Development of Air/Oil-Cooled Motorcycle Engine Using Thermal and Fluid Analyses

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SAE TECHNICALPAPER SERIES 2007-01-0538

Development of Air/Oil-Cooled MotorcycleEngine Using Thermal and Fluid Analyses

Yoshitsugu Gokan and Yasushi TakahashiHonda R&D Co., Ltd. Motorcycle R&D Center

Makoto InayoshiDesign Network Co., Ltd.

Tsuneaki Ishima and Tomio ObokataGunma University

Reprinted From: Thermal Systems & Management Systems(SP-2132)

2007 World CongressDetroit, MichiganApril 16-19, 2007

Downloaded from SAE International by Automotive Research Association of India, Friday, August 01, 2014



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2007-01-0538

Development of Air/Oil-Cooled Motorcycle EngineUsing Thermal and Fluid Analyses

Honda R&D Co., Ltd. Motorcycle R&D Center

Design Network Co., Ltd.

Gunma University

Copyright © 2007 SAE International

ABSTRACT

To achieve power output and cooling performance inmotorcycle air-cooled engines equivalent to those inwater-cooled engines, an engine utilizing air-flowgenerated by the moving motorcycle and a newoil-cooling system for an air-cooled engine was studied.The engine temperature distribution was obtainedbased on a CFD (Computational Fluid Dynamics)analysis of cooling-air/oil behavior using an inlinefour-cylinder 900cm 3 engine. As a result of this study,engine temperature was sufficiently decreased and thedifference in temperature among cylinders wasreduced; the engine was tested in a prototypemotorcycle.

INTRODUCTION

Although the demands of motorcycle customers haveseen much diversification in recent years, many ofthem still want motorcycles equipped with a simpleair-cooled engine. Harley-Davidson, BMW, DUCATI,and MOTO GUZZI all prefer to use air-cooled enginesfor their machines, from smaller to larger models, andthey are still offering innovative design. Somemotorcycles manufactured by Suzuki have engineswith an air-cooling system, and the company alsoutilizes an oil-cooled engine.

The most significant issue with air-cooled engines isthat they have no cooling mechanism other than theair-flow generated by the moving motorcycle itself.

Along with controlling total heat discharge, correctingthe difference in temperature among cylinders inmulti-cylinder engines is important. In addition, topursue cooling performance deeply means seeking forthe functional beauty of air-cooled engines; this is alsoimportant from a design point of view. Add to thesefactors the growing importance of dealing with exhaustemissions and fuel economy for motorcycles, andcontrolling engine temperature becomes a veryimportant challenge.

Formerly, in aircraft reciprocating air-cooled engines, asingle radial engine or a double row radial engine (1)

with radially-disposed cylinders around a crankshaftwas employed. This is because at high cruisingspeeds, the wind velocity at the engine cooling-fin wasoverwhelmingly high and sufficient coolingperformance could be maintained.

In automobiles, both Porsche and VolksWagenemployed air-cooled engines. However, since theengines need a large amount of lubrication oil ascooling-oil, it can be said that these engines utilized anoil-cooling system as well. Similarly, in the late 1960sand early 1970s Honda sold N360, N600, and H1300models that were equipped with air-cooled two orfour-cylinder engines. However, since an air-cooled

engine doesn't have a cooling-water jacket,mechanical noise and combustion noise generated bythe engine is directly emitted outside. This leads toinferior performance in terms of noise when comparedto water-cooled engines.

Conventionally, development of an air-cooled enginerelies on the visualization of cooling air-flow in awind-tunnel or on temperature measurement resultsobtained in actual driving tests. The quantitativeanalysis of air-flow behavior and temperatureestimation in the early design stages has been difficult.Recently, attempts to examine cooling air-flowbehavior using CFD have been reported in some

papers(2), (3)

. We reported some effective findingsrelated to fin shapes, oil cooler locations, and theshape of the inlet duct in a previous paper (4). Otherstudies (5), (6) of an oil-cooling system aimed atimproving the thermal efficiency of an engine werepublished.

