Diesel injection system modelling. Methodology and ...

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Diesel injection system modelling. Methodology and application for a rst-generation common rail system R Payri1*, H Climent1, F J Salvador1 and A G Favennec2 1CMT-Motores Termicos, Universidad Politećnica de Valencia, Valencia, Spain 2PSA Peugeot-Citroe¨ n, La Garenne-Colombes, France

Abstract: This article details a method for modelling the most critical parts of an injection system. It focuses on the most important component of the system, the injector itself. As a clear example of this methodology, the modelling of a rst-generation common rail injection system is carried out using a commercial code. The proposed methodology for modelling the injection system is based on two types of characterization: a detailed dimensional characterization and a hydraulic characterization of the di erent internal parts of the injector. The dimensional characterization is based on the use of a ne detail measuring technique applied to all the constituents of the injector. These include the passages and internal lines of the injector, internal volumes, calibrated ori ces, nozzle springs, clearances between moving sections of pistons, etc. The second type of characterization makes reference mainly to the hydraulic characterization of the nozzle and injector control ori ces, which together with dimensional information makes it possible to determine the discharge coe cient. In this case, special emphasis is placed on the detection of critical cavitation conditions and repercussions of this on the ow. This is a typical phenomenon in control ori ces and also in nozzles subject to strong pressure gradients. Once the model is obtained, it is tested and validated. Following this, the values of the experimental injected mass and rate of injection at di erent operating points are compared with the model results. Keywords: diesel, injection, modelling, cavitation NOTATION A CN C c C d d DP D m G I L1 n N1, N2, N3 O OV1, OV2, OV3, OV4

surface of the outlet ori ce cavitation number contraction coe cient discharge coe cient diameter of the spire of a spring pressure di erence DP =P P 1 2 mean diameter of a spring shear modulus inlet ori ce line between rail and injector holder considered in the model number of spires lines feeding to the nozzle considered in the model outlet ori ce variable-section ori ces considered in the model

The MS was received on 7 May 2003 and was accepted after revision for publication on 4 September 2003. * Corresponding author: CMT-Motores Termicos, Universidad Politećnica de Valencia, Camino de Vera s/n, Valencia 46022 Spain. D07903 © IMechE 2004

P1, P2, P3, P4, P5, P6, PN P v P 1 P 2 R1, R2 Re V1, V2, V3, V4, V5, V6, VN

pistons considered in the model

r

gas oil density

vaporization pressure ori ce upstream pressure downstream pressure (back pressure) springs of the electrovalve Reynolds number volumes considered in the model

Subscripts crit max 1

critical conditions maximum value

INTRODUCTION

The necessity to reduce consumption and emissions in order to comply with increasingly strict regulations means that new methods of research have to be found. Proc. Instn Mech. Engrs Vol. 218 Part D: J. Automobile Engineering

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R PAYRI, H CLIMENT, F J SALVADOR AND A G FAVENNEC

Within these investigations is the study of the injection rate, which greatly in uences the combustion process and pollutant formation [1–3]. In recent years, new electronic command injection systems have been developed that enable the injection rate to be controlled. Numerous studies have been completed to model the behaviour of these last-generation hydraulic injection systems [1–9]. The ow modelling is always based on the momentum and mass conservation, with simpli ed hypotheses and a further relation between pressure and density. The e ect of cavitation and fuel property variations has also been considered in recent work [1–3, 5, 6, 8, 9] allowing more accurate models to be obtained. For example, some papers [2, 3] discuss the treatment of non-linear bulk modulus to include temperature and pressure variations for the two-phase ow. These recent models also require further modelling of elements closely related to the injection system and elements of the electronic control system. From an experimental point of view, it is di cult to obtain information about the internal hydrodynamic behaviour of the system. Using these developed models, if they are su ciently accurate, information can be obtained easily. Furthermore, they make it possible to reproduce injection rates without the use of an injection test bench. Another important point concerning these models is that they can be used to carry out sensitivity studies [1, 4–6 ]. This means that the in uence on the injection rate can be determined both qualitatively and quantitatively, testing di erent design parameters such as volumes, sti ness of springs, control diameter ori ces, etc. The success of obtaining models of injection systems that accurately reproduce reality is founded on using methods that produce a good characterization on both a dimensional and a hydraulic level. The term accurate model means a model that is able to reproduce the same values for the mass owrate and injection rate as measured experimentally within real operation ranges. This accurate characterization is a basic necessity for the correct modelling of the injection system as it produces the input data for the calculation code used for the simulation. The rigour of this characterization, together with the potential of the calculating code, which is able to reproduce physical phenomena and observations, guarantees an accurate model. Even if the models presented in the literature are based on complex equations that take account of pressure wave propagation and fuel properties, none of them explains in detail the dimensional and hydraulical characterization of the fuel injection system which is the base for a suitable model. The present article details a method for modelling the most critical parts of a rst-generation common rail injection system using the potential of a commercial code. The work of modelling focuses on the most important component of the system, the injector itself. The Proc. Instn Mech. Engrs Vol. 218 Part D: J. Automobile Engineering

