T R A N S P O R T A N D ROAD RESEARCH L A B O R A T O R Y Department of Transport

RESEARCH REPORT 10

D Y N A M O M E T E R TESTS OF THE EFFICIENCY OF A V A N

T R A N S M I S S I O N SYSTEM

by T W I L L I A M S , J R A M S H A W and I C P S I M M O N S

Any views expressed in this Report are not necessarily those of the Department of Transport

Vehicle Engineering Division Vehicles and Systems Assessment Department Transport and Road Research Laboratory Crowthorne, Berkshire, RG11 6AU 1985

ISSN 0266-5247

Ownership of the Transport Research Laboratory was transferred from the Department of Transport to a subsidiary of the Transport Research Foundation on I st April 1996.

This report has been reproduced by permission of the Controller of HMSO. Extracts from the text may be reproduced, except for commercial purposes, provided the source is acknowledged.

CONTENTS

Page

Abstract 1

1. Introduction 1

2. Test method 2

3. Instrumentation 4

4. Axle tests 5

4.1 The test axle 5

4.2 Calibration 5

4.3 Test oils 7

4.4 Test procedure--effects of viscosity 7

4.5 Test procedure--effects of friction modifiers (in a synthetic oil) on churning losses 8

4.6 Track tests 8

5. Tests of two gearboxes 8

5.1 The test gearboxes 8

5.2 The test rig 9

5.3 Calibration and analysis of measurements 9

5.4 Test procedure 10

6. Results 11

6.1 Dynamometer tests of the back axle-- conventional oils 11

6.2 Dynamometer tests of the back axle-- synthetic oils 11

6.3 Track tests of the effects of different axle oils 12

6.4 Dynamometer tests of two gearboxes 13

7. Discussion 15

7.1 Axle lubricants 15

8. Conclusions 16

9. Acknowledgements 16

10. References 16

11. Appendix 17

11.1 Explanation of some technical terms used 17

© CROWN COPYRIGHT 1985 Extracts from the text may be reproduced, except for

commercial purposes, provided the source is acknowledged

D Y N A M O M E T E R TESTS OF THE EFF IC IENCY OF A V A N T R A N S M I S S I O N S Y S T E M

ABSTRACT

Measurements have been made of some of the parameters affecting energy losses in back axles and gearboxes. The work was mostly carried out on a dynamometer but a limited number of track tests were used to corroborate the test rig results. The axle and gearboxes tested were standard Ford Transit units. There was a large change in axle efficiency with oil temperature, efficiency improving with increasing temperature. Under low temperature, high viscosity conditions, power loss in the axle increased approximately linearly with rotational speed and efficiency improved as transmitted power increased. However even under high power, low viscosity conditions the axle absorbed at least 6 per cent of the power transmitted and under low power, high viscosity conditions the power loss was as high as 40 per cent of transmitted power. Gearbox efficiency was generally above 90 per cent, however tests with a Ford Transit van on the test track showed that efficiency changes are significant in terms of fuel consumption.

1 INTRODUCTION Light commercial vans use about 10 per cent of all the fuel burnt by road vehicles in Britain. For such vehicles

about one third of the fuel used is converted to mechanical power, with the other two thirds being rejected as heat from the radiator, exhaust and engine. The mechanical power is required to overcome the aerodynamic drag of the vehicle, the rolling resistance of the tyres, hub bearing losses and transmission losses. A computer simulation model has been developed to study the effects of improving component efficiencies on the overall fuel consumption of commercial vehicles and it has become apparent that insufficient information is available on the losses in van transmissions. Work by Smith (1970) has shown that these losses are about 10 per cent of transmitted power at full load, and proportionately larger at low power inputs. Hobson (1979) has looked in detail at the effects of lubricants on the eff iciency of axles.

This report describes the measurement of the power losses in the gearbox and back axle of a Ford Transit van. Most of the work was done on a dynamometer, but limited tests of the back axle were also carried out on a vehicle on the TRRL test track. The results which confirm that transmission losses are particularly substantial at part power and low temperature, are for use in a computer simulation of the fuel consumption of light vans.

In addition a limited number of tests were carried out to assess the efficiency of a Ford J type gearbox. This

Belt

=_. I

Slave axle

Auxiliary shaft -~-_

Gearbox ~ and E" )

dynamometer "_

Test axle b"

Input

torque Engine and gearbox

iiil"~llllllllllllUlllll I IIIllllllllllllllllll I~

Belt Output torque

Fig. 1 Layout of axle test rig

size of gearbox is f itted to the Ford Transit light goods vehicle, so these tests complement those of the Transit back axle.

2 TEST M E T H O D

The test method for the axle was selected to suit the facil it ies in the existing dynamometer test house. Various methods were considered such as the four- square balanced beam (see Appendix), heat rejection and measurement of input and output torque and

speed. The procedure selected was that of measuring input and output torque over a range of rotational speeds, and careful measurements were needed as the torque loss being measured was the small difference between the large input and output torques (Figure 1). A one per cent error in the measured input and output torques would lead to a 20 per cent loss being read as anything between 18 and 22 per cent. The instrumentation layout for the measurement of axle efficiency is illustrated in Figure 2.

The test rig dynamometer was incapable of absorbing the power at the speed of the axle hubs so a slave

7 N2

~.~ / 7 ~

" ' - - - ' I ~,,. ,,,,,j~==j ~ " I 3 . T1

~ 4 3 2 1

! I I I I I I i I I

N1 / / ~ I P"-,,.

