Ride Index

7/31/2019 Ride Index

1/15

Experimental and Analytical Ride Comfort Evaluation of a Railway Coach

K. V. Gangadharan

Mechanical Engineering Department, National Institute of Technology Karnataka, Surathkal,Srinivasnagar 575025, India. Email: [email protected] or [email protected]

C.Sujatha*

Machine Dynamics Lab., Applied Mechanics Department, Indian Institute of TechnologyMadras, Chennai, 600 036, India. Email: [email protected] or [email protected]

V. Ramamurti

Machine Dynamics Lab., Applied Mechanics Department, Indian Institute of TechnologyMadras, Chennai, 600 036, India. Email: [email protected]

(* Corresponding author)

ABSTRACT

The dynamic performance of a railroad vehicle as related to safety is evaluated in terms of specific performanceindices like ride quality. The ride quality is interpreted as the capability of the railroad vehicle suspension tomaintain the motion within the range of human comfort and or within the range necessary to ensure that there isno damage to the cargo it carries. The recent trend in dynamic analysis is to make use of finite element (FE)models, which are closer to the real systems than conventional rigid body models. Experimental and analyticalevaluation of Sperlings ride index and ISO 2631 ride comfort analysis of a suburban railroad vehicle using anelaborate finite element model is presented in this paper. The track power spectral densities (PSDs) were used asinputs to the system and the response predictions were carried out using random vibration theory. Measurementswere also made at different parts on the coach and bogie of an electrical multiple unit trailer (AC/EMU/T) runningon a broad gauge (1676 mm) suburban track to obtain the dynamic response under normal operating conditions.

A comparison of predicted ride index and measured ride index has been presented. A through parametric studyhas brought out possible design modifications for better ride index and ride comfort.

KEY WORDS

Railroad vehicle, ride index, ride comfort, dynamic response, finite element modeling, track irregularities, powerspectral density

NOTATION

a peak acceleration Sp(f) force power spectral density (if) receptanceB acceleration weighting factor Sx(f) displacement power spectral density rollf frequency (Hz) Wz Sperlings ride index pitch[K] stiffness matrix {x} displacement vector [M] mass matrix z bounce


2/15

1 INTRODUCTION

The dynamic performance of a railroad vehicle as related to safety is evaluated in terms of specific performanceindices. The quantitative measure of ride quality is one of such performance indices. Ride quality is interpreted asthe capability of the railroad vehicle suspension to maintain the motion within the range of human comfort and orwithin the range necessary to ensure that there is no damage to the cargo it carries. The ride quality of a vehicledepends on displacement, acceleration, rate of change of acceleration and other factors like noise, dust, humidityand temperature. EERI report B 153/RP21 [1] gives a detailed comparison of different methods used forevaluation of ride index. There are two approaches generally used to evaluate the ride quality of a vehicle: theride index method and the fatigue time method. The Sperlings ride index (Wz) is the ride index used by IndianRailways to evaluate ride quality and ride comfort. Ride comfort implies that the vehicle is being assessedaccording to the effect of the mechanical vibrations on the human body, whereas ride quality implies that thevehicle itself is being judged. According to International Standards Organisation (ISO), the influence of vibrationon the human body is expressed by the fatigue time T. ISO 2631 (1985) [2] defines three levels of fatigue time T:

fatigue decreased proficiency, exposure limits and reduced comfort boundary.

Safety, ride comfort and economy are the major factors behind any successful transportation business.Elmaraghy [3] presented a computer program developed in FORTRAN, to find the ride quality of a rail vehicleconsidering analytical track PSDs as input. Lyon [4] described the experimental methods used for dynamicmeasurement in rail vehicle research and emphasised the importance of response measurement along the lengthof the coach for accurate ride quality calculation. Sujatha et al. [5] conducted experimental and analytical studieson a typical Indian bus to evaluate whole body vibration exposure of the occupants and the driver. Finite element

(FE) models were used for the analytical studies. Suzuki [6] presented a review of various ride comfort evaluationmethods and relevant research carried out in Japan. Hariharan [7] estimated the ride index of a high power diesel

multiple unit rail coach of Indian Railways. He used the multibody simulation package, ADAMS/Rail for ride indexestimation.

