Ringpak Comparison of Predicted

8/13/2019 Ringpak Comparison of Predicted

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A COMPARISON OF PREDICTED AND MEASURED JOURNALBEARING PERFORMANCE CHARACTERISTICS

B.PlayfootRicardo Consulting Engineers Ltd.

The crankshaft bearings of a modern automotive engine have to operate under arduousloading conditions, often at high temperatures. Their reliable operation is dependant on acomplex interaction between hydrodynamic, structural and metallurgical effects. For theengine designer it is important to have easy to use proven analytical design tools to assist himin the task of designing suitable bearings. For many years Ricardo have used software basedon the Mobility method without necessarily going to the complexity of using an Elastro-hydrodynamic (EHD) model such as that used in the Ricardo program ORBIT. This articlepresents work originally performed by Ricardo and Leeds University to develop analytical toolsto predict the performance characteristics of journal bearings. In this original study the bearingloads were derived assuming the crankshaft and crankcase were rigid. This article reviews

the results of the original study and considers the effect of including the stiffness and dynamicbehaviour of the crankshaft. ENGDYN was used to predict the performance characteristicsand the results were compared with experimental data.

INTRODUCTION

The crankshaft bearings of a modern automotive engine have to operate under arduousloading conditions, often at high temperatures. Their reliable operation is dependant on acomplex interaction between hydrodynamic, structural and metallurgical effects.

The purpose of this study is to compare the predicted main bearing performance

characteristics from ENGDYN with experimental data.

EXPERIMENTAL SET-UP

RICARDO HYDRA ENGINE

The experimental data presented in this paper was collected during a research programcarried out by G.A.Clayton Ref.[1]. The experimental work was carried out at RicardoConsulting Engineers. A Ricardo Hydra single cylinder gasoline engine was used for the test,fitted with two fully grooved main bearings. Figure 1 shows a section through the engine. Thisengine has been specifically designed for research purposes and has several points that

make it particularly suitable for this type of work.

1. The engine has only one cylinder and hence the load sharing between the crankshaft mainbearings is greatly simplified.

2. The bearing housings and the crankcase are extremely rigid.3. The main bearings are fixed in rigid bearing housings, which, in turn fit into the crankcase.

They may be easily removed and this facilitates their instrumentation.


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Figure 1 Section of the Ricardo Hydra

INSTRUMENTATION

For this study, the front main bearing was instrumented. The use of the front main bearingensures that the dynamic flywheel effects are minimised, therefore simplifying the bearing

loading. In the original engine design, the front main bearing was used to supply oil to thecrankshaft main bearing. The crankshaft was therefore modified to enable oil supply from therear main bearing. Measurements were taken of the following parameters.

1. Inlet oil temperature and pressure2. Engine speed3. Outlet oil temperature4. Bearing temperature at a number of locations5. Oil flow rate through the bearing6. Cylinder pressure

Temperature Measurement

Eight rows of four thermocouples were positioned so that the weld beads lay 0.5mm below thebearing surface. Figures 2 and 3 illustrate the axial and angular locations of the K-type(Nickel-Chromium/Nickel-Aluminium) thermocouples, respectively. It can be seen from Figure2 that the axial positions 2 and 4 were equally spaced on the lands either side of the centralgroove. This enabled an assessment of the degree of bearing misalignment to be made,reflected as a difference in the temperature of the two lands. In addition, the bearing oil outlettemperature could be compared for both sides of the bearing at axial positions 0 and 01.


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Figure 2 Axial Bearing Temperature Measurement

Figure 3 Angular Bearing Temperature Measurement

The ENGDYN model does not currently incorporate a bearing temperature gradient model. Inorder that the temperature gradient can be neglected, an effective bearing temperature must

HighestTemperatures

LowestTemperatures

Mid-RangeTemperatures

Mid-RangeTemperatures


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be determined that enables the bearing performance to be accurately predicted. Claytonsstudy considered the following three ways of determining an effective bearing oil temperature.

1. Assume the oil outlet temperature2. An average of the thermocouple readings at angular positions 1, 5, 6 and 8. This provided

and average of the highest (8) and lowest (5) bearing temperatures, with an additionalweighting towards the middle range of bearing temperatures.

3. An average of the temperatures of the thermocouples at angular position 5. This gave aneffective temperature which is biased towards the lower end of the bearing temperaturerange.