[ 1 , " , r - ~ ~ I I

\ I

1 Gearbox (and engine, not shown)

2 Universal coupling(s)

3 Slip ring(s)

4 Drive shaft (with strain gauges)

5 60 too th wheel and speed transducer

6 Drive coupl ing (and cal ibrat ion arm temporary mount ing)

7 Hub (and cal ibrat ion arm temporary mount ing)

8 Hal f -shaf t (both strain gauged)

9 Thermocouple (oil temperature)

10 Test axle

N 1 = R.H. hub speed

N 2 = L.H. hub speed

N 3 = Drive shaft speed

T 1 = R.H. hub output torque

T 2 = L.H. hub output torque

T 3 = Input torque

N 1 = N 2 = N 3 Axle ratio

Fig. 2 Axle efficiency measurement

2

Plate 1 Axle test rig

Neg. no. B703/80

Neg. no. B882/82

P la te 2 T r a n s d u c e r o u t p u t s a n d d r i v e f o r s l ave a x l e

axle was interposed between the test axle and the dynamometer driven by toothed belts from hubs attached to the wheel locations (Figure 1 and Plates 1 and 2). This system allows power to be absorbed at the input shaft of the slave axle within the operating speed range of the dynamometer. The belt drives to the slave axle were also linked by a common shaft which ensured that the two hubs of the test axle were turning at the same speed, as the differential was not required to simulate cornering motions.

The mechanical efficiency of the axle was calculated as (referring to Figure 2):--

Power output = 27rN1T1 +27rN2T2 = Power input 2~rN3T3

Where N1, N2, N3 = hub and input drive shaft speeds ml, m2, T3 = hub and input drive shaft torques

T~ +"1"2 = total output torque

and 2N3 - axle ra t io=R N~+N2

T1 + T2 If N1 and N2 are the same, efficiency (77)-~

The test method for the gearbox consisted of comparing the torque reaction of the gearbox with either the input or output torque. This makes possible the measurement of small losses in efficiency and was preferred to the more usual way of measurement of torque and speed at the input and output shafts. This latter method although acceptable for the measurement of back axle efficiency losses may give large errors in the estimation of efficiency when the difference between input and output torque is small, as for instance in direct ratio at a low power input.

An annotated diagram of the test installation .is given in Figure 3, while Plate 3 shows a view of the gearbox mounting arrangement. The gearbox was separated from its normal attachment to the engine bellhousing and mounted between bearings in a way which allowed it to partially rotate about the engine crankshaft axis.

Plan view

A drive-shaft was fitted between the engine and the gearbox. Movement of the gearbox was restricted by an arm from the gearbox, the end of which was attached to the test bed through a strain-gauged proof ring. This was used to measure the torque reaction on the gearbox due to the applied input torque and gear reduction (if in use).

The output torque of the gearbox was measured using the same strain-gauged drive-shaft and instrumentation system as in the axle tests (see Figure 2 and Plate 1). An additional display was provided for the output from the strain-gauged proof ring. A sheathed thermocouple was fitted to the gearbox sump to measure oil temperature.

Track" tests of the axle were carried out using a Ford Transit 115 van fitted with a Transflo fuel consumption meter, on the TRRL track; the vehicle was run from a cold start at constant speed to determine the time taken for the axle lubricant to reach a stable temperature. This speed was maintained while the vehicle was driven to complete six successive laps of the 7.5 km reversed figure of eight test cycle (Williams, Simmons and Jacklin, 1981). Fuel consumption was noted at the completion of each lap. At the end of the test laps the oil was drained while still warm then the axle was flushed with the next test oil before finally refilling with the new oil. The oil containing a friction modifier was used last to avoid any problem in cleaning the differential surfaces; the friction modifiers in some oils can be tenacious in clinging to the metallic surfaces. Ambient temperature was also recorded.

3 INSTRUMENTATION For the tests of the back axle, the instrumentation was designed to measure the input and output torques, and the drive-shaft and half-shaft speeds, to an

Side view

Drive to Strain gauged Slip Support Test axle propshaft ring bearing gearbox

J

~ etreZoPceruap~r e "~"~"~ 60 Tooth wheel and - ~ proximity transducer

Support Drive from bearing engine

Strain gauged proof ring

Fig. 3 Test gearbox installation

4

Plate 3 Gearbox

accuracy of 1 per cent. The torque in each of the shafts was measured using conventional full bridge strain-gauges (see Figure 2). Initially slip rings were used to transmit the hub output signals and a telemetry system used for the higher speed rotating drive-shaft. Later, to improve accuracy, the telemetry system was changed to a slip ring device. The signals were amplified and displayed on digital panel meters wi thout converting them to engineering units. This simplified the electronic system and avoided errors due to signal conditioning (Figure 4). The half-shaft and drive-shaft speeds were measured using toothed wheels (60 teeth) and proximity detectors. The outputs from the transducers were fed to frequency counter modules which provided a direct readout of revolutions per minute. During the tests the signals were monitored on a desk-top console (Plate 4) which had the facility to hold displayed values while readings were taken.

The slip-rings chosen for this work were of very high quality and produced an electrical noise level of less than 15 #V/mA. The amplifiers were based upon the National LH0038 true instrumentation amplifier which offered a very low offset drift of 0.25 #V/°C and an equally low noise figure. This device included a guard drive signal which helped to reduce noise pick-up. The displays consisted of modular digital panel meters for voltage measurements and frequency counting. The strain-gauged shafts were designed to provide an output at least one thousand times larger than the noise generated by the instrumentation. The axle oil temperature was measured using a sheathed copper- constantan thermocouple inserted through a gland in the normal filler plug and in contact with the oil.

efficiency rig

4 A X L E T E S T S

Neg. no. B632/83

4.1 THE TEST AXLE The test axle was of the type f i t ted as standard to the Ford Transit single wheel van series. The axle was secondhand wi th a differential and bearings in good condit ion; the mileage run wi th the axle was unknown. It had a ratio of 4.625 to 1 wi th a hypoid crown wheel and pinion wi th a four pinion differential and was of three-quarter f loat ing type (see Appendix) .

