Turbine Throttle Loss Recovery using a Variable Geometrymate.tue.nl/mate/pdfs/11534.pdf ·...

ABSTRACTTwo of the most pressing challenges of the automotive sectorare reduction of fuel consumption and correspondingemission of greenhouse gases, especially when taking intoaccount the growing degree of luxury in modern passengercars, which increases the auxiliary load on the engine.Preferably, this increase in auxiliary load is compensated bythe recovery of waste energy. To accomplish this, atechnology called WEDACS (Waste Energy Driven AirConditioning System) is being developed to recover throttlinglosses. WEDACS uses a turbine to induce provide the enginewith the same air mass flow rate as a throttle valve whileproducing mechanical energy and cold air. An alternatorcoupled to this turbine converts mechanical energy intoelectrical energy and the cold air is used to cool A/C fluid.This way the load of both the engine mounted alternator andA/C compressor is reduced or eliminated, resulting in higherefficiency. A previous paper [ 1 ] provides a proof ofprinciple, using a turbine from a turbocharger, but alsodiscusses a challenge in the form of a limited operatingrange). The present paper focuses on addressing thischallenge. To expand the control range of the engine, aturbocharger with variable nozzle turbine is used. Due tolimitations in the variable nozzle mechanism, the range islimited to higher engine powers. It is shown that between 50W and 1.3 kW of energy can be recovered from a 2 literengine, depending on the operating point. A secondturbocharger's variable nozzle mechanism is adapted toenable control of a 2 liter engine from idle to about 50 %engine power. With decreasing engine power, the energyrecovery efficiency eventually drops to zero. The root causefor this is identified and an attempt is made to improveefficiency. Finally the drive cycle model from the previouspaper is expanded and a new drive cycle simulation shows afuel consumption improvement over the NEDC of about 5 to8 % for a mid-sized passenger car with a 2 liter engine.

INTRODUCTIONThe spark ignition engine is popular for its low cost andsimplicity, especially versions operating on stoichiometriccombustion, which enable removal of harmful emissions by alow-cost catalyst. In such engines, power is controlled by theamount of air let into the cylinders, which is usuallyaccomplished using a throttle valve. Associated throttlelosses, for a 2 liter engine, can be as high as 25 % ofdelivered engine power at 50 km/h [1]. Some solutions to thisproblem have been developed and brought to the market, likeBMW Valvetronic and Fiat Multiair, but these technologiesare integrated in the engine and cannot be retrofitted onexisting engines.

Another issue with modern passenger cars is the increasingauxiliary- and especially electrical load of vehicles. Even themost economical vehicles are equipped with A/C (airconditioning) and all kinds of electrical comfort enhancers.The alternator has to provide increasing electrical power anddoes this with low efficiency; a maximum of 60 % [2].

WEDACS (Waste Energy Driven A/C System) is a newsolution to these problems. A turbine throttles the air andrecovers part of the throttle losses. A high speed and highefficiency alternator coupled to the turbine converts therecovered energy to electricity. In addition, air cools downover the turbine as a result of expansion. This cool air cancool A/C fluid in a heat exchanger, therefore assisting the A/C pump. This way the engine has to produce less power andfuel is saved.

In a previous paper [1], throttling losses on a 2 liter Nissanengine were quantified and it was shown that at cruisingspeeds of 50, 80 and 120 km/h throttle losses were 1.2, 1.1and 1.25 kW respectively. A conventional turbocharger wasinstalled on this engine to prove the possibility of recoveringthrottle losses and investigate the possibility of engine control

Throttle Loss Recovery using a Variable GeometryTurbine

2010-01-1441Published

05/05/2010

R.H.L. Eichhorn, M.D. Boot and C.C.M. LuijtenTechnische Univ. Eindhoven

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by turbocharger speed. This kind of engine control did notwork properly, besides the turbocharger could only controlthe engine at high engine power. As a consequence, anoperating point of 155 km/h cruising speed was chosen where1.1 kW of throttle losses was recovered, which correspondsto a fuel efficiency improvement of 4 %. Finally a drive cyclewas simulated based on the experimental data. The result wasa fuel efficiency improvement of 15 to 19 % for a turbineefficiency of 40 to 60 %.

In this paper, first the operating range problem is addressed; aVGT (Variable Geometry Turbocharger/Turbine) is installedon the same 2 liter Nissan test engine to investigate the rangein which the engine can be controlled. Subsequently, anotherVGT mechanism is modified to extend the engine operatingrange down to idle and an attempt is made to improve turbineefficiency. Finally, based on the experiments, a drive cycle issimulated to investigate system performance. For thispurpose, the model from the previous paper is extended toimprove the accuracy of the results. An explanation is givenfor the difference between results in this paper and thosereported in the previous paper.

THEORYTURBINEBefore explaining the turbine theory, it is appropriate to treatthe energy that is available to the turbine. This is the energythat is usually lost in the throttle valve. This throttle loss canbe calculated with:

(1)

Where indices a refer to ambient conditions, or conditionsjust before the throttle valve, i refers to conditions in theintake manifold behind the throttle valve. For a moreelaborate explanation, see [1].