proposed methodology for modelling of injection systems is based on two types of characterization: a detailed dimensional characterization and a hydraulic characterization of the di erent internal parts of the injector. For the dimensional characterization a ne detail measuring technique applied to all the constituents of the injector is used. Possible dimensional measurements include the passages and internal lines of the injector, internal volumes, calibrated ori ces, nozzle springs, clearances between moving sections of pistons, etc. The hydraulic characterization makes reference mainly to the characterization of the mass owrate for di erent pressure conditions of the nozzle and injector control ori ces, which together with dimensional information allows the discharge coe cient to be determined. In this case, special emphasis is placed on the detection of critical cavitation conditions and repercussions of this on the ow. This is a typical phenomenon in control ori ces and also in nozzles subject to strong pressure gradients. The model of the injector is divided into three parts: injector holder, nozzle and electrovalve. The two types of information, dimensional and hydraulic, constitute the input data for the components of the di erent libraries of the calculation code employed. Once the model is obtained, it is tested and validated. Following this, the values of the experimental injected mass and rates of injection at di erent operating points are obtained in an injection rate test rig, and are compared with the model results. The comparison shows how the model has the ability to predict the injection rate. 2

EXPERIMENTAL TOOLS

The experimental tools used for the dimensional and hydraulic characterization of the injection system are as follows: (a) determination of the internal geometry, (b) the cavitation test rig, (c) the injection rate test rig. As detailed below, each of the above are explained, with special emphasis placed on their accuracy for characterizing injection systems with a view to creating an e cient model.

2.1 Determination of the internal geometry In order to analyse the internal characteristics of the injector, such as volumes, calibrated ori ces, nozzle ori ces and other channels of the injector, a special methodology was followed [10]. This methodology consists of making silicone moulds of the di erent sections of the injector and observing them with a digital camera or D07903 © IMechE 2004

DIESEL INJECTION SYSTEM MODELLING

electronic microscope depending on the size of the sample. With the camera or microscope, a series of photographs are taken and then processed with the help of an image treatment program. To obtain the dimensions of the di erent parts of the injector, computer aided design (CAD) software is used [11]. Pictures of the speci c geometries obtained using the electronic microscope come with a reference dimension (magni cation factor). With this reference dimension it is possible to load the pictures in the CAD software with the appropriate scale factor and from this point obtain accurate dimensions of the part of the injector being characterized. This methodology is applicable when determining the geometry of the volumes and internal channels as well as the dimensions of the control and nozzle ori ces. This will be explained in the characterization of the present system. For example, in Fig. 1, on the left, the mould of the high-pressure tting of the injector holder is shown, with the lines feeding the control volume and the nozzle, and on the right the control volume can be seen, where the upper part of the rod is placed. In this gure the inlet ori ce and outlet ori ce of the control volume can also be seen.

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the grade of cavitation, quantifying this against a cavitation number. The objective of the cavitation test rig is hydraulically to characterize the nozzle in order to determine the coe cient of discharge in cavitating and non-cavitating conditions. The critical pressure conditions at which cavitation appears are determined when choked ow due to cavitation is being produced in the ori ces of an injector with steady ow conditions, as the one-dimensional theory [12] establishes. For this purpose, the injector nozzle to be tested is installed in a modi ed injector holder, whose needle can be kept at its maximum lift, thus allowing the inlet of the ori ces always to be wide open without the need for an electrical signal input. The injection pressure is controlled by means of the pressure regulator of the standard injection system and can be xed at di erent values, up to a maximum depending on the maximum mass ow that the pump is able to supply. The injector is mounted in a 1-litre vessel lled with fuel, where pressure is controlled, xing the injector ori ce outlet pressure. A sketch of the installation is shown in Fig. 2. The pressure di erence across the ori ce can be kept constant and, since the geometry is also xed, steady

