Precise instrumentation of a diesel single-cylinder research engine

Measurement 44 (2011) 1261–1278

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Precise instrumentation of a diesel single-cylinder research engine Usman Asad ⇑, Raj Kumar, Xiaoye Han, Ming Zheng Mechanical, Automotive & Materials Engineering, University of Windsor, 401 Sunset Ave., Windsor, Ontario, Canada N9B 3P4

a r t i c l e

i n f o

Article history: Received 22 September 2010 Received in revised form 3 February 2011 Accepted 28 March 2011 Available online 31 March 2011 Keywords: Single-cylinder engine Diesel Instrumentation Engine setup Engine testing Research engine

a b s t r a c t The accuracy of any empirical result is a direct consequence of the quality of experimental setup and the strict control over testing conditions. For internal combustion engines, a large number of parameters that also exhibit complex interdependence may significantly affect the engine performance. Therefore, this work describes the essentials required to establish a high-quality diesel engine research laboratory. A single-cylinder diesel engine is taken as the fundamental building block and the requirements for all essential sub-systems including fuel, intake, exhaust, coolant and exhaust gas recirculation (EGR) are laid out. The measurement and analysis of cylinder pressure, and exhaust gas sampling/conditioning requirements for emission measurement are discussed in detail. The independent control of EGR and intake boost is also highlighted. The measurement and analysis techniques are supported with empirical data from a single-cylinder diesel engine setup. The emphasis is on providing the necessary guidelines for setting up a fully-instrumented diesel engine test laboratory. Ó 2011 Elsevier Ltd. All rights reserved.

1. Introduction The subject of engine development and testing has changed considerably in the last few decades, from a purely mechanical task, into the realm of an inter-disciplinary field, incorporating mechanical, electronics, and computer hardware/software engineering, to name a few. The instrumentation of any engine now incorporates a wide range of sensors and actuators, with varying signal types and communication protocols (analog voltage, digital pulse I/O, TTL pulse I/O, pulse-width modulated signals, controller area network, RS-232, GPIB, etc.). This demands a broad range of knowledge and skills from a test engineer to undertake any aspect of engine testing and successfully meet the test requirements. Moreover, the full implementation of the research plan may not be possible because of the inbuilt limitations within the standard engine control unit (ECU) as well as restricted access to the engine calibration data. Therefore, it becomes imperative at times to ⇑ Corresponding author. Tel.: +1 519 253 3000x4154; fax: +1 519 973 7007. E-mail address: [email protected] (U. Asad). 0263-2241/$ - see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.measurement.2011.03.028

bypass the ECU to enable unrestricted access to the necessary hardware such as the fuel injection system to explore the requisite engine operation [1,2]. The fundamental tool for undertaking research and development in the field of internal combustion engines is the single-cylinder engine because it allows segregation of the performance within an individual cylinder (compared to an averaged performance with multiple cylinder configurations) while also eliminating the often complex inter-cylinder interactions [3]. More importantly, unlike a conventional engine where trade-offs between the simultaneous availability of intake boost and EGR (exhaustturbocharger limitations) severely restrict the exploration of borderline engine conditions, a critical advantage of a single-cylinder engine is to enable the test engineer to independently investigate the effects of intake boost pressure, intake temperature and gas composition, exhaust back pressure and EGR on the engine operation and performance. Single cylinder versions of multi-cylinder production engines are used extensively to minimize testing times and to carry out high-quality research (including evaluation of combustion chamber designs, valvetrain

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Nomenclature (dp/dh)max maximum rate of cylinder pressure rise (bar/ °CA) BDC bottom dead centre CA crank angle (°CA) CA50 crank angle of 50% heat released (°CA) CO carbon monoxide (ppm, g/kW h) CO2 carbon dioxide (%) ECU engine control unit EGR exhaust gas recirculation (%) EOC end of combustion (°CA) EPA Environmental Protection Agency FIP fuel injection pump FPC fired pressure coefficient FPGA field programmable gate array FSN filter smoke number HC hydrocarbons (ppm, g/kW h) HCCI homogeneous charge compression ignition HRR heat release rate (J/°CA, 1/°CA) I/O input/output IMEP indicated mean effective pressure (bar) IVC intake valve close (°CA) kS/s kilo samples per second LTC low temperature combustion

configurations and air-system boundary condition development) that can be readily transferred and applied to multi-cylinder configurations. Single-cylinder engines are also increasingly being used for analyzing the fuel and lubricant impacts on the intake valve, piston and combustion chamber deposits [4]. The installation and instrumentation of a single-cylinder research engine is generally more complicated than that of a multi-cylinder engine because most of the engine sub-systems including lubricating, cooling and fuel systems have to be externally established and controlled. Moreover, the flows in a single-cylinder configuration (intake, exhaust) are highly pulsating and require special setup and instrumentation considerations to provide reliable and accurate measurements. Although a number of publications on the subject of engine testing and development are now available [1,2,5]; however, most of them provide generalized information pertaining to setting up an engine testing system and no such publication provides a comprehensive treatment for the setup and instrumentation of a single-cylinder research engine. In this paper, a single-cylinder diesel engine coupled with a dynamometer is taken as the fundamental building block. The authors have made an attempt to address the specific subject of setting up a diesel single-cylinder research engine, including detailed schematics for the various sub-systems and common testing practices adhered to by engine research laboratories in government, academia (including the authors’ own laboratory) and the automotive industry. The measurement of cylinder pressure, fuel flowrate, exhaust emissions and their analysis, and the independent setup of intake, exhaust and exhaust gas recirculation (EGR) systems are described in detail along

MPC MS/s NMHC NOx O2 PC PCV PDR PID PM pmax ppm rpm RT SAE SOC SOI TDC THC TTL UHC VCV

motored pressure coefficient mega samples per second non-methane hydrocarbon (ppm, g/kW h) oxides of nitrogen (ppm, g/kW h) oxygen (%) personal computer pressure control valve pressure departure ratio Proportional–Integral–Derivative particulate matter maximum cylinder pressure (bar) parts per million revolution per minute real-time Society of Automotive Engineers start of combustion (°CA) start of injection (°CA) top dead centre total hydrocarbon (ppm, g/kW h) transistor–transistor-logic unburned hydrocarbon (ppm, g/kW h) volume control valve

with the recommended instrumentation and measurement techniques. Integration of the data acquisition devices to synchronize and record the engine test data is also suggested. It is pertinent to mention here that single-cylinder engine systems are commercially available including a complete package of equipment and services on a turnkey basis. However, these systems can also be assembled from commercially available components by the test engineers and the focus of this work is on the setup of such self-assembled systems. Moreover, this paper does not deal with the subjects of laboratory layout, mounting of the engine, dynamometer test bed mounting and coupling to the engine, vibration and noise containment which are more than adequately addressed in numerous publications [1,2,5,6].

2. Empirical approaches and research methodology The setup and instrumentation of a single-cylinder diesel engine are covered under the following twelve subheadings. 2.1. 2.2. 2.3. 2.4. 2.5. 2.6. 2.7. 2.8. 2.9. 2.10.

Cylinder pressure measurement Cylinder pressure analysis Fuel flow measurement Intake, exhaust and EGR systems Exhaust emission measurement and analysis Carbon mass-balance for estimating fuel flow rate Temperature and pressure measurements Lubricating and cooling systems Fuel injection system Data acquisition system integration

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2.11. Measurement accuracy and calibration frequency 2.12. Multi-cylinder engine conversion to single-cylinder operation

2.1. Cylinder pressure measurement The analysis of the combustion process pivots on the accurate measurement of the cylinder pressure [7]. A number of commercial engine indicating systems comprising both hardware and software, are widely available. However, these systems can also be assembled from commercially available components and the focus here is on the setup and use of such assembled systems. An overview of a cylinder pressure measurement system is shown in Fig. 1.