In Appendix A, it is explained how a turbine works and whatpossible implications the modifications made to the turbinehave. From the throttle losses available, the turbine willproduce a certain amount of power:

(2)

where Cp is the specific heat of air at constant pressure (1005J/kg.K), the numerical indices refer to stagnation properties atlocations in the turbine in figure 19 and Pcond is the powerabsorbed by the condensation of water vapor. Because thetemperatures are measured at a distance from the turbine inletand exit, flow velocities were low (in the order of 10 m/s) andthe influence of this speed on the measured temperatures can

be neglected. During experiments, turbine power can bedetermined by measuring the mass flow rate and temperaturesfrom equation 2. However, when low temperatures exist atthe outlet of the turbine (T5), condensation effects must beconsidered. When the temperature of air decreases, itscapacity to hold water vapor decreases too. This results in anincrease in relative humidity and finally in condensation ofwater vapor. The associated heat release causes the air to heatup, resulting in a higher temperature at the turbine exit. Whenthe relative humidity ψ (%) at the turbine inlet is measured,the specific humidity (kg water vapor / kg air) can becalculated using:

(3)

Where p is the total pressure and pg is the saturated vaporpressure. The number 0.622 is a result of the ratio of gasconstants from water vapor and air. At the turbine exit, bothtemperature and pressure are decreased and the specifichumidity is calculated back to relative humidity again. Themass of water vapor in excess of a specific humidity of 100%condenses. Per unit of mass, condensation of water vaporcauses a certain amount of energy to be released and the air atturbine exit heats up accordingly. For a more comprehensiveexplanation with examples, see [1].

After calculating turbine power, turbine efficiency can beobtained by dividing it by the amount of available throttlelosses from equation 1:

(4)

DRIVE CYCLE MODELIn order to predict a realistic efficiency gain of a vehicleequipped with WEDACS, it is desirable to simulate a drivecycle using a model. Turbine performance data is readilyavailable from the experiments. Integration in the vehicle,such as conversion of energy recovered by the turbine toelectricity and the corresponding efficiency gain will have tobe modeled. In the preceding paper [1], such a model wasalready presented. For a certain vehicle, engine operatingconditions were calculated from the road load. Theseconditions were used to find the fuel consumption in anengine fuel consumption map. Total fuel consumption overthe (NEDC) drive cycle was calculated using the followingequation:




(5)

Where mf,t is fuel consumption from the engine fuelconsumption map and the numbers above and below thesigma represent seconds of the drive cycle. The assumptionwas made that all energy recovered was saved at thecrankshaft and that the A/C was switched on. Because theengine would have to produce less energy, fuel consumptionwas reduced. Fuel consumption of a vehicle equipped withWEDACS was calculated using the ratio between recoveredpower and brake engine power:

(6)

Where the numerator is recovered power and the denominatoris brake engine power. For a more accurate estimate of fuelconsumption improvement, in addition to brake enginepower, some power losses should also be considered. Theselosses are pumping work, friction and auxiliaries and theelectrical load of the vehicle.

An SI engine friction model from Heywood [3] was used tocalculated pumping, friction and auxiliary power. Some typesof friction which are included in this model are friction fromthe piston rings, crankshaft bearings and valvetrain, pumpinglosses from the throttle valve and intake and exhaust valves,oil and coolant pump and alternator idle work.

In addition to friction, the engine also has to provide thevehicle with electricity to run the lights, entertainment systemand a growing number of actuators. These actuators varyfrom electrical windows to electrical seats and powersteering. In 2000, an average electrical power consumption of750 to 1000 W is claimed, depending on the vehicle and itsaccessories [4]. Therefore, an electrical power consumptionof 1500 W is chosen for the drive cycle simulation. Higher(2000 W) and lower (1000 W) average electrical loads arealso investigated.

To deliver this load, the engine has to drive the alternator,which has an efficiency that is dependent on rotational speed.In the preceding paper [1] this efficiency was assumed to beat its maximum value over the whole speed range. Here thespeed dependency will be taken into account, which can beseen in figure 1. An alternator speed of 1800 rpm correspondsto engine idle speed and 6000 rpm corresponds to cruisingspeed. Now power drain of the alternator from the engine is:

(7)

Where Palt is power removed from the engine by thealternator, ηalt is alternator efficiency and Pel is the electricalload of the vehicle. For each point in the drive cycle thisalternator power is calculated to torque and added to thetorque requested from the engine which is used to look upfuel consumption in equations 5 and 6.

Figure 1. Experimentally determined alternatorefficiency [2] and resulting alternator input power for an

electrical load of 1500 W.

Substituting these losses and power demand in equation 6yields:

(8)

In the denominator, the pumping, friction, auxiliary andalternator powers are now added to the brake engine power.In the numerator the recovered power is multiplied by theratio of new (WEDACS) alternator efficiency andconventional alternator efficiency. The conventionalalternator is a low speed AC type alternator that has a lowefficiency, see figure 1. The new alternator is a DCpermanent magnet or switched reluctance high speedmachine, which has a high efficiency [5]. Equation 5 and 6can be used to calculated fuel consumption over the drivecycle. In case of braking, it is assumed that the engine isidling, as is often the case in practical driving conditions.




To investigate the additional fuel efficiency improvementcaused by assistance of the A/C by cooling A/C fluid with thecold air, the A/C coefficient of performance (COPac) isintroduced to equation 8:

(9)

For a more elaborate explanation of this alternator and A/Cassist, see [1].

