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2)On doctoral leave from Universidad Technológica de Pereira (Colombia). (Received 2 August 2007; Revised 14 April 2008). ABSTRACT−One of the major ...
International Journal of Automotive Technology, Vol. 9, No. 4, pp. 447−458 (2008) DOI 10.1007/s12239−008−0054−1

Copyright © 2008 KSAE 1229−9138/2008/041−09

ASSESSMENT OF THE INFLUENCE OF DIFFERENT COOLING SYSTEM CONFIGURATIONS ON ENGINE WARM-UP, EMISSIONS AND FUEL CONSUMPTION A. J. TORREGROSA1), A. BROATCH1), P. OLMEDA1)* and C. ROMERO2) CMT-Motores Térmicos, Universidad Politécnica de Valencia, Camino de Vera s/n, E-46022 Valencia, Spain 2) On doctoral leave from Universidad Technológica de Pereira (Colombia)

1)

(Received 2 August 2007; Revised 14 April 2008) ABSTRACT−One of the major goals of engine designers is the reduction of fuel consumption and pollutant emissions while keeping or even improving engine performance. In recent years, different technical issues have been investigated and incorporated into internal combustion engines in order to fulfill these requirements. Most are related to the combustion process since it is responsible for both fuel consumption and pollutant emissions. Additionally, the most critical operating points for an engine are both the starting and the warming up periods (the time the engine takes to reach its nominal temperature, generally between 80ºC and 90ºC), since at these points fuel consumption and pollutant emissions are larger than at any other points. Thus, reducing the warm-up period can be crucial to fulfill new demands and regulations. This period depends strongly on the engine cooling system and the different strategies used to control and regulate coolant flow and temperature. In the present work, the influences of different engine cooling system configurations on the warm-up period of a Diesel engine are studied. The first part of the work focuses on the modeling of a baseline engine cooling system and the tests performed to adjust and validate the model. Once the model was validated, different modifications of the engine coolant system were simulated. From the modelled results, the most favourable condition was selected in order to check on the test bench the reduction achieved in engine warm-up time and to quantify the benefits obtained in terms of engine fuel consumption and pollutant emissions under the New European Driving Cycle (NEDC). The results show that one of the selected configurations reduced the warm-up period by approximately 159 s when compared with the baseline configuration. As a consequence, important reductions in fuel consumption and pollutant emissions (HC and CO) were obtained. KEY WORDS : Cooling strategies, Emissions, Fuel consumption, Diesel engine

1. INTRODUCTION

thermostats (Chanfreau et al., 2003; Chalgren and Barron, 2005), in which the potential offered by computerized control technologies is exploited, to the implementation of different techniques, such as precision cooling (Robinson et al., 1999), nucleate boiling (Campbell et al., 1997), evaporative cooling (Porot et al., 1997), thermal insulation of walls (Jaichandar and Tamilporai, 2003), intake air heating (Payri et al., 2006a), and flame heaters (Park, 2007). In addition, the importance of both coolant and inlet charge temperature on performance and emissions of a Diesel engine were investigated recently (Torregrosa et al., 2006b) under steady state running conditions, i.e., the engine was already heated. The introduction of strategies aimed at shortening the engine warm-up time have gained special attention, since the shorter the time the engine is operating inefficiently and with components below their nominal temperatures, the lower the wear experienced by the engine, more fuel can be saved, and lower amounts of pollutant emissions will be exhausted (Broatch et al., 2008). Finally, numerical simulation of cooling systems allows the estimation of the thermal behaviour of engine compo-

The design of engine cooling systems is based on the necessity of providing heat dissipation capacity at high power operating conditions, despite the fact that most of the time vehicle engines run at partial load conditions. Thus, by being oversized for partial load regimes of operation, traditional cooling system designs are responsible for unnecessary heat losses, particularly during the warm-up period of the engine. Technological measures to make the cooling system adaptable to engine operating conditions can lead to an improvement in engine efficiency (i.e. lower fuel consumption), and also to a reduction of harmful emissions. In recently published work the benefits of flexible controlled cooling system configurations are demonstrated (Campbell et al. 2000; Allen and Lasecki, 2001; Pang and Brace, 2004). Proposed improvements range from the introduction of electronically controlled coolant pumps (Cortona and Onder, 2000; Cho et al., 2005), fans, shutters, and *Corresponding author. e-mail: [email protected] 447