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The cylinder pressure transducers are, in almost all instances, based on the piezoelectric principle. The correct mounting of the transducer in the engine and its integration within the measurement system are vital ingredients for obtaining accurate combustion data. If space and material are available in the cylinder head structure, the transducer is usually installed directly by modifying the cylinder head; otherwise, special adapters in the shape of a glow plug are generally used to provide access to the combustion chamber and also to support the transducer. The transducer needs to be installed such that its tip (the sensing surface) is flush with the cylinder head (combustion chamber) surface. With typical glow-plug sizes around /5 mm, the smallest pressure transducers generally cannot be flush-mounted without modifying the cylinder head.

Fig. 1. System layout for cylinder pressure measurement & analysis.

Fig. 2. Pressure pulsations typically observed with a non-flushed cylinder pressure transducer.

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Non-flush-mounted sensors can be severely affected by the pressure wave action within the narrow access passages, resulting in high frequency pressure pulsations that may obscure the actual combustion characteristics [8,9]. A representative case of pressure data from a non-flushmounted pressure transducer is shown in Fig. 2. The quality of the pressure data from such a transducer can be improved by accounting for the access passage resonance frequency given by

fn ¼

c 4L

ð1Þ

where fn is the passage resonance frequency in Hz, c is the speed of sound in m/s and L is the length of the access passage in meters. To compensate for this resonant frequency, the pressure data needs to be analyzed in the frequency domain (typically using a Fast Fourier Transform). However, this also requires that the transient angular-speed variations must be accounted for, to convert the data from the ‘crank angle domain’ to the ‘time domain’ which can add significantly to the computational overhead. Alternately, the data can be filtered using a low-pass filter but this may result in a possible loss/change of some important characteristics in the pressure data including the maximum rate of pressure rise and the combustion phasing parameters. It can be seen that without a considerable amount of post processing, it is difficult to use the raw data to carry out heat release analysis or for example, estimate the start of combustion (SOC), crank angle of 50% heat released (CA50) and end of combustion (EOC). In comparison, the pressure data from a flush-mounted pressure transducer (shown in Fig. 3) allows unambiguous identification of combustion parameters while also reducing the need for data filtering and additional computational resources. It is also important to ensure that the transducer temperature limits are not violated. Sensors are also available with integrated cooling circuits to reduce the temperature drift and improve durability. These sensors require filtered, deionized water and a separate flow control system to ensure reliable operation [1,2]. The signal from the piezoelectric transducer is conditioned using a charge amplifier which converts the charge input into an amplified analog voltage output. The calibra-

Fig. 3. Pressure data from a flush-mounted pressure transducer.

tion data, provided by the manufacturer is used to match the charge amplifier and the transducer. Some important issues that must be accounted for in any piezoelectric measurement system are as follows: 1. The time constant of the charge amplifier needs to be adjusted depending on the type of the test engine. Most charge amplifiers have a user selectable ‘short’ or ‘long’ time constant. A ‘short’ time constant is generally used for the modern diesel engines (high speed, light to heavy-duty). However, for large engines (marine engines for example) that run at very low speeds, a ‘long’ time constant is necessary to prevent significant errors in the measurement. 2. Every piezoelectric principle-based measurement system is characterized by an inherent drift in the output signal. Signal drift is an undesirable change in the output signal over time that is not a function of the measured variable, i.e. cylinder pressure in this case. While not totally eliminated, the drift is limited by operating the amplifier in the ‘short’ time constant mode and/or by the use of electronic drift compensation circuitry within the charge amplifier. However, further compensation can be done in several ways, including using dual pressure transducers, with one high-pressure transducer located in the combustion chamber and the second low-pressure transducer located in the liner at approximately mid stroke position [10], by calculating and correcting the polytropic index for the compression stroke to about 1.35–1.37 or by referencing the pressure at some key point, for example the intake valve closing (IVC) with the pressure value from an additional sensor in the intake manifold. 3. Most charge amplifiers also allow the user to filter the raw signal to remove undesirable frequencies (noise). Low pass filters are usually used to remove high frequency signal contents such as interference from other electrical sources or structure-borne noise from the engines that are transmitted to the transducer. However, it is important to understand that every electrical filter causes a certain amount of phase shift which needs to be accounted for. For example, a 1 °CA shift in the pressure data over an engine cycle can cause an error of up to 5% in the indicated mean effective pressure (IMEP) calculation. Since the frequency contents of the pressure data vary with engine speed, the usual practice in the industry is to set the minimum filter cut-off frequency to at least 100 times the cylinder firing frequency [2]. Since the signal from the pressure transducer is of a very low level, the charge amplifier should be installed as close as possible to the pressure transducer to avoid the addition of external noise to the low level signal as it travels down the input wires to the amplifier. Moreover, proper shielding from electro-magnetic interference (EMI) should be ensured and ground loops prevented by maintaining all the components at a common ground reference. The cylinder pressure data is almost always acquired in the crank angle domain [11,12]. For this purpose, an optical incremental rotary encoder is installed at the free-end of

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the crankshaft to provide positional data. The encoder shaft is aligned with the crankshaft axis, mostly by using a flexible coupling between the encoder shaft and a special mounting fitted onto the crankshaft. To accurately align the cylinder pressure data with the corresponding cylinder volume (piston position), an optical encoder provides two different outputs as follows: 1. A digital square wave that corresponds to a fixed angular rotation of the crankshaft. This signal is typically called ‘Index A’ or the clock signal. Encoders with an incremental resolution of 0.1 °CA are commonly used nowadays. The benefit of using a fine resolution like 0.1 °CA is that the signal can be downsampled during post processing to obtain equivalent data at a lower resolution (0.2, 0.5 °CA, etc.) if required. 2. A digital square wave signal that occurs once every revolution and indicates a discrete angular position in the engine revolution. This signal is used as an event trigger (to start the pressure-data acquisition) and is called ‘Index Z’ or simply the trigger signal. The discrete position is usually taken as the top dead centre (TDC) or the bottom dead centre (BDC), which requires the accurate determination of the true TDC and the alignment of the encoder with this position. A number of techniques for estimating the true TDC position can be found in the literature [2,8,9]. The amplified signal from the charge amplifier and the encoder signals are input to a high-speed data acquisition card-PC system. Data acquisition cards with sampling speeds up to 1 MS/s and 16 bit resolution are commonly available. For modern diesel engines, a 500 kS/s sampling rate is more than sufficient since the data throughput at 5000 rpm with an encoder resolution of 0.1 °CA is 300 kS/ s. The acquisition should be externally triggered (using encoder’s Index Z) and externally clocked (using encoder’s Index A) to ensure data synchronization with the engine rotation [12]. The pressure signal is typically acquired using an analog input channel with ±10 V range. For a 16 bit card, the measurement resolution is 0.61 mV that for instance, corresponds to a pressure resolution of 0.012 bar with the amplifier gain set as 1 V = 20 bar. The acquired data should preferably be stored using binary file format (32-bit single precision or 64-bit double precision hexadecimal floating point representation) – the use of text ASCII files is not recommended since the file size is generally 3–4 times larger than a binary file for the same amount of data. The data should be duly timestamped and marked to allow archiving and retrieval at a later stage.