EXPERIMENTSTEST BENCHExperiments are conducted on a test bench, which wasoriginally used for the development of control strategies forCVT's. The powertrain consists of a SI 2 liter engine andCVT from a 2000 Nissan Primera, relevant properties of thisengine are listed in Appendix B. This powertrain drives aneddy current brake. Controllers adjust the brake power andthrottle valve position to reach a user specified engineoperating point. In the inlet of the engine, a turbocharger isinstalled in such a way that the inlet air is flowing through theturbine. This turbocharger has its own oil circuit. Thefollowing sensors are installed:

• 2x T-type thermocouple to measure temperature before andafter the turbine, T1 and T5, accuracy 3 %,

• 2x Pressure transducer to measure pressure before and afterthe turbine, p1 and p5, accuracy 2.25 %,

• The original Bosch HFM5 hot film sensor signal to theengine was measured and calibrated [1] to measure the airmass flow rate with an accuracy of 5 %,

• Turbocharger rotational speed was measured using anACAM Picoturn sensor, nt, accuracy is 800 rpm.

In figure 2, the engine inlet trajectory and location of thesensors are indicated. Relative humidity is measured withinthe test cell between series of measurements. For each datapoint, 60 seconds of measurement data is averaged.

Figure 2. Inlet trajectory of engine; F is air filter, C andT are compressor and turbine, I is engine inlet manifold.

OPERATING RANGEThe first experiment was conducted to investigate the engineoperating range a VGT would provide and how much loss isrecovered. From previous experiments it was known that theturbine should be as small as possible, consequently thesmallest available VGT was used. This is the BorgWarnerBV35, as used in the Opel/Fiat 1.3 1. diesel enginesintroduced in 2005. The VGT actuator is removed and themechanism is actuated by a Woodward L-series rotationalcontroller. During experiments it was made sure that theengine's throttle valve was open and in addition pressure wasalso measured downstream of the throttle valve to guaranteeit did not influence measurements.

The aim of the second experiment was to prove that the VGTcould be used to control engine power all the way down toidle and to quantify power recovery. To attain this, the vaneshave to be closed, which requires modification of theturbocharger. The BorgWarner model could not bedisassembled without damage so another turbocharger had tobe found that did allow dis- and reassembly. Because themanufacturer provided data of the turbocharger and requestednot to release the turbocharger brand and type, thisturbocharger is called “VGT 2”. It has a slightly larger turbinewheel, but the possibility of closing the vanes would allowmuch lower air mass flow rates. The minimum air mass flowrate using the BorgWarner VGT was 12.2 g/s, removing thestops of the VGT 2 yielded 7.8 g/s. There was room forfurther improvement of the VGT mechanism; the guidingrollers of the unison ring blocked the mechanism, see figure3. Different rollers yielded a mass flow rate of 4.2 g/s. Finallythe leakage flow around the blades was decreased by




shortening the spacers, which yielded 3.5 g/s. The result wasan idle speed of 800 rpm with a cold engine. When the enginewarm, it uses only 3.2 g/s at idle. This way the engineoperating range provided by the VGT is expanded all the wayto an engine speed of 1200 rpm at no load.

Figure 3. VGT mechanism, a circle is drawn around oneof the three rollers.

EFFICIENCYThe third and final experiment is aimed at investigatingturbine efficiency. By varying the turbocharger rotationalspeed, turbine efficiency can be influenced. This is achievedby installing a throttle valve at the compressor exit, as wasdone in [1]. By closing the throttle valve, the compressoroperating point moves from high mass flow rate and lowpressure difference to low mass flow rate and high pressuredifference. This causes an improvement in compressorefficiency and thus lower energy dissipation. A newequilibrium between energy recovery by the turbine andenergy dissipation by the compressor and bearings is found ata higher rotational speed. Turbine speed is varied for anumber of engine speeds at a number of VGT blade angles. Itwas found that for very low VGT openings the efficiencygoes to zero, caused by the degeneration of the nozzle shapeand the increase of the vaneless space, see figure 19. A fixednozzle ring, see figure 4 addressing both these problems, wasdesigned and constructed in an attempt to improve efficiency.

Figure 4. Turbine and nozzles; bottom: VGT blades; top:custom nozzle ring.

REPEATABILITY ANDREPRODUCABILITYBefore discussion of the results, some words on repeatabilityof the measurements are in order. To assess repeatability, aseries of measurements is done on two engine operatingpoints. Two different operating points were chosen to includethe influence of the engine. One operating point is 2000 rpmand high torque, chosen because the engine seemed to runless stable at this point, the other is 3000 rpm and low torque,called ‘low’ and ‘hi’ respectively in figure 5, referring toengine speed. For each operating point, around 15measurements are done with engine restarts between some ofthe measurements. From left to right, variance seems toincrease for every variable. The reason for this is that theright variables are influenced by the left ones. For examplethe temperature drop is a function of the pressure drop, whichin turn is a function of engine speed for a fixed VGT angle.Depending on the measured variable, the maximum deviationfrom the mean value is 0.02 to 1.5 %. This indicates a verygood repeatability.




Figure 5. Normalized values of measured variables;engine speed, air mass flow rate, pressure downstream of

the turbine, temperature drop over the turbine,turbocharger rotational speed.

RESULTSOPERATING RANGEThe processed results from the experiment with theBorgWarner BV35 are presented below. Figure 6 representsthe engine operating range with torque expressed inpercentage of maximum torque for that speed. The dash-dotted lines represent constant VGT opening. Black isolinesindicate power recovered by the turbine and the blue isolinesindicate turbine efficiency. Now a large area of the engineoperating range is covered. In figure 6a, operating points ofconstant speeds are indicated, which show that the area is toomuch inclined towards high power operating points. It ispossible to drive 120 km/h, but 50 and 80 are unreachableusing only the VGT.