This paper presents the ride index evaluation and comparison of reduced comfort boundary curves withanalytically and experimentally determined dynamic response of the Electrical Multiple Unit/Trailer (EMU/T) coachand also presents a parametric study to evaluate possible design changes to improve ride comfort of a suburbanrailroad vehicle.

2 MATHEMATICAL MODELS OF VEHICLE /TRACK SYSTEM

Sub structure

Car body

Bogie

Primary suspension

Wheel and axle

Track

Figure 1 A railway vehicle component description

Secondary suspension


3/15

An EMU/T coach running on broad gauge (1676 mm) suburban Indian track has been modelled. The vehicleconsists of a car body, two bogies per car, four wheels and two axles for each bogie (Figure 1). The vehicle hastwo tier suspension bogies. The car body is connected to the bogies through the secondary suspension, whichhas four sets of coil springs and two dampers in the vertical direction. Wheels and axles are connected to thebogies through the primary suspension system. Four coil springs, two on either side of the wheel constitute theprimary suspension. These springs are vertically guided using dashpots fixed in the centre of the coil springs. Twoapproaches used for modelling the track vehicle system are those using the finite element model and the rigidbody model.

2.1 Finite element models

Three different FE models were generated with increasing levels of sophistication. The simplest is the modelwhere the underframe alone was modelled by lumping the superstructure mass and inertia. The next level wasachieved by adding the superstructure framework on to the underframe model, neglecting the sheet metalcovering the framework. The last and most elaborate FE model is the one where the superstructure frameworkand the sheet metal panels were also modelled. Details of the model developed and relevant dynamic parametershave been presented by the one of the authors in an earlier work [8]. A brief explanation of the FE model usedhas been presented in this paper.

In the FE model, the rail has been treated as a beam on elastic foundations (BEF), ie. the track has beenconsidered as an Euler Bernoulli beam resting on Winklers foundation with vertical translation and rotation about

lateral axis being present. The primary and secondary suspensions were modelled as springs with verticaldegrees of freedom (dof) in the case of the vertical model and as springs with vertical and lateral dof in the caseof the combined vertical and lateral model. The axle, bogie frame and underframe with sole bar, cross bearer andsuperstructure framework were modelled with 3D beam elements with all the six dof present. All sheet metalpanels were modelled using triangular plate elements. The model has 576 nodes, 866 beam elements and 768triangular plate elements. It has 3080 active dof and a bandwidth of 216. Mass and stiffness matrices were

assembled in banded form to save core and labour. The problem size is 3080 216 for the vertical and lateralcombined model (Figure 2)

Z

X

Z

Y

Y

X

Figure 2 FE model of vehicle / track system UFBP model.


4/15

2.2 Rigid body model

A rigid body model for vertical dynamics studies was developed, consisting of seven rigid bodies ie. car body, twobogies, four wheels and axles and stiffness of primary and secondary suspensions. Degrees of freedom

considered are the bounce (z), pitch () and roll () of the car body and bogie. Besides, the wheel and axle'sbounce and roll are also taken into account adding up to 17 degrees of freedom (Figure 3). The equations ofmotion were written for all the masses and moments of inertia and are rearranged in matrix form [9]

[ ]{ } [ ]{ } { }0xKxM =+&& (1)

zc

zb1

zw1

c

b1

w1

b1

c

Figure 3 Seventeen dof vertical rigid body model of vehicle / track system

3 SPERLINGS RIDE INDEX (Wz)

The ride index Wz was introduced by Sperling, in order to evaluate ride quality and ride comfort of a railroadvehicle [10]. The vehicle itself is judged by the ride quality and ride comfort implies that the vehicle is assessedaccording to the effect of mechanical vibration on the occupant.