Journal Orbit Measurement

Four proximity transducers were used to determine the journal centre orbit within the bearing.The axial position of the transducers relative to the crankshaft can be seen in Figure 2.

Oil Supply Circuit

Figure 4 shows a schematic representation of the oil supply from the engine sump to thebearing. The dotted line encircles the equipment that was added for the purpose of theexperimental investigation.

The supply pressure to the bearing was set by the position of the manually operated valve andwas measured using a dial gauge. To safeguard against accidentally starving the bearing ofoil a low pressure cut out was positioned as indicated.

Figure 4 - Schematic Representation of the Oil Supply System


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EXPERIMENTAL INVESTIGATION

During the tests, the influences of the oil supply pressure, engine speed and effective oiltemperature (and hence viscosity) on the oil flow rate through the bearing were investigated.Specific tests were undertaken to investigate:

1. The influence of the supply pressure in the oil flow rate for constant engine speeds.2. The effect of engine speed on the oil flow rate and oil temperature, for a constant supply

pressure.

For the purposes of this study, data is presented for just one engine oil, Rotella X30.

TECHNICAL BACKGROUND

ENGDYN OVERVIEW

ENGDYN is a fully-coupled crankshaft, oil film and cylinder block dynamic simulation package.The program provides a number of solution techniques for predicting engine dynamics usingmodels of varying sophistication. The crank train and cylinder block models can be modelledas rigid, compliant or dynamic. In its simplest form, ENGDYN can be used to perform astatically determinate solution, whilst in its most sophisticated form it enables the time-domainresponse of the 3-dimensional crank train and cylinder block vibrations with non-linear oil filmsat each of the bearings.An ENGDYN model consists of two components, the crank train and the cylinder block. Thesecomponents are defined using models of varying degrees of accuracy dependent on the

application. Each model can be specified from a finite element model or as user definedlumped masses. ENGDYN reduces the finite element model into a simplified mass elasticsystem, calculating the system stiffness and distributing the mass at node points.

The input loads due to the reciprocating motion of the piston and connecting rod, the rotatingmotion of the connecting rod and the force due to cylinder pressure are applied to each crank-pin and cylinder of the engine. The reaction forces at each of the main journal bearings and atthe thrust bearing are calculated according to several solution options.

ENGDYN enables three solution types:

Statically Determinate The reactions at the main bearings are calculated assuming the

crankshaft is pin-jointed at each of the bearings. The journal orbits are calculated fromthese reactions using the Mobility Method.

Statically Indeterminate The crankshaft and cylinder block are assumed to be compliant.The reactions and journal orbits at each of the main journal bearings plus the quasi-staticdisplaced shape of the crankshaft (and cylinder block) are calculated using the MobilityMethod at each main bearing. This method accounts for load sharing between bearings.


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Dynamic The crankshaft is assumed to be dynamic and includes the non-lineargyroscopic effects. The reactions and the journal orbits at each of the main journalbearings and the 3-dimensional vibratory behaviour of the crankshaft are calculated.

OIL FILM MODEL

The journal oil film model used by ENGDYN is a fully non-linear model based on methods byBooker and Kikuchi.

For radial motion (axis parallel to the bearing axis) of the shaft in the bearing, the oil film modelis based on the mobility method of Booker and using the short bearing approximation. In itsusual form, the method is used to determine the shaft eccentricity and attitude angle for aknown force. This is the procedure used for the quasi-static solution methods and for thedynamic solution when the cylinder block is compliant. However, when the cylinder block isrigid or dynamic the process is inverted so that the oil film load and its direction can bedetermined from a known shaft position and velocity.

For rotational motion (shaft axis tilting relative to the bearing axis) the short bearingapproximation is used to determine the stiffness and damping of the oil film. This is based onthe work of Kikuchi who modelled the oil film as a set of radial and rotational springs anddampers.

The instantaneous oil flow rate through the bearing and the power dissipated by the oil throughthe bearing action are calculated by ENGDYN for each journal bearing. These instantaneousvalues are integrated numerically to estimate the overall cyclic oil flow rate and power loss.

ENGDYN also offers the ability to perform solutions where the bearing oil temperature is

specified, or alternatively, calculated by performing a thermal balance. For the thermallybalanced solution, ENGDYN calculates the power loss at each angular position of the journalthen integrates for a complete engine cycle to calculate the total power loss in the bearing. Itis then assumed that this energy is used in heating the oil as it flows through the bearing. Theoil temperature is used to calculate the operating viscosity of the oil in the bearing. ENGDYNiterates until a converged solution is obtained.