2.2 Cavitation test rig The cavitation phenomenon is likely to occur in certain internal parts of the injector, such as the discharge ori ces of the nozzle and the ori ces of the control volume. It is very important to have tools available that allow the characterization of the conditions at which this phenomenon is going to appear and take into account the e ects of this phenomenon on the ow. The cavitation test rig is based on the one-dimensional ow theory established by Nurick in 1976 [12] and proved su ciently valid by Schmidth and Corradini [13]. This theory allows the experimental detection of cavitation and establishes the in uence of this phenomenon on the coe cient of discharge as a function of

Fig. 1 D07903 © IMechE 2004

Fig. 2

Sketch of the cavitation test rig

Samples of silicone moulds taken of the injector Proc. Instn Mech. Engrs Vol. 218 Part D: J. Automobile Engineering

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ow conditions are achieved. In order to carry out the test at di erent pressure di erences, the injection pressure, that is, the pressure at the ori ce inlet, is kept constant and the discharge pressure in the vessel is varied. For wide variation in the ow conditions in the injector ori ce, tests are conducted at di erent values of injection pressure, i.e. an ori ce inlet pressure of 10 and 20 MPa, while pressure at the outlet is varied between a minimum value of 0.1 MPa and the injection pressure. After a short stabilization time, when steady ow conditions are achieved, the mass owrate across the injector is measured. Once the mass owrate is obtained for each of the pressure conditions, it is possible to obtain the discharge coe cient combining the Bernoulli equation and the mass conservation equation b m

C = (1) d Aã 2r(P P ) 1 2 b is the mass owrate, P and P are the upstream where m 1 2 nozzle pressure and the pressure at the ori ce outlet respectively, r represents the liquid density and A is the geometric cross-section of the ori ce which can be determined taking into account the geometric section previously determined by the silicone technique [10]. Cavitation conditions in an injector can be represented by using some of the di erent cavitation numbers proposed in the literature. These are non-dimensional parameters that make it possible to establish whether the relevant ow conditions in the injector nozzle, that is to say, the pressure di erence, are favourable or not to the occurrence of cavitation. De nitions of this parameter vary throughout the literature, but they are mainly based on the pressure di erence across the injector ori ce. In this work the cavitation number de nition used by Soteriou et al. [14] is considered P P 2 CN= 1 (2) P P 2 v where P is the vapour pressure. v For cavitating nozzles, the critical cavitation number is de ned as CN , corresponding to the pressure drop crit for which cavitation starts in the injector ori ce. This phenomenon occurs at a given value of the injection pressure, and it is detected by the stabilization of the mass owrate across the ori ce, in spite of the further decrease in discharge pressure (choking ). Hence, cavitation will not be produced unless the cavitation number corresponding to these pressure conditions is higher than the critical value, CN . crit The discharge coe cient increases with the Reynolds number, Re, up to a certain value of Re (the critical Re), which strongly depends on the ori ce geometry [15]. At this point the discharge coe cient achieves a value close to its maximum (0.95C ), from which the rate of dmax increase is negligible. Taking into account that the Proc. Instn Mech. Engrs Vol. 218 Part D: J. Automobile Engineering

Reynolds number is a function of the ow velocity in the ori ce, and therefore is dependent on the square root of the pressure drop, the Reynolds number and the square root of the pressure drop are directly proportional [16]. Under cavitating conditions, the discharge coe cient only depends on the cavitation number, not on the Reynolds number [14], following the equation C =C d c

S

1+

1 CN

(3)

where C is the contraction coe cient in the ori ce c owing to cavitation phenomena. 2.3 Injection rate test rig The injection rate test rig makes it possible to take measurements of the injection rate in order to compare with the modelling results. This allows the obtained model to be validated. The test rig where the measurements of injection rate were carried out is a commercial system called the injection discharge rate curve indicator (IDRCI ), depicted in Fig. 3, and it enables the display and recording of data describing the chronological sequence of an individual fuel injection event. The measuring principle is the Bosch method [17], which consists of a fuel injector feeding into a fuel- lled measuring tube. The fuel discharge produces a pressure increase inside the tube, which is proportional to the increase in the mass of fuel. The rate of this pressure increase corresponds to the injection rate. A pressure