Figure 6. Engine operating map (BV35) showing VGTrange and a.) recovered power and recovery efficiency,

b.) temperature drop over turbine and turbine rotationalspeed.

The amount of recovered power is calculated using equation2. Between 100 and 1300 Watts of throttle losses arerecovered. Recovery efficiency is calculated with equation 4and efficiency is between 25 and 73 %. At 120 km/h, 500Watts of power is recovered with an efficiency of 40 %, afterconversion to electricity this is equivalent to the alternatordraining 1125 W from the crankshaft. In figure 6b, thetemperature drop over the turbine and turbine speed areplotted. The temperature drop varies from 2 to 40 °C. Turbinerotational speed varies from 16.000 to 122.000 rpm and isaround 80.000 rpm near common operating points.

From these results it is clear that the engine operating rangeprovided by the VGT is not large enough yet. The range hasto be expanded towards lower engine powers (air mass flowrates). Also efficiency is observed to drop when engine powerdecreases, this causes recovered power to decrease too.




To extend the engine operating range, the VGT 2 wasmodified as described in the experimental section. In figure7a, the dash-dotted lines are lines of constant VGT vaneangle. Contrary to previous results, the lowest dashed linedoes not indicate minimum engine power. Lower powers arepossible but very hard to attain on the engine test stand,because the VGT actuation mechanism is very sensitive inthis range. Now all common engine operating points can beprovided by the VGT. Compared to figure 6a, the engineoperating area provided by the VGT is expanded all the wayto the lower left corner. The maximum torque line is slightlylower. Recovered power is almost the same as in figure 6a,but slightly lower, except for the upper left corner, where it ishigher. It is easily seen that this is caused by lower recoveryefficiencies. When the requested engine power decreases, sodoes recovery efficiency, dropping to almost zero for thelowest engine power. At a steady 120 km/h, almost 500 W isrecovered with an efficiency of almost 40 %. At 80 km/h, 300W is recovered with an efficiency of 25 % and at 50 km/h,110 W is recovered with an efficiency of 12 %. In figure 7b,temperature drop over the turbine and turbine speed areplotted. Comparison of the temperature drop of figures 6band 7b is exactly the same as the recovered powercomparison above because recovered power is calculatedfrom the temperature drop. Turbine rotational speed rangesfrom 19.000 to 100.000 rpm, with around 60.000 nearcommon engine operating points. Rotational speed is lowerbecause the VGT 2 turbine wheel has a larger diameter thanthe BV35 one (39 mm versus 35 mm, respectively).

From these results it can be concluded that the VGT canprovide a useable engine operating range, so it can beoperated without the need for a throttle valve in series withthe VGT, as was done in Ref. [1]. However, when requestedengine power (air mass flow rate) decreases, recoveryefficiency decreases to unacceptable levels.

Figure 7. Engine operating map (VGT 2) showing VGTrange and a.) recovered power and recovery efficiency

and b.) temperature drop over turbine and turbinerotational speed.

The operating range provided by the turbine was expandeddramatically since the first experiment. In figure 8, operationranges are plotted for all turbines used. The Garrett GT1238was a non-VGT turbine, so its operating range consists of aline. The BorgWarner provided a usable range but it wasinclined to higher powers. Finally, the modified VGT 2 wasable to provide the right range.




Figure 8. Engine operating map with operating rangeprovided by each turbine.

TURBINE EFFICIENCYBecause the recovery (turbine) efficiency can be influencedby rotational speed, see figure 25 and equation 18, and inWEDACS the turbine speed is not compromised bycompressor efficiency an additional experiment is conducted.A number of operating points are chosen for which theturbine rotational speed is varied. Because this wasaccomplished using a throttle valve at the compressor exit,the rotational speed range is relatively small.

Below measured turbine efficiency from the experiments isplotted against blade speed ratio for the wide open VGTposition. To get an impression of the significance of thesemeasurements, this data is compared to manufacturer turbineefficiency data that is recalculated to the efficiency - bladespeed ratio format. In figure 9, this manufacturer data can beseen as blue dots and the data from the experiment is plottedas black circles. A fit through all data in the shape of the U/Ccharacteristic, see figure 25, is seen as a solid black line.Manufacturer data consisted of 41 measurement points andmeasurement data of 25 points. Both data clouds are roughlyin line, indicating that the measurements complement themanufacturer data. Figure 10 and 11 show similar results fordifferent operation points. The point cloud from themeasurement data is wider, this is caused by two things; firstthere was a difference in efficiency between differentpressure ratios and secondly the turbocharger was installed onan engine, causing more uncertainties and irregularities. Inthe explanation of figure 25 (blade speed ratio versusefficiency plot), it was claimed that peak efficiency occurs ata blade speed ratio of 0.7. This is under the assumption of aperfect turbine where the exit whirl velocity is zero, whichmeans that all available energy is taken from the air. In thefigures below, the optimum blade speed ratio appears to bebelow 0.7. The most likely cause of this is a non-zero exitwhirl velocity.

Figure 9. Manufacturer and measurement blade speedratio versus efficiency data of the VGT 2 turbine at 100%

VGT blade opening.