2.2. Cylinder pressure analysis The analysis of the cylinder pressure can provide a wealth of information about the combustion process, including both qualitative and quantitative data. A number of publications are available that provide an in-depth treatment of the subject [2,5,7–10,13–15]. Therefore, this section will focus on the calculation of two important com-

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bustion parameters, that is, the heat release rate (HRR) and the IMEP. The process of heat release, in combination with certain additional parameters (such as the exhaust oxygen concentration, engine load level or ignition delay), can adequately describe the performance and emission characteristics of a diesel engine. The basis for the modeling of the heat release is the first law of thermodynamics for an open system [10,15–17]. A number of simplified heat release models of reduced levels of complexity have been derived from the first law model, with the most common being the apparent or net heat release model. Assuming the cylinder charge is contained in a single-zone, neglecting the heat transfer, crevice volume, blow-by and the fuel injection effects, and using the ideal gas law, the apparent or net heat release during combustion, dQapp on a crank angle basis is given by:

  dQ app dQ gr dQ ht 1 dV dp cp þ V ¼  ¼ dh dh dh dh dh c1

ð2Þ

where dQgr/dh is the heat release rate by combustion, dQht/ dh is the cylinder charge-to-wall heat transfer rate, dQapp/ dh is the apparent or net heat released, h is the engine crank angle, c is the ratio of specific heats, and p and V are the instantaneous cylinder pressure and volume, respectively. The procedure for heat release calculation for the closed engine cycle (intake valve closing to exhaust valve opening) is as follows: 1. The volume-crank angle data for the engine is used to calculate the rate of change of volume per crank angle interval of 1 °CA (dV/dh). 2. The rate of change of pressure per degree crank angle (dp/dh) is calculated from the acquired cylinder pressure data. 3. The HRR is then estimated using Eq. (2). A constant value of the specific heat (normally 1.37) or a temperature-dependent value can be used. 4. The cumulative apparent heat release is obtained by summing the incremental values from step #3 over the combustion period. For online processing of the cylinder pressure or for calculation of feedback for fuel injection control purposes, a diesel pressure departure ratio (PDR) algorithm has been shown by the authors to provide fast feedback on the combustion characteristics with sufficient accuracy for a multitude of advanced combustion strategies including HCCI, low temperature combustion (LTC) and post-injection for aftertreatment [15]. The PDR is given by:

PDRðhÞ ¼

pðhÞ þ FPC 1 pmot ðhÞ þ MPC

ð3Þ

where p(h) is the fired cylinder pressure data, pmot(h) is the motored cylinder pressure data, FPC is the ‘fired pressure characterization’ coefficient, and MPC is the ‘motored pressure characterization’ coefficient. The MPC is a correction for the polytropic relationship between the cylinder volume and the motoring pressure while the FPC is a correction for the difference between the change in pressure caused by the volume change, for the fired and the

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motored cycles. The coefficients, FPC and MPC are constants for a given engine configuration and are largely not affected by the boost pressure, EGR, etc. A detailed description of the PDR algorithm including engine validation tests and the procedure for determining the FPC and MPC is given in the authors’ previous publications [15,18]. An estimate of the mass fraction burnt is then obtained by normalizing the PDR from Eq. (3) with its maximum value PDRmax as follows:

MFBPDR ðhÞ ¼

PDRðhÞ PDRmax

ð4Þ

In case of a single-cylinder research engine, the IMEP is the parameter of choice for power output since the auxiliary equipment such as the fuel injection pump (FIP) are normally sized for the production engine and therefore, the brake power output is of less relevance. The IMEP over a complete thermodynamic cycle is numerically expressed as:

H

IMEP ¼

pav g dV Vd

ð5Þ

where pavg is the average pressure over the crank angle interval, Vd is the displacement volume and dV is the rate of change of cylinder volume over the crank angle interval. The IMEP can be reported either on a gross or a net basis. The gross IMEP takes into account only the work done during the compression and expansion processes, while the net IMEP accounts for the cumulative work output over all the four strokes, including intake and exhaust processes (pumping loop). The exploration of advanced combustion regimes such as LTC and the need for compliance with the diesel emission norms has significantly increased the limits for the intake pressure, exhaust back pressure and EGR. Intake pressures up to 4 bar are now commonly being utilized with up to 70% EGR. To drive such high amounts of EGR, the required exhaust back pressure can be quite high and may significantly affect the pumping work. This can also affect the engine performance calculations as explained later in Section 2.6. Therefore, it is recommended as a good engineering practice to include the pumping loop in the analysis and to report the net IMEP as the default parameter for power output. With modern PCs, the cylinder pressure data can be processed online on a cycle-by-cycle basis to analyze the combustion process that can include displaying the cylinder pressure and rate of pressure rise diagrams, calculation of the heat release rate and other parameters such as IMEP and CA50 (Fig. 1) [12–15]. This concurrent display of the engine performance during testing greatly aids the test engineer to not only exercise a strict control over the test conditions, but also to modify them to explore new, interesting research areas. 2.3. Fuel flow measurement The measurement of the liquid fuel consumption for a single-cylinder research engine is one of the most difficult as well as the most critical of all the measurements. A stable fuel supply temperature is a prerequisite for accurate fuel consumption measurement for the duration of any experiment since diesel fuel has a high coefficient of volu-

metric expansion (5–10 times higher than that of water). This can mean a reduction of 0.1–0.2% in the mass of fuel delivered per degree rise in fuel temperature [2]. Chen [19] examined the effect of fuel temperature on engine performance and emissions in a direct-injection diesel engine both experimentally and numerically. The effects of fuel temperature on injection parameters can be attributed to changes in the bulk modulus and the fuel density. An increase in the fuel supply temperature predicted the nozzle fuel-injection-start timing to be relatively retarded and a reduced rate of injection. Moreover, an increase in CO, HC and smoke emissions was observed at higher fuel supply temperatures [19]. Therefore it is important to regulate the inlet liquid fuel temperature within a narrow band to attain repeatable results. The measurement of the fuel flowrate in a single cylinder setup is further complicated by the following facts: 1. The fuel has to lubricate as well as cool the FIP and injector components; therefore, most of the high-pressure fuel flows back through the return loop. 2. The temperature of the returned fuel from the highpressure system can be as high as 90–100 °C at high injection pressures (>1800 bar). Additionally, the fuel returned from the engine may contain vapor bubbles and pressure pulsations. This implies that the fuel volume and density can vary within the measurement system and this variability is the main source of errors in the flow measurement. Therefore, the returning fuel has to be cooled ideally to the same temperature as that of the supply fuel (control temperature) so that variations in density and therefore, in volumetric flowrate are minimized. 3. At idle/low load operation, the fuel flow can be as low as 0.01 g/s. The fuel system can broadly be divided into the lowpressure and the high-pressure sub-systems. The main components of the low-pressure system are the fuel storage tank, fuel filters, the fuel flow detector, low pressure pump, pressure regulator and the low-pressure inlet/outlet of the FIP. The high-pressure side consists of the FIP output, common rail, connecting fuel lines and the injector. Two options for measuring the fuel flowrate are shown in Fig. 4. The standard configuration commonly used employs a single fuel flow detector in the supply line. The fuel in the return path is cooled to ensure that the fuel return temperature, TR is equal to the fuel supply temperature, TS – the control temperature of the measurement system, thereby minimizing variations in density and therefore, volumetric flowrate. The pressure regulator should be set to provide the requisite inlet pressure, PS for the FIP, generally in the range of 35–100 kPa (5–15 psig). Moreover, the pressure in the return line, PR should be almost zero. Although, a single filter is shown in the figure after the low pressure pump, in practice, it is beneficial to use several filters to dampen out the fluctuations in the flow as well as provide a larger volume in the closed fuel loop. The standard configuration also shows a glass bulb of 30–50 ml capacity. This can serve two purposes: first, the glass bulb can be used to quickly check the accuracy of the fuel flow detector

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Fig. 4. System configurations for fuel flow measurement.

