Engdyn Rover Kseries

8/6/2019 Engdyn Rover Kseries

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Crankshaft Durability of Rover K-Series Engine:

Comparison of ENGDYN Analysis with DynamicMeasurements

Roger B. Dailly David J. BellBMW Group, Birmingham, UK Ricardo Consulting Engineers

Figure 1: MGF with 1.8L K-series VVC engine

Abstract

This paper describes the technique used to perform a dynamic simulation of the Rover K-

Series crankshaft with the aim of predicting multi-axial stresses and Fatigue Factors of Safetyusing ENGDYN from Ricardo. This production crankshaft had been previously analysed andvalidated using in-house software techniques and contracted instrumentation work.ENGDYN was used to predict bearing loads and oil film thicknesses across the runningrange of the engine. Torsional vibration predictions were then compared to actual dynamicmeasurements of the crankshaft and in-house software results. Multi-axial stresses andF ti F t f S f t th d t t i t ENGDYN


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1 Introduction

The Rover K-Series engine has now been in

production for over 10 years, initially of 1.4 L

capacity. Since then 1.1L, 1.6L, 1.8L and 1.8L

VVC (Variable Valve Control) have beenintroduced with an accumulative build of two

million engines since 1989. The 1.8L engine is

also built under official licence for Lotus andCaterham, and also in the MGF motorsport

series. The K-Series is an attractive buy due toits reduced cost and low weight, which is anecessity for sport cars.

This report aims to validate Ricardo ENGDYN

software with respect to Rover K-Series 1.8

Litre VVC crankshaft durability. The softwarewill be used to output the behaviour of thecrankshaft under as realistic conditions as

possible. The oil film thickness and bearingload characteristics of the crankshaft throughthe running range of the engine will be

compared to results obtained from in-house

software. The torsional and bending vibrationoutput from ENGDYN will then be compared

with results obtained from dynamic

measurements. Finally, the crank stress anddurability results from ENGDYN will be

compared to strain gauge measurements at

comparative points on the crankshaft.

ENGDYN is a computer program used foranalysing the dynamics of the engine, and in

particular the crankshaft and its interaction with

the cylinder block. In this analysis the software

will be used to predict the time-domainresponse of the 3-dimensional vibration of the

crankshaft coupled to the block by way of a

non-linear oil film. When this loading andmotion has been calculated the software can

perform a fast Fourier transform to break down

2 Method of Analysis

2.1 Engine Specifications

Configuration: in-line 4

Fuel: GasolineCylinder bore: 80 mm

Piston stroke: 89.3 mm

Swept volume: 1.8L Crankpin

Peak Power: 107 KW @ 7000 rpmPeak Torque: 174 N/m @ 4500 rpm

Engine running range: 750-7200 rpm

2.2 Component Modelling

2.2.1 Crankshaft

To perform the analysis within ENGDYN two

crankshaft models were created. These

included a complete stiffness representation of

the crank (excluding the crank nose hub and theflywheel), and a detailed model of the crank

from main bearing 4 to main bearing 5, with

mesh density increased around the fillets.

ENGDYN can however perform crank analysisof any portion of the crank as long as the model

incorporates at least two main journal bearings.Features such as bolt holes and oil drilling were

omitted on both models, which were meshed

using solid tetrahedral elements.

The stiffness model of the crank as shown in

Fig. 2 was used to generate the mass andstiffness matrices of the crank within

ENGDYN. The flywheel and torsional

vibration damper assemblies were added withinENGDYN as lumped masses concentrated at

the appropriate points. The detailed crank


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was treated as a lumped mass which did not

contribute to the stiffness of the crank.

The torsional vibration damper was defined inENGDYN by the hub and annulus properties.

This flywheel and vibration damper areeffectively added to the FE model of the

crankshaft so that ENGDYN can generate mass

and stiffness matrices of the full crank. It ismore accurate to define these components in

the FE model, but for this analysis these

components were not previously modelled.

2.2.2 Cylinder Block

The cylinder block was received as an IDEAS

model file and required some modification

before being read into ENGDYN. The modelof the block was very large and it was

necessary to create a master-degree-of-freedomset, which represented a condensed mass andstiffness matrix of the engine block. These

matrices were derived in MSC/Nastran by

static condensation performed in a normal

modes solution.