This paper describes a method for calculating thetemperature distribution in an engine (cylinder headand cylinder barrel) with CFD, based on its heattransfer coefficients. Using the method, we examinedthe temperature of an inline four-cylinder 900 cm 3

engine with an oil-cooling system. Furthermore, some

Yoshitsugu Gokan and Yasushi Takahashi

Makoto Inayoshi

Tsuneaki Ishima and Tomio Ob okata




effects of the oil-cooling system were tested andconfirmed using a prototype motorcycle.

OUTLINE OF ENGINE LAYOUT

ENGINE DEVELOPMENT GOALS AND MEANS FOR ACHIEVING THEM

In the early design stages of a new air-cooled enginedesign with high performance, the established goal

was to achieve cooling performance comparable towater-cooled engines in terms of power outputperformance, fuel economy, noise reduction, durability,etc. When designing a multi-cylinder engine, reducingthe temperature difference among cylinders is alsoimportant in terms of improving durability andstabilizing quality. Figure 1 shows the concretetechnological means used to pursue this goal. Theyinclude improvements such as pulling heat from theheat source in the cylinder head, effective use of thecooling-air, cooling-oil around the combustion chamber,and FI (Fuel Injection) system for corresponding to riseof engine temperature.

To achieve the development goal, CFD analysis todetermine the effective utilization of cooling-air wasfirst considered, along with more efficient cooling-finshapes and oil cooler locations, as discussed in ourprevious paper (4).

The oil-cooling system shown in the next section(Figure 2) was employed to achieve a coolingperformance comparable to water-cooled engines. Toeffectively decrease the temperature, a predictivecalculation of the temperature of engine parts wasmade through CFD analysis, and effective findings of,for example, the cooling-oil passage structure wasexamined. To reduce the difference in temperature

among cylinders, the amount of cooling-oil flow wasadjusted by setting orifices in the oil passage of thecooling system. For fuel economy and exhaustemissions, a temperature sensor was installed in thecooling-oil passage, and correction of fuel injectionquantities was controlled by the FI system. In addition,an oil thermostat was installed before the oil cooler toimprove warm-up performance at cold temperatures.

3. Oil-cooling aroundcombustion chamber

Temperature distribution at same levels as in water-cooledenginesReduce temperature difference among cylinders

Development goals

5. Fuel Injection (FI) systemfor corresponding to rise ofengine temperature

1. Effective useof cooling-air

2. Improvement of oilcooler performance

4. Improvement ofpulling heat from heatsource in cylinder head



Development goals


Development goals


1. Effective useof cooling-air

2. Improvement of oilcooler performance

4. Improvement ofpulling heat from heatsource in cylinder head

Figure 1: Development goals and means to achievecooling performance

In addition, the length of the exhaust port was reducedto minimize the cylinder head heat source.

These examinations were attempted to bring the CFDanalysis into line with actual vehicle experiments.

OIL-COOLING SYSTEM

Figure 2 (a) shows an oil passage diagram for theoil-cooling system. As shown in the figure, oil supplied

from an oil pump is divided into two passages. Oneforms a lubrication oil passage (the green line in Figure2), and the other is an oil passage exclusively forcooling (red line in Figure 2). An oil jacket for coolingwas placed near the cylinder head combustionchamber. With an oil-cooling system, stagnation of oilflow causes excessively high oil temperatures at thepart, and degradation of the oil is a concern. Inaddition, too large surface area inside the oil passageincreases heat receiving and cooling-oil temperatures.Therefore, the shape of the oil cooling jacket wasdesigned differently from that of a water coolingsystem, which covers the entire combustion chamber.

A passage was formed between the intake and

exhaust ports, as shown in Figure 2(b), to cool a heatspot intensively while avoiding stagnation of thecooling-oil. The shape was also considered in terms ofpotential production or casting problems.

Oil line for cooling

Oil line for lubrication

Oil jacket aroundcombustion chamber

Oil thermo sensor forFuel Injection control

Thermostat valve fortemperature control

Oil pump

Oil line for cooling

Oil line for lubrication

Oil jacket aroundcombustion chamber

Oil thermo sensor forFuel Injection control

Thermostat valve fortemperature control

Oil pump

(a) Oil passage diagram for the oil-cooling system

(b) An oil jacket around the combustion chamber

Figure 2: Diagram of the oil-cooling system




CALCULATION METHOD AND PROCEDURE

CFD ANALYSIS ITEMS

Figure 3 shows the relationship between the designstudy items and CFD analysis items for anair/oil-cooled engine in the early design stage.