TECHNICAL APPROACH

ENGDYN Model

A finite element model of the Hydra crankshaft was generated and read into ENGDYN. This isshown in Figure 5. The data properties for the equivalent reduced mass-elastic system werethen calculated by ENGDYN. The Hydra bearing housings and crankcase are extremely stiff,consequently a block model has not been incorporated in the analysis and is assumed to berigid. The main engine parameters are described in Appendix 1.


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Figure 5 Finite Element Model of Crankshaft

For this study, all three solution types have been considered (statically determinate, statically

indeterminate and dynamic). Each successive solution level offers greater potential accuracybut with a corresponding computational overhead penalty. Since this study is concerned witha single cylinder Hydra engine, there is expected to be little difference between the staticallydeterminate and indeterminate solutions.

Analyses have been carried out using both set and thermally balanced bearing temperaturesolutions. For the specified temperature solution, the bearing oil temperature was set equal tothe experimentally measured bearing oil outlet temperature.

DISCUSSION OF RESULTS

Figures 6 to 8 show the results of the determinate solution.

Figures 6a and 6c compare the effective temperatures measured by Clayton. This shows thatthe average temperature (Tave) is approximately 1.5% lower than the bearing outlettemperature (Tout) across the entire engine speed range. As previously discussed, theaveraged temperature at angular position 5 (T5) is at the lower end of the temperature range,and is lower than the outlet temperature by up to 5%.

A comparison of the ENGDYN predicted flow rates using these temperatures is shown inFigures 6b and 6d. As can be seen, the use of temperature T5enables a very good correlationbetween the experimental and predicted oil flow rates.

The oil temperature in the bearing rises rapidly across the bearing lands. The most restrictiveregion to oil flow is therefore likely to be close to the groove where the oil is cooler andtherefore more viscous. Consequently, the use of the lower effective oil temperature (T5)results in a better oil flow rate correlation.

However, the location of the lowest bearing temperature (T5) is largely dependant on thepresent geometry and loading conditions. Consequently, for the purposes of a thermallybalanced solution ENGDYN assumes the effective bearing temperature to be equal to theoutlet temperature. This approach allows a standard prediction approach to be maintained for


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all bearings. For a realistic comparison between the thermally balanced and set temperaturesolutions this paper therefore uses Toutfor the set effective bearing temperature.

Bearing temperatures and oil flow rates were measured across the engine speed range for oilsupply temperatures of 55, 70, 85 and 95 oC. A comparison between the ENGDYN set oil

temperature and thermally balanced analyses using the determinate solution is made in Figure7 for all four bearing oil temperatures. It can be seen that the ENGDYN thermally balancedsolution slightly under predicts the oil flow rate at all supply temperatures and speeds. The settemperature solution over predicts the oil flow rate by an increasing magnitude of error withspeed. The maximum percentage error for each of the four temperatures using both solutiontypes is compared in Table 1.

Supply Temperature 55 70 85 95

Set Temperature -21.6% -14.4% -8.9% -6.3%Thermally Balanced -13.3% -12% -9.3% -7.9%

Table 1 Oil Flow Rate Percentage Error

As can be seen, the percentage error reduces with an increase in supply temperature. Themaximum error for the thermally balanced solution being only 13.3%.

Figure 8 compares the variation of predicted and measured oil flow rate with oil supplypressure at three engine speeds (1500, 2500 and 3500 rev/min), with the bearing temperaturespecified and using a thermally balanced solution. It can be seen that a good correlation isachieved with both solution types, particularly at the lower engine speeds. The largest errorsare 15% for the set temperature solution at 3500 rev/min and 2.4 bar oil supply pressure and21% at 2500 rev/min and 3.1 bar.

Figure 9 compares the predicted bearing temperatures and oil flow rates for both determinateand indeterminate analysis solutions. This shows both analyses produce very similar results,with the indeterminate solution being slightly more accurate across the engine speed range.An improvement in accuracy would be much more noticeable with the indeterminate solutionwhen analysing a multi-cylinder engine, where load sharing between the main bearingsbecomes apparent.