2.2.3 ENGDYN Solution Technique

The software previously used by CAE Design

Analysis for analysing engine bearing

behaviour was Engine Bearing Analysis(EBA). This software represented a rigid

cranktrain with appropriate lumped masses

modelled within a rigid block. The oil film

being modelled using the Booker MobilityMethod. For direct comparison it was therefore

necessary to solve an indeterminate solution

within ENGDYN(1)

against a rigid block. Forthis analysis the oil characteristics and cylinder

firing pressure maps across the speed range

the journal bearings. The centrifugal effects

were also taken into account due to the mass

and inertia of the crankshaft. 10W30 oil wasused along with pressure maps for a full load

engine condition. A no-load condition ismore detrimental to crank life since the inertia

torque is not relieved by the gas torque,

however accurate cylinder pressure maps werenot available for this load case and hence a

full-load condition was used for comparison.

Bearing force and eccentricity results were thenoutput and compared with EBA for full load

and inertia only running conditions.

With the FE models of the crank and block

defined in ENGDYN, a dynamic/compliantsolution was then performed to evaluate the

torsional and bending response of the

crankshaft. The dynamic solution generatesmass and stiffness matrices of the crank andincludes the non-linear gyroscopic effects in

the solution. The reactions and journal bearing

orbits are then calculated for each pin and journal bearing, with the 3-dimensional

vibration behaviour also evaluated. The block

was set to compliant, which considered only

the stiffness of the block. This means the blockwas modelled such that the natural mode

shapes did not contribute to its interaction with

the crankshaft. The results from this solutionwould then be compared to those from

experimentally measured results. A dynamic

solution of the crank against a rigid block wasalso done so that the results could be compared

with the Rover Torsional Vibration (TV)program. The TV software uses a mass-elasticsystem to represent the inertia and stiffness

between each web and journal along the crank.


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was already known that the fillet in main

bearing five closest to web 8 suffered the

greatest loading. Failure had occurred in thisregion during durability tests at 7500 rpm.

Since the stress model used in this analysisconsisted of webs 7 and 8, then stresses were

only resolved in this area. ENGDYN splits the

model at cut planes through the journal centres.Each split would extend from the centre of a

main journal bearing to the centre of the next

immediate pin journal bearing. Taking eachsplit part in turn ENGDYN applies unit force

loads in the four orthogonal directions (+Y,-

Y,+Z,-Z) at both the main and pin journal

centres. In addition unit moment loads in thefour orthogonal directions are applied at the

main journals. For this dynamic solution, which

takes into account bending vibration, additional

unit moment loads were applied at the mainbearing centres. Unit torsional and axial loads

were applied at what would be the nose of thecrank. In this case the nose is represented by

the start of main bearing 4, as shown in Fig.3.

A body load equivalent to a constant angular

velocity about the crankshaft axis was alsoapplied so as to include the centrifugal effects

of the crankshaft. For the model analysed here,

the total number of unit loads amounted to 23.

ENGDYN then writes out a file, which can beconverted to an MSC/Nastran deck and solved

to produce stress per unit loading. The stresses

due to quasi-static (a snap shot in the timecycle of crank rotation) bending, torque and

vibration loading are calculated such that anumber of combined loadcases are created ateach crank angle. It is the variation of the

bearing loads through the crank cycle that

causes the stress to change, and hence a newfactor is required so that the unit loads can be

speed and Goodman diagrams to be created.

The resulting file containing all of the factored

stresses for each fillet can then be convertedand read into IDEAS for post-processing the

stress and FSF plots. For this solution the FFSwere calculated for the speed range 5000 rpm

to 7000 rpm.

3 Results & Discussion

3.1 Normal Modes

Table 1 shows the frequencies and mode shapedescriptions for a free-free crank with

conventional flywheel attached as produced by

the Lanczos method using MSC/Nastran. Theflywheel was attached to the crank palm with a6 degree-of-freedom spring with stiffness

equivalent to that of the flywheel. The

installed modes were calculated with thecrankshaft fitted with the torsional vibration

damper, flywheel and masses equivalent to the

con-rod big-end mass plus piston mass addedto the centre of the pin journals. The entire

crank was then grounded with linear springs of

5x108

N/mm stiffness at all four of the main

journal bearing centres.

Natural

Frequency

Mode Shape Description

175 Hz Vertical Bending224 Hz Lateral Bending

358 Hz 2nd

Vertical Bending

509 Hz Torsional Vibration

547 Hz 2nd

Lateral Bending


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Table 2 shows that the torsional vibration

damper created two smaller torsional vibrations

at 321 Hz and 541 Hz. Without the TV damperthe torsion mode would create a larger peak

and in this case was excited at 509 Hz.