Fig. 3

Sketch of the injection rate test rig D07903 © IMechE 2004


sensor detects this pressure increase, and a data acquisition and display system further processes the recorded data and renders it visible. A proportional relationship exists between the integral value of the pressure and the amount of fuel injected. This relationship, however, cannot be expressed by means of a constant multiplication factor. Instead, particular relationships exist between the absolute quantity of fuel, the pump rotation speed and the fuel temperature. In order to determine the momentary mean quantity, a cumulative measurement process must therefore take place downstream of the IDRCI. The mean quantity corresponds to the mean surface area of the integral pressure for a recorded sequence of injection events. Three injection pressure values, xed at the common rail, were used: 30, 80 and 135 MPa. The back pressure or pressure at discharge of the injector was 4 MPa. For all the pressure values, four injector energizing times were considered: 0.25, 0.5, 1 and 2 ms. Furthermore, in order to validate the behaviour of the model with multiple injections, four points with pilot and main injection were also tested with di erent pilot and main pulse duration and at di erent rail pressures. The injection rate, fuel injection pressure and energizing time intensity signals were recorded using a YOKOGAWA data acquisition and display system and the in-house developed software TRATASA to obtain the injection rate curves.

Fig. 4 D07903 © IMechE 2004

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PROPOSED MODEL

For the model, the commercial calculation code AMESim was used [18]. Within this calculation code, a series of libraries of mechanical, hydraulic and control elements is available. The combination of these elements allows the simulation of any mechanical–hydraulic system, and in particular these elements allow the simulation of an injection system such as the one detailed in this paper. The proposed model has been divided into three parts: the injector holder, the nozzle and the electrovalve, connected mechanically and hydraulically, as described below.

3.1 Model of the injector holder The proposed model of the injector holder is shown in Fig. 4. In this model a pressure source is considered, simulating the pump. This pressure source feeds a volume of 20 cm3 which represents the rail. The rail is connected to the injector holder by a line ( line L1). At the entrance of the injector holder (high-pressure tting) there is a restriction which simulates the edge lter. Following this, there is a separation into two lines. The rst line feeds directly to the nozzle model ( line N1). The second line feeds the upper part of the injector, leading to the inlet ori ce of the control volume where the upper part of the rod is located. The e ective press-

Model of the injector holder Proc. Instn Mech. Engrs Vol. 218 Part D: J. Automobile Engineering

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In the modelling of the actual spring, the sti ness is obtained from a theoretical formula taking into account the number of spirals and the internal and external diameters. The formula used for the characterization is the same as that used by Bosch [19] Gd4 (4) 8nD3 m where G is the shear modulus, d is the diameter of the spire, n is the number of spires and D is the mean m diameter of the spring. The initial deformation and the pretension force were determined experimentally. The ori ces of the control volume are characterized in two ways as before: K

Fig. 5

Scheme of the upper part of the rod

ure section on the upper part of the rod can be decomposed into two crowns, separated by an ori ce of variable section (OV4) and depending on the rod lift. The pressure on these two crowns is modelled using two pistons connected to volumes 5 and 6 ( V5 and V6), as represented in Figs 4 and 5. Following volume 6, the outlet ori ce of the control volume is found (O), the opening of which is controlled by the command piston (electrovalve). After this ori ce, the return channel is located. For the characterization of volumes 5 and 6, photographs taken with the electronic microscope of the silicone mould were used ( Fig. 1 on the right). These photographs, together with the photograph of the upper part of the rod, make it possible to determine volumes 5 and 6. The pistons P5 and P6 are connected to the rod mass. The model of the rod is considered as two masses, with a total mass equal to the total mass of the rod, separated by a ctitious spring which allows the simulation of the deformation e ects due to the pressure. The upward movement of the upper mass is limited by the value of the maximum lift of the rod, in this case 200 ím. The sti ness of this spring is calculated from a theoretical formula presuming that this spring deforms in the same manner as the rod under the pressure at its extremes.