Figure 10 compares the results of the dynamic analysis using set temperature and thermallybalanced solutions with the experimental and determinate analysis results. This shows thatalthough the dynamic analysis thermal balance solution predicts a similar bearing oiltemperature to the determinate solution this analysis method under predicts the oil flow rate atall speeds and lower than that predicted by the determinate analysis. However, the slope ofthe curve for the dynamic analysis is more consistent with the experimental result than the

determinate analysis at all speeds. This result suggests, given that the predictedtemperatures are similar, that the predicted oil film is significantly different between the quasi-static and dynamic analyses. Figures 11 and 12 shows this to be the case. Figure 11compares the journal orbit for the dynamic and determinate analyses at 3000 rev/min with theoil supplied at 85oC. The orbit for the indeterminate analysis is very similar to the determinateresult and has not been presented for clarity. Figure 12 compares the maximum pressure andminimum film thickness for the oil for each of the three analyses against speed at the same oil


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supply temperature. The larger minimum oil film thickness for the dynamic analysis isconsistent with the lower predicted oil flow rate shown in Figure 10.

SUMMARY

A good correlation between the determinate and indeterminate analyses and the experimentalresults has been presented. The results have shown that ENGDYN over-predicts the oil flowrate when the effective bearing oil temperature is specified. Likewise, with the thermallybalanced solution, ENGDYN under-predicts the oil flow rate at all engine speeds. TheENGDYN predicted bearing oil temperatures are generally lower than measured, thedifference increasing with engine speed. For the statically determinate analysis, very closeagreement is obtained between predicted and measured oil flow rates as the oil supplypressure is varied.

The results of the indeterminate analysis show a slight improvement in accuracy.

The dynamic analysis predicts lower oil flow rates than the quasi-static analyses but showssimilar trends to the experimental measurements. The predicted lower oil flow rates are due tothe larger oil film thickness predicted using this method.

REFERENCES

1. Engine Bearing Analysis and Design, G. A. Clayton, PhD Thesis, Department ofMechanical Engineering, University of Leeds, February 1990.


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Figure 6 Comparison of Set Temperatures Determinate Solution

70oC Oil Supply Temperature, 3.103 bar

6

8

10

12

14

16

18

1500 2500 3500 4500 5500

Engine Speed (rev/min)

OilFlow

Ra

te

(cm

3/s)


10

11

12

13

14

15

16

17

18

19

1500 2500 3500 4500 5500


OilFlow

Ra

te

(cm

3/s)


60

65

70

75

80

85

90

95

100

105

1500 2500 3500 4500 5500


Bearing

Tempera

ture

(oC)


80

85

90

95

100

105

110

1500 2500 3500 4500 5500


Bearing

Tempera

ture

(oC)

a

dc

b

1020

15

00

25

00

35

00

45

00

55

00

Experiment (Tout)

ENGDYN - Set Temperature (Tout)

ENGDYN - Set Temperature (T5)

ENGDYN - Set Temperature (Tave)


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55oC Oil Flow Rate

3

4

56

7

8

9

10

1500 2500 3500 4500 5500Engine Speed (rev/min)

OilFlow

Rate(cm

3/s)

55oC Oil Supply Temperature

5055

6065

70

7580

8590

1500 2500 3500 4500 5500


Bearing

Temperature

(oC)

70oC Oil Flow Rate

56

789

101112131415


OilFlow

Rate(cm

3/s)

85oC Oil Flow Rate

89

101112131415161718

1500 2500 3500 4500 5500


OilFlow

Rate(cm

3/s)


6065

707580859095

100105


Be

aringTemperature

(oC)


80

85

90

95

100

105

110

1500 2500 3500 4500 5500


BearingTemperature

(oC)

95oC Oil Flow Rate

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1500 2500 3500 4500 5500


OilFlow

Rate(cm

3/s)


90

95

100

105

110

115


BearingTemperature

(oC)

4090

0 2000 4000 6000

ENGDYN - Set Temperature

ENGDYN - Thermally BalancedExperimental

Supply Temp

Figure 7 Comparison of Thermally Balanced and Set Bearing Temperatures for DeterminateSolution (3.103 bar Oil Supply Pressure).