To effectively utilize the air-flow generated by themoving motorcycle, visualization of the cooling air flow,

cooling-fin shape, oil cooler location, etc. were treatedas CFD analysis items. In a wind-tunnel test, only thevisualization of air-flow is possible. On the other hand,since CFD analysis aims to study cooling air flowbehavior under actual driving conditions and calculatethe local heat transfer coefficients, calculation of thetemperature distribution in the engine main body andheat strain analysis became possible.

Study items forair-cooled engines CFD analysis items

1. Effective use of coolingair

2. Improvement of oil coolerperformance


4. Improvement of pullingheat from heat source incylinder head


Vibration control rubber

shape between cooling fins

Air-flow channel shape incylinder head

Shape of exhaust port

Cooling fin shape of engine

Oil cooler location

Cooling fin shape in air-flowchannel

Oil Jacket shape in cylinderhead

Pathway of cooling oil

Study items forair-cooled engines CFD analysis items

1. Effective use of coolingair

2. Improvement of oil coolerperformance


4. Improvement of pullingheat from heat source incylinder head


Vibration control rubber

shape between cooling fins

Air-flow channel shape incylinder head

Shape of exhaust port

Cooling fin shape of engine

Oil cooler location

Cooling fin shape in air-flowchannel

Oil Jacket shape in cylinderhead

Pathway of cooling oil

Figure 3: Relationship between an air/oil-coolingengine design and CFD analysis

THE PARTIAL CELLS IN CARTESIAN COORDINATEMETHOD (7)

From the viewpoint of the design process, a key pointin CFD analysis is mesh generation.

Especially, when analyzing a whole motorcycle, meshgeneration is a big issue. The Partial Cells in Cartesiancoordinate (PCC) method was developed tosubstantially reduce the mesh generation processwhen a flow field around complicated objects isanalyzed. Figure 4 (7) shows the concept of input datafor the PCC method and the general Boundary FittedCoordinate (BFC) method. With BFC method, acomputation mesh is required in accordance withshape data, so a mesh must be generated each time ashape is changed. On the other hand, with the PCCmethod, shape data and mesh data are setindependent of each other, so regeneration of themesh is not required every time the shape changes.

Object shape(solid model)

Cartesiancoordinate

mesh

Input data

CFD code with PartialCells in Cartesiancoordinate method

Object shape

Meshgeneration

Input data

CFD codewith BFC


Cartesiancoordinate

mesh

Input data



Cartesiancoordinate

mesh

Input data


Object shape

Meshgeneration

Input data

CFD codewith BFC

Object shape

Meshgeneration

Input data

CFD codewith BFC

Figure 4: Comparison of data flow between the PCC

and BFC methods

When the surface of the object shape does notcoincide with the grid, a partial fluid cell which is apart of a hexahedron cell that has been clipped by thewall may be generated near the wall. This is called"a partial cell". Partial cells can take several shapes,as shown in Figure 5. To make calculations inaccordance with the shape of any partial cell,discretization equations were introduced based on theFinite Volume Method (FVM). In addition, since thecomputation mesh that goes along the boundary wasnot available, an original near-wall turbulence model (8)

was employed to consider turbulent flow behavior in

the low Reynolds number area near the wall whileusing a k model and wall function. A divided cell, asshown in Figure 5, was not taken into account. Thisdecision is based on the definition "state quantity in acell is indicated by a single value". When such a caseappears, the mesh near the cell is further divided as acountermeasure.

fluid

wall

Partial cell

Wall

Hexahedron cell

fluid

wall

fluid

wall

Partial cell

Wall

Hexahedron cell

Partial cell

Wall

Hexahedron cell

Partial cell

Wall

Hexahedron cell

Figure 5: Examples of partial cells and the exclusion ofpartial cells divided by the wall

Verification of this calculation method was conductedfor several kinds of applications, for example withpipes (9), around a rectangular cube (10) , with perforatedplates (11) , and in the intake ports of engines (8), (12), (13) .