Ride quality, ( ) 10/13896.0 faWz= (2)

Ride comfort, ( )[ ] 10/13 )(896.0 fFfaWz= (3)

where a is the peak acceleration in cm/s2, f the frequency in Hz and F(f) the frequency dependent weighting factor

that expresses human vibration sensitivity. F(f) is different for vertical and horizontal vibration components. Table1 gives the ride index evaluation scale. Equations (2) and (3) were rewritten without changing their contents, tosuit the use of electronic instrumentation directly to evaluate the ride index.

( ) 10/133BaWz=

( ) 67.6/122BaWz= (4)


5/15

where a is the amplitude of acceleration in cm/s2

and B the acceleration weighting factor (Figure 4). Thefrequency weighting factors are defined for ride quality and ride comfort in different directions as follows.

The weighting function B for vehicle ride quality is

( ) ( ) ( )[ ]

( ) ( )[ ]( )

2/1

22322

2222

55.3100444.0547.1252.01

35.30645.0056.0114.1

++

+=

ffff

ffB (5)

The weighting factor B for ride comfort in the horizontal direction is given by

( )( ) ( )

2/1

2322

222

0368.0563.1277.01

25.0911.1737.0

+

+=

fff

ffBw (6)

The weighting factor B for ride comfort in the vertical direction is given by

( )

( ) ( )

2/1

2322

222

0368.0563.1277.01

25.0911.1588.0

+

+=

fff

ffBs (7)

Hence

sw BB 25.1= (8)

0 5 10 15 20 25 300.0

0.2

0.4

0.6

0.8

1.0 Ride comfort horizontal

Ride comfort vertical

Ride quality

WeightingfactorB

Frequency Hz

Figure 4 Sperlings ride index frequency weighting curves


6/15

The vehicle body vibration is not at a single frequency, but encompasses a whole spectrum of frequencies inwhich the natural frequencies of the vehicle are very much pronounced. In such cases, the ride index calculationhas to be done for the entire spectrum. The Wz ride factor is determined for each individual frequency fromEquation (4) and the total Wz factor is calculated as

( ) 10/110103

10

2

10

1 ntotal WzWzWzWzWz ++++= (9)

or

( ) 67.6/167.667.63

67.6

2

67.6

1 ntotal WzWzWzWzWz ++++= (10)

The vibration spectrum of the vehicle is not discrete, but a continuous function of frequency. The total Wz rideindex can be obtained for a continuous spectrum by integration over the given range of frequencies.

10/1

332

1

=

f

f

dfBaWz (11)

where f1 and f2 are the lower and upper end of the range of frequencies considered.

Table 1 Ride evaluation scales ride quality and ride comfort [10]

Ride index Wz Ride quality

1 Very good

2 Good

3 Satisfactory

4 Acceptable for running

4.5 Not acceptable for running

5 Dangerous

Ride Index Wz Ride comfort

1 Just noticeable

2 Clearly noticeable

2.5 More pronounced but not unpleasant

3 Strong, irregular, but still tolerable

3.25 Very irregular

3.5 Extremely irregular, unpleasant, annoying;prolonged exposure intolerable

4 Extremely unpleasant ; prolonged exposure

harmful

3.1 Ride Index Analytical Response

The ride index (both ride comfort and ride quality) was calculated from the predicted dynamic response. A railroadvehicle can be treated as a system with eight random loadings, i.e. random input disturbances due to the trackirregularities at each of the eight rail wheel contact points. If the input from the left rail is completely correlated


7/15

with that of the right rail, then the system can be simplified to a case of four random loadings. For a single randomloading to a system, the response is given by the following equation.

)()()(2

fSiffS px = (12)

Equation (12) gives the simple relationship between the spectral densities of excitation and response. Here S x(f)

is the output displacement power spectral density, Sp(f) is the input force power spectral density and (if) is the

ratio of displacement at any point to a unit sinusoidal force as the input. This equation can be extended for fourrandom loadings. The dynamic response of different points on the coach floor was predicted using the FE modelwith the track irregularity PSD as input. Track PSDs of the Indian rail tracks were obtained from Iyengar andJaiswal [11], the origin of which can be traced to Research and Design Standards Organisation (RDSO),Government of India, Ministry of Railways. The track PSDs used for the present work have been shown in Figure5 as spatial PSDs.