during filling of the bulb. Secondly, the fuel flowrate of the engine can be estimated by switching the fuel source to the filled glass bulb (gravity feed) and then by measuring the time taken for the fuel to be consumed by the engine. An alternate configuration for determining fuel consumption utilizes fuel flow detectors in both the supply and the return lines – the difference between the two flowrates then provides the engine fuel consumption. The calibration of the detectors can also be easily checked with the engine shut-off. This method provides for fast measurement times even with very low fuel consumption. The fuel storage tank should be installed such that even the minimum level of fuel inside the fuel storage tank can provide a pressure head that is greater than the flow losses in the tubing as well as the pressure drop across the fuel flow detector. Furthermore, the glass bulb can only be filled if the fuel is at a higher level inside the fuel tank. The fuel flow detectors commonly provide both an analog voltage as well as a digital pulse output. The voltage, using the detector transfer function gives the instantaneous flowrate while the digital pulse indicates the flow of a fixed amount of fuel per pulse. This information can be summed over a predetermined time period (1 s to 1 min) to provide the average flowrate. 2.4. Intake, exhaust and EGR systems A typical layout for the intake, exhaust and EGR system for a single-cylinder research engine is schematically

shown in Fig. 5. The intake (combustion) air is usually supplied from a clean, oil-free source, for which a dedicated air compressor with adequate storage capacity is generally installed. The compressor system must have an oil separation unit, air filters and water removal unit to provide clean, dry combustion air. It is important to control the moisture content of the combustion air since it can adversely affect the engine performance. A higher moisture content implies less ‘free’ oxygen available for combustion than the oxygen available in a similar volume of dry air under the same conditions of temperature and pressure. Moreover, reduction in the ‘free’ oxygen content becomes more pronounced at higher intake temperatures. Another significant effect of the moisture is on the formation of NOx in diesel engines. If the NOx measurement is done with dry air, EPA & SAE recommend adding corrections to the NOx readings (explained later in Section 2.5). A relative humidity of 30% at 25 °C is specified in the existing standards for combustion air [2]. The pressure in the engine intake system needs to be regulated with a pressure regulator, either of mechanical or electronic type. Electro-pneumatic regulators are commonly used as they provide accurate and quick regulation of the intake pressure. The necessity of a higher intake pressure, for example, for implementing alternate combustion regimes or emission control techniques, dictates a control range of 1–5 bar abs for the intake pressure. Moreover, the control system should be able to vary the intake pressure, typically with a resolution of 0.5–1 kPa.

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Fig. 5. Setup and instrumentation for the intake, exhaust and EGR systems.

Over-pressure protection mechanisms need to be implemented to ensure safety of both the personnel and the equipment. Since the flow in the intake manifold is highly fluctuating, a large surge tank is also required to reduce the cyclic pulsations (dampen the pressure wave action) that would otherwise cause difficulties in the measurement of the flowrate and pressure. The volume of the surge tank is recommended to be at least 50–100 times the engine (cylinder) displacement based on the authors’ experience. The air flow measurement is required for determining the air/fuel ratio, converting the raw emissions in ppm into specific emission values and for estimating the EGR flow/ ratio. While production engines commonly employ hotwire anemometer based mass-air-flow sensors upstream of the turbocharger, these cannot be used under pressurized environments. The common types of flow measuring devices more suited to laboratory use are the venturi, turbine flow and positive-displacement meters such as the roots blower. Since these devices measure the volumetric flow, the pressure and temperature are also measured to convert the volume flowrate into the corresponding mass flowrate. A surge tank in the exhaust system, similar to the one for the intake, is also recommended to reduce the flow pulsations. To enable EGR, a flow restriction valve in the exhaust system is required to control the exhaust backpressure so that a positive pressure differential always exists across the exhaust-intake systems. The EGR loop consists of an EGR cooler and a EGR flow control valve. Three strategies/possibilities exist for controlling the quantity of EGR as follows: 1. A constant exhaust backpressure is maintained (generally up to 20 kPa) and the EGR quantity is changed by adjusting the opening of the EGR valve. Since the exhaust backpressure is primarily affected by the EGR valve opening, exhaust temperature, engine speed and load, this requires frequent adjustments of the exhaust backpressure valve. Furthermore, since the typical EGR valves operate with a stepped response, a small adjustment in the EGR (1–2%) is generally not possible.

2. The EGR valve opening is fixed (typically 15–25%) and the EGR quantity is changed by adjusting the exhaust backpressure. This is the simplest solution for controlling EGR since the backpressure valve is the only element to be adjusted. Based on the authors’ experience, EGR can be adjusted to within 1% resolution. For steadystate testing including sweeps of EGR or injection timing, this option provides reliable control over the EGR. 3. The quantity of EGR is changed with a combination of the EGR valve opening and the exhaust backpressure. Although the control system is complex (a closed-loop automated system is generally required), this option is necessary if the turbocharger response needs to be replicated during the tests or transient engine testing is required to be done. A flow restriction device (not shown in the figure) is also sometimes installed after the intake surge tank. This can serve two purposes. First, it allows simulating the pressure drop across an air-filter so that its effect on the engine performance can be studied. Second, it can be used to enable EGR with low-pressure EGR loops or to create a pressure differential between the exhaust and the intake in case of restrictions on the maximum back-pressure, or the non-availability of a backpressure valve. A number of pressure and temperature sensors are required for monitoring the intake, exhaust and EGR systems as well as to provide the necessary feedback for the EGR control system. These are explained in a later section. The amount of EGR is generally evaluated with sufficient accuracy by measuring the carbon dioxide (CO2) concentration in the intake and exhaust [16–18,20] as follows:

EGRFraction ¼

ðCO2 Þintake  ðCO2 Þatm Intake CO2  ðCO2 Þexhaust  ðCO2 Þatm Exhaust CO2

ð6Þ

2.5. Exhaust emissions – measurement and analysis The emissions that are presently monitored in the test cells can be divided into regulatory emissions and

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diagnostic emissions. The regulated emissions for diesel engines include oxides of nitrogen (NOx), smoke, carbon monoxide (CO) and unburned hydrocarbon (UHC) in North America as per regulations of Environment Canada and US Environmental Protection Agency (EPA). The measurement of other exhaust constituents such as carbon dioxide (CO2) and oxygen (O2) concentrations in the intake and exhaust can provide valuable information about the combustion process and is primarily for diagnostic purposes. Previous work including that of the authors has shown that regulated emission quantities such as NOx, CO and UHC are a strong function of the intake charge dilution (intake oxygen concentration). The measurement of these diagnostic quantities can help to prevent the engine from entering unstable combustion modes such as high EGR or high soot regions. The exhaust emission measurement in the test-cell consists of two parts; one for smoke measurement and the other for the rest of the emissions (gaseous emissions). The list of emission measurement devices and their working principle is given in Table 1. For gaseous emission measurement, an emission sample conditioning system is an absolute necessity. The purpose of the emission conditioning system is to provide a clear and dry exhaust stream to the analyzers. The conditioning system removes the particulate matter (PM) and water content from the exhaust stream before it reaches the analyzer bench. A suggested layout of the emission analyzing system is shown in Fig. 6. The raw exhaust gas is transported from the sampling location to the conditioning unit in a heated sampling line. The heated sampling line prevents any water/UHC condensation in the line. As a standard practice, it is required to

Table 1 Emission measurement instruments and their working principles. Emission type

Working principle

Carbon monoxide (CO) Carbon dioxide (CO2) Oxides of nitrogen (NO, NO2) Unburned hydrocarbons (UHC) Exhaust oxygen Particulate matter

Non-dispersive infrared analyzer Non-dispersive infrared analyzer Chemiluminescence detector Flame ionization detector Paramagnetic detection analyzer Filter paper method