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nents, thus reducing experimental work. In fact, 1-dimensional models are a fast and efficient way to analyze this problem. However, in some cases, these simple models require information from more complex models (CFD). In recently published works (Strupp et al., 2006), an attempt to couple both types of models are reported, but the interface between them involves high computing costs. The aim of the present work is to explore ways of reducing engine warming up time over a driving cycle (transient operation), by changing only the cooling system configuration, while preserving the system components. The proposed modifications are simple and the reliability of the new components is high, so a quick implementation on serial engines could be possible. It was also intended to evaluate the benefits in fuel consumption and pollutant emissions obtained with the changes made in the cooling configurations. With this purpose, the following methodology was followed: • First, a model of a baseline engine cooling system was developed. The required information was obtained as follows: – The hydrodynamic characteristics of the cooling system components, i.e., the pressure drop as a function of engine flow rate for all engine components, were obtained by testing different steady state running conditions with the baseline cooling system configuration.. – The thermal characteristics of engine components, i.e., resistance to heat flow, thermal inertia, and heat exchanged, were estimated from the results obtained at a constant running point (speed and torque). • Next, the model was used to study the influence on the warm-up period of several modifications to the baseline cooling system. These modifications consisted of both the reduction of the coolant mass flow and the total volume of coolant by means of allowing or avoiding flow through different branches using different valves. The use of a model at this stage allows a substantial reduction in the number of engine tests required. This is important for economical reasons, and also because in the New European Driving Cycle (NEDC) homologation tests, the time to start a new test, when all material temperatures are totally stabilized at 20ºC, is close to seven hours. • Finally, the most favorable strategy for reducing the warm-up time was selected, and the model performance and the impact on fuel consumption and pollutant emissions were then checked on the test bench. The paper is organized following the described methodology. The first part focuses on the description of the fundamentals of the model developed for the simulation of engine cooling systems. Additionally, a basic outline of the baseline engine cooling system and a brief comment on the required experimental information are given. The second part is centered on the experimental data (thermal and hydrodynamic) obtained from the baseline configuration and used to adjust the model. In the third part, the utilization of the model to predict engine thermal performance

when different modifications are introduced in the cooling system is described. The fourth part discusses the experimental validation of the modifications defined with the model. In addition, the influence of these modifications on fuel consumption and pollutant emissions is analyzed. Finally, the main conclusions of the work are outlined.

2. ENGINE THERMAL MODELING 2.1. Description of the Engine Cooling System In order to study the influence of the different cooling configurations on engine thermal response, a four cylinder direct injection Diesel engine was evaluated. Technical data of the engine are summarized in Table 1. The layout of the engine cooling system, hereafter referred to as the baseline cooling system, is presented in Figure 1, where the air cooling circuit is omitted as it is beyond the scope of this work. From Figure 1, the behaviour of the baseline cooling system can be described. First, the coolant exits from the water pump; then it flows through the engine block and through the cylinder head. Some of the coolant is diverted to the oil cooler, where it rejoins the main flow at the water box, placed at the engine outlet. The water box is a mixing volume that collects the coolant coming from the radiator when the engine is already warmed. From the water box, the coolant is conveyed to the engine inlet through four parallel branches: to the radiator (if the thermostat is opened, i.e., the engine is at nominal temperature), to the water tank, to the heater core connected in series with the EGR cooler, and to the bypass, which is the shortest way to the engine inlet. The thermostat has two main functions; on one hand, it restricts the coolant flow to the radiator at low operating temperatures, while on the other hand, it keeps coolant temperatures between pre-defined limits. 2.2. Engine Cooling System Model The engine cooling system model consists of heat and hydrodynamic sources and sinks connected by means of lumped resistance elements (Bohac et al., 1996), which are mathematically modelled to represent both steady state and transient behaviour. The model development comprised the construction of two interacting models: a hydrodynamic model and a thermal model. The hydrodynamic model is based on the conservation Table 1. Technical data of the engine used for the experiment. Configuration

4 cylinder in line

Engine displacement

1998 cm3

Bore

85 mm

Stroke

88 mm

Valves per cylinder

4

ASSESSMENT OF THE INFLUENCE OF DIFFERENT COOLING SYSTEM CONFIGURATIONS

Figure 1. Schematic of engine cooling system. FM represents flow meter, TM represents thermostat; dots and squares represent temperature and pressure sensors. of mass and momentum laws. The momentum law can be expressed as a relationship between the pressure drop through each component, ∆p, and the respective coolant flow rate m· , as shown in equation (1). The constant in this expression can be fitted by engine testing under stationary conditions for each of the components of the engine cooling system. In the case of the coolant pump a series of similar equations are obtained for the different working speeds, i.e., the working map of the pump is obtained. 2 ∆p = K m· (1) f

Pressure loss characteristics for a particular branch in a network are obtained by adding up single component pressure losses at equal fluid flow rate values. The momentum law is complemented by Kirchoff ’s law, expressing the pressure balance in any closed loop in a hydrodynamic network as:

∑ ∆p

i

=0

(2)

Mass conservation implies that the total mass flow rate into a node is equal to the total mass flow rate out of the node: n