The amount of tooth contact wi l l affect the boundary fr ict ion of a lubricant but for this first series of tests no adjustments were made to the crown wheel and pinion backlash, or to the differential bearing pre-load.

There were several problems that arose during the preliminary test phase. The outputs from sets of hub- mounted strain gauges (not shown in Figure 2), designed as an alternative means of measuring output torque, suffered from f luctuat ions caused by the normal eccentr ici ty of the wheel hubs. This was improved by machining the hubs and incorporat ing an eight gauge bridge to compensate for small bending effects. These f luctuat ions were not apparent in the outputs from the gauges on the axle half-shafts, so in subsequent tests the half-shafts were used to measure the torque outputs.

4.2 CALIBRATION The drive belts were removed and the drive-shaft uncoupled. One output shaft and the axle input f lange

5

MEASURED PARAMETER

METHOD

Half-shaft speed

Toothed wheel with p rox im i ty detector

OUTPUT DISPLAY

Modular digital panel

meter

Prop-shaft speed

Toothed wheel with p rox im i ty detector

P Modular

digital panel meter

Oil temperature

Sheathed Cu/Con. thermal- couple

Strain-gauge power supply

( lOV)

Analogue meter

L.H. Half-shaft torque

Strain-gauge bridge

R .H. Half-shaft torque

Strain-gauge bridge

Prop-shaft torque

Strain-gauge bridge

Modular digital panel

meter

Modular digital panel

meter

Modular digital panel

meter

Fig. 4 Axle efficiency, instrumentation system

were locked. A lever was clamped at right angles to the other output shaft to allow torque to be applied to it, and weights hung on the lever via a load hanger and knife edge (see Figure 2)° The electrical signals from the output shaft strain-gauges were calibrated against the applied torque. The second output shaft

and input drive-shaft were similarly calibrated. Several Ioadings were applied and the calibration calculated for the loading and unloading curves. Figure 5 shows the calibration curves for the drive-shaft, half-shaft and gearbox; the calibrations are linear and without hysteresis.

! .

/

41'

Plate 4 Control console

Neg. no. B878/82

The shaft speeds were checked against a cal ibrated digital meter and a comparison made from the axle ratio. The oil temperature thermocouple was calibrated against a mercury thermometer before f i t t ing.

4.3 TEST OILS All tests in the initial phase were with a standard SAE 85W-140 oil as used at TRRL, formulated to API GL5* quality. A series of oils were tested at a later stage and specifications are as shown in Table 1.

test oil being used as the f lushing oil pr ior to the actual test; the axle was f lushed and run tw ice before any measurements were taken.

The forego ing axle tests w i th the dynamomete r rig were fo l lowed by a short series of track tests to a t tempt to measure the ef fect of a f r ic t ion modi f ier in axle oil on the fuel consumpt ion of a van. These tests concerned an SAE 90EP oil, and a SAE 90EP oil plus fr ict ion modif ier , wh ich were compared wi th the SAE 85W-140 base oil.

4.4 TEST P R O C E D U R E - - E F F E C T S OF V I S C O S I T Y

During the preliminary tests it was evident that oil temperature and hence oil viscosity were important parameters when making measurements of axle efficiency. The test house temperature, which was uncontrol led, varied from 0 to 10°C. The axle temperature increased rapidly under loaded running condit ions, so a cool ing fan was used to control the temperature. The input torque to the axle was set at pre-determined values and the output torques and axle oil temperature measured for a range of speeds. Small differences in temperature were al lowed for by interpolat ion of the results. The first test oil was a SAE 85W-140 as used in TRRL vehicles. Subsequent ly the series of oils listed in Table 1 were tested, each

TABLE 1

Speci f icat ions of au tomot i ve gear oils tested

Kinemat ic v iscosi ty (cSt) Viscosi ty*

Grade 40°C 100°C Index

SAE 75W 32.18 5.46 105 SAE 80W 86.97 10.25 99 SAE 80W-90 130.4 14.19 107 SAE 90 190.6 16.78 92 SAE 85W- 140 312.5 24.4 99 SAE 140 516.8 30.5 86

Formulated to Amer ican Petro leum Ins t i tu te - -Gear Lubricant 5 (API-GL5) speci f icat ion *See Append ix

4.5 TEST PROCEDURE--EFFECTS OF FRICTION MODIFIERS (IN A SYNTHETIC OIL) ON CHURNING LOSSES

Because of the claims made for the fuel saving properties of synthetic oils which include materials to reduce fr ict ion, a short comparative test of a synthetic and a conventional oil of the same viscosity (SAE 80W-90) was carried out using the axle test rig described in Section 2. The conventional oil was a high qual i ty petroleum based oil to specification API-GL5; the synthetic oil was diester based with a specification exceeding API-GL5. Properties of both oils are summarised in Table 2. It was decided to concentrate on measuring the churning losses in the differential under no-load conditions to determine whether the synthetic oil gave the same results as the conventional oil. A fuller understanding of the effects of the synthetic oil could only be obtained from a more comprehensive series of tests than was possible on this occasion.