CALCULATION PROCEDURE

Figure 6 shows the calculation procedure. With anumerical wind tunnel, shape data is created by3D-CAD systems (Figure 6 (a)) such as CATIA, andthen written into a file in STL (Stereo Lithography) dataformat, which is commonly used in the CAM field. Inthis format, the surface of a solid model is expressedas an aggregation of triangles; it is a standard dataformat among almost all CAD systems. However, the




For boundary conditions of the heat conductivecalculation, the heat transfer coefficients of cooling-airflow and cooling-oil flow were obtained through CFD. Itis also necessary to consider the cooling effect bylubrication oil and heat flux in the combustion chamber,cylinder wall, and inner wall of the intake/exhaust ports,but these values are not definite. The objective was toexamine the cooling performance of the air/oil-cooledengine, especially the effect of the oil jacket. Therefore,actual measurement values were substituted for

temperature at the bottom surface of the cylinder block(the surface contacting the crankcase); heat flux at thecombustion chamber, cylinder wall, and inner wall atthe exhaust ports was assumed based on reference (16) .Then the values were adjusted in comparison totemperature measurement data from a test engine sothat they roughly coincided with each other.

Since the calculation grid for heat conductivecalculation is the same with CFD, it is relatively rough.On the other hand, since some parts of the cylinderhead have a minimum thickness of a few millimeters, itis often necessary to set two heat transfer coefficients,for example cooling-air and cooling-oil in a cell asshown in Figure 7. In addition, a contact with acombustion chamber requires boundary conditions ofheat flux. To make calculation possible even in suchcases, shape data (STL data) for heat conductivecalculation are divided into groups depending onboundary conditions. In this way, it is possible to setmultiple boundary conditions in one calculation cell.

combustion chamber

intakeport

exhaust port

air flow channel

oil jacket

combustion chamber

intakeport

exhaust port

air flow channel

oil jacketoil jacket

Figure 7: Section of a computation grid for heatconductive calculation

HEAT TRANSFER COEFFICIENTS OF OIL JACKET

To calculate the heat receiving value at the oil jacket,the heat transfer coefficients of cooling-oil were

calculated through CFD. An example of the calculationresult is shown in Figure 8. The calculation grid wasset to correspond to the oil jacket size. Since both inletand exhaust widths between ports are inevitablynarrower, the flow rate increases and the heat transfercoefficient rises.

Unlike water, the viscosity of oil varies substantiallydepending on its temperature (T oil). Because of this,the heat transfer coefficient ( ) also varies greatlydepending on the oil temperature. Therefore, therelations between amount of oil flow, oil temperature,and heat transfer coefficients were obtained. Figure 9shows the result. The heat transfer coefficients in the

figure are average values of the area above the headgasket surface of the oil jacket. From the figure, it canbe seen that heat transfer coefficients at 160 °C are 70- 80% higher than at 120 °C.

A pedestal formanufacturing

Narrower shapes ateach port

A pedestal formanufacturing

Narrower shapes ateach port

Figure 8: Heat transfer coefficients of oil jacket

0

1

2

3

0 1

Amount of oil flow (per one cylinder)

H e a

t t r a n s

f e r c o e

f f i c i e n

t ( )

10 3 W/m 2K

10 -4 m3/s

#1 Cylinder head(CFD analysis result)

2

Toil=160Toil=140Toil=120

t2

t1

0

1

3

0 1 2


H e a

t t r a n s

f e r c o e

f f i c i e n

t ( )

10 3 W/m 2K

10 -4 m3/s

#1 Cylinder head(CFD analysis result)

2

Toil=160Toil=160Toil=140Toil=140Toil=120Toil=120

t2

t1

Figure 9: Relationship between oil flow amount andheat transfer coefficients

With water cooling, when water temperature rises, thedifference in temperature ( T) between the coolingliquid and the inner wall of the jacket gets smaller, socooling performance decreases. However, with oilcooling, the increase of (heat transfer coefficients) islarger than the decrease of T when the oiltemperature rises, so the increase of oil temperature

actually results in greater cooling performance. Forexample, assuming that the temperature of the innerwall of the oil jacket is 250 °C, when oil temperaturerises from 120 to 160 °C, T increases by 0.7, from130 to 90 °C; however, increases 1.7 times, and

T becomes 1.19 times larger.

1301202501 t T , 901602502 t T




7.013090

1

212

t

t t t T

T T

from Figure 9:

07.11 t , 83.12 t

7.107.1

83.1

1

212

t

t t t

19.17.17.01212 t t t t T

where Tt1 is T at T oil=120 °C, Tt2 is T atToil=160 °C, t1 is at T oil=120 °C, and t2 is atToil=160 °C.

