design of engine mounts and their effects on

vehicle NVH performance. Sachdeva [et.al] studied

effect of engine mounting strategy on vehicle

NVH. In the study couple of design strategies

involving powertrain rigid body mode coupling and

de-coupling are examined. This study compares

three such strategies that are commonly used. All

decoupled modes, coupled bounce pitch modes,

and coupled bounce-roll modes strategies are

compared by looking at the response of a vehicle

model, to wheel and powertrain inputs. In the study

it was concluded that for all three input cases

studied, the best system was one of the coupled-

mode systems. It shows, however, that different

modal coupling/decoupling characteristics will

favor different inputs and system responses. There

is not one magical modal characteristic setup that

will be ideal for all inputs and all responses. The

real case of powertrain mounting design for trucks

Paulo [etc.al] studied mounting design and

manufacturing constraint, the simulation modelling

basis, inputs required to perform the computational

simulation, the experimental work required, also

the method used to determine the centre of gravity

and rotational inertia of the powertrain and a

general mounting tuning strategy. In the study it

was concluded that the computational simulation

procedure used in this work, supported by

experimental powertrain properties determination

and all experimental work associated, is presented

as a reliable tool in a mounting system

development. The powertrain computational

simulation provides the mount stiffness

optimization to meet the basic design requirements

and NVH improvements. A computerized

optimization method of engine mounting system

Liu [etc.al] studied a method for optimization

design of an engine mounting system subjected to

some constraints. The engine center of gravity, the

mount stiffness rates, the mount locations and/or

their orientations with respect to the vehicle can be

chosen as design variables, but some of them are

given in advance or have limitations because of the

packaging constraints on the mount locations, as

well as the individual mount rate ratio limitations

imposed by manufacturability. A computer

program, called DynaMount, has been developed

that identifies the optimum design variables for the

engine mounting system, including decoupling

mode, natural frequency placement, etc. The degree

of decoupling achieved is quantified by kinetic

energy distributions calculated for each of the

modes. Several application examples are presented

to illustrate the validity of this method and the

computer program.

III. THRORITICAL BACKGROUND

A – THORY

For the simplest analysis of an engine mounting

system the structures that support the engine are

treated as a rigid object with infinite mass and

stiffness. This allows the performance of the engine

isolation system to be estimated with a simple

matrix analysis. At each mounting location

between the powertrain and vehicle support

structure is modelled as a spring/damper with

stiffness and damping properties in three direction

in space. An engine mount must satisfy two

essential but conflicting criteria. First, it should be

stiff and highly damped to control the idle shake

and engine mounting resonance over 5-30 Hz.

Also, it must be able to control, like a shock

absorber, the motion resulting from quasi-static

load conditions such as travel on bumpy roads,

abrupt vehicle acceleration or deceleration, and

braking and cornering. Type are engine Mounting,

1. Rigid (solid mounted) 2. Semi-Rigid 3. Isolation

(soft mounted) Rigid (Solid mounted): Primarily

used for large engines in stationary applications

with a sub base that is placed on a massive

foundation. Mounting allows only minimal motion

for thermal expansion and operating deflections of

the engine block. Alignment to driven devices can

be controlled very precisely. Also possible with

small, structural engines. For rigid mounting to

work effectively the subbase structure must be truly

‘rigid’ under common forces to the point of excess.

Isolation pads are used to attach the subbase to

foundation for larger engines. These isolators

effectively ‘stiffen’ the subbase and damp

vibrations. Block attachments must allow for

expansion and normal deflection under firing

forces. This is why slip joints are used on front of

very large engines. Rigid mounting of small

structural engines is very advanced the adapter

design must be coordinated between product

engineering and OEM structural engineers. Semi-

Rigid: used when only limited isolation of engine is

required or isolator subframe is used. Very

common in industrial equipment with massive

frames and for engines with 12 or more cylinders.

Semi-rigid mounting degree of isolation varies

wide. Often involves major subframes that allow

two levels of isolation. Often the only practical

choice for large engines in mobile equipment.

Mounting provisions and load levels may not lend

themselves to simple isolation mounting.

Considerations of shock loading and frame

deflection may be as important as isolation

performance. May also require isolation of specific

machine components such as cab, control boxes,

etc. Isolation (soft mounted): desirable for mobile

equipment and with smaller, high speed engines.

The engine and direct attached components are

totally isolated and allowed to move in response to

normal engine vibration excitations. Isolation

mounting, this is the most typical mounting for

mobile equipment with engines of 8 cylinders or

less. Isolation mounting creates separate systems

which are isolated from each other. The engine and

directly attached components vibrate as a unit and

all components of this system must be capable of

handling the vibration levels. This can include

exhaust, intake and cooling components.