VGT blade opening.





VGT blade opening.

Because of modifications to the VGT mechanism, the bladescould be closed to “below 0 % opening”, with the originalsituation as a reference. This data could not be compared tomanufacturer data and will be presented separately fordifferent turbine pressure ratios, see figure 12. The negativeVGT blade opening has the unmodified VGT as a reference,indicating the blades are almost closed. At −100% the bladesare fully closed. Maximum measured efficiencies are 11.5 %for both pressure ratios. At a pressure ratio of 2.7 this is a realmaximum while the efficiency seems to rise for an additionalincrease of blade speed ratio at a pressure ratio of 2.1. Theselow efficiencies are caused by the fact that the ring nozzlesare becoming more important by a decrease in reaction ratedue a decrease in VTG opening, while at the same time thisprevents the nozzles from functioning properly due to adistorted nozzle area and large vaneless space.

Figure 12. Blade speed ratio versus efficiency for twodifferent pressure ratios of a modified VGT 2 turbine at -

60% VGT blade opening.

In figure 13, blade speed ratio versus efficiency in case of thecustom nozzle ring is plotted. Air mass flow rates through theturbine are exactly the same as in the situation above (figure12). The pressure ratios are close to those above. For thepressure ratio of 2.0, maximum measure efficiency is 17 %,which, according to the fitted line is the maximum even ifblade speed ratio could be increased further. For pressureratios of 2.4 and 2.9, maximum efficiencies are 19.5 and 21% respectively. The pressure ratio of 2.4 seems to yield alittle higher efficiency for higher blade speed ratios. Theresult is an efficiency improvement of 50 to 80 %. Becausethe nozzle ring was not optimized and the surface was rough,a consequence of rapid prototyping, there is room for furtherimprovement.

Figure 13. Blade speed ratio versus efficiency for threedifferent pressure ratios of the VGT 2 with a custom

nozzle ring.

In figure 14, the efficiency gains made by changing theturbine rotational speed and custom nozzle ring are indicated.The most significant gains are made at low engine powers,where up to 10 efficiency percentage points are gained.




Figure 14. Turbine efficiency improvement (inpercentage points) in the engine operation map.

DRIVE CYCLETo predict the system performance when installed in avehicle, a drive cycle is simulated. A Nissan Primera issimulated over the NEDC, the relevant vehicle parameterscan be found in Appendix B. The speed pattern of this cycleis plotted in figure 15.

Figure 15. NEDC speed pattern.

For the case of an electrical load of 1500 W, idling duringbraking (as opposed to brake energy regeneration) and no A/C, the drive cycle is simulated using a map for the turbineefficiency that is based on figure 7a and has the turbineefficiency improvements in it that were realized above. Thismap is plotted in figure 16. Note the other expression oftorque; Nm instead of percentage of maximum torque, whichcauses the different shape of the range. In this figure, drivecycle operating points are plotted, showing that during thelast acceleration in figure 15, some operating points falloutside the turbine operating range. In practice, even higher

accelerations occur so it is desirable to extend the rangetowards higher engine powers. The cumulative fuelconsumption over the drive cycle can be seen in figure 17,where a fuel efficiency improvement of 5 % is made. Moreinsight is gained when efficiency improvements for stationaryspeeds are plotted; this is done in figure 18. The strangecourse in the data is an effect of shifting gear, which causes asudden and big change in both engine operating conditionsand turbine efficiency.

Figure 16. Modified turbine efficiency map and drivecycle operating points.

Figure 17. Cumulative fuel consumption over theNEDC.




Figure 18. Fuel consumption improvement at variousstationary vehicle speeds.

Turning on the A/C and assisting it with WEDACS onlyyields and additional 0.3 - 0.4 % improvement over thestationary vehicle speed range. The fuel consumptionimprovement over the NEDC is 5.2 %, only 0.2 %improvement over the case without A/C. The reason for thedifference in impact of assisting the electrical system and A/C is in the efficiency of the system. The alternator on theengine has an efficiency of 60 % and lower, while thealternator used in WEDACS has an efficiency of 90 %, so forevery Watt converted to electricity by WEDACS, at least 1.5Watts are saved at the crankshaft. The A/C system on theother hand has a coefficient of performance of 2.5, so forevery Watt of cooling power, only 0.4 Watts are saved at thecrankshaft. For a more efficient A/C system, this effect onlygets stronger. When the turbine efficiency is raised to valuesclaimed by the manufacturer, efficiency improvements of 7.2% (without A/C) and 7.8 % (with A/C) are achieved.

When the electrical load of the vehicle is varied, it has aneffect on engine operating points and efficiencyimprovement. For example, when the electrical load isincreased to 2 kW, efficiency improvements without and withair conditioning drop to 4.3 and 4.6 %. When the electricalload is reduced to 1 kW on the other hand, efficiencyimprovements without and with air conditioning are raised to5.6 and 6.1 %. A lighter electrical load increases the ratio ofrecovered energy to engine power. The extra efficiencyimprovement by assisting the A/C also increases because thelower electrical load decreases total engine power.

Table 1. Simulation results.