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clean the sampling line at regular intervals by blowing out with compressed air to prevent clogging with PMUHC–water deposits that may skew the test results. Chemical solvents may also be used for cleaning the sampling lines if permitted by the manufacturer. In the conditioning system, the exhaust gases first pass through the heated filter where all the PM in the exhaust is removed (recommended filter element size is 100 ID  700 length, 0.01 or 1 lm rating). Next, the sample passes through the heated inline pump to the chiller/dryer (refrigeration unit) where any suspended water droplets and moisture are removed. Finally, the clear and dry gaseous sample enters the emission analyzers. For the smoke emissions, no separate conditioning unit is typically used. The smoke meter is normally a self-contained unit and does not require any additional conditioning system. All the emission analyzers generally require no back pressure at the outlet/exhaust port which should be ensured to avoid erroneous emission readings. The heated sampling line, the heated pump and the heated filter are typically maintained at a temperature of around 190–200 °C (375–400 °F). The heated filter element should be replaced after about 20 h of operation under moderately sooty conditions (an engine operating with smoke less than 2.0 FSN can be considered to be moderately sooty). The filter should be changed more frequently if the engine exhaust contains more soot. Although the selection of the analyzers is outside the purview of this work, however, it is important to highlight the emission reporting for UHC emissions. In North America, the term non-methane hydrocarbons (NMHC) is often used to distinguish from the total hydrocarbons (THC) that also account for the methane concentration in the exhaust (THC = NMHC + methane concentration). However, in Europe and Japan, the reported UHC emissions are the total hydrocarbons. Therefore, the selection of the HC analyzer is dependent on the type of measurement being performed – NMHC or THC. The range of these emission analyzers depends on the type of the combustion research being performed. The measurement range required for conventional diesel tests can be different by an order of magnitude from other combustion modes such as LTC. For conventional diesel

Fig. 6. Typical exhaust measurement system showing exhaust conditioning and analyzer system.

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Table 2 Suggested emission analyzer ranges. Emission

Prior to Euro V or US EPA 2007

Euro VI or EPA 2010 without aftertreatment

Euro VI or EPA 2010 with aftertreatment

CO CO2 NOx (NO, NO2) UHC Exhaust O2 PM

0–500 ppm 0–20% 0–3000 ppm 0–500 ppm 0–21% 0–5 FSN

0–20,000 ppm 0–20% 0–1000 ppm 0–3000 ppm 0–10% 0–10 FSN

0–1000 ppm 0–20% 0–1000 ppm 0–500 ppm 0–21% 0–5 FSN

combustion, the NOx emissions are moderately high with low smoke, CO and UHC emissions. However, for LTC it is possible to attain simultaneous low-NOx and low-smoke emissions, but the CO and UHC emissions are significantly higher (as high as 1% for CO and 5000 ppm for UHC). A suggested range for the various emission analyzers is given in Table 2. 2.5.1. Estimation of indicated/brake-specific emissions The emission values reported by the emission analyzers are typically the raw emissions in ppm (volumetric percentage of the exhaust). However, these values need to be expressed in terms of mass-based specific emissions (g/kW h or g/hp h) for legislative considerations. Moreover, since it is relatively easier to make a measurement from a dry sample of the exhaust gas, therefore, most of the emission analyzers report raw emissions on a dry sample basis (hence the need for the dryer unit in the sample conditioning system). However, the difference between the wet and dry emission values can be up to 10% as shown in Fig. 7. The important modifiers to the raw (dry) emissions are the humidity and the wetness correction factor (accounting for the moisture contents of the exhaust gas). As per the emission regulations, these raw values need to be corrected for the moisture in the exhaust gas and the ambient humidity to get the wet emission values. Most of the automotive companies in North America follow the procedure given in Electronic Code of Federal Regulations Title 40, Part 89 created by the US EPA for reporting specific emissions on a wet basis [21]. The important steps in estimating the specific emissions are as follows [21]: 1. Conversion of dry emission concentration to wet emission concentration: This is done by estimating the correction factor Kw which relates the wet and dry

emissions as Concwet = Kw  Concdry. This correction needs to be applied to all the concentrations that are measured on the dry-basis. The term Kw is evaluated as shown below: a. Obtain the dew point temperature and estimate the partial vapor pressure, Pv (kPa). b. Obtain the barometric pressure Pamb (kPa). 2. Calculate the humidity ratio (H)

H¼

622  Pv ðg of vapor per kg of dry airÞ P amb  Pv

ð7Þ

3. Calculate the dry fuel–air ratio (FAR)

FARdry ¼

M fuel M air;dry

FARdry ¼

Mfuel M air;wet  ð1  H=1000Þ

ð8Þ

4. Calculate the hydrogen mass percentage of the fuel (ALF). If a is the H/C mole ratio of the fuel, then

ALF ¼

1:008a 12:01 þ 1:008a

ð9Þ

5. Kw is then given by

K w ¼ ð1  FFH  FARdry Þ  K w1

ð10Þ

where

  1 FFH ¼ ALF  0:1448 1 þ FARdry and

 K w1 ¼ 1:608

H 1000 þ 1:608H



6. For NOx emissions, EPA recommends an additional correction based on the intake air temperature and humidity given by the following formula:

KH ¼

1 1  0:0182ðH  10:71Þ

ð11Þ

7. For NOx, CO and HC emissions measured on a dry-basis, the corrected wet values for all the three species are calculated as follows:

NOxwet ¼ K w  K H  NOxdry COwet ¼ K w  COdry Fig. 7. Difference in NOx values based on the dry-sample or wet-sample basis.

HCwet ¼ K w  HCdry

ð12Þ

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8. To obtain the specific emission values, the mass flowrate for each species is calculated as follows:

M exh;wet NOxwet  1000   MW NOx MW exh 106 Mexh;wet COwet CO ðg=hÞ ¼  1000   MW CO MW exh 106 M exh;wet HCwet HC ðg=hÞ ¼  1000   MW HC MW exh 106 NOx ðg=hÞ ¼

ð13Þ

_ f Þ þ Fresh Air ðm _ a Þ þ EGR ðm _ EGR Þ Reactants : Fuel ðm _ f þm _ aþm _ EGR Þ Products : Exhaust ðm The following calculations are shown for a hydrocarbon fuel with no oxygen (c = 0) , i.e. CaHb. It is assumed that all the carbon in the reactants comes from the fuel and therefore, the mass of carbon in the fuel can be written as:

_f Mass of Carbon in Fuel ¼ m

where Mexh,wet is the mass flowrate of exhaust in kg/hr, MWexh is the molecular weight of exhaust calculated as mole-fraction average of the component molecular weights, MWCO is the molecular weight of CO, MWHC is the molecular weight of fuel using the composition of fuel as C1Ha, MW NOx is the molecular weight of NO2. NOx is treated as NO2 as per EPA guidelines. 9. The emission flowrates in (g/h) are divided by the indicated or brake power in kW or hp to obtain the indicated-specific or brake-specific emissions respectively.

The mass of carbon in the exhaust can be written as follows:

Mass of Carbon in Exhaust _ f þm _ am _ H2 O Þ  ðyCO2 þ yCO þ ayHC Þ 12:011ðm ¼ MW exh Combining these expressions, the carbon mass-balance for the combustion is given by the following equation:

am_ f MW f

2.6. Carbon mass-balance for estimating fuel flowrate The measured exhaust emissions can be used to provide a fairly accurate estimate of the engine air/fuel ratio and the fuel flow rate. The calculation of the air/fuel ratio from the composition of the exhaust gases is based on the general equation for the complete combustion of a hydrocarbon fuel with air, suitably modified to account for CO and HC production. It is often assumed that air is a simple mixture of oxygen and nitrogen. Actually, air contains a number of other gases, notably about 0.934% argon and 0.035% CO2. The inert gases do not take part in the combustion and can be lumped together with the nitrogen. The CO2 quantity is small and ignoring it causes a 0.2% error in the calculated results [22]. The general equation for the combustion of a hydrocarbon fuel with air, modified to account for the actual exhaust composition (CO and HC only) can be written as:

  b c Ca Hb Oc þ k a þ  ðO2 þ 3:76N2 Þ 4 2 ! CO2 þ H2 O þ ðCO þ Hb Ca Þ   b c þk aþ  3:76N2 þ O2 4 2

_ f þm _ am _ H2 O Þ  ðyCO2 þ yCO þ ayHC Þ ðm MW exh

ð15Þ

MW exh ¼ MW CO2  yCO2 þ MW CO  yCO þ MW HC  yHC þ MW O2  yO2 þ MW N2  yN2 Similarly, the hydrogen balance can be written as:

_f _ f þm _ am _ H2 O Þ  bðyHC Þ 2mH2 O bm ðm ¼ þ MW f 18:016 MW exh

ð16Þ

Solving Eqs. (15) and (16) and after necessary simplifications, an estimate of the air/fuel ratio is arrived at as:

¼

where k is the air excess ratio. Note that the mole quantities of carbon, hydrogen and oxygen are not balanced in Eq. (14). The carbon contained in the fuel that is added to the system should be quantitatively equal to the carbon that is contained in the engine exhaust. Therefore, by measuring the actual concentrations of the CO2, CO and HC in the exhaust, the air/fuel ratio and the fuel flowrate can be determined. It is important to note that the application of EGR results in the recirculation of the exhaust gas back into the intake stream and the composition of the EGR is the same as that of the exhaust gases. Therefore, the application of EGR does not affect the calculations using the carbon balance method.