The dynamic response at four points on the central line of the coach as indicated in Figure 6 has beendetermined. Table 2 shows the ride index (vertical direction) at four different points on the coach. The table showsboth ride comfort and ride quality at different speeds. As shown in Figure. 7, the response at different points onthe coach varies at different speeds. This variation is reflected in the ride index also. In all the cases, the extremeend of the coach has higher vibration and hence higher ride index. The pattern of variation of vertical accelerationresponse along the length of the coach is not same at all speeds. This is seen from the variation of ride index at

different points too. For the coach analysed, upto 60 kmph, ride comfort value is less than 2, i.e. in the justnoticeable to clearly noticeable region. From 75 kmph to 150 kmph, ride comfort is in the range of morepronounced but not unpleasant to strong, irregular, but still tolerable region. Similarly, ride quality upto 60 kmphvaries from good to satisfactory and 75 kmph to 150 kmph, it is in the range of satisfactory to acceptable forrunning. It is to be stressed that the coach used for the analysis is a suburban coach which runs at an averagespeed of 45 kmph. From the lateral acceleration response, ride index in the lateral direction was calculated andtabulated in Table 3. The lateral acceleration levels were higher than the vertical acceleration and at speedsabove 75 kmph, the vehicle showed acceleration levels and ride indices higher than the acceptable limits. Thelateral rms acceleration does not vary along the length of the coach and hence same ride indices were found atdifferent locations on the coach floor.

0.01 0.1 1

10-1

100

101

102

103

Track vertical profile PSD

Verticalprofilemm

2 /

(c/m)

Frequency c/m0.01 0.1 1

10-2

10-1

100

101

102

103

104

Track alignment PSD

Alignmentmm

2 /(c/m)

Frequency c/m

0.01 0.1 110

-2

10-1

100

10

1

102

103

104

Track Cross level PSD

Crosslevelm

m2 /(c/m)

Frequency c/m0.01 0.1 1

10-2

10-1

100

10

1

102

103

104

Track gauge PSD

Gaugemm

2 /(c/m)

Frequency c/m

Figure 5 Geometrical track irregularity PSDs in spatial domain (mm2/(cycle/m) vs cycle/meter)


8/15

1 3 5 7

2 4 6 8

Longitudinal

Lateral

Figure 6 Measurement points on the floor of the coach (Top view of the floor)

0 4000 8000 12000 16000 20000

0.00

0.02

0.04

0.06

0.08

0.10

150120

9075

6045

3015

Vertical rms acceleration along solebar length

Spee

dkm

ph

Distance along solebar mm

Verticalrmsaccelerationg

Figure 7 The vertical rms acceleration response along solebar length

Table 2 Ride index predicted (vertical direction) at different speeds

Ride quality(at locations 2, 4, 6 and 8 on floor)

Ride comfort(at locations 2, 4, 6 and 8 on floor)Speed

kmph 2 4 6 8 2 4 6 8

15 1.26 1.22 1.30 1.36 1.43 1.36 1.46 1.56

30 1.90 1.38 2.03 1.47 2.31 1.51 2.46 1.66

45 1.84 1.56 1.96 1.64 2.20 1.64 2.37 1.82

60 1.84 1.81 1.84 1.87 2.01 1.91 1.94 2.02

75 2.15 2.10 2.24 2.22 2.38 2.23 2.51 2.37

90 2.56 2.51 2.72 2.72 2.83 2.70 3.03 2.92

120 2.80 2.78 2.94 3.20 3.03 2.99 3.17 3.24

150 2.73 2.65 2.79 2.84 3.07 2.96 3.14 3.18


9/15

Table 3 Ride index predicted (lateral direction) at different speeds

Ride quality(at locations 2, 4, 6 and 8 on floor)

Ride comfort(at locations 2, 4, 6 and 8 on floor)Speed

kmph 2 4 6 8 2 4 6 8

15 2.26 2.26 2.26 2.26 2.58 2.58 2.58 2.59

30 2.31 2.31 2.31 2.31 2.65 2.65 2.65 2.65

45 2.06 2.06 2.06 2.06 2.40 2.40 2.40 2.4060 2.02 2.02 2.02 2.02 2.29 2.29 2.29 2.29