NaturalFrequency

Modes Shape Description

213 Hz Crank Nose Bending

238 Hz Crank Nose Bending

321 Hz 1st

Torsional Vibration

349 Hz Crank Palm Bending

394 Hz Crank Palm Bending

478 Hz 2nd Vertical Bending

541 Hz 2nd

Torsion Vibration

644 Hz Flywheel Axial

778 Hz Swashing of Flywheel

786 Hz Swashing of Flywheel

Table 2: Installed modes of crankshaft

3.2 Bearing Loads & Oil Film EccentricitiesFig. 4a illustrates the ENGDYN solution for

the characteristics of main bearing 5 under afull load condition at 4000 rpm. Fig. 4b shows

how the EBA results for the same bearing

compare It can be seen from the graphs that

of the oil film whereas the EBA calculation

performs a more simple calculation. A similar

comparison was made for an inertia onlycondition at an engine speed of 7500 rpm. Fig.

5a/b show results for main bearing 4 for aninertia only engine condition. As the no load

pressure maps were not available the

ENGDYN solution was run with only theinertia loading of the crankshaft. This would

account for some differences in the bearing

eccentricities as EBA is a no load solution.However the maximum bearing load compares

exactly.

3.3 Torsional Vibrations

The ENGDYN results for a dynamic solution

of the crankshaft supported in a compliantblock are shown in Fig. 6. Node 1 represents

the node at the centre of the hub on the crank

nose. The figure shows the response of thisnode in the rotational direction along the crank

axis. The main orders of vibration are plotted

and these are relative to a rotating axis set. Thespeeds at which these resonant peaks occur

compare quite well to those obtained throughexperimental measurement as shown in Fig. 7.The amplitudes of these peaks depends very

much on the damping coefficient of the damper

rubber and the overall cylinder damping withinthe engine. For this analysis the damping

coefficient of the damper rubber was calculated

to be 1.77 N.m.s/rad and the cylinder damping

was set to 1000 Nms/rad (default inENGDYN). The torsional vibration results are

tabulated in Table 3 and shows how the

predicted values compare to the measured andin-house results. It can be seen that the

amplitudes of the predicted harmonic peaks


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order peak. This indicates that the FE model of

the crankshaft does not represent the real

crankshaft accurately enough.

Improved results can be obtained by

incorporating actual FE models of the torsional

vibration damper and flywheel to moreaccurately model these components in the

solution.

HarmonicOrder ENGDYNResults MeasuredResults

4.0th Order

Speed (rpm)

Freq. (Hz)

Amp. (Deg.)

4400

293

0.15

4750

317

0.15

6.0th Order

(I,II)

Speed (rpm)

Freq. (Hz)

Amp. (Deg.)

3000 & 5000

300 & 500

0.07 & 0.06

3400 & 5125

340 & 512

0.09 & 0.07

3.5 Order

Speed (rpm)

Freq. (Hz)

Amp. (Deg.)

5000

292

0.10

5625

328

0.10

Table 3: Comparisons of measured and

predicted torsional vibrations.

Fig. 8b shows the ENGDYN results for adynamic/rigid solution. The dominant orders

of vibration are plotted and show that 4.0 th

order vibration is greatest around 4400 rpm.

The sixth order peaks occur at 3000 rpm and

3.4 Bending Vibration & Strain Gauge

Measurements

The motion of the node representing mainbearing 5 (closest to crankshaft palm) was

analysed. The results provided a comparison of

bending vibration to the strain gaugemeasurements. Fig. 9a shows the axial

displacement of this node in the X-direction.

The graph shows that 2.0 order (relative to the

crank) vibration is dominant at high engine

speed and reaches a maximum around 7000rpm. This amplitude equates 0.185 mm 2.0order and 0.228 mm in the time domain. At a

lower engine speed the crank is excited at 5400

rpm by 2.5 order and 6100 rpm by 1.5 order. Asimilar result was obtained for the movement

of the same node in the vertical Y-direction as

shown in Fig. 9b. Inspection of the results

relative to the fixed datum show that the 1.5and 2.0 order resonances correspond to the

forward flywheel whirl mode whilst the 2.0

order resonance corresponds to the reversewhirl mode of the flywheel. A photograph of

the instrumented crankshaft is shown in Fig.10a. The location of the strain gauge is

illustrated in Fig. 10b. The precise location ofthe gauge corresponds to the area of highest

stress in the fillet, found from previous

analysis. Areas of high strain will coincidewith areas of high stress and hence low factors

of safety. The strain gauge measures

predominantly axial and vertical strain causedby bending of the fillet, the results of which are

shown in Fig. 11. Fig. 12 compares the 1.5 and2.0 order measured strain amplitudes ofFig. 11with those derived from the ENGDYN

analysis. This shows good correlation of the

resonant frequencies for each of these harmonicorders. Inspection of the measured 1.5 order


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thought that this vibration absorber mechanism

may be due either to the effect of the oil film at

the adjacent main bearing, the flexibility of thecrankshaft web counterweight or more likely

the flexibility of the flywheel which was notincluded in this model (but can be included in

ENGDYN if required.) The correlation of the

1.5 order strain amplitude therefore requiressome further investigation. The predicted 2.0

order strains correlate well with the

measurements although the predicted amplitudeis lower than the measured value.