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Figure 8 Relationship between Oil Supply Pressure and Flow Rate (Determinate Solution

1500 rev/min, 100oC Oil Temp out

Set Temperature

02468

101214161820

0 1 2 3 4

Oil Supply Pressure (bar)

OilFlow

Rate(cm3/s)


Set Temperature

0

1

2

3

4

5

6

7

8

9

0 1 2 3 4


OilFlow

Rate(cm3/s)


Set Temperature

0

1

2

3

4

5

6

7

8

9

0 1 2 3 4


OilFlow

Rate(

cm3/s)


Thermally Balanced

02468

101214161820

0 1 2 3 4


OilFlow

Rate(cm3/s)


Thermally Balanced

0

1

2

3

4

56

7

8

9

0 1 2 3 4


OilFlow

Rate(cm3/s)


Thermally Balanced

0

1

2

3

4

5

6

7

8

9

0 1 2 3 4


OilFlow

Rate(cm3/s)

020

0 5

ENGDYN

Experimental


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70oC Oil Flow Rate

56789

10111213

1500 2500 3500 4500 5500Engine Speed ( rev/min)

O

ilFlow

Rate(cm3/s)

85oC Oil Flow Rate

89

10111213141516

1500 2500 3500 4500 5500

Engine Speed ( rev/min)

OilFlow

Rate(cm3/s)


6065

707580859095

100105


BearingTemperature

(oC)


6065707580859095

100105110

1500 2500 3500 4500 5500


BearingTemperature

(oC)

55oC Oil Flow Rate

3

4

56

7

8

9

10


OilFlow

Rate(cm3/s)


50556065

70

75808590


BearingT

emperature

(

oC)

95oC Oil Flow Rate

12

13

14

15

16

17

18

19

20


OilFlow

Rate(cm3/s)


90

95

100

105

110

115


BearingTemperature

(oC)

1415161718

0 5000

ENGDYN -IndeterminateENGDYN - Determinate

Experimental

Figure 9 Comparison of Indeterminate and Determinate Solutions(3.103 bar Oil SupplyPressure).


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70oC Oil Flow Rate

56789

10111213


OilFlow

Rate(cm

3/s)

85oC Oil Flow Rate

89

10111213141516

1500 2500 3500 4500 5500


OilFlow

Rate(cm

3/s)


6065

707580859095

100105

1500 2500 3500 4500 5500Engine Speed ( rev/min)

BearingTemperature

(oC)


6065707580859095

100105110


BearingTemperature

(oC)

55oC Oil Flow Rate

3

4

56

7

8

9

10


OilFlowR

ate(cm

3/s)


5055606570

75808590


BearingT

emperature

(

oC)

95oC Oil Flow Rate

12

13

14

15

16

17

18

19

20


OilFlow

Rate(cm

3/s)


90

95

100

105

110

115


BearingTemperature

(oC)

95oC Oil Flow Rate

121314151617181920

150

0

250

0

350

0

450

0

550

0


ENGDYN - Dynamic (balanced)

ENGDYN - Dynamic (set)

Experiment

ENGDYN - Determinate

Figure 10 Comparison of Determinate and Dynamic Solutions (3.103 bar Oil SupplyPressure)


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Figure 11 - Comparison of Journal Orbits at 3000 rev/min (85O

C Oil Supply Temperature,3.103 bar Oil Supply Pressure))

0

10

20

30

40

50

1500 2500 3500 4500 5500


MaximumF

ilmP

ressure(MPa)

0

1

2

3

4

5

6

7

1500 2500 3500 4500 5500


MinimumF

ilmT

hickness(microns)

05

10

1500

4500

Determinate

Indeterminate

Dynamic

Figure 12 Comparison of Oil Film Pressure and Thickness (85OC Oil Supply Temperature,3.103 bar Oil Supply Pressure)


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APPENDIX 1

TABLE 1 - Ricardo Hydra TestEngine Data

UnitsEngine

Bore mm 85.7Stroke mm 88.9

Piston

Mass kg 0.3566

Connecting Rod

Length mm 0.158Mass kg 0.9162Inertia about c.g. kgm2 5.08E-03

Big end to c.g. distance mm 52.67

Flywheel

Mass kg 31.2Inertia XX kgm2 0.45

Main Bearings

Length mm 27.96Diameter mm 57.139Radial Clearance mm 0.044Groove Width mm 5.4Feed Type - Full Groove

Pin Bearings

Length mm 20Diameter mm 47.66Radial Clearance mm 0.03

Feed Type - Oil HoleOil Hole Diameter mm 6Oil Hole Angular Position deg 25

Crankshaft

Material - Steel

Table 1 Engine Parameters

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