From these results several conclusions can be drawn. First ofall, assisting the A/C system does not yield a significantefficiency improvement. Besides, adding a heat exchangerwould add to the complexity and space requirement of thesystem. This heat exchanger however could possibly addfunctionality in the form of a torque boost, a possibility thathas not yet been fully explored. The fuel efficiencyimprovement, which is based on experimental data, issignificant, especially when the turbine efficiency could befurther increased. This increase in efficiency is notinconceivable, considering this would mean usual turbineefficiencies. Efficiency improvements in this paper are moremodest than those in the last paper, the root cause of this isfound in the drive cycle model. Engine friction, pumpinglosses, auxiliary and alternator power are now included inequation 9, decreasing the ratio of recovered power andengine power. Two differences that tend to increaseimprovement percentages as opposed to the previous paperare present. Previously, the alternator efficiency was assumedto be a constant 60 %, now a speed-dependant alternatorefficiency is introduced, resulting in a lower alternatorefficiency for most operating points. Secondly, in theprevious paper it was assumed that no power was recoveredat engine idle and during deceleration, in the present paper,power recovery at idle and deceleration is possible.

CONCLUSIONSTo reduce fuel consumption of gasoline engine vehicles, asystem is introduced that recovers throttle losses by replacingthe throttle valve with a turbine coupled to a generator. Thealternator and A/C compressor on the engine are assisted bythe system, reducing fuel consumption. Two problems, whichwere revealed in the previous paper [1], are listed; very smalloperating range and the need for a more accurate drive cyclemodel. To solve the problem of a small operating range, aVGT is installed. This VGT enabled control in a high enginepower range only. 100 to 1300 W of throttle losses wererecovered. A different VGT was modified to extend theoperating range to idle, solving the range problem. Duringexperiments, a new problem of low turbine efficiency wasidentified. An attempt was made to improve turbineefficiency by changing the turbine speed and by installing acustom vane ring. Both methods yielded higher efficiency.Accordingly, a different variable VGT mechanism design isneeded, which has the features of the custom nozzle ring:good nozzle geometry at low vane angles and a smallvaneless space at all times. Speed variation increased




efficiency by 0 to 9 % and the combination of custom nozzlering increased efficiency further from 11.5 to 21.5% for oneVGT blade opening position.

Finally, using a drive cycle simulation based of this data, it isshown that the system can improve fuel efficiency of amidsized family car. For the turbine efficiency accomplishedduring the experiments, the efficiency improvement is 5 and5.3 % without and with A/C. With the turbine efficiencyincreased to levels as reported by the manufacturer underoperating conditions typical for turbochargers, the efficiencyimprovement could be increased to 8 %. Higher efficiencyimprovements are possible when turbine efficiency is furtherimproved. Using WEDACS to assist the A/C appears not tobe interesting.

FUTURE WORKThe final goal of this project is a test vehicle with a fullyfunctional prototype installed. To accomplish this, some stepshave to be taken:

• Design, build and test a VGT mechanism that enables bothhigh efficiency and a wide operating range

• Optimize turbine efficiency via variable alternator load

• Design a combination of a dedicated alternator, bearings,seals and turbine to enable electricity conversion and test thisprototype

• Design power electronics to process the electricity from thealternator and design a control system to control thealternator and VGT mechanism

• Install the system into a test vehicle.

• Explore additional benefits of the system, like a torqueboost function.

REFERENCES1. Eichhorn, R.H.L., Boot, M.D., and Luijten, C.C.M.,“Waste Energy Driven Air Conditioning System(WEDACS),” SAE Int. J. Engines 2(2):477-492, 2009.

2. Perreault, David J., Caliskan, Vahe, Automotive PowerGeneration and Control, IEEE transactions on powerelectronics, vol. 19, no. 3, may 2004.

3. Sandoval, D. and Heywood, J.B., “An Improved FrictionModel for Spark-Ignition Engines,” SAE Technical Paper2003-01-0725, 2003.

4. Kassakian, John G., Miller, John M., Traub, Norman,Automotive electronics power up, IEEE Spectrum, May2000, Vol. 37, No. 5.

5. Drury, W., Drury, Bill, London 2001, The controltechniques drives and controls handbook, The Institution ofElectrical Engineers, ISBN 0852 96793 4.

6. Watson, N., New York 1982, Turbocharging the internalcombustion engine, Wiley, ISBN 0471 87072 2 and 033324290 4.7. Çengel, Yunus A., Boles, Micheal A., “Thermodynamics,an engineering approach”, (2002), McGraw-Hill, ISBN0-07-121688-X

DEFINITIONS/ABBREVIATIONSSubscripts01

Static unit at turbine entry

1Turbine entry

5Turbine exit

WEDACSWEDACS

aAir or ambient

acAir conditioning

altConventional alternator

alt, newWEDACS alternator

auxAuxiliary

condCondensation

elElectrical load

fFuel

frFriction




http://www.sae.org/technical/papers/2003-01-0725

iIntake manifold

pPumping

recoverRecovery

tFor the time instant or turbine

throtThrottle

Greek symbolsη

Efficiency

κRatio of heat capacities

ξSpecific humidity

πPressure ratio

ρDensity

φFlow function

ψRelative humidity

ωRotational speed

Roman symbolsAeff

Effective flow area

CAir speed attained by isentropic expansion

CdDrag coefficient

CpHeat capacity at constant pressure

COPACCoefficient of performance of the air conditioning

PPower

RGas constant of air

TTemperature

TbBrake engine torque

TeEngine torque

UTurbine blade tip speed

V̇Volume flow rate

hEnthalpy

ṁMass flow rate

pPressure

pgSaturated vapor pressure

rRadius




uφTangential speed




APPENDIXAPPENDIX A

TURBINE THEORYAt present, advanced diesel powertrains are equipped withVGT's. Compared to a conventional turbocharger, the VGTdramatically improves transient response of the powertrain.In addition, VGT's are usually used to control the EGR-rate.This is made possible by a number of movable vanes placedaround the turbine wheel. A turbine, as is used in automotiveturbochargers, can be divided into two areas; the nozzle areaand an impeller.