TABLE 2 Technical properties of the test oils

Synthetic Petroleum Oil type (Diester base) based

SAE classification 80W-90 80W-90 Density (15.5/15.5°C) 0.93 0.905 API gravi ty 21.2 -- Flash point °C 210 141 Viscosity 100°C (cSt.) 14.25 15.4 Viscosity 40°C (cSt.) 110.0 152.0 Viscosity - 18°C (cPs) 9,200 -- Viscosity index 125 103 Pour point °C - 40 - 30 Channel point - 51 -- Copper strip l b -- Pentone ° insolubles Nil - - Sulphur % 2.1 -- Phosphorus % 0.14 -- Nitrogen % 0.05 --

The drive belts connecting the output shafts of the test axle to the slave axle were removed, so that the power required to drive the axle was that absorbed in churning plus a small amount lost due to tooth and bearing fr ict ion. The axle sump was flushed thoroughly using a flushing oil, then wi th the test oil and finally drained and refilled with the correct quanti ty of the t e s t oil. Torque input and drive-shaft speed was measured using the same instrumentation shown in Figure 2. Power input was calculated for a range of drive-shaft speeds at oil temperatures from the minimum possible (5 to 10°C) to 70°C.

4.6 TRACK TESTS The van was warmed up and run over the 7.5 km cycle on the TRRL test track at a steady 72 km/h and six

readings taken of the fuel consumed using the SAE 85W-140 base oil in the axle. The oil was drained and the axle refilled with the SAE 90EP grade and the vehicle warmed up. The oil was drained a second time and the axle refilled with SAE 90EP and the test procedure repeated. The same test procedure was followed for the SAE 90EP plus friction modifier and finally for the SAE 85W-140 base oil, as the base oil tests were repeated to provide a check on the initial test results and on the consistency of the test procedure.

5 T E S T S OF T W O G E A R B O X E S

5.1 THE TEST GEARBOXES A limited number of tests to assess the efficiency of a Ford J type gearbox were carried out. This size of gearbox is fitted to the Ford Transit light goods vehicle, so these tests complement those of the Transit back axle. Two gearboxes of the same type were tested. The first unit was a secondhand Ford J type gearbox which had been removed from a relatively low mileage vehicle. The test results from this part-worn gearbox were compared with the results of similar tests on a new unit.

Both gearboxes had four forward ratios of 4.412:1, 2.353:1, 1.505:1, 1.000:1 and one reverse ratio. The gearboxes were of the constant mesh type having helically cut gears with synchromesh engagement on all forward ratios. The gear arrangement and drive train

2.0

1.5

o v 1.0

O

0.5

N Half-shaft calibration

Drive shaft calibrations /

Gearbox / ~ /

I I I I I 10 20 30 40 50

Load {kgf)

Fig. 5 Drive shaft, half shaft and gearbox p r o o f ring cal ibrat ions

60

3rd and 4th (top) gear synchroniser

Mainshaft

Countershaft

Main drive

~ Drive

4th (top)

Mainshaft bearing

3rd gear Main drive bearing

~ Drive

]

Countershaft gear

3rd

1st and 2nd gear sy nch r o n i s e r

1st gear

r

2nd gear

.

rive

2nd

1st and 2nd gear synchroniser

3rd and 4th (top) gear synchroniser

Main drive I gear

rive

Neutral

Fig. 6 Gearbox arrangement and drive train

is shown in Figure 6 (reverse gear is not shown for clarity). The main drive gear and mainshaft are mounted on ball bearings and the countershaft and mainshaft spigot on needle rollers. When driven in neutral, the main drive gear and countershaft gear revolve and first, second and third gears revolve on the mainshaft. Gears are engaged by moving the particular gear synchroniser along the splined mainshaft to engage dog teeth on the selected gear. Fourth ratio (top) is a direct drive, the main drive gear being locked to the mainshaft by the forward movement of the third and fourth gear synchroniser.

5.2 THE TEST RIG The test rig which was described in Section 2, enabled the torque reaction of the gearbox to be compared with either the input or output torque. The gearbox

was allowed to rotate about the input and output shafts and was restrained by a strain-gauged proof ring which was used to measure its torque reaction.

5.3 CALIBRATION A N D ANALYSIS OF MEASUREMENTS

The proof ring was calibrated by attaching a horizontal arm carrying a load hanger to the gearbox and progressively increasing the load on the hanger to apply torque to the gearbox. The calibration is shown in Figure 5; the strain-gauged drive-shaft was calibrated in a similar manner.

The sketch below shows the torques acting on a

gearbox:

9

Output

T2

/ T3

T1

TI A kJ

Input

Sketch to show torques acting on a gearbox when driven in direct ratio

]'3 is the torque reaction of the gearbox through the Ioadcell mounted on the arm and is the net effect of two opposing moments on the gearbox, one created by internal fr iction, churning etc and acting in the direction shown, the second due to gear reduction and opposite handed. In direct gear the torque due to churning etc is small and proportional to the losses in the gearbox, and the torque due to gear reduction is zero, so T3 acts in the direction shown. In an indirect gear the torque due to the gear reduction is comparatively large and the torque due to churning etc is relatively small and opposite in hand. Thus, referring to the sketch, in an indirect ratio T3 usually has a negative value.

For equil ibrium, T1 = ('1"2 + ]'3) and, by definit ion, T1NtT/ = T2N2

where N~ = input speed N2 = output speed

and r/ = gearbox efficiency

_ T2N2 T2 s o , , 1 -

where the gearbox N~ reduction ratio r N2

But T1 = (T2+T3) hence 7/ -1"2

-- r(T2 + T3)

and gearbox efficiency can be determined by measurement of the output torque and the torque on the gearbox case.