75 2.81 2.81 2.81 2.81 3.22 3.22 3.22 3.22

3.2 Ride Index Measured Response

The ride comfort and the ride quality were calculated from the measured acceleration responses at differentpoints on the coach. The measurements were carried out at different stretches of suburban track at about 45kmph speed as explained in [12]. Table 4 shows the comparison of the ride index calculated from the analyticalresponse and the measured response in the vertical direction and Table 5 shows the comparison of lateral rideindices. The ride index obtained from measured response has the same pattern of variation as the analytical ride

index with 10 to 30 % higher magnitudes. The measured dynamic response also has the same level of highervalues; this can be attributed to incomplete inputs used for analytical prediction, system parameter variation andpossible small speed variations during measurement. Along with the track irregularities, irregularities on wheelslike wheel flats might have been present in the normal running condition. The coaches used for measurementwere the same as the one modelled, but a difference in system properties due to extensive use might have beenpresent. The track PSD used for analysis was measured on a continuously welded main line track [11], whereasthe response measurements were carried out on suburban tracks with short welded rails. It has been found thatthere are variations in the response amplitude for different stretches of the track. Tables 4 and 5 give two sets ofride indices calculated from the dynamic response measured on different stretches of the track.

The FE model was used to predict the ride index of the railroad vehicle at different locations of the coach floor.There exist considerable variation of the ride indices calculated from one end of the coach to the middle of thecoach (Table 2). The measured response and corresponding ride indices verified these analytical findings. A rigidbody model cannot predict the variation in the ride index along the length of the coach. This emphasises the needfor using an FE model for ride index analysis. The measured ride indices show that the ride comfort at 45 kmph isin the range of clearly noticeable to more pronounced but not unpleasant. It has been found that the end of thecoach has higher vibration levels than the rest of it. The analytical studies show that the car cg can have higher orlower amplitude of vertical acceleration response compared to other points on the coach depending on thelocation in the coach and speed (Figure 7). This has been clearly brought out in the analytical ride indices shownin Table 2.

Table 4 Comparison of analytical and experimental ride indices at 45 kmph vertical

Analytical Experimental*Locationon thecoach

Ride quality Ride comfort Ride quality Ride comfort

2 1.84 2.20 2.582.59

2.822.84

4 1.56 1.64 2.012.02

2.212.26

6 1.96 2.37 2.362.16

2.712.23

8 1.64 1.82 2.192.02

2.442.26


10/15

Table 5 Comparison of analytical and experimental ride indices at 45 kmphlateral

Analytical Experimental*Locationon thecoach

Ride quality Ride comfort Ride quality Ride comfort

2 2.06 2.40 3.093.07

3.343.32

4 2.06 2.40 3.12

3.10

3.37

3.366 2.06 2.40 3.00.3.15

3.113.38

8 2.06 2.40 3.013.06

3.143.21

* Two sets of values for two similar stretches of track indicated

4 COMFORT GUIDELINES ISO 2631

The evaluation of ride comfort involves assessment of human sensitivity to vibration, which depends not only onthe physiological and biomechanical response of the human body, but also on a number of psychological andenvironmental factors. The human reaction to vibration is a function of the amplitude and frequency ofacceleration applied to the body, the direction (vertical and horizontal) and character of the motion (linear or

rotation). Extensive research has been conducted on human sensitivity to vibration and its results have beenreviewed in literature [13]. The International Standards Organisation [2] has specified numerical values for limits

of exposure to vibrations transmitted from a solid surface to the human body in the frequency range of 1 80 Hz.These limits cover human sensitivity to vertical and lateral vibration to periodic vibration exposure time rangingfrom 1 minute to 24 hours.