¼

_ f is the mass flow rate of fuel (g/s), m _ a is the mass where m _ H2 O is the mass flow rate of flow rate of fresh air (g/s), m water removed from the exhaust sample (g/s), yi is mole fraction of the exhaust species ({i = CO2, CO, HC}), MWf is the molecular weight of the fuel and equal to 12.011a + 1.008b (g/mol), MWexh is the molecular weight of the exhaust gas (g/mol) and equal to

AFR ¼

ð14Þ

12:011a MW f

_a m _f m 9:008b  ðyCO2 þ yCO Þ þ a  MW exh  MW f  ðyCO2 þ yCO þ ayHC Þ MW f  ðyCO2 þ yCO þ ayHC Þ

ð17Þ Using the measured fresh air mass flowrate, the fuel _ f can be estimated from Eq. (17). flowrate, m The moisture in the ambient air is typically 1% and to account for the moisture, the dew point temperature is normally measured in the test-cell, which gives the partial-pressure of water-vapor (pvap). Knowing the ambient pressure (pamb), the moisture in the ambient air can be estimated by calculated the mole fraction of water-vapor as pvap/pamb. It is important to highlight here that the carbon massbalance algorithm presented here is an approximate method, since the accuracy of the fuel flowrate estimation using this method is dependent on the fuel species that show up in the engine exhaust. Any fuel (or UHC) that is drained to the oil sump, gets condensed and collected in the sample conditioning system will not reach the emission analyzers

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and therefore, will NOT be included in the estimation. Therefore, it is important to realize that the calculated fuel flowrate will always tend to be underestimated or at best, approach the actual flowrate. Therefore, caution should be exercised when reporting performance metrics such as the thermal efficiency since the underestimation in the fuel flowrate tends to erroneously exaggerate the engine performance. This is all the more important when the engine performance is being gauged over the compression and expansion processes only, as in the case of gross IMEP that can further exacerbate the errors in the performance metrics. It is also pertinent to mention here that the carbon mass-balance method is more suited for exploration/research strategies where the emphasis is on achieving a substantial improvement (10–20%) in the engine performance parameters. This method is not recommended for performance benchmarking tests where an absolute accuracy of typically 1% or less is normally required. A more comprehensive treatment of the carbon mass-balance method can be found in the literature [22,23]. 2.7. Temperature and pressure measurements In a well-instrumented engine, pressures and temperatures are monitored at all the important locations. The temperature and pressures can be classified into two categories – functional and safety. The safety related temperature and pressures include the oil temperature and pressure, whereas the intake temperature and pressure fall under the list of functional category. A list of critical temperatures and pressures along with the suggested ranges is given in Table 3 for reference. 2.7.1. Intake system For the single-cylinder engine, the boost and the EGR systems are de-coupled and can be implemented independently. The manifold air temperature and the manifold air

Table 3 Suggested range for critical temperatures & pressures. Parameter Intake System a. Engine Inlet Air

b. Dew point temperature EGR System a. EGR Gas Inlet b. EGR Gas Outlet c. EGR Cooler Liquid In d. EGR Cooler Liquid Out Exhaust Gas Safety a. Lubricating Oil b. Engine Coolant c. Fuel

Suggested Range Manifold Air Temperature < 80 °C for conventional diesel (<200 °C for diesel HCCI / LTC) Manifold Air Pressure < 5 bar

TEGR,in < 800 °C TEGR,out < 140 °C TC,in < 90 °C TC,out< 200 °C Texh < 900 °C Toil < 120 °C Poil < 7 bar Tcoolant < 100 °C Pcoolant < 2 bar Tfuel,in < 40 °C Pfuel < 2 bar

pressure are the important parameters that must be monitored. The manifold temperature can have a significant impact on the combustion and emissions. The minimum manifold temperature at high EGR rates must be limited to prevent water condensation in the intake manifold. The lowest limit of manifold temperature is decided by the condensation limit and if the intake temperature is cooled below the condensation limit, water formation in the intake manifold occurs. It is therefore important to monitor the dew point temperature to prevent condensation and also to implement corrections to the raw emissions. 2.7.2. EGR system For the EGR system, the gas temperature and pressure at the entrance and exit of the EGR cooler are monitored. The desired EGR stream temperature at the exit of the EGR cooler is decided primarily by the intake temperature considerations. For conventional high temperature diesel combustion, the manifold air temperature has a significant impact on the emissions. For the same EGR, both the NOx and smoke are lower at lower intake temperatures. The intake temperature requirements for homogeneous charge compression ignition (HCCI) combustion or LTC are more varied than for the conventional diesel combustion and at low load HCCI, the intake temperature should be high to improve the combustion stability whereas at higher loads, a lower intake temperature helps to prevent premature auto-ignition. Three strategies are generally employed to regulate the temperature of the EGR stream (see Fig. 5): – Bypass the EGR cooler at very light-loads – at very lightloads such as idle operation, no EGR cooling may be required and the use of hot-EGR may help to stabilize the combustion process. The presence of a by-pass loop can readily help the implementation of this feature. – EGR cooler with engine-coolant as the cooling medium at mid-to-high loads – this is the typical EGR cooling method in a conventional production engine and replicates the multi-cylinder engine operation. – EGR cooler with city water (or chilled coolant) as the cooling medium at mid-to-high-loads – useful to study the effects of cooled EGR on combustion and emission performance. A high pressure-drop in the EGR cooler indicates clogging of the EGR cooler and typically results in degraded cooling performance. To reduce clogging, a small section of a diesel oxidation catalyst can be added to the EGR loop (before the cooler) to oxidize the UHC and CO in the recycled exhaust as shown in Fig. 5. 2.8. Lubricating and coolant systems The coolant and lubricating oil conditions can have a significant effect of the engine performance including the fuel consumption and engine-out emissions. It is therefore important to tightly regulate the coolant and lubricating oil conditions to improve the repeatability of test results. For the lubricating oil, both the temperature and pressure

U. Asad et al. / Measurement 44 (2011) 1261–1278

needs to be regulated during the engine tests (oil temperature within ±1 °C and the oil pressure within ±0.1 bar). The coolant conditioning system should also be able to maintain the coolant temperature at the inlet of the engine within ±1 °C of the set-point temperature at steady-state engine operating condition. The availability of an overflow bottle in the coolant circuit allows for the thermal expansion of the coolant when the coolant temperature increases. It also presents a location where the coolant may conveniently be added to the system. The cooling system is generally pressurized from 0.5 to 1 bar (8–15 psi). There are two approaches to setting the values of the coolant temperature as follows: In the first approach, the coolant temperature is set at a fixed value independent of engine speed and load condition. A typical value for the coolant temperature at the engine inlet is 80–90 °C. The second approach is to replicate the multi-cylinder engine operation where the coolant temperature at the engine inlet is allowed to vary as a function of the engine load and speed. In this method, the single-cylinder engine is operated at the peak power condition of the representative multi-cylinder engine. At this engine operating condition, the ball valve for the cooling water valve is adjusted to achieve the desired coolant temperature at the engine inlet as shown in Fig. 8. The cooling water valve is then kept fixed at this position for all other engine operating conditions. Therefore, the coolant temperature varies at the engine inlet during the engine test, similar to the actual multi-cylinder engine. Similarly, the engine can be run at a fixed engine oil temperature and pressure or it can be varied as a function of engine operating condition. An oil pressure of about 4 bar and a temperature of about 80–90 °C can be taken as representative values if detailed oil temperature and pressure maps are not available. 2.9. Fuel injection system A brief description of the fuel control system is presented here. Other hardware details about the fuel injection and delivery system can be found in numerous references such as the ‘‘Bosch Handbook for Diesel-Engine

