Abstract
Currently, turbocharged common rail high pressure direct injection diesel engines are regarded as state-of-the-art. The use of the turbocharging technique in gasoline engines is also increasing, in order to achieve further fuel consumption reductions via downsizing. As the specific power outputs of both diesel and gasoline engines rise, the low-end torque behavior of such engines and turbo-lag are becoming increasingly critical. This is primarily a result of the specific characteristics of turbochargers and internal combustion engines themselves.
When it comes to matching a turbocharger to a given engine, the compressor map over the entire operating area has to be known with sufficient accuracy, especially at low turbocharger speeds corresponding to the engine low part loads (i.e. urban traffic). This map is established assuming the adiabatic behavior of the compressor. While this assumption is acceptable at rather high speeds, it is no longer valid for low speeds, and for that reason, the compressor map in this area is not provided by the turbocharger manufacturer.
Worldwide, there are no standard guidelines for the correct measurement and calculation of turbocharger maps at low speeds. In collaboration with a French automotive manufacturer, a special method was therefore designed and applied within the laboratory LGP2ES at Cnam Paris in order to obtain the compressor low speed map. A special torquemeter was fitted in a cold turbocharger test bench, affording measurements from 30,000 rpm to 120,000 rpm. The experimental results presented in this paper show the combined effect of the lubricating oil temperature and pressure on the compressor performance, expressed in terms of compression ratio, compressor power, isentropic efficiency and mechanical efficiency. These results afford a better estimation of the compressor map at low speeds.
Keywords
Compressor; Torquemeter; Isentropic efficiency; Mechanical efficiency; Friction losses
Nomenclature
cpspecific heatkspecific heat ratioNturbocharger rotational speedPcompressor powerpi1compressor inlet total pressurepi2compressor outlet total pressurePffriction power lossPmmechanical power at the turbocharger shaft (measured with the torquemeter)qmmass air flow rateTi1compressor inlet total temperatureTi2compressor outlet total temperature after real compressionTi2’compressor outlet total temperature after isentropic compressionToillubricating oil temperatureWisisentropic work used for the isentropic compression of airWrealreal work used for the real compression of airΔTisisentropic temperature variation between compressor inlet and outletΔTrealreal temperature variation between compressor inlet and outletηiscompressor isentropic efficiencyηmturbocharger mechanical efficiencyτccompressor compression ratio
1. Introduction
Downsizing automotive engines is currently considered as one of the most promising ways to improve fuel economy with an acceptable cost-to-benefit ratio. The challenge is to reduce the engine displacement while keeping the same performance in terms of torque and power as the initial larger engine and to simultaneously ensure an improvement in engine efficiency. This cannot be done without increasing the density of the air introduced into the engine. Turbocharging is the boosting technology generally used in today’s market and is the subject of extensive research, which seeks to overcome its drawbacks, such as low-end torque, turbo-lag [1] and [2] and compressor surge [3] and [4].
When it comes to matching a turbocharger to a given engine, one problem is the operation of these two machines which have highly different characteristics. The turbocharger is a continuous flow machine whereas the reciprocating engine is a discontinuous flow machine, which generally operates over a wide range of speed and torque [5] and [6].
Consequently, improving the operation of highly downsized, turbocharged engines requires knowledge of the turbine and compressor maps over the entire operating area. At low part loads corresponding to urban traffic according to the NEDC (New European Driving Cycle) test, the turbocharger speed is less than 100,000 rpm [9] and [16]. Unfortunately in this area, the compressor map is not provided by the turbocharger manufacturer. If the maximum turbocharger speed is 150,000 rpm, characteristic curves are not given for speeds less than 70,000 rpm, or 90,000 rpm if the maximum speed is 240,000 rpm. The reason is that these maps are based on the assumption of adiabatic behavior of the compressor, an assumption that is no longer valid at low turbocharger speeds [7].
The isentropic efficiency of the compressor is deduced from measurements of compressor inlet and outlet parameters such as pressure and temperature. At low speeds, the inlet–outlet variation in temperature is small, so the inherent measurement errors become important and adversely influence the accuracy of the calculation of compressor power and isentropic efficiency. Moreover, in this case, thermal transfer cannot be neglected and standard calculations can no longer be used, [7] and [8], especially on hot gas stands which are widely used by turbocharger manufacturers.