Therefore, it was preferable to have a simple oil jacketshape to improve the velocity of the fluid flow, unlikethe water-cooled jacket shape which entirely coversthe combustion chamber. Moreover, the surface areain the inner wall of the oil jacket was reduced as much

as possible to prevent raising the oil temperature.

HEAT CALCULATION RESULTS ANDDISCUSSION

COMPARISON WITH SURFACE TEMPERATUREMEASUREMENT

Figure 10 compares the heat conductive calculationresults for the engine and the experimental resultsfrom the vehicle dynometer test (obtained with aninfrared thermometer [thermo-viewer]). For calculationand measurement conditions, the air-flow velocity andengine load were equivalent to 80km/h. In calculation,the heat transfer coefficients were obtained usingequation (2). The heat conductive calculation wasmade based on the assumption that cooling airtemperature was a constant 25 °C.

As seen in the calculation results, the temperature ofthe cooling-fin at the exhaust side was lower than theintake side. However, the cooling-fin temperature inthe experimental results was approximately the sameat the exhaust and intake sides. It can be said that thiswas because in the actual engine, air temperaturearound the exhaust pipe was high, and so T (thetemperature difference between the air and fin)became small.

In addition, with the actual engine, it was recognizedthat the fin tip had been cooled especially well due tohigher air-flow velocity. However, the calculation resultdid not show this because the size of the calculationgrid was larger than the actual shape, and temperaturein the cell was averaged.

Therefore, though there are some points that shouldbe improved, these are reasonable calculation resultsfor temperature distribution in an engine.

Experimental

result

Calculationresult

Experimental

result

Calculationresult

0

130

Figure 10: Comparison of temperature (at 80 km/h)

THE EFFECT OF THE OIL-COOLING SYSTEM

To examine the effect of the oil-cooling, heatcalculations with/without the oil-cooling were made.The calculation conditions simulated driving at 230km/h.

Figure 11 shows the calculation results without oil-cooling by assuming that the heat transfer coefficientsof the oil jacket were zero. The figure shows thatthough the fin tip was cooled, temperature of themiddle wall of the exhaust port of the middle cylinderswas high. Figure 12 (a) shows the cylinder headtemperature around ignition-plug seat (T plug ). Thedifference in T plug among the cylinders was about40 °C; T plug of the middle cylinder was slightly below300 °C.

Figure 11: Temperature distribution without oil-cooling




T e m p e r a

t u r e ,

#4 #3 #2 #1

4070(a)

(d)

(b)

(c)

(a) Estimation results without oil-cooling(b) Estimation results with oil-cooling(c) Measurement results

(d) Estimation results with oil-cooling(oil flow quantity in cylinders # 1 and #4is 1/4 that in cylinders #2 and #3)

Cylinders

T e m p e r a

t u r e ,

#4 #3 #2 #1

4070(a)

(d)

(b)

(c)

4070(a)

(d)

(b)

(c)





Cylinders

(a) perspective view

(b) bottom view

(c) top view

Figure 13: Temperature distribution with oil -cooling 160 °C

(a) perspective view

(b) bottom view

Figure 14: The distribution of the temperaturedifference reduction among cylinders

(Toil=160 °C)

Figure 12: Calculation of T plug (at 230 km/h)

Next, in the calculations to verify the effect ofoil-cooling, it was assumed that the heat transfercoefficient of the cooling-oil jacket was 160 °C, and theoil flow rate was 1x10 -4 m 3/s per cylinder. Figures 13(a),(b), and (c) show temperature distributions withoil-cooling. The temperature was generally low,

confirming the effect of oil-cooling. It was expected thattemperatures could be reduced by approximately70 °C at T plug of the middle cylinders, which had thehighest temperature, as shown in Figure 12 (b).

Here, a comparison with the actual measurements wasmade. The actual measurement values are shown inFigure 12 (c). Though they are not absolutely accurate,Tplug at the outer and inner side cylinders was almostthe same in the calculations and in the experimentalresults.

The examination became possible by the calculationfor reducing the difference in temperature among

cylinders.