5.4 TEST P R O C E D U R E In measuring gearbox efficiency, the output torque was set at predetermined values by adjusting the power absorbed by the dynamometer. Gearbox reaction (in terms of output from the strain-gauged proof ring) and temperature were measured for a range of speeds. An external fan was used to control oil temperature which rose due to power dissipation in the gearbox. Small differences in the test temperature from the nominal value were allowed for by interpolation of the results. Churning, pumping and residual bearing and frictional losses in the gearbox under no-load conditions were measured by noting the torque reaction in the gearbox with the output shaft disconnected. Churning should

account for the majority of the power absorption by the gearbox, with no output load.

The test oil was a conventional one of SAE 85W-140 grade in normal use in the gearbox; the viscosity of the oil was 345 cSt at 40°C and 25 cSt at 100°C. No other oils were tested.

Before any test work was carried out with the new gearbox it was run-in for approximately 20 hours at moderate speeds and loads.

No load - - . - - 1 lkW input power . . . . 20 kW input power

Oil-SAE 85W-140

v

O

O o_

/ /

/ A

Oil temperature ( ° C ) 4 ~ 0 , / ~ / ' ~

80 ~ I l I

0 1000 2000 3000 Drive-shaft speed (revs/min)

Fig. 7 Power loss (kW) with speed and oil temp temperature of an axle (0 ,11 and 20kW) input power)

10

6 RESULTS

6.1 D Y N A M O M E T E R TESTS OF THE BACK A X L E - - C O N V E N T I O N A L OILS

Figure 7 shows the effect of the input drive-shaft speed and oil temperature on the power consumption of the Ford Transit back axle filled with an SAE 85W-140 oil, under no-load conditions and with input powers of 11 kW and 20 kW. At the no-load and 11 kW input power conditions there was an almost linear increase with speed of the power dissipated for any given temperature, suggesting that power lost was due to oil churning and pumping, increasing proportionately with speed. With no output loading on the axle there was a large change in power loss due to temperature change. For example, at 2,800 rpm the power loss dropped from 2.7 kW at 40°C to approximately 1 kW at 80°C; the change was not quite so significant at 11 kW or at 20 kW except at speeds around 3,000 rpm.

Plotting the effect of temperature and power on efficiency for the SAE 85W-140 oil at one speed, 2,400 rpm, again showed the dependence of power loss on temperature (Figure 8). At the higher temperatures, such as 100°C, the differences in efficiency were small for a range of different input powers, but at 40°C there was a large drop in efficiency as the no-load power consumption assumed a large proportion of the lower power inputs. Figure 9 shows the effect of three measured variables (power, temperature and speed), on the efficiency of the axle for the SAE 85W-140 oil. Again the lowest efficiencies occur at the lowest temperatures, highest speeds and lowest power inputs.

Changing the oil to a SAE 75W grade gave a worthwhile improvement in efficiency. This is shown

100

90 E

t~

E

._U

.., 8 0 -

70 0

OiI-SAE 85W-140 Input power (kW) 23

1 4 11

J J Input speed 2400 rpm

- - - [

4O

Fig. 8

l l I 60 80 100 120

Oil temperature (°C)

Effect of temperature and power on efficiency of an axle

A

¢Z

LU

1 oo

90

80

70

0

100

90

80

70

6o 0

1 oo

Input speed 2800 (rpm) ~ --

2 4 0 0

Oil-85W-140 Oil temperature - 90°C

I I 10 20 30

Input speed

1200" 1800 J 2800 _ ~ _ _ . ~

10 20 30

Oil temperature - 60°C

I

90

80

70

6O 0

1200 ~ 1800 2 / ~ 8 0 0 ~

,o, 2400///13200 Oil tem:erature - 40 C

10 20 30 Power (kW)

Fig. 9 Effect of oil temperature, power and speed on efficiency of an axle

in Figure 10 where the eff iciency of the axle with speed and oil temperature for the base oil and for SAE 75W is plotted. The improvement was greatest at the highest speed and lowest temperature. These results suggested that viscosity was largely responsible for the changes in efficiency with temperature for a particular oil.

To examine the effects of viscosity, the axle was tested with six different oils at an input power of 11 kW and the efficiency loss* plotted against a logarithmic plot of viscosity (Figure 11). A further curve on the figure shows the results for a higher power of 20 kW at 3,000 rpm; the efficiency improved with increased power but was still viscosity dependent.

6.2 D Y N A M O M E T E R TESTS OF THE BACK A X L E - - S Y N T H E T I C OILS

Figure 12 shows the churning losses under no-load conditions of the back axle filled with either a fully synthetic or a conventional petroleum base 80W-90

*Efficiency loss= (100--efficiency) (per cent)

11

40

I 11kW input power I - - = -- 20kW input power

A

c 30

C~

20 0

• (2- 10 iml

OiI-SAE 85W-140

Oil t e m p e r a t u r e ~

60 ~ . . . - ~ - ~ - ~ -~ - .~. ~ 80 4 0 f . - - - s . . . . ~ - - - / ~ - ~ .~ ""

/ 60 / 8O

I I 1000 2000 3000

40

I =

8 30

e~

o 20

u

u . l

0 0

OiI-SAE 75W

Oil temperature ( °C ) 4 0 ~ _ ~ - ~ ~ 50-----~ ~ ~ - - - - ' - " ~ " - - 1

60-80 ~ 4 0 ~ - -

50" / I i 60 /70 ; I

1000 2000 3000

Drive-shaft speed (revs/min)

Fig. 10 Loss in efficiency with speed and oil temperature for two oils and two input powers

4.0

3.0

v

~o 2.0 O .