Figure 8 shows the predicted vertical rms acceleration response of car cg in onethird octave bands at differentspeeds, superimposed over ISO curves for 1 hour and 8 hour reduced comfort boundaries. It can be seen fromthe figure that upto 90 kmph, the response acceleration levels are below the 1 hour curve. The frequency range ofthe analytical response increases with speed as the input temporal track PSD has larger frequency componentsat higher speeds. Predicted vertical acceleration response of four different points on the coach floor has been

plotted along with ISO curves in Figure 9. The rms acceleration values in onethird octave bands of the measuredacceleration response in the vertical direction, of four different points on the coach floor were calculated andplotted along with ISO curves in Figures 10


11/15

2 5 12.5 31.50.001

0.010

0.100

2 5 12.5 31.50.001

0.010

0.100

2 5 12.5 31.50.001

0.010

0.100

2 5 12.5 31.50.001

0.010

0.100

2 5 12.5 31.50.001

0.010

0.100

2 5 12.5 31.50.001

0.010

0.100

ISO 1 Hr

ISO 8 Hr

30 kmph

Frequency Hz

ISO 1 Hr

ISO 8 Hr

90 kmph

Rms

accelerationg

Rmsac

celerationg

Frequency Hz

ISO 1 Hr

ISO 8 Hr

120 kmph

Rmsaccelerationg

Frequency Hz

ISO 1 Hr

ISO 8 Hr

45 kmph

Rmsaccelerationg

Rmsaccelerationg

Frequency Hz

ISO 1 Hr

ISO 8 Hr

60 kmph

Rmsaccelerationg

Frequency Hz

ISO 1 Hr

ISO 8 Hr

150 kmph

Frequency Hz

Figure. 8 Analytical vertical response (1/3 octave rms acceleration) and ISO curves for reduced comfort boundaryfor 1 hour and 8 hours at various speed.


12/15

2 5 12.5 31.51E-3

0.01

0.1

2 5 12.5 31.51E-3

0.01

0.1

2 5 12.5 31.51E-3

0.01

0.1

2 5 12.5 31.51E-3

0.01

0.1

ISO 1 hr

ISO 8 hr

45 kmph

Rmsaccelerationg

Verrtical rms acceleration at point 2

R

msaccelerationg

Frequency Hz

ISO 1 hr

ISO 8 hr

45 kmph

Rmsaccelerationg

R

msaccelerationg

Frequency Hz


ISO 1 hr

ISO 8 hr

45 kmph

Frequency Hz


ISO 1 hr

ISO 8 hr

45 kmph

Frequency Hz


Figure. 9 Analytical vertical response (1/3 octave rms acceleration) of different points on a coach floor and ISOcurves for reduced comfort boundary for 1 hour and 8 hours

2 5 12.5 31.51E-3

0.01

0.1

2 5 12.5 31.51E-3

0.01

0.1

2 5 12.5 31.51E-3

0.01

0.1

2 5 12.5 31.51E-3

0.01

0.1

ISO 1 hr

ISO 8 hr

45 kmph

Rmsaccelerationg


Rmsac

celerationg

Frequency Hz

ISO 1 hr

ISO 8 hr

45 kmph

Rmsaccelerationg

Rmsaccelerationg

Frequency Hz


ISO 1 hr

ISO 8 hr

45 kmph

Frequency Hz


ISO 1 hr

ISO 8 hr

45 kmph

Frequency Hz


Figure. 10 Measured vertical response (1/3 octave rms acceleration) of different points on a coach floor and ISOcurves for reduced comfort boundary for 1 hour and 8 hours


13/15

5 PARAMETRIC STUDY

A parametric study was undertaken to find the influence of different suspension parameters on the dynamicresponse or ride index. This study was carried out using the 17 dof vertical rigid body model and track PSD asinputs. The parameters considered for the analysis were primary stiffness, primary damping, secondary stiffness,

secondary damping and car mass (tare gross).