A detailed examination of Figure 13 (a), confirms thatin every place temperature decreased greatlycompared with Figure 11. However, the difference intemperature among cylinders was still not completelyeliminated. This was because sufficient cooling-air wasnot flowing at the front center of the engine, which wasobstructed by the front wheel, front forks (suspension),or headlight.

IMPROVEMENT OF DIFFERENCE INTEMPERATURE AMONG CYLINDERS

Figure 13 (b) shows the temperature distribution of theengine viewed from underneath. The temperature ofthe combustion chamber on the exhaust side washigher than on the intake side. Especially, it was clearthat the temperature between the two divergedexhaust ports was higher. A comparison of thecylinders revealed that the outer cylinders were cooledwell, however, irregularities in temperature on thecylinder periphery were larger at the outer cylinder.

To reduce the difference in temperature amongcylinders, the amount of cooling-oil flow per cylinderwas adjusted and several types of calculation weremade. Accordingly, orifices were used to decrease the




cooling-oil flow into the outer side cylinders toapproximately one-fourth; these calculation results areshown in Figures 14 (a), (b), and Figure 12 (d). Theyindicate that the difference in temperature amongcylinders can be reduced.

CONFIRMATION OF THE OIL-COOLINGSYSTEM EFFECT IN THE ACTUAL VEHICLETEST

DECREASE OF CYLINDER HEAD TEMPERATURE AROUND IGNITION-PLUG SEAT

The cylinder head temperature around the ignition-plugseat (T plug ) was used as an indicator to evaluate thecooling performance of the engine.

Figure 15 shows the relationship between thecooling-oil flow rate and T plug with the oil-coolingsystem. According to the graph, the optimal value forcooling-oil flow is 1x10 -4 m 3/s to each cylinder at anengine speed of 7000 rpm.

As a result, it became possible to minimize the

discharging volume of oil pump from the viewpoint ofdecreasing friction.

Ne = 7 000 r/minWOTToil = 130


C y l i n

d e r

h e a

d t e m p e r a

t u r e

a r o u n

d i g n

i t i o n - p

l u g s e a

t ( T

p l u g

)

0 110 -4 m3/s

2

50

Ne = 7 000 r/minWOTToil = 130


C y l i n

d e r

h e a

d t e m p e r a

t u r e

a r o u n

d i g n

i t i o n - p

l u g s e a

t ( T

p l u g

)

0 1 210 -4 m3/s

50

Figure 15: Optimization of oil flow amount (per onecylinder)

Figures 16 (a) and (b) show the effect of decreasingTplug by employing the oil-cooling system.

Figure 16 (a) shows the correlation between indicatedmean effective pressure (IMEP) and T plug . The plots

shown by solid circles in the figure are T plug values forour typical mass-produced models. The square orangecolored plot is the result for this air/oil-cooled engine.In a water-cooled engine, compared to an air-cooledengine, the area around the combustion chamber of iscovered with coolant (cooling water). Therefore, thetemperature of the inner wall of the combustionchamber is as low as the temperature around theignition-plug seat. On the other hand, in an air-cooledengine, the temperature tends to rise dramatically inproportion to IMEP.

Accordingly, we were able to decrease temperature(Tplug ) by approximately 70 °C in an air/oil-cooledengine that generates a similar IMEP.

Figure 16 (b) shows the correlation between theamount of heat dissipation (kJ) and T plug . This figurewas derived from Figure 16 (a). As shown in Figure 16(a), a decrease in temperature (T plug ) of approximately70 °C has been realized. Each plot in the figurerepresents the temperature of our typical mass

production models. Qt (kJ) was calculated by thefollowing equation (3).

321 t t t t QQQQ (3)

where Q t1 is heat dissipation of combustive heat, Q t2 isheat dissipation of radiant heat, and Q t3 is heatdissipation of frictional heat.