Fig. 6

stiffness

=

(a) a dimensional characterization, (b) a hydraulic characterization. For the dimensional characterization, photographs from the electronic microscope are taken from silicone moulds containing these ori ces. By processing these photographs it is possible to obtain their exact dimensions. In Fig. 6 there are two photographs, one of the inlet ori ce ( left) and one of the outlet ori ce (right). Also appearing in the photograph of the outlet ori ce is the cone supporting the small sphere that controls the opening of the ori ce by means of the electrovalve. The functional characterization consists of the determination of the mass owrate as a function of the working conditions (the pressure di erence) as well as the critical cavitation conditions. This characterization, together with the dimensional information, allows the determination of the discharge coe cient law as a function of the pressure di erence without cavitation. Furthermore, the critical cavitation conditions (onset of cavitation) and the discharge coe cient behaviour under cavitation conditions can be determined according to the one-dimensional ow theory. For the functional characterization it is necessary to control the upstream and downstream pressures of the ori ces. In order to do this, a special modi cation is made to the cavitation test rig. In particular, a new test rig is made where the part of

Silicone mould of the ori ces of the control volume: left, inlet ori ce; right, outlet ori ce

Proc. Instn Mech. Engrs Vol. 218 Part D: J. Automobile Engineering

D07903 © IMechE 2004


the injector that contains both the inlet and outlet ori ces of the control volume is placed (see Fig. 7). This test rig has three access lines. When individually characterizing the ori ces, one of the three lines has to be blocked (see Fig. 7). For characterization of the inlet ori ce ( Fig. 7, the lower section on the left), the upper access to the test rig is blocked, feeding through the pump and rail by the lateral ori ce where the feed pressure is controlled ( point 1). The volume passes through the inlet ori ce and exits through the lower part (point 2). The lower part is connected to the cavitation test rig where the discharge pressure is controlled. For the characterization of the outlet ori ce (Fig. 7, the lower section on the right), the access to the lateral ori ce is blocked. The test rig is fed from the lower part (point 1) where the feed pressure is controlled. The ow passes through the outlet ori ce which is connected to the cavitation test rig where control of the backpressure is carried out. Figure 8 shows the results obtained for the inlet and outlet control ori ces. The mass owrate as a function of the square root of the pressure di erence is presented in the same gure. The experiment was carried out for two injection pressures, 10 and 20 MPa, and the discharge pressure was varied from these values to atmospheric pressure. The pressure di erence at the point where choking of the mass owrate occurs indicates the onset of cavitation (shown by the vertical dotted lines), and for these conditions it is possible to determine the

Fig. 8

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Hydraulic characterization of the control volume ori ces

critical cavitation number using equation (2). The values for the critical cavitation number obtained at 20 MPa are 1.84 and 1.36 for the inlet and outlet ori ces respectively. With the results of the owrate as a function of the pressure di erence, it is possible to determine the evolution of the discharge coe cient using equation (1). The evolution of the discharge coe cient under cavitation conditions comes from equation (3). The model of the ori ces of the control volume considers the increase in the discharge coe cient as a function of the root of the pressure di erence (or the Reynolds number) up to a maximum value, as well as the decrease [equation (3)] in this value when cavitation occurs. The input data of these ori ces in the model are the geometrical diameter, the maximum discharge coe cient without cavitation and the critical cavitation number. 3.2 Model of the nozzle Figure 9 shows a scheme of the nozzle with the di erent internal lines and volumes. The most important part

Fig. 7

Part that contains the control ori ces (upper part) and test bench for the control ori ce characterization ( lower part)

D07903 © IMechE 2004

Fig. 9

Scheme of the nozzle lines and volumes considered Proc. Instn Mech. Engrs Vol. 218 Part D: J. Automobile Engineering

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from the point of view of the model is the con guration of the needle seat. Based on this detailed con guration, Fig. 10 shows the proposed model of the nozzle. This model is connected to the model of the injection holder by means of a hydraulic connection (connection of line N1 in the injector holder to line N2) and a mechanical connection (the connection needle-rod ). In the upper part of the nozzle there is the line N2, feeding the volume VN connected to the piston PN, which allows the simulation of the action of the pressure force in the upper part of the needle. Following the volume VN there is the line N3 with a section equivalent to the clearance between the needle and the internal part of the nozzle where the needle is placed. This line feeds the di erent volumes V1, V2, V3 and V4 associated with the four pistons that simulate the existing volumes between the needle and its seat and the crowns under the action of the pressure forces. For the characterization of these volumes and the geometric sections of each one of these pistons, the superposition of a photograph of the needle and a photograph of the silicone mould of the seat was used ( Fig. 11). These four volumes are sep-