1.0

Temp (°C)

40

100

Oil viscosity and key

Viscosity (centistokes)

Pet. base Synthetic 80W- 90 80W-90

152 110

15.4 14.25

Output load -- nil

Input speed (revs/min)

3 0 0 0 \

\

50

4 0

(3L

3 0 0

~, 20

LU

10

0 10

Oil 75W

11 kW input power

O ~ 20kW input power

Oil temperature 40°C

80W-90 85W-140 80W I 90 140

I I L Input (rpml

~ 3000 2600

2200

1800 3000

s ~ 1400

I I I 50 1 O0 500

Viscosity (centistokes)

Fig. 11 Variation of axle efficiency loss with oil viscosity

0 20 40 60

Oil temperature (°C)

Fig. 12 Churning loss in an axle of a synthetic SAE 80W-90 oil compared to a petroleum base SAE 80W-90 oil

gear oil, at drive-shaft speeds of 1,000 to 3,000 rpm, over the oil temperature range of 10 to 70°C. Churning loss increased with decrease in the oil temperature and increased with increasing drive-shaft speed. At any test condition in the speed and temperature range covered, the synthetic oil had churning losses which were lower than those using the petroleum base gear oil. For example, at an oil temperature of 20°C, churning losses with the synthetic oil over the speed range showed a saving of about 20 per cent over the petroleum base oil and at 40°C, around 35 per cent.

6.3 TRACK TESTS OF THE EFFECTS OF DIFFERENT AXLE OILS

Figure 13a shows the axle oil temperature measured on a Ford Transit 115 van driven around the TRRL

12

100

U

3

80

60

40

20

(3. Q . O O

E: E

E E 0 0

,,, < <

Start

0 I I I I I I I I I I I 0 20 40 60 80 100

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Fig. 13a Van axle oil temperature during track test

track 7.5 km test cycle at a steady 72 km/h. This showed that from a cold start the oil warmed up to a temperature of just over 40°C in about half an hour. Ten minutes after the run finished the temperature had increased and after 50 minutes the temperature had dropped to about 20°C. These results were obtained with an ambient air temperature of 16 to 18°C. At a lower ambient temperature of around 7°C the oil temperature reached 39°C under similar test conditions. Trying to relate the effects of fuel efficient oils on fuel savings in a moving vehicle therefore requires an appreciation of the likely effect of ambient temperature to compare with the effects observed on dynamometer tests (Dobson, 1981).

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Figure 13b shows the results of a track test comparison between a van whose axle was filled successively wi th three different oils, a SAE 85W-140, a SAE 90EP and a SAE 90EP with fr ict ion modifier. The test was run at the 15 to 20 kW power range at axle input speeds approaching 3,000 rpm. The results showed an improvement of around 1.5 per cent due to the friction modifier, and improvements wi th the SAE 90EP and SAE 90EP plus modifier over the SAE 85W-140 base oil of around 2.5 and 4 per cent respectively.

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Vehicle -- Ford Transit Cycle length-- 7.5km Speed -- 72km/h

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Fig. 13b Track tests to compare fuel used with different axle oils

6.4 D Y N A M O M E T E R TESTS OF TWO GEARBOXES

The two Ford J type gearboxes tested, one new and one secondhand, gave rather dif ferent results. These are shown as representative of equipment that might be tested by manufacturers during development and of equipment in service.

Figure 14 shows the power loss in the new gearbox running at zero output power in the three forward ratios and neutral. These losses are mainly due to churning increasing wi th gearbox speed and wi th reducing oil temperature (or increasing viscosity).

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115 "

1000 2000 3000 1000 2000 3000 Engine speed (revs/min) Engine speed (revs/min)

Fig. 14 Power losses in 3 gear ratios and neutral of a new Ford J type gearbox at 3 oil temperatures (output load -- nil)

Figure 15 shows the eff iciency of the new and secondhand boxes in 2nd, 3rd and direct (4th) gears at an oil temperature of 50°C. One effect w i th in the secondhand gearbox caused eff ic iency to fall w i th transmitted power, whi le another caused it to rise. In fourth ratio, the gearbox showed expected trends; eff iciency increased wi th increase in input power or wi th decrease in input (engine) speed. The secondhand gearbox showed levels of ef f ic iency approaching 100 per cent wh ich are usual in this ratio; the lower efficiencies of the new gearbox were attr ibuted to the need for a longer running- in period to relieve manufactur ing tolerances. Tests in third ratio in the secondhand gearbox demonstrated the existence of a possible fault in the gearbox wh ich caused eff ic iency to fall as input power was increased and the effect of

13

SECONDHAND GEARBOX NEW GEARBOX

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Gearbox input power (kW)

100

Input speed (rprr I)

1200 1800 2700

95

90 -

4th ratio (direct)

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Gearbox input power (kW)

Ef f ic iency o f new and secondhand gearboxes in 3 ratios (SAE 85W - 140 oi l at 50 °)

30

14

this fault was seen more clearly in second ratio, whereas in both third and second ratios the efficiency of the new gearbox generally showed the normal trends.

Figure 16 shows that, on the three forward gears tested, the efficiency of the secondhand box improved as the oil temperature rose (the same was found for the new gearbox). The difference between the new and secondhand box is emphasised by Figure 17 which shows the efficiency of the boxes as a function of output shaft speed and input power.