Concerning turbine performances, on a hot gas stand, the turbine obviously cannot be considered adiabatic and a common practice is to give the efficiency as the product of mechanical efficiency and isentropic efficiency of the compressor, which implies doing the usual measurements on the compressor during turbine tests.
Worldwide, there are no standard guidelines for the correct measurement and calculation of turbocharger maps at low speeds.
To achieve a better understanding of turbocharger performances, tests can be done on a cold test rig. This enables the isentropic efficiency of the turbine to be determined. In the present study, this is the approach adopted. Similar experiments have been conducted at low speed [8], using a water-cooled turbo to minimize heat exchange.
Another way to improve our knowledge of turbocharger efficiency is to assess mechanical power losses. In our case, this has been accomplished thanks to a torquemeter: if power losses are known, we are able to calculate the power given to the air flow in spite of heat exchange. Work is also in progress on bearing losses, using a combined experimental and numerical approach [9].
Experiments on measurements of bearing losses have been done by Honeywell Turbocharger Technologies [10]. The turbine was carefully insulated (adiabatic conditions) and experiments were performed at 100 °C. The compressor blades were removed, so compressor power could be neglected. The laboratory at Stuttgart University [11] has designed a specific test bench for the direct study of bearing losses. Turbocharger wheels were removed and axial forces were generated by an electromagnetic device. Torque was measured with a high degree of precision for a rotary strain gauge torque sensor. Results are expressed in terms of percentage of power and torque; while this does not provide access to the real value, the results of power or torque evolution are in agreement with our experiments.
An interesting method based on turbocharger inertia and measurements of speed deceleration has been proposed by the University of Hanover [12]. Unfortunately, it seems that the friction power determined by this method is overestimated.
The calculations presented in the above-mentioned papers highlight the difficulties of approximating bearing losses on the basis of empirical calculations.
In collaboration with a French automotive manufacturer, a special method was therefore designed and applied within the laboratory LGP2ES at Cnam Paris in order to obtain the compressor low speed map. A special torquemeter was fitted on a standard turbocharger test bench, affording measurements from 30,000 rpm to 120,000 rpm [7] and [14].
2. Turbocharger test rig setup
Turbocharger maps are usually acquired on hot test benches [13], or cold test benches [7] and [8]. In the former case, the heat flux between the hot turbine and the cold compressor causes overestimation of the calculated compressor power and underestimation of isentropic efficiency. This error, due to the assumption of adiabatic behavior of the compressor, becomes even greater at low turbocharger speeds, as stated before [7], [8] and [13].
As there is no standard that provides detailed descriptions of the correct measurement and calculation method to obtain turbocharger maps, experiments were conducted in the LGP2ES laboratory at Cnam Paris on a standard cold turbocharger test rig fitted with a torquemeter specially designed for such applications. Torque was measured from shaft twist, which was deduced from the phase difference between two toothed wheels located at either end of the shaft [14]. The same device gives the rotational speed and hence the power. The main characteristics of the torquemeter are as follows:
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Maximum speed: 120,000 rpm
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Shaft diameter: 2.46 mm
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Maximum power: 5 kW
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Accuracy: ±0.0016 N m
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Grease lubricating ceramic ball bearings
The test rig layout is shown in Fig. 1 and a picture of the torquemeter and the turbocharger in Fig. 2.
Schematic of test rig.
Fig. 1.
Schematic of test rig.
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Test bench.
Fig. 2.
Test bench.
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The turbine is fed by dry compressed air under steady flow and in this application is used only for driving the compressor. The air source is a 700 m3 tank under 25 bar. The turbine flow rate is controlled by a valve and a second valve is used to modify the operating conditions of the compressor. The compressor map can thus be found by adjusting these valves. Both center housings are fed by the lubricating unit with SAE 15–30 W oil. Oil inlet temperature and pressure are adjustable, respectively from 20 to 120 °C and from 0.5 to 4 bar.
3. Determination of compressor performance
As mentioned above, usually the compressor is presumed to have an adia