Fig. 10

Model of the nozzle

Fig. 11

arated by variable-section ori ces (OV1, OV2 and OV3). The section of these ori ces depends on the needle lift. The ve pistons considered are connected to the needle mass. In this model, the needle is considered as two masses, with a total mass equal to that of the needle, separated by a ctitious spring. As in the case of the rod, this spring makes it possible to model the deformation of the needle that is caused by the pressure forces at the extremes of the needle. As with the inlet and outlet ori ces of the control volume, one of the most critical aspects for modelling the nozzle is the characterization of the discharge ori ces, dimensionally as well as hydraulically. In this case, through the use of the silicone method, the most important dimensions of the needle seat can be found as well as the diameters of the ori ces. From the point of view of the one-dimensional model, the most important parameters are the number of ori ces and their respective diameters, since they de ne the permeability (discharge capacity) of the nozzle. For the nozzle characterized, the diameter of the six ori ces was determined. The mean value was 174 ím, with a standard deviation of ±2 ím. For the functional characterization, with the help of the cavitation test rig, the nozzle was tested under feed pressure conditions of 10 and 20 MPa. Proceeding as in the case of the control ori ces, the discharge coe cient as a function of the pressure di erence and the critical cavitation conditions were found. Figure 12 shows the hydraulic characterization obtained for the nozzle of the injection system of the model. Shown in this gure are the results for the mass owrate obtained for the injection pressures tested as a function of the square root of the pressure di erence. As can be seen, the mass owrate increases proportionally to the square root of the pressure di erence until the point where it then stabilizes owing to the appearance of the cavitation phenomenon. The point at which the mass owrate starts to stabilize is associated with the critical cavitation number which in this case is CN =5.51 for crit a pressure of 10 MPa and CN =4.24 for a pressure of crit 20 MPa. As can be seen, the critical cavitation number

Photographs of the needle and the mould of the needle seat


D07903 © IMechE 2004


Fig. 12

Hydraulic characterization of the nozzle

depends on the injection pressure, or more generally on the Reynolds number, as found in the literature [20]. The value of the discharge coe cient is obtained by using equation (1) and taking into account the geometric section of the ori ces. As in the case of the control ori ces, the model of the nozzle ori ces takes into account the cavitation phenomena, and equation (3) governs the behaviour under cavitation conditions. 3.3 Model of the electrovalve Figure 13 on the left shows the physical scheme of the electrovalve and on the right shows the model created with the code. The moving elements of the electrovalve are the command piston and the armature. The movement of these two elements depends on the clearances that have been estimated experimentally. The ascending movement of the command piston owing to the electric current of command opens the outlet ori ce of the control volume by displacing the small sphere. Therefore, the displacement signal of the command piston is the only connection that exists with the model of the injector holder. The armature absorbs the inertia of the command piston during valve closure through the spring R1. The

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spring R2 opposes the opening of the control piston and maintains the piston and the small sphere closing the outlet ori ce while there is no excitation current. The characterization of the springs was carried out by estimating their sti ness using equation (4). The most critical aspect of the electrovalve model is the determination of the opening force. This force is a function of the current signal, the spirals of the coil, the magnetic permeability and the air gap which is a function of the command piston lift. The determination of this force is carried out taking into account theoretical considerations of the continuity of the magnetic ow. A transfer function is used which has as input parameters the electronic pulse and the position of the armature (variation in the air gap), and as the output parameter the force that overcomes the force of spring R1 and produces the lift of the command piston.

3.4 Fuel properties The properties of the fuel (density, kinematic viscosity, bulk modulus and vapour pressure) are considered constant during each simulation. Nevertheless, these properties have been estimated for each validation point as a function of the temperature measured in the IDRCI and the rail pressure. This estimation has been made using the equations proposed by Rodriguez-Anton [21] for a similar fuel (diesel oil CEC RF-73-A-93). The standard diesel fuel CEC RF-06-99 has been used in this work.