1 0 0

A 98

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~ 9 4 w

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9kW input power at 1800 revs/min 13.5kW input power at 2400 revs/min

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Fig. 16 Effect of oil temperature, input power and speed on efficiency of a secondhand Ford J type gearbox

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I 10kW inpu t power J

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Fig. 17 Efficiency of new and secondhand gearboxes with output shaft speed, for 2 input powers and 3 gear ratios (Ford J type gearbox) (SAE 85W-140 Oil at 50 C)

7 D I S C U S S I O N

7.1 AXLE L U B R I C A N T S It is well known that friction losses in gearboxes decrease with increasing temperature (Hobson 1979; Oliver, Reuter and Sendra 1981). At low temperatures, hydrodynamic lubrication effects predominate and the friction decreases with increasing temperature because of decreasing oil viscosity. This assumes that moving surfaces are not in contact, and the friction losses are due to the viscosity of the lubricant. Under boundary lubrication conditions, when surface asperities begin to make contact, wear will take place and at some stage an increasing oil temperature wil l lead to increasing friction and wear. It is important to maintain protection against wear as it is possible that some of the lower viscosity oils are deficient in extreme- pressure protection. There is a trade-off between fuel economy effects and axle durabil i ty and this is a topic to which the manufacturers of oils have devoted a considerable amount of effort. Various friction modifiers have been evolved which do reduce friction losses. Difficulties arise when attempts are made to

15

assess a relationship between oil viscosity and fuel savings (Dobson 1981). When a high viscosity oil (eg SAE 140) is used in a vehicle axle, and the vehicle is left overnight at a low temperature and then used on a short journey, the axle ineff iciency wil l contribute to a relatively high fuel consumption. In such a case it might be worth considering the use of a suitable lower viscosity axle and gearbox oil, bearing in mind the need to maintain adequate wear protection. When starting a vehicle from cold, until each component reaches working temperature there wil l be increased energy losses in the engine, in the remainder of the transmission system and in the tyres. An American study (Blackmore and Thomas 1977), using cars, suggested that around 20 km of travel (at an ambient temperature of 21°C) was required to attain 90 per cent of the ful ly warmed-up fuel economy; on a 1.7 km journey f rom a cold start, a Ford Escort car used twice as much fuel as in the warmed-up state and on a 3.4 km trip, approximately 60 per cent more (Waters and Laker 1980).

Dynamometer tests of an axle indicated that when a conventional petroleum based oil was replaced by a synthetic oil the churning losses in the axle were reduced by between 15 to 35 per cent, when the axle oil temperature varied between 20 and 40°C in average British ambient temperatures. This would represent a saving in fuel consumption of perhaps 3 or 4 per cent and an additional saving should be possible if a synthetic oil is used to replace a conventional gearbox oil, although no attempt was made during the study reported here to check the durabil i ty of the synthetic oil or the degree of protection from wear that it provided.

Corroborative tests with a Ford Transit van on the TRRL test track gave similar results to the dynamometer tests showing that efficiency changes are significant in terms of fuel consumption. The use of lower viscosity oil and oil with friction modifier in the back axle gave fuel savings of 2 to 4 per cent at a steady 72 km/h.

Two gearboxes of the same model were tested, one new and the other secondhand. They produced differing results which demonstrate the efficiencies of gearboxes before and after lengthy service. In fourth (direct drive) ratio, both gearboxes produced the expected trends in that the efficiency increased with increase in input power or oil temperature or with a decrease in input speed. Lower levels of efficiency in the new gearbox in this ratio suggested a need for further running-in to relieve manufacturing tolerances. There was an anomalous effect in the secondhand gearbox in second and third ratios, in that efficiency fell with increase in power and this demonstrated that there might be a fault in the gearbox due to wear. Churning losses at zero input power in the gearbox(es) increased with input speed and with reducing oil temperature.

At low engine power and when both gearbox and axle oil viscosities were high, as when running.a vehicle from a cold start, the combined effect noted from the reported test results would be a markedly low transmission efficiency. When this effect was combined with a low engine efficiency and high tyre rolling resistance, this would go some way to explain why vehicle fuel consumption should be so high on short trips in cold weather.

8 C O N C L U S I O N S Dynamometer tests have been made of the power loss in a Ford Transit back axle and gearbox. In addition, a small number of track tests have been made to measure the effect of axle eff iciency on fuel consumption.

There was a large change in axle efficiency wi th axle oil temperature, mainly due to change in viscosity of the oil wi th temperature; under given conditions, efficiency improved with reducing viscosity.

Under low temperature, high viscosity conditions power loss in the axle increased approximately linearly with rotational speed. This loss was probably due to oil churning and pumping. Because the power loss was relatively independent of the transmitted power, the eff iciency improved as the transmitted power increased.

Even under high power, low viscosity conditions the axle absorbed at least 6 per cent of the power transmitted; under low power, high viscosity conditions the loss could be as high as 40 per cent of the transmitted power.

9 ACKNOWLEDGEMENTS The work described in this report was carried out in the Vehicle Engineering Division of the Vehicles and Systems Assessment Department of TRRL.

The authors wish to acknowledge the assistance of Mr D J Jacklin and Mr A Mumford with the test work.

10 REFERENCES BLACKMORE D R AND THOMAS A, (1977), Fuel economy of the gasoline engine; fuel, lubricant and other effects, The Macmillan Press Ltd., London

DOBSON G R (1981), The prediction of fuel efficiency of engine oils, Co-ordinating European Council International Symposium, Rome.

HOBSON D E (1979), Axle efficiency; test procedures and results, SAE Paper No. 790744, Society of Automotive Engineers.