Components that are separated by the isolators are

not subject to engine excitation frequencies and can

often be less massive and rigid. Any connections

between the two systems must allow for free

motion. Engine Vibration Motions. 1. Yaw 2. Roll

3. Pitch. An isolated engine package is free to

move in three directions and three rotations.

Fig 1.1

A four cylinder engine has significant second order

shaking forces in the vertical direction. This will

cause pure vertical motion of the engine as well as

some pitch rotation depending on CG location.

Fig 1.2

Most engines have significant motion due to firing

torque reaction to engine block at firing frequency.

Engine roll about centre of least inertia close to

crank axis. Estimated as axis through the CG of

major components.

Fig 1.3

Engine isolator location – ideal isolator location

would be as close to roll axis as possible for best

roll isolation for given mount stiffness. Actual

location is usually dictated by mounting pad

location and bending moment limitations. Long

cantilevers off mounting pads must be avoided,

causes – excess moment loads and poor isolation

the engine is subjected to various vibratory

disturbances some external to it, other internal.

Random shocks from the road, transmitted through

the suspension, shake it; so do periodic shaking

forces from the universal joints in the propeller

shaft. Any rotating unbalances in the engine. The

mounts must isolate all of them; in addition, they

must support the static weight of the engine and

restrain it from fore-and-aft movement during

acceleration and breaking. Design consideration for

engine mounting system 1. Torque Axis 2. Centre

of Percussion 3. Mount configuration 4. Decouple

the Pitch and Bounce Mode 5. Specify the values of

the dynamic rates in compression and share for

each of the mounts. 6. Make the final tuning on the

complete vehicle by variation in the rubber

specification. Torque axis – in an engine the axis of

torque application is the crankshaft and its seldom

parallel to a principle axis. The output torque has

periodic oscillations that cause the engine to

oscillate in reaction. We can minimize the

disturbance to the rest of the vehicle by allowing

the engine mass to oscillate about this natural

torque axis. Torque axis of an engine is dependent

on the magnitude of the principal axes of inertia

and on their location relative to the crankshaft. The

torque axis must pass through the centre of gravity

of the engine. Centre of percussion mounting – if

the mass is free in space and an oscillating force is

applied to the mass at its centre of gravity, the mass

will translate without rotation in the direction of the

force, but if the force is not applied at the centre of

gravity there will be rotation as well as translation

and there will be a centre of oscillation. This centre

is the centre of percussion for that particular force

location. Thus if a disturbance is applied to the

front engine mount we would like to have the rear

mount mount located at the centre of percussion so

there would be no reaction forces on it. Mount

Configuration – arrange the mounts in their planes

so that the elastic centre will fall on the torque axis.

Establish the relation between compressions and

shear rates of the mount. Decouple the Pitch and

Bounce Mode – this is achieved by making the

product of front mount vertical rate and the

distance along the elastic axis to the centre of

gravity equal to the similar product for the rear

mount. Dynamic rate of rubber – the amplitude and

frequency of a vibrating system is determine by

the mass of system, the stiffness of its elastic

elements, the damping resent and the magnitude

and frequency of the excitation. The purpose of

damping in an isolator is to reduce or dissipate

energy as rapidly as possible. Damping is also

beneficial in reducing vibration amplitudes at

resonance. Resonance occurs when the natural

frequency of the isolator coincides with the

frequency of the source vibration. The ideal isolator

would have as little damping as possible in its

isolation region and as much as possible at the

isolator’s natural frequency to reduce amplification

at resonance. Damping however can also lead to a

loss of isolation efficiency. As damping is

increased, the curve (Fig 2.1) of transmissibility is

flattened, so that in the region near to resonance,

the curve is reduced, but in the region where

isolation is required, the curve is increased. The

curves show that if there is a significant amount of

damping in an isolator, its natural frequency has to

be reduced to retain the desired degree of isolation

at the frequency ratio of concern. Damping

provides energy dissipation in a vibrating system. It

is essential to control the potential high levels of

transient vibration and shock, particularly if the

system is excited at, or near, to its resonant

frequency. To achieve efficient vibration isolation

it is necessary to use a resilient support with

sufficient elasticity so that the natural frequency fn

of the isolated machine is substantially lower that

the disturbing frequency fe of vibration. The ratio

fe/fn should be greater than 1.4 and ideally greater

than 2 to 3 in order to achieve a significant level of

vibration isolation.