3.5 Crankshaft Durability

The stresses due to quasi-static and vibration

loading for every 10 through a cycle of 720

were calculated for each engine speed underfull load. These stresses were then combined

using an Alternative Goodman Criterion toproduce the fatigue safety factors through the

desired speed range, as shown in Fig. 13.These results are shown for the fillet at the rear

of web 8, closest to the crankshaft palm. Under

full engine load at 7000 rev/min the maximum

stress was found at 570 of crank rotation, witha value of 198 MN/m2. The stress distribution

at this condition is shown in Fig. 14a and themaximum stress corresponds to a fatiguesafety factor of 1.283. Similarly at 5000

rev/min as shown in Fig. 14b the maximum

stress occurs at 20 crank rotation. It can beseen therefore at higher engine speeds the shear

stress due to the gas torque rotates the high

stress region in the direction of crank rotation.

The ENGDYN factor of safety results compare

very well with those obtained throughexperimental measurement and follow a similar

trend. The general decrease in fatigue safety

factor with increasing engine speed is due to

rev/min.. The discrepancy at the lower engine

speeds is not easily explained since the

measured and predicted strains for thedominant 1.5 order resonance at these speeds,

as shown in Fig.12, correlate well. Howeverthe predicted 2.0 order strains at these speeds

are lower than the measured values. Further

work is required to explain the differences infatigue safety factor at these speeds. The

discrepancy in fatigue safety factors at speeds

between 6000 and 6500 rev/min is consistentwith the comparison of measured and predicted

strains as shown in Fig.12. As previously

discussed, this showed that the 1.5 order strain

amplitudes were over-predicted by ENGDYN.

4 Conclusions The results have shown that ENGDYN couldbe used with confidence to predict the bearing

load and oil film eccentricities of pin and main

journal bearings. The mathematics behind theoil flow rate calculation within ENGDYN is

more accurate than the EBA solution.

With respect to the torsional vibrationsENGDYN predicted slightly lower frequenciesfor 6.0 and 3.5 order vibration when compared

to measured data. However, when compared to

in-house software the frequencies matched for

all orders of vibration. The analysis did showhowever that the damping values used in the

model were a good approximation. The

amplitudes of the predicted peaks were close tothe measured values.

The ENGDYN results describing the motion

of main bearing 5 compared well with the

experimental results, in that both showed aid i i 2 0 d ib ti t d


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measurement and follow a similar trend. The

most significant discrepancy is the over

prediction of the 1.5 order strain amplitudewhich results in lower predictions of fatigue

safety factor. Further work is required toinvestigate this difference.

5 References

(1) Ricardo Software: Engine Dynamics

Simulation ENGDYN Users Manual,Revision 1.2

(2) Perkins Technology Consultancy:Predication of Durability for a 1.8L I4

Crankshaft Using Three Flywheel

Models, March 20, 1998

(3) Perkins Technology Consultancy: AnInvestigation into 1.8L K-SeriesFlywheel Whirl Activity, June 16, 1998


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Figure 2: FE Model of crankshaft for Mass and Stiffness Matrix Formulation

X

Y

Z


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Figure 7: Torsional vibrations measured at TV damper on running crankshaft(2)


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Main bearing 5, closest

to crankshaft palm.

Strain gauge location at

predicted high stress region

Figure 10a: Photograph of Crankshaft Instrumented with Strain Gauges(2)

Pin bearing 4

Gauge 1 and Gauge 2

(backup)


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Figure 11: Strain gauge results for web 8 fillet on main bearing 5(3)


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Figure 13: Fatigue Factor of Safety resolved for the worst stress in the fillet at eachcrankshaft speed


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5th Ricardo Software International User Conference 23

Figure 14a: Stress plot around fillet,

showing maximum stress at 570 crankangle at 7000 rpm full load

Figure 14b: Stress plot around fillet,

showing maximum stress at 20 crankangle at 5000 rpm


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5th Ricardo Software International User Conference 24

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