These parameters were varied by 25% of their present values in steps of 5%, except for the car mass, which

was varied from tare mass to gross mass, in order to find the influence of loading on ride index. The vehiclespeeds considered for the analysis were 15 to 150 kmph in steps of 15 kmph. The performance indicesconsidered for the study were RMS acceleration at car cg and bogie cg, RMS displacement at car cg and bogiecg and ride index at car cg

The results of the parametric study have been plotted in terms of performance index versus vehicle parameter.Figures 11 and 12 show the variation of different performance indices as a function of primary stiffness andsecondary stiffness respectively. Remaining parameters and its influence have been summarised in the nextsection (graphs are not presented)

1800 2000 2200 2400 2600 28000.00

0.05

0.10

0.15

0.20

0.25

Boige cg

Car cg

Primary stiffness N/mm

Rms acceleration at bogie cg vs. primary stiffness

RMSaccelerationg

1800 2000 2200 2400 2600 28000.00

0.05

0.10

0.15

0.20

Car cg kmph15

30

45

60

75

90

105

120

135

150

Rms acceleration at car cg vs. primary stiffness

RMSaccelerationg


1800 2000 2200 2400 2600 2800

20

30

40

50

Boige cg


RMSdisplacementmm

Rms displacement at car cg vs. primary stiffness Rms displacement at bogie cg vs. primary stiffness

1800 2000 2200 2400 2600 280012

14

16

18


kmph

15

30

45

60

75

90

105

120

135

150RMSdisplacementmm

Figure. 11 The influence of primary stiffness


14/15


15/15

6 SUMMARY AND CONCLUSIONS

The Sperlings ride indices were determined for a railroad vehicle both experimentally and analytically using thetrack PSD as input. The Sperlings ride quality and the ride comfort were predicted analytically at different speeds.Using ISO 2631 guidelines, reduced comfort boundaries for 1 hour and 8 hours have been superimposed over thepredicted response and measured response.

An extensive parametric study has been carried out with emphasis on better ride index and reduced dynamicresponse. The parametric study has brought out possible design changes required in different parameters todeliver better ride index. It should be noticed that the parametric study was carried out to suggest designmodifications to improve ride index, but other dynamic behaviour like stability, curve negotiation capability etc.were to be considered when implementing the design modifications.

REFERENCE

1. EERI B 153/RP 21Application of ISO standard 2631 to railway vehicles.Utrecht, 1993

2. ISO 2631 (1985) Evaluation of human exposure to whole body vibration. International StandardsOrganisation.

3. Elmaraghy, W. H (1987) Ride quality and dynamics of rail vehicle models for microcomputers. Vehicle

System Dynamics,16, 193 211.4. Lyon, D. (1987) Dynamic measurements in the research and development of rail vehicles. Vehicle System

Dynamics,16, 149 165.

5. Sujatha, C., P. V. Phaskara Rao, and S. Narayanan (1995) Whole body vibration exposure in Indianbuses. Heavy Vehicle Systems, International Journal of Vehicle Design,2 (2), 160 173.

6. Suzuki, H (1998) Research trends on riding comfort evaluation in Japan, Proceedings of Institute ofMechanical Engineers Part F, 212, 61 72.

7. Hariharan, T. (2000) Estimation of Ride Index of High Power Diesel Multiple Unit Rail Coach. M. Techthesis, IIT Madras, India.

8. Gangadharan, K. V. 2001Analytical and Experimental studies on dynamics of railroad vehicles, PhDthesis, Indian Institute of Technology Madras

9. Gangadharan, K. V., Sujatha, C. and Ramamurti, V., 1999, Railroad vehicle dynamics comparison ofrigid body model and FE model. Proceeding of Asia pacific vibration conference, Singapore, December1999, 939 944.

10. ORE reports C116/RP 19 /EC 19711978, Interaction Between Vehicle and Track. ORE, Utrecht, 1978.

11. Iyengar, R. N. and O. R. Jaiswal 1995 Random field modelling of railway track irregularities, Journal ofTransportation Engineering July/ August 303 308.

12. Gangadharan, K.V., Sujatha, C. and Ramamurti, V., 2001 Railroad vehicle dynamics Experimentalstudies on Mass Rapid Transit System (MRTS) at Chennai. Proceedings of the 12

th Indian Society of

Mechanical Engineers (ISME) Conference on Mechanical Engineering, January. 2001, 357363.

13. Nigam, N.C. and S. Narayanan.Applications of Random Vibrations. Narosa Publishing House, New Delhi,

1994.

Documents

Ride Index