Fig. 1.4 Four Point Mounting

B- MATHMETICAL MODEL

The mounting system effectiveness is commonly

measured as the Transmissibility. The inertia force

developed in a reciprocating engine or unbalanced

forces produced in any other rotating machinery

should be isolated from the foundation/Mounting

so that the adjoining structure is not set into heavy

vibrations. Transmissibility is the amount of engine

vibration force which is transmitted through the

mounting system to the vehicle structure as a

percentage. A transmissibility of 0.4 or less of

engine idle speed is necessary for a good mounting

system. In the region of attention rather than

referring to the transmissibility, we use the

isolation efficiency as a measure reduction of

vibration input usually as a percentage value

occurring for a particular disturbing frequency.

This is defined by Isolation efficiency % isolation =

[1-Tr] x 100. Tr > 1 = Increased Transmitted

vibration TR <1 = Vibration Isolation, TR = 1

No Vibration Isolation

2

22 20

1 2

T

1 2

ntrr

n n

F

F

ωω

ω ωω ω

+ ζ

= = − + ζ

Where, ζ = Damping Factor = C/Cc or Damping /

Critical damping, ω = Exciting frequency (fx2xπ),

ω n = System natural frequency, T = Transmission

Fig. 2.1 Transmissibility v/s frequency ration for

different amounts of damping

The engine mounting system effectiveness is

usually measured with the term “Vibration

Transmissibility” the vibration transmissibility is

the amount of engine vibration which is transmitted

through the mounting system to the vehicle

structure. Vibration transmissibility TR > 1 means

that the engine mounting system is actually

transmitting more vibration into the vehicle

structure than is coming from the engine. This is

possible if the natural frequency of the mounting

system is close to the frequency of the engine

vibration, resulting in the mounting system

operating at or near resonance, resulting in

magnification of the input vibration. If vibration

transmissibility TR<1 indicate that the mounting

system is actually transmitting only a fraction of

the vibration input from the engine, so is isolating

the vehicle from engine vibration. As engine speed

increases, the firing frequency and therefore

vibrational input frequency to the mounting system

also increases. This means that the engine

mounting system has the worst (highest) vibration

transmissibility at lower engine speeds. Since idle

speed is the lowest engine speed commonly used,

it’s the most critical speed for design of engine

mounting system. A vibration transmissibility of

0.4 or less at engine idle speed is necessary for

good mounting system. In the region of attenuation

rather than referring to the transmissibility, we use

the isolation efficiency as a measure reduction of

vibration input usually as a percentage value

occurring for a particular disturbing frequency.

The basic rule of thumb is that isolation levels will

generally be improved (transmissibility will be

lowered) by increased rigidity and mass in the

supporting structures and by lower stiffness in the

isolators. The mounting system performance

calculations assume that the mounting system is

attached to a rigid base rather than a flexible

vehicle frame. The stiffer the frame, the closer the

mounting system will perform to the theoretical

calculations. After an acceptance design is chosen,

vehicle testing will reveal whether the system will

meet the desired performance in-vehicle when

attached to the vehicle frame and subjected to road

as well as power train vibrational inputs.

Table – 1 Variation in Isolation Efficiency for

Different values Damping Factor & Frequency

Ratios.

Damping

Factor (ξ) Frequency Ratio R fe/fn

C/Cc 1.5 2 2.5 3 3.5 4 4.5 5

0.05 20 66 80 87 91 93 94 95

0.1 19 64 79 85 89 91 93 94

0.15 17 62 76 83 87 90 91 93

0.2 16 59 74 81 85 87 89 91

0.3 12 52 67 75 80 83 85 87

% Isolation Efficiency

% Isolation V/S Damping Factor

0

10

20

30

40

50

60

70

80

90

100

0.05 0.1 0.15 0.2 0.3

Damping factor

% I

sola

tio

n E

ffic

ien

cy

R=1.5 R=2 R=2.5 R=3

R=4 R=4.5 R=5

Graph – 1 Variation in Isolation Efficiency for

Different values Damping Factor & Frequency

Ratios.

Graph is showing variation % Isolation for

different values of damping factor (for given

frequency ratios) curves shows that nature of

variation of % Isolation for all values of damping

factor and frequency ratio is same but the point to

be noted here is that when the disturbing frequency

is nearer to the natural frequency the isolation

efficiency is very less. For higher values [fe/fn ≥ 2]

the frequency isolation efficiency doesn’t vary

much. This study shows that system should be

operated well above the ratio value of 1 to 1.5.

Mounting Reactions & Bending Moment at RFOB

& FFOB:

Fig 2.2 Bending Moment Calculation

All engine installations must be designed to limit

the vertical bending moment at rear face of the

block (RFOB) below the value listed on the engine

datasheet. Fig 2.2 & 2.3 illustrates the method for

calculating the bending moment at the rear face of

the block for a typical powertrain with transmission

and rear tail support (R3). The calculation method

can be applied for rear mounts on flywheel housing

or transmission housing or a subframe.