4

VALIDATION

In order to validate the model, several experimental measurements were performed using the injection rate test bench. The experimental measurements correspond to rail pressures of 30, 80 and 135 MPa. For each of these pressures, four di erent durations of electronic pulse are used: 250, 500, 1000 and 2000 ís. Moreover, in order to validate the accuracy of the model with mul-

Fig. 13 Scheme of the electrovalve and model D07903 © IMechE 2004


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tiple injections, four multiple injection points were also examined. These measurements were compared with the results obtained by the model. In Fig. 14, in the upper part, the experimental mass owrates for an injection pressure of 80 MPa and the four electronic pulses are compared with the results of the model. On the lower part the same results are compared for the case of 135 MPa. In Fig. 15, the total injected mass of the model and the experimental measurements are also compared for the 12 points tested. In Fig. 16, the experimental mass owrate for a multiple injection, with a pilot injection of 200 ís and a main injection of 800 ís for a pressure of 138 MPa, is compared with the results of the model. In Figs 14 to 16 it is possible to see the ability of the model to predict the experimental results accurately.

Fig. 16

5

Comparison of the experimental mass owrate with the results of the model for a multiple injection

CONCLUSIONS

This article presents a methodology for modelling standard common rail diesel injection systems. This methodology is based on two types of characterization: 1. A dimensional characterization of all the internal parts of the injector, which is of high quality and detail. This characterization is based on the acquisition of silicone moulds of the internal parts, and then the visualization of these silicone moulds using an electronic microscope. 2. A hydraulic characterization of the ori ces and internal elements of the injector (nozzle and control ori ces). In this case, a special emphasis has been placed on the detection of the critical cavitation conditions and their repercussions on the ow (discharge coe cient).

Fig. 14

Comparison of the experimental mass owrate with the results of the model for 80 and 135 MPa

With this information and the help of a one-dimensional calculation code, a model of this system has been created. The comparison of the injection rate proportioned by the model with the experimental data shows the good performance of the model and therefore the ability of the model to predict the injection rate. ACKNOWLEDGEMENTS This work has been nanced by PSA Peugeot-Citroen Automobiles. AMESim is a registered trademark of Imagine S.A. REFERENCES

Fig. 15

Comparison of the experimental injected total mass with the results of the model


1 Bianchi, G. M., Falfari, S. Parotto, M. and Osbat, G. Advanced modeling of common rail injector dynamics and comparison with experiments. SAE paper 2003-01-0006, 2003. D07903 © IMechE 2004


2 Lee, H. K., Rusell, M. F. and Bae, C. S. Development of cavitation and enhanced injector models for diesel fuel injection system simulation. Proc. Instn Mech. Engrs, Part D: J. Automobile Engineering, 2002, 216, 607–618, 2002. 3 Lee, H. K., Rusell, M. F. and Bae, C. S. Mathematical model of diesel fuel injection equipment incorporating nonlinear fuel injection. Proc. Instn Mech. Engrs, Part D: J. Automobile Engineering, 2002, 216(D3), 191–204. 4 Desantes, J. M., Arre`gle, J. and Rodrg´guez, P. Computational model for simulation of Diesel injection systems. SAE paper 1999-01-0915, 1999. 5 Amoia, V., Ficarella, A., Laforgia, D., et al. A theoretical code to simulate the behaviour of an electro-injector for diesel engines and parametric analysis. SAE paper 970349, 1997. 6 Ficarella, A., Laforgia, D. and Landrsicina, V. Evaluation of instability phenomena in a common rail injection system for high speed diesel engines. SAE paper 1999-01-0192, 1999. 7 Augugliaro, G., Bella, G., Rocco, V., Baritaud, T. and Verhoeven, D. A simulation model for a high pressure injection system. SAE paper 971595, 1997. 8 Favennec, A.-G. and Lebrun, M. Models for injector nozzles. Sixth Scandinavian International Conference on Fluid Power, SICF’99, Tampere, Finland, 26–28 May 1999. 9 Payri, R., Salvador, F. J., Plazas, A. H. and Gimeno, J. Hydrodynamic modelling of diesel common rail injection systems. National Mechanical Engineering Conference, Ca´diz, December 2002. 10 Maciań, V., Bermu´dez, V., Payri, R. and Gimeno, J. New

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