Figure 19. Components of a radial inflow turbine(Copied from [6])

Consider the conventional radial inflow turbine with fixedvanes in figure 19. The nozzle area consists of a scroll (Inletcasing) and a set of vanes (Nozzles) that direct the air to theimpeller wheel (Rotor).

Because the flow area of the scroll and the vanes decrease,the air speeds up. The angle of the vanes, with minorinfluence of the scroll, results in a certain direction of theflow, while the flow area determines the magnitude of thespeed and mass flow. It is possible to estimate the turbinemass flow rate by modeling the turbine as an adiabaticnozzle:

(10)

Where Cd is the discharge coefficient, Aeff is the effectiveflow area of the nozzles around the impeller, subscript 01indicates stagnation properties at the turbine exit, R is the gasconstant of air and φ is the flow function:

(11)

Where π is the pressure ratio over the turbine and κ is thespecific heat ratio of air. In principle the flow equationdetermines that the mass flow rate increases with increasingpressure ratio until the nozzle chokes at a certain pressureratio.

According to Watson the adiabatic nozzle model is not agood representation of a turbine flow characteristic becausethe latter one chokes at higher pressure ratios and is(rotational) speed dependent to some extent [6]. The turbinecould be considered as two nozzles in series, the nozzle androtor passages. The centrifugal field created by the rotorintroduces a rotational speed dependence.

The impeller rotates inside these vanes and is driven by theair leaving the nozzles. This air hits the rotating blades of theimpeller at a certain angle. The shapes of the impeller and theblades form additional nozzles.

The rotational pressure field causes the pressure to risebetween both nozzles because of the rotational speed of theimpeller. This pressure rise can be calculated by a simplifiedform of the Navier-Stokes equation in the radial direction,assuming no variation of radial speed in all directions andtime. Friction is not considered.

(12)

where the tangential speed can be expressed as:

(13)

Additionally, using the third isentropic relation for an idealgas [7], the density can be expressed as:

(14)




Where subscript 1 indicates conditions at the inside radius ofthe impeller and subscript r indicates conditions at radius r.Subscript r will be omitted in further deduction of thepressure gradient. Substituting equations A.4 and A.5 inequation A.3 yields

(15)

Integration of

(16)

yields

(17)

which is written as:

(18)

The pressure at the edge of the rotational pressure field canbe calculated.

(19)

By dividing this by p1 the ratio of pressures inside andoutside the rotational pressure field can be calculated.

(20)

The density can be calculated from the temperature andpressure using the ideal gas law.

(21)

Substitution into equation 20 yields the final equation.

(22)

Using equation 22, the pressure gradient by the rotatingimpeller can be calculated. The nozzle and centrifugalpressure field are ingredients of a turbine mass flow modelwhich looks like figure 20.

Figure 20. Turbine mass flow model; N1 is the nozzlering, CPF is the centrifugal pressure field and N2 is the

impeller nozzle.

Consider the case of a VGT where the turbine has movablevanes. Now the effective area in equation 10 can be adjustedby changing the vane angle. In figure 21, the solid linerepresents the ring (VGT) nozzles only. This line has a shapethat is typical for nozzles and starts choking at a pressureratio of 1.85. This is why a single nozzle is not a goodrepresentation of a turbine, which chokes at higher pressureratios. Also it is clear that every pressure ratio corresponds toa single mass flow rate. This fact is used to add the impellernozzle to the model. The pressure before the turbine isknown. First equation 10 is used to calculate the mass flowrate through the ring nozzles for a number of pressure ratios.Because of conservation of mass, the mass flow rate throughboth nozzles has to be equal. So now for every pressure ratioover the ring nozzles, equation 10 can be used again tocalculate the correct mass flow rate and correspondingpressure ratio over the impeller nozzle. The result is thedashed line in figure 21. Now both nozzles start choking at amuch higher pressure ratio of 3. Addition of the centrifugalpressure field yields the dash-dotted line. The rotatingimpeller causes an increased pressure behind the first nozzle.When the air travels inward towards the second nozzle, thepressure decreases without doing any work. The pressuredifference over the first nozzle decreases. At high rotationalspeeds and thus at high pressure ratios over the turbine thecentrifugal pressure field can cause the air mass flow todecrease again.




Figure 21. Turbine map produced by the model afteradding elements.

To assess the validity of the model, it is compared tomanufacturer data, see figure 22. Since the drag coefficientand leakage effects are not incorporated into the model, thenozzle flow areas are slightly (around 10%) adapted from themeasured areas to fit the manufacturer data. The rotationalspeeds in the model were set to those specified by themanufacturer. As seen in figure 22, the model correspondsvery well to the manufacturer data. The shape of both lines isvery similar and choking occurs at the same pressure ratio. Itcan be concluded that the model covers the main phenomenaof mass flow through a turbine.

Figure 22. Comparison of model and manufacturer data.