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INSTITUTE OF PETROLEUM (1976), IP Standards for Petroleum and its Products, Pt. 1, Methods for analysis and testing, Section 1, IP Method 166, Applied Science Publishers Ltd., London, 35th. ed.

OLIVER C R, REUTER R M AND SENDRA J C (1981), Fuel efficient gasoline engine oils, Lubrication, Texaco Services Europe, Vol 67, No 1.

SMITH G L (1970), Commercial vehicle performance and fuel economy, SAE Paper No. 700194, Society of Automotive Engineers.

WATERS M H L AND LAKER I B (1980), Research on fuel conservation for cars, TRRL Report 921 Transport and Road Research Laboratory, Crowthorne, Berks.

WILLIAMS T, SIMMONS I C P AND JACKLIN D J (1981), Fuel consumption testing of heavy goods vehicles, TRRL Supplementary Report 687, Transport and Road Research Laboratory, Crowthorne, Berks.

11 APPENDIX

11.1 EXPLANATION OF SOME TECHNICAL TERMS USED

Axle, three-quarter floating

Axle hub bearings can be of three types: semi-floating, three-quarter floating and fully floating. The three- quarter floating type uses only one bearing at each hub end of the axle. The bearing is mounted on the outside of the axle casing and is secured by a lock nut. The outer part of the bearing carries a cup which has a studded exterior flange to receive the half-shaft flange (and brake drum flange in normal service).

On the TRRL axle test rig, the half-shafts were subjected to a bending load due to the tension present in the belt drive to the slave axle but the effect was relatively small and did not affect the accuracy of measurement of half-shaft torsion using the strain- gauge bridge on the shaft.

The four-square balanced beam (a type of axle efficiency test rig)

In the four-square design of axle efficiency test rig, power is transmitted in a closed loop, the initial power source being external to the loop. The power may be transmitted through a variable speed drive and Morse roller chain or an alternative arrangement. The test rig is composed of two axles face to face and connected by shafts to form a closed transmission loop which forms a square seen in plan view, hence the design name.

A propeller shaft drives a test axle which has its differential locked so that power can be taken from one hub only and transferred into a drive-shaft parallel

to the driven propeller shaft. This forms the third side of the square, the fourth side being composed of a slave axle driven through one hub (this differential also being locked) which is connected to the propeller shaft. The power input is made at this point but a feature of the design is that the power required is only that necessary to overcome losses in the loop.

The test axle is mounted in a rectangular frame arranged to pivot about the axis of the propeller shaft and normally maintained in a level position by initial adjustment of a balance weight with reference to a spirit level on the frame.

Although the power output from the test axle is taken through one half-shaft only, both half-shafts are mounted in bearings f i t ted at the ends of the frame so that the axle is free to rotate about the half-shaft axis. A lever having an" adjustable balance weight is fitted at the rear of the frame at a fixed distance from the propeller shaft axis and extending in the same plane. The torque reaction of the axle to that of the input torque through the propeller shaft is balanced by the addition of suitable weights to the lever arm. Because the frame is balanced initially and the lever arm position is f ixed and known, input torque can be calculated. In a similar way, output torque from the axle is counter-balanced by adjustment of the position of the previously applied weights on the lever arm and may be calculated.

Degrees of torsion can be locked into the power loop by jacking the back of the slave differential by a lever and screw arrangement.

Hydrodynamic and boundary lubrication

Hydrodynamic lubrication is the condition which exists when two surfaces are moving in relation to one another, yet are completely separated by a f i lm of oil. The only friction developed under this circumstance is due to the viscosity of the oil as it resists the surface movement and produces internal friction in the oil film. Additives in the oil under these circumstances have no effect, except by altering the viscosity.

If the oil f i lm becomes thin enough, irregularities on each surface will make contact and cause fr ict ion and surface wear. This situation is known as boundary lubrication and additives may be used in the oil to reduce the friction and wear.

Factors influencing whether either hydrodynamic or boundary lubrication is present are the sliding speed of the surfaces, the oil f i lm viscosity and the pressure between the surfaces tending to squeeze the oil film. An increase in the sliding speed or the oil viscosity or a decrease in the surface pressure will help to establish hydrodynamic lubrication.

Gear lubricant 5 (GL 5)

This lubricant designation is part of a classification scheme of the American Petroleum Institute (API) intended to describe the service for which the lubricant

17

is suitable. Thus API -GL 5 refers to service for gear lubrication (particularly hypoid gears) in passenger cars and other vehicles which may be operated under high speed, shock load; high speed, low torque; and low speed, high torque conditions.

Ex t r eme p ressure (EP) lubricants

These are lubricants formulated to possess increased load carrying ability over and above that of a straight mineral oil. The term extreme pressure lubricant is not intended to convey an indication of how great the load carrying ability is; a classification system defines the load carrying capacity which may be rated from very low to very high. Classification can be achieved by a standard labratory test ( INSTITUTE OF PETROLEUM, 1976), such as the load carrying capacity test for oils

using the IAE gear machine where the performance of test oils under load in test gears may be examined and compared with reference oils.

Viscosity index

An arbitrary number indicating the effect of change of temperature on the kinematic viscosity (a measure of the resistance to gravity flow) of a petroleum product or oil. Viscosity index is calculated by comparing the viscosity/temperature characteristics of the fluid in question against two reference oils having the same viscosity as the test fluid at 100°C. A minimum change of kinematic viscosity with temperature is desirable; the higher the viscosity index the smaller will be the change in viscosity with change in temperature.

Printed in the UK for HMSO (1679) Dd8222661 11185 HP Ltd So'ton G426

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