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Management’’ [24]. A general layout of the fuel injection and control system along with the location of all the feedback sensors and control actuators is shown in Fig. 9. The fuel injection system requires the implementation of separate control systems for the common-rail FIP and the piezo/solenoid injector. The FIP generally has two valves called the ‘Pressure Control Valve’ (PCV) and the Volume Control Valve (VCV) to regulate the pump output pressure. The PCV and VCV are solenoid actuated valves and the operating voltage is mostly 12 V DC. The PCV and the VCV are pulse-width modulated valves where the duty-cycle of the pulsetrain determines the opening of the valve. To control the fuel pump, a practical application is to fix the value of the VCV to about 30% opening and vary the PCV to control the pressure. The 30% opening of the VCV ensures that the pump has enough fuel at all times to raise the pressure as well as to provide adequate cooling of the pump components as the fuel can heat up to 100 °C when the rail pressure exceeds 1800 bar. The PCV can be satisfactorily controlled using a Proportional–Integral– Derivative (PID) or a PI controller with the rail pressure sensor as the feedback. The frequency of the digital pulse train will depend on the manufacturer but a frequency of 200–300 Hz for the VCV and about 1–1.2 kHz for the PCV will generally provide acceptable controllability. The duty-cycle of the control pulsetrain should be updated at a speed of 1–40 kHz, to achieve a stable output pressure. Since these valves can draw about 1–3 Amps of current, therefore, a driver circuit such as H-bridge is used to convert the control pulsetrain from the data acquisition card (0–5 V TTL) into a high power 0–12 V pulsetrain. It is preferable to use a deterministic system (real-time) to ensure safety. It is advisable to limit the rate of change in the setpoint as well as the maximum attainable pressure during the tuning of the PID controller to prevent damage to the equipment and to ensure safety of the personnel. The fuel injection control is much more complex since both the geometric and the time domains have to be tracked – the start of injection is decided based on the geometry (crank angle) and the time for which the injector remains opens is usually measured in micro seconds (time domain). A minimum of three signals are required to ascertain the absolute as well as the relative position of the

Fig. 8. Schematic for the coolant system.

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Fig. 9. Schematic for the fuel injection and control system.

crankshaft in case of 4-stroke engines. The Index Z provides the absolute position reference (TDC or BDC) while the Index A provides the relative position of the crankshaft with respect to the Index Z. The third signal is needed to identify the compression stroke so that the injection is not done at the end of the exhaust stroke (corresponding to the exhaust TDC). This is usually done by using a cammounted hall-effect sensor which provides a signal at the beginning of the compression stroke. An in-cylinder pressure sensor can also be used to identify the compression stroke but should be used as a last resort. The fuel injector cannot be operated directly using the data acquisition hardware and therefore, a special power drive is required which would generate the required voltage and current (voltage 6200 V, current 620 Amps) to open the injector. A TTL pulse train based upon the encoder Index A, Index Z and the cam sensor is generated. The rising edge of the TTL pulse should correspond to the crank angle at which the injector is commanded to open (SOI) and the width of the pulse is the duration for which the injector is desired to remain open (injection duration). The actual injector opening slightly lags the commanded setpoint due to inertia of the mechanical components. Similarly, the injector closing also takes a finite time. To ensure absolute determinism in the injection control, a real-time (RT) operating system is essential (RT-FPGA system is recommended) to implement the fuel injection control. The authors have successfully used a real-time controller embedded with a FPGA device to generate the necessary control signals for dynamically and precisely commanding the fuel injection strategies with a control resolution of 0.1 °CA (0.0056 ms at 3000 rpm). The system can be configured to provide up to 12 independent injections/cycle (injection timing, pulse width) and generates a TTL pulse

train that serves as the input for the injector power drive. The details can be found in the authors’ previous publications [18,25,26]. The necessity for such a precise control over the fuel injection events is highlighted by the test results shown in Figs. 10 and 11. Retarding the SOI by only 0.1–0.2 °CA caused a significant change in the combustion phasing (represented by CA50) and the maximum rate of pressure rise. The NOx emission decreased by 57 ppm and the THC increased by 76 ppm as a result of a 0.2 °CA delay (364.8–365 °CA) in the SOI. Retarding the SOI further by 0.1 °CA (beyond 365 °CA) resulted in engine mis-fire. The benefits of a fine resolution for the injection timing are obvious once the engine is operated at the borderlines of combustion exploratory regimes, for example, during unstable operation frequently encountered in LTC modes.

Fig. 10. Sensitivity of combustion to the fuel injection timing.

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Fig. 11. Effect of injection sensitivity on the exhaust emissions.

2.10. Data acquisition system integration The principal function of an engine research and testing laboratory is to produce high-quality, reliable data. The collection, validation, manipulation, display and storage of these data should be prime considerations in the design and operation of such a laboratory. Compared to a commercial all-in-one ‘black-box’ system, a custom-built system for data acquisition and control offers not only a high degree of flexibility in operation (logging frequency, sampling rates, displays) but also allows for easily adapting the system to different testing requirements [18]. A proposed engine test system, based on the authors’ own laboratory setup is shown in Fig. 12.

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The critical system parameters related to engine safety can be grouped together to provide both audible and visual warnings once set limits are exceeded. These parameters will usually include the temperatures across the engine and its sub-systems, fuel flowrate, fuel, oil and coolant pressures, and engine speed and load (from the dynamometer controller). The emissions, flowrates and boost/backpressure parameters can be acquired, displayed and recorded on separate PCs or grouped together, if limited by the availability of hardware. However, it is advisable to carry out cylinder pressure acquisition and online analysis of the combustion on separate systems because of the high rate of data throughput. Each PC can and should record the data independently, duly date- and time-stamped with proper annotation of the information. However, it is of utmost importance to not only synchronize the data among the different PCs but also to gather it at one location for ease of use at a later stage. The ‘Sync manager’ should time-stamp the data with a common reference (PC’s internal clock), log the data both continuously (for example at 1–2 Hz) and whenever a snapshot record of all data is required corresponding to a discrete operating point (by pressing a record button on the ‘Sync Manager’, for example). It can also provide for computation of derived parameters and analysis of the raw data as well as graphical and/or tabulated plotting of these data. The authors’ have used the data socket/shared variable technology built into National Instruments’ LabVIEW software to successfully implement such a system [18]. It is also pertinent to mention here that output from commercial systems used for engine fueling control can

Fig. 12. Integrating the data acquisition systems.