In figure 23, the nozzle functions of the ring and impellernozzles are plotted for three different VGT openings: open(1), medium (2) and closed (3). The influence on air massflow rate is clearly visible. Another thing that stands out isthe fact that the impeller nozzles become less important whenVGT opening is decreased. In the circles in figure 23,

corresponding pressure drops of both nozzles are shown. Forexample for the open VGT position the pressure ratio overthe nozzle ring is 1.16 and that over the impeller is 1.33. Sopressure might drop from 1 bar to 0.86 bars over the nozzlering and from 0.86 bars to 0.65 bars over the impeller. Thecorresponding pressure drops are 0.14 bars and 0.21 bars, thelatter is higher, meaning that more energy is recovered in theimpeller. The ratio of energy transfer due to change in staticpressure in the rotor to the total energy transfer in the stage iscalled the degree of reaction [6]. Ignoring efficiency, thedegree of reaction in the situation in ellipse 1 in figure 23 asdescribed above is 60 %. The degree in reaction in ellipse 2and 3 are 41 % and 10 % respectively. These may not be thetrue degree of reaction values as found in practical operationof the turbine, but they do indicate that the degree of reactiongoes to zero for very low VGT openings found in themodified turbine (see the experiments Chapter). When thedegree of reaction of a turbine is 0 %, it is of the impulsetype, meaning that all energy transfer comes from an array ofnozzles blowing air onto a rotating wheel.

Figure 23. Nozzle ring and impeller nozzle functions forthree different VGT openings: solid line, open, dashed

line, half way open, dotted line, closed.

In the case of extremely low VGT openings, the VGT vanesalmost touch each other, see figure 24. Now the nozzle shapeis little more than a crevice, failing to aim the air at theimpeller at the right angle. Also the vaneless space is verylarge, causing the gas jet from the nozzle to distort. Becausethe turbine operates in the impulse regime at this VGTopening, it relies on gas jets driving the impeller wheel.Distortion of the gas jets prohibits the turbine fromfunctioning properly and thus decreases turbine efficiency.To solve this problem, a nozzle ring with better nozzle shapesand a smaller vaneless space will have to be made.




Figure 24. Deteriorated nozzle function at very low VGTopening.

As said in impulse turbines, all air is accelerated to a certainspeed through a single array of nozzles. This speed can beused to assess turbine efficiency. A good predictor of turbineefficiency is the ratio of rotational speed to the speed the airwould attain by full isentropic expansion [6]:

(18)

Where U is the tangential speed of the rotor blade tip and thedenominator is the velocity equivalent of isentropicexpansion, with Cp the specific heat of air and subscript 01indicating total turbine inlet and 5 indicating turbine exitconditions. In figure 25, a typical relationship between thespeed ratio and turbine efficiency is shown. The efficiency isrelatively insensitive to pressure ratio, as shown in the upperfigure. The turbine efficiency peaks at U/C = 0.7. At lownozzle angles the exit kinetic energy becomes significant,relative to turbine power. At low mass flow rates and highrotational speeds (high U/C), the flow becomes insufficient toovercome the centrifugal pressure field, this explains thesteep pressure drop when U/C increases above 0.8. When fora certain turbine operating point, mass flow rate and pressure,

the rotational speed is changed, according to equation 18, Uand thus U/C are also changed. Unlike in a turbocharger, theturbine rotational speed is not determined by a compromiseof the turbine and compressor efficiencies. Consequently theturbine efficiency can be chosen by adjusting the rotationalspeed.

Figure 25. Variation of turbine efficiency with bladespeed ratio for a turbine with fixed stator nozzles.

The Engineering Meetings Board has approved this paper for publication. It hassuccessfully completed SAE's peer review process under the supervision of the sessionorganizer. This process requires a minimum of three (3) reviews by industry experts.

All rights reserved. No part of this publication may be reproduced, stored in aretrieval system, or transmitted, in any form or by any means, electronic, mechanical,photocopying, recording, or otherwise, without the prior written permission of SAE.

ISSN 0148-7191

doi:10.4271/2010-01-1441

Positions and opinions advanced in this paper are those of the author(s) and notnecessarily those of SAE. The author is solely responsible for the content of the paper.

SAE Customer Service:Tel: 877-606-7323 (inside USA and Canada)Tel: 724-776-4970 (outside USA)Fax: 724-776-0790Email: [email protected] Web Address: http://www.sae.orgPrinted in USA




http://dx.doi.org/10.4271/2010-01-1441

Table 2. Engine specifications

Table 3. Vehicle model parameters

The Engineering Meetings Board has approved this paper for publication. It hassuccessfully completed SAE's peer review process under the supervision of the sessionorganizer. This process requires a minimum of three (3) reviews by industry experts.

All rights reserved. No part of this publication may be reproduced, stored in aretrieval system, or transmitted, in any form or by any means, electronic, mechanical,photocopying, recording, or otherwise, without the prior written permission of SAE.

ISSN 0148-7191

doi:10.4271/2010-01-1441

Positions and opinions advanced in this paper are those of the author(s) and notnecessarily those of SAE. The author is solely responsible for the content of the paper.

SAE Customer Service:Tel: 877-606-7323 (inside USA and Canada)Tel: 724-776-4970 (outside USA)Fax: 724-776-0790Email: [email protected] Web Address: http://www.sae.orgPrinted in USA

APPENDIX B

ENGINE AND VEHICLESPECIFICATIONS




http://dx.doi.org/10.4271/2010-01-1441

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