IMEP (MPa)

Air-cooled engine domain

only air-cooling

with oil-cooling system

Water-cooled engine domain

T 70

C y l i n

d e r

h e a

d t e m p e r a t u r e

a r o u n

d i g n

i t i o n - p

l u g s e a t

( T p

l u g

)

1.2 1.3 1.4 1.5 1.6 1.7

50

IMEP (MPa)

Air-cooled engine domain

only air-cooling

with oil-cooling system C y l i n

d e r

h e a

d t e m p e r a t u r e

a r o u n

d i g n

i t i o n - p

l u g s e a t

( T p

l u g

)


T 70

50

1.2 1.3 1.4 1.5 1.6 1.7

(a) IMEP vs. T plug

Amount of heat dissipation (kJ)

C y l i n d e r

h e a

d t e m p e r a

t u r e

a r o u n

d i g n

i t i o n - p

l u g s e a

t ( T

p l u g

) Air-cooled engine domain only air-cooling



T 70

50

0 0.4 0.8 1.2 1.810 4

Amount of heat dissipation (kJ)

C y l i n d e r

h e a

d t e m p e r a

t u r e

a r o u n

d i g n

i t i o n - p

l u g s e a

t ( T

p l u g

) Air-cooled engine domain only air-cooling



T 70

50

0 0.4 0.8 1.2 1.810 4

(b) Amount of heat dissipation vs. T plug

Figure 16: Effect of oil-cooling




IMPROVEMENT OF KNOCKING LIMITS

Figure 17 shows the improvement of knocking limitswhen the oil-cooling system is employed in anair-cooled engine. At engine speeds of 2000 rpm andabove, the ignition timing at which knocking occurswas improved by more than +5 degrees. It is supposedthat abnormal heat spots in the combustion chamberwere prevented, enabling a uniform temperature overthe surface of the combustion chamber wall.

Furthermore, after the durability test, the effects ondeflection and degradation of the valve seat wereimproved, and blow-by gas emissions were alsoimproved.

0

10

20

30

40

50

0 2 000 4 000 6 000 8 000 10 000

Engine speed (r/min) I g n

i t i o n

t i m i n g o

f k n o c k

i n g g e n e r a

t i o n

( d e g r e e

)

Measuring cylinder ; #1(WOT)

only air-cooling


0

10

20

30

40

50

0 2 000 4 000 6 000 8 000 10 000

Engine speed (r/min) I g n

i t i o n

t i m i n g o

f k n o c k

i n g g e n e r a

t i o n

( d e g r e e

)

Measuring cylinder ; #1(WOT)

only air-cooling


Figure 17: Effect of oil-cooling on engine knockinglimits

CONCLUSIONS

1) Heat conductive calculations using heat transfercoefficients obtained by CFD analysis of the cooling-airmade it possible to visualize engine temperaturedistribution.

2) The effect of an oil-cooling system in an air-cooledengine was examined utilizing the above method. As aresult, it was seen that a sufficient cooling effect couldbe expected. Furthermore, it was found that it wasfeasible to improve the temperature balance amongcylinders by controlling cooling-oil quantity.

3) Tests with a prototype vehicle confirmed that

cylinder head temperature around the ignition-plugseat was equivalent to that in a water-cooled engine.Furthermore, it was shown that the oil-cooling systemimproved knocking resistance.

As discussed above, employing the oil-cooling systemmade it possible to design an air-cooled engine thathas potential output performance and environmentalfriendliness (fuel economy and exhaust emissions)equivalent to water-cooled engines.

ACKNOWLEDGMENTS

We would like to express our appreciation to thepeople concerned at Honda R&D Sun Co., Ltd. fortheir support in creating 3D solid models.

REFERENCES

1. Gunston, B., World Encyclopedia of Aero Engines,4th edition, Patrick Stephens Ltd., 1998.

2. Kiura, T., Kushibiki, T., Shiomi, K., “Research onPrediction of Cooling Air Flow in Small Air-CooledUtility Engines”, Small Engine TechnologyConference & Exhibition, SAE paper2005-32-0023, 2005.

3. Garg, M., Kumar, O. H., Prasad, N., Swamy, V.,“Experimental and CFD Simulation-Based

Analytical Optimization of Air-Cooling System fora Small 4-Stroke Scooter Engine”, Small EngineTechnology Conference & Exhibition, SAE paper2005-32-0026, 2005.

4. Takahashi, Y., Gokan, Y., "CFD Analysis of AirFlow of Air-cooled Motorcycle Engines" SmallEngine Technology Conference & Exhibition atNovember 13-16, 06SETC-25, 2006

5. Kiryu, M., “Development of Oil-cooled 750ccMotorcycle Engine” (in Japanese), Journal ofSociety of Automotive Engineers of Japan , Vol.40,No.9, p.1154-1158, 1986.