Fig 2.4 Inclination Angle of Front Engine Mounts.

Fig 2.5 Front Engine Mounts.

Distance L4 is the distance to the rear mount

isolator centres from the rear face of block.

Fig 2.6 Rigid body engine model

Engine Mounting Calculations –

1. Dry weight of engine (weight flywheel

and alternator, but less starter and air

compressor) - WTD

2. Wet weight of engine - WE

3. Weight of starter and air compressor - WA

4. Total weight of Engine - WTE

5. CG Dist. From FFOB – h cg

6. CG Dist. Above crank centre line – h ccl

7. Dry weight of Transmission - WDT

8. Wet weight (oil tank capacity) - WOT

9. Clutch cover assembly weight (cover

bearing + Adaptor ring + Inner plate) - WC

10. Weight of dics - WD

11. Total weight of Transmission - WTT

12. Transmission CG (X,Y,Z) – h TCG

Fig 2.5 Loading Diagram mounting of Flywheel

Housing

Where,

X1 = Dist. Between engine front mount to FFOB

X2 = Dist. between FFOB to total weight of engine

(C.G. of Engine)

X3 = Dist. Between engine front mount to total

weight of engine.

X4 = Engine block distance.

X5 = Dist. Between engine front mount to RFOB.

X6 = Dist. Between engine front mount to engine

rear mount.

X7 = Total length of span of drive line.

X8 = Dist. between RFOB to engine rear mount.

X9 = Dist. Between flywheel housing face to total

weight of transmission (CG of Transmission).

R1 = Reaction force of engine Front mount

R2 = Reaction force of engine Rear mount

Calculation of Reaction Forces:

R2 x X5 = Total weight of Engine x X3 + Total

weight of Transmission x X7

Reaction force on each rear mount = R2 / 2

R1 = (Total weight of Engine + Total weight of

Transmission) – R2

Reaction force on each front mount = R1/2

Bending Moment at RFOB = R1 x X5 - Total

weight of Engine x (X4 - X2) or Total weight of

Transmission x ( X7-X6+X8) – R2 x X8

This bending moment should not exceed

recommended bending moment of Engine at

RFOB. If not than fifth mounting is required or

change position of mounts is required to meet the

recommended bending moment of Engine.

Vibration Isolation Six Cylinder Engine:

calculation of torsional vibration excited by firing

disturbances, exciting frequency by firing

disturbance is given by,

2. .

60.

n if Hz

C=

Where,

n = Engine idle RPM

i =Number of Cylinder

C=Engine Cycle (4-Stroke or 2-Stroke)

πω = 2. .ƒ , ω = Engine forced frequency =

rad/sec

Calculation of natural frequency of Front/Rear

mounts

.n

k g

mω =

Where,

nω = Natural frequency of the front mount

k = Stiffness of the front mount, Kg/mm

g = Acceleration duet to gravity, m/s2

m= Reaction force on front mount, Kg

Frequency Ratio =

n

ωω

Transmissibility ( Tr ) graph for the engine mounts

is obtained by iteration method. Transmissibility is

calculated for different values of frequency ratios.

Frequency ratios are obtained on the basis of test

results and design guide lines.

IV. CONCLUSION

The conclusion of the present study is as follows.

1) Comparison of different methods to

analyze NVH performance of engine

mounts.

2) Study of mathematical models of engine

mounts.

3) Graphical representation of different NVH

parameters like (stiffness, damping,

isolation efficiency & vibrational

transmissibility)

4) Exhaustive study of vibrational effects for

engine mounts.

V. REFERENCES

[1] Hamid Mir, “Focused 4-Mount Concept

Evaluation”, Technical report, NVH Development

and Engineering, DaimlerChrysler Corporation,

2001.

[2] Yunhe Yu et al.”Automotive Vehicle Engine

Mounting Systems: A Survey”, Journal of Dynamic

Systems, Measurement, and Control, June 2001,

Vol., 123, pp.186-194.

[3] Thomas D. Gillespie, Fundamentals of Vehicle

Dynamics, SAE, 1992

[4] Akinon Matsuda, Yasutaka Hayashi and Junzo

Hasegawa, Vibration Analysis of a Diesel Engine

at Cranking and Idling Modes and Its Mounting

System, SAE Proceedings, 870964, 139-146, 1987

[5] Stuecklschwaiger, W. and Ronacher, A.;

“Optimization of Engine Mount Parameters by

Simulation and Statistic Techniques”, European

ADAMS Users Conference, 1994.