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also be easily incorporated into the users’ own data acquisition system. 2.11. Measurement accuracy and calibration frequency The reliability and accuracy of any test result is a direct consequence of the accuracy of the instruments used to measure the operating parameters. The accuracy of a measurement system is a combination of the accuracy of the data acquisition device, the sensor/transducer (and any other device like an amplifier), and how precisely that

instrument has been calibrated. The US EPA recommends an overall calibration accuracy and calibration frequency (inclusive of the data acquisition system) for the important engine test parameters as given in Table 4 [21]. 2.12. Multi-cylinder engine conversion to single-cylinder operation A single-cylinder engine is typically coupled with a motoring dynamometer that can ‘motor’ the engine for start-up, for conducting friction tests or for research work

Table 4 EPA recommended calibration accuracy & frequency. Item

Calibration accuracy (inclusive of data acquisition system)

Calibration frequency

Engine speed Torque Fuel consumption (raw measurement) Air consumption (raw measurement) Coolant temperature Lubricant temperature Exhaust gas temperature Air inlet temperature Atmospheric pressure Humidity (g of water/kg of dry air) Fuel temperature HC analyzer CO analyzer NOx analyzer CO2 analyzer

±2% Larger of ±2% of point or ±1% of engine maximum ±2% of engine MAXIMUM ±2% of engine maximum ±2 °C ±2 °C ±15 °C ±2 °C ±0.5% ±0.5 ±2 °C ±2% ±2% ±2% ±2%

30 days 30 days 30 days As required As required As required As required As required As required As required As required Monthly or as required Once per 60 days or as required Monthly or as required Monthly or as required

Fig. 13. Schematic representation of the single-cylinder engine setup with independent control on intake boost and exhaust backpressure at University of Windsor.

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with the engine not firing. The most commonly used motoring dynamometers are the DC or AC type. Typically, single cylinder versions of multi-cylinder engines are quite expensive. However, a multi-cylinder engine can also be converted into an equivalent single-cylinder engine as shown in Fig. 13 [18,27]. A modern four cylinder Ford Puma diesel engine was modified by the authors to run in the single-cylinder mode using a three cylinders–to–1 cylinder configuration. The three cylinders–to–1 cylinder configuration is an exploration strategy that enables investigation of unstable combustion regimes like low temperature combustion with a non-motoring dynamometer. The three cylinders are operated in the conventional high temperature combustion mode at low load for stable engine operation. The combustion in cylinder #1 is then pushed into the low temperature combustion cycles by independently controlling the EGR, boost, exhaust backpressure and fuel scheduling for cylinder #1. An alternate to this strategy can be the use of a motoring dynamometer with the three cylinders under motoring condition and only the first cylinder is fired during the tests. The engine was modified by separating the cylinder #1 from the other three cylinders. This requires the intake and exhaust system of the single cylinder to be separated from the other three cylinders. The modifications were carried out so that the engine can operate under both one- and four-cylinder modes. A new intake manifold was made with provision for running cylinder #1 with either an independent intake system or using the same intake as for the rest of the cylinders. A similar provision was also made on the exhaust side by separating the exhaust stream of cylinder #1 from the rest of the cylinders. This is necessary since the exhaust gas from the single cylinder is required for measuring the emission performance during the tests [27].

3. Concluding remarks The single-cylinder engine is a reliable and efficient platform for carrying out high-quality combustion research that exhibits good correlation to multi-cylinder tests. It also allows the test engineer to explore engine performance outside the standard engine operation regime imposed by the production ECUs. In this work, the authors have documented the essential instrumentation required to establish a high-quality diesel engine research laboratory. Taking a single-cylinder diesel engine coupled with a dynamometer as the fundamental building block, the cylinder pressure, fuel flowrate, exhaust emissions, intake, exhaust and EGR systems have been described in detail along with the recommended instrumentation and measurement techniques. Detailed schematics for the essential engine sub-systems and common testing practices adhered to by engine research laboratories in Government, academia and the automotive industry have also been laid out. A data synchronization system to integrate the various data acquisition devices, and to efficiently display and record the engine test data has also been suggested. The main

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emphasis of this work has been on providing the basic guidelines for any engineer to successfully setup a highquality, fully-instrumented, single-cylinder diesel engine research laboratory. Acknowledgments The research at the Clean Diesel Engine Laboratory is sponsored by the Canada Research Chair program, NSERC, CFI, OIT, AUTO21, the University of Windsor, Ford Motor Company and other OEMs. References [1] R.D. Atkins, An Introduction to Engine Testing and Development, SAE International, USA, 2009. [2] M.J. Plint, T. Martyr, Engine Testing: Theory & Practice, third ed., Butterworth Heinemann, UK, 2007. [3] O. Laguitton, C. Crua, T. Cowell, M.R. Heikal, M.R. Gold, The effect of compression ratio on exhaust emissions from a PCCI diesel engine, in: 19th International Conference on Efficiency, Cost, Optimization, Simulation and Environmental Impact of Energy Systems, Crete, Greece, 2006. [4] AVL List GmbH, AVL Single Cylinder Research Engines, 2010. [5] B. Challen, R. Baranescu, Diesel Engine Reference Book, Butterworth Heinemann, UK, 1999. [6] C.F. Taylor, The Internal Combustion Engine in Theory and Practice, Combustion, Fuels, Materials, Design, vol. 2, The MIT Press, USA, 1985. [7] C.A. Amman, Cylinder-pressure measurements and its use in engine research, SAE Paper 852067, 1985. [8] B.D. Hsu, Practical Diesel-Engine Combustion Analysis, SAE International, USA, 2002. [9] D.R. Rogers, Engine Combustion: Pressure Measurement and Analysis, SAE International, USA, 2010. [10] R.S. Benson, N.D. Whitehouse, Internal Combustion Engines, Robert Maxwell-Pergamon Press, UK, 1979. [11] B.R. Brown, Combustion Data Acquisition and Analysis, M. Eng. Project Report, Loughborough University, UK, 2001. [12] U. Asad, M. Zheng, D.S-K. Ting, R. Kumar, S. Banerjee, G.T. Reader, J. Tjong, Real-time heat release analysis towards on-fly combustion control for diesel engines, in: Proceedings of Combustion Institute/ Canadian Section, 2006 Spring Technical Meeting, University of Waterloo, Canada, 2006. [13] M.F.J. Brunt, H. Rai, A.L. Emtage, The calculation of heat release energy from engine cylinder pressure data, SAE Paper 981052, 1998. [14] T.J. Callahan, D.M. Yost, T.W. Ryan III, Acquisition and interpretation of diesel engine heat release data, SAE Paper 852068, 1985. [15] U. Asad, M. Zheng, Fast heat release characterization of a diesel engine, Int. J. Thermal Sci. 47 (2008) 1688–1700. [16] J.B. Heywood, Internal Combustion Engine Fundamentals, McGrawHill, USA, 1988. [17] R. Stone, Introduction to Internal Combustion Engines, Palgrave Macmillan, UK, 1999. [18] U. Asad, Advanced diagnostics, control and testing of diesel low temperature combustion, PhD Dissertation, University of Windsor, Canada, 2009. [19] G. Chen, Study of fuel temperature effects on fuel injection, combustion, and emissions of direct-injection diesel engines, J. Eng. Gas Turbines Power 131 (2009) 022802 (8 pages). [20] C.E. Bowen, An experimental investigation into the use of exhaust gas recirculation for diesel engine NOx control, PhD Thesis, University of Calgary, Canada, 1998. [21] EPA Electronic code of Federal Regulations, Title 40, Part 89, Control of Emissions from New and In-use Nonroad Compression-ignition Engines, . [22] W.M. Silvis, An algorithm for calculating the air/fuel ratio from exhaust emissions, SAE Paper 970514, 1997. [23] J. Brettschneider, Extension of the equation for calculation of the airfuel equivalence ratio, SAE Paper 972989, 1997. [24] Robert Bosch GmbH, Diesel Engine Management, fourth ed., Wiley, UK, 2006.

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[25] M. Zheng, Y. Tan, G.T. Reader, U. Asad, X. Han, M. Wang, Prompt heat release analysis to improve diesel low temperature combustion, SAE Paper 2009-01-1883, 2009. [26] U. Asad, M. Zheng, Efficiency & stability improvements of diesel low temperature combustion through tightened intake oxygen control, SAE Int. J. Engines 3 (2010) 788–800.

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