6. Danno, Y., Kamada, H., Kitada, T., “Effect andFeasibility of Cylinder Block Oil Cooling forPassenger Car Engine” (in Japanese), Journal ofSociety of Automotive Engineers of Japan , Vol.42,No.8, p.1058-1064, 1988.

7. Takahashi, Y., Fukuzawa, K., Fujii, I., "Numerical

simulation of flow in intake ports and cylinder ofmulti-valve S.I. engine using PCC method",Proceedings of International Symposium onCOMODIA 94 , p.529-534, 1994.

8. Takahashi, Y., Nomura, T., Ishima, T., Obokata,T., "Numerical analysis on steady statein-cylinder-flow by Partial Cells in CartesianCoordinate method and verification with PIVmeasurement" (in Japanese), Transactions of theJapan Society of Mechanical Engineers , Series B,Vol.71, No.706, p.1694-1701, 2005.

9. Takahashi, Y., Gokan, Y., Inayoshi, M., Ishima, T.,Obokata, T., “Numerical prediction of heat transfercoefficient by Karman's analogy and anapplication for air-cooled motorcycle's engines” (inJapanese), Transactions of the Japan Society ofMechanical Engineers , Series B, Vol.72, No.724,2006.

10. Takahashi, Y., Nomura, T., Ishima, T., Obokata,T., "Analysis on flow around a rectangular cube bythe numerical simulation of PCC method with LDAverification (in the case of inclined cube)" (inJapanese), Transactions of the Japan Society ofMechanical Engineers , Series B, Vol.68, No.666,p.331-337, 2002.

11. Nomura, T., Takahashi, Y., Ishima, T., Obokata,T., "Analysis of flow over the perforated plate in




the duct by means of LDA and numericalsimulation" (in Japanese), Transactions of theJapan Society of Mechanical Engineers , Series B,Vol.70, No.691, p.693-700, 2004.

12. Nomura, T., Takahashi, Y., Takeda, C., Ishima, T.,Obokata, T., "LDA and PIV analyses of in-cylindersteady flow and verifications of numericalsimulation results with Partial Cells method" (inJapanese), Transactions of the Japan Society ofMechanical Engineers , Series B, Vol.70, No.694,

p.1618-1625, 2004.13. Takahashi, Y., Murakami, Y., "Establishment of an

intake port design support tool using CFD" (inJapanese), Honda R&D Technical Review , Vol.15,No.1, p.103-108, 2003.

14. 3DataExpert http://www.deskartes.com

2006/1/1615. Tachibana, F., Akiyama, M., Okamoto, Y.,

Morishita, T., Fundamentals of Heat Transmission,(in Japanese), Colona, Tokyo, 1965, p.44.

16. Hoshi, M., "Heat transfer and heat fluidmechanism for application to water / air cooledsystem" (in Japanese), Seminar text of Engineheat engineering and cooling technology,2002/6/13-14.

NOMENCLATURE

q Heat transfer quantity per unit area, W/m 2

Tw Temperature on the wall, °C

Tf Gas temperature defined at a pointsufficiently far from the wall, °C

Heat transfer coefficient, W/m 2K

Density, kg/m 3

Cp Specific heat at constant pressure, J/ (kg K)

w Wall shear stress, kPa

P r Prandtl number

Us Mean velocity at outer edge of boundarylayer, m/s

U2 Mean velocity at the boundary betweenlogarithmic layer and transition layer, m/s

T Difference between cooling liquidtemperature and temperature of inner wall of

jacket, °C

Tt1 T at 120 °C oil temperature, °C

Tt2 T at 160 °C oil temperature, °C

Tt2/t1

1

2

t

t

T T

t1 at 120 °C oil temperature, W/m 2K

t2 at 160 °C oil temperature, W/m 2K

t2/t1

1

2

i

t

IMEP Indicated mean effective pressure kPa

Q t Amount of heat dissipation, kJ

Q t1 Heat dissipation of combustive heat, kJ

Q t2 Heat dissipation of radiant heat, kJQ t3 Heat dissipation of frictional heat, kJ


Development of Air/Oil-Cooled Motorcycle Engine Using Thermal and Fluid Analyses

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