∑ m· = 0

(3)

i=1

Here, n denotes the total (inlet and outlet) number of flow paths connected by that node. The coolant flow rate in the engine cooling system is determined by the simultaneous solution of the pump headcoolant flow rate equation and the total system pressure drop-coolant flow rate equation (the flow rate is estimated by the intersection of the pump curve and the system resistive curve). Thus, the calculation procedure is as follows: as the pump is mechanically driven by the engine, the pump speed is obtained from the engine speed. Then, the pump curve for this speed is obtained from the pump chart. In addition, an equivalent of equation (1) for the whole hydrodynamic network is obtained. Next, the intersection between these two curves gives the total coolant flow and pressure loss. Finally the equations are solved for each individual element. The energy conservation law applied to each component,

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in combination with the energy balance at each node, is the basis for obtaining the coolant temperature at the different nodes of the cooling system. In this way, the engine coolant thermal model can be divided into two subsets: one for the internal circuit and the other for the external circuit. The internal circuit coolant temperature model includes heat rejection from the engine enclosure to the coolant flowing through the water jacket and exchanging heat with the external part of the block. The external coolant temperature model includes energy exchange along the bypass, expansion tank, and heater core branches in series with the EGR cooler incorporated with the coolant flow in circulation. These two models interact as the engine coolant temperature evolves. While considering the engine internal cooling circuit, the temperature at the engine outlet can be calculated if the coolant mass flow, the heat rejected by the engine, and the temperature of the coolant entering the engine are known. Similarly, if the temperature of the coolant entering the external cooling system, the global coolant mass flow, the hydrodynamic impedances of the external branches of the cooling system, and the thermal characteristics of the heat exchangers are all known, the coolant temperatures downstream of each branch component can be determined. Therefore, for the simulation of the cooling system a heat transfer model is required in order to predict engine wall and coolant temperatures after each engine operating cycle. The heat transfer model receives the speed, load and in-cylinder main parameters, such as instantaneous pressure − either measured or modelled by means of a computer program (Payri et al., 2006b) −, gas temperature, intake and exhaust averaged mass flow rates, pressures, and temperatures, and computes the heat exchanged between the engine and coolant assuming the Woschni correlation (Woschni, 1967) to calculate the in-cylinder heat transfer coefficient. In the following paragraphs, a more detailed description of the fundamentals and the equations used during the development and adjustment of the model is presented. 2.2.1. Engine internal cooling circuit The fundamentals of the model are similar to those found in (Kaplan and Heywood, 1991) and (Cortona et al., 2000). With the approach of a lumped parameter model, the engine is considered as a compound heat exchanger split into three different zones, as shown in Figure 2. The first zone represents the cylinder, the second zone represents the coolant fluid, and the third zone represents the thermal behaviour of the engine block. First zone The first zone is characterized by the cylinder wall and the cylinder mass, whose temperature Tw is obtained after establishing the heat balance presented in equation (4): · · Q g , w – Q w, c dT --------w- = ------------------------m w c pw dt

(4)

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· w Aw Q w, c = k---------- ( T w – T w, c ) = h A c ( T w, c – T c , m )

δ

(6)

Here, kw is the wall material thermal conductivity and δ is the thickness of the wall. Since both the inlet and outlet coolant temperatures are already known, the mean coolant temperature in the engine passageways can then be estimated as: in

out

Tc + Tc T c, m = -------------------2

(7)

In the case of liquid cooling systems, the heat transfer inside the cylinder block can be calculated by assuming the following relationship (Kays and London, 1974): 0,8 0,4 k h c = c ⎛ Re Pr ----⎞ ⎝ D⎠

Figure 2. Schematic of model zones. · Here, Q g, w is the cycle-averaged heat flow from the cylinder enclosure through the combustion chamber walls, including the heat transferred from the burning gases and also part of the heat transferred to the cylinder liner through piston rings and oil due to engine friction, mw accounts not only for the cylinder liner mass, but also for the total “sensible mass” of the· engine enclosure in contact with the gases and coolant, Q w, c represents the heat transferred from the walls to the coolant, and cpw is the specific heat of the wall material. · The cycle-averaged heat flow ( Q g, w ) for a particular operating point is predicted using thermodynamics and heat transfer-based cycle calculations (Torregrosa et al., 2006a) as: · Q g, w = h g A w ( T g – T w )

(5)

where hg is the cycle-averaged heat transfer coefficient, Tg is the cycle-averaged gas temperature, and Aw is the wall area in contact with in-cylinder gases. In equation (5) it is assumed that, for a given engine running condition, neither the cycle-averaged heat transfer coefficient nor the cycle-averaged temperature change with combustion chamber wall temperature. The heat transfer to the cylinder wall can be obtained from the cycle calculation only under steady-state operating conditions. It is thus necessary to update the wall temperature after a selected time interval in order to compute the transient heat flow and wall temperature during the engine warm-up period. The heat flow from the cylinder enclosure to the wall may then be recalculated with the new wall temperature. · The heat flow from the wall to the coolant ( Q w, c ) is calculated considering the heat transfer due to conduction through the cylinder wall area in contact with the coolant, and it depends on the magnitude of the cylinder wall temperature on the coolant side, Tw,c , the coolant side heat transfer coefficient, hc, and the convection area, Ac:

(8)

Here, Pr is the Prandtl number, c is a correlation constant to be adjusted from the experimental heat fluxes, m· cyl is the coolant mass flow, k is the coolant conductivity and D is a characteristic length (in this case the engine bore). The Reynolds number (Re) was calculated by applying the relationship proposed by (Lehner et al., 2001) and shown in equation (9): 4m· cyl Re = -----------------(9) ρc νc π D where ρ c and ν c are the coolant density and dynamic viscosity, respectively. Because of the difficulties posed by the calculation of the coolant heat transfer coefficient inside the cylinders for different engine operating conditions, it was decided to adopt an estimated area-averaged value of hc = 4,500 W/m2 °C as suggested in (Robinson et al., 2003). Second zone The second zone is characterized by the coolant. The temperature in this zone is assumed to be equal to the temperature of the coolant leaving the engine, whose variation is described by the differential equation: · · out dT c w, c – Q c, exbl --------------------------------------- = Q (10) m c c pc dt · where Q c, exbl represents the heat exchanged between the coolant and the external part of the engine block, and · out in Q c = m· c c pc ( T c – T c ) , with m· c and cpc the coolant mass flow and specific heat, respectively, is the power exchanged between the coolant and the cylinder wall. Third zone Finally, assuming that the external part of the block and the oil are at the same temperature, the engine block temperature Texbl is obtained by: · · dT exbl Q c, exbl – Q bl, ex ------------- = -----------------------------------------(11) m exbl c pbl + m oil c poil dt · The term Q bl, ex in equation (11) represents the heat exchanged between the external part of the block and its

ASSESSMENT OF THE INFLUENCE OF DIFFERENT COOLING SYSTEM CONFIGURATIONS

surroundings, including the exhaust manifold interface. This term also includes the remaining fraction of the heat resulting from the friction work · which was not previously accounted for in the term ( Q g, w ) representing the heat transferred from the cylinder in equation (4). 2.2.2. Engine external cooling circuit During the warm-up process, since temperatures are lower than the nominal temperatures, there is no flow through the radiator branch. Therefore, the external circuit comprises only the bypass valve, the passenger heater core, the EGR cooler, and the leaking branch through the expansion tank. The percentage of coolant flowing through each of these branches depends on their hydrodynamic impedances. Given the total coolant mass flow at the outlet of the engine block, the performance characteristic of the pump and the overall system hydrodynamic resistance, conservation of mass applied to the water box volume (Figure 1) in a quasi-steady approach gives: m· c = m· by + m· hc + m· ex, tk (12) where m· by , m· hc and m· ex, tk represent the coolant mass flow through the by-pass, the heater core and the expansion tank, respectively. Connections between components are characterized by both their volume and their pressure drop. The thermal delays in the system components are determined by their volume (V) and by the volumetric flow ( v· ) passing through them, as: V τ = ---v·-

(13)

Connections and component walls are all considered to be adiabatic (the heat losses through these elements are smaller than the other elements heat losses) and therefore any heat losses through them are neglected. Junctions were modelled as simple conveying nodes (i.e., pressure drop across them is neglected). The heat exchangers (EGR cooler and heater core) are modelled as lumped systems, with the general expression: · in · out dT hx Q hx – Q hx --------- = ----------------------m hx c pc dt

(14)

with the heat flow entering to the heat exchanger being given by: · in Q hx = m· c, hx c pc ( T hxin – T out hx )

(15)

The coolant temperatures at the inlet of the heat exchangers are functions of the coolant temperature at the engine outlet (water box), and depend on the delays caused by connections. The description of· out the heat exchanged through the heat exchangers walls ( Q hx ) is based on the effectiveness − NTU − Cr method described by (Kays and London, 1984).

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The NTU, defined in equation (16), represents the number or transfer units; Cr, given by equation (17), is the ratio of flow stream heat capacity rates, and ε , defined in equation (18), is the heat exchanger effectiveness: Aa Ua NTU max = -----------------( m· c p ) min

(16)

Here, Aa is the characteristic area of the heat exchanger for which the overall heat transfer coefficient Ua is defined. ( m· c p ) min (17) C r = ------------------( m· c ) p max

Q ε = ---------Q max

(18)

The bypass and passenger heater core branches join before entering the engine, so a mixing occurs at this point. Hence, a mean weighted temperature is calculated at the engine coolant inlet as shown in equation (19), since it is assumed that there are no heat losses through connecting elements: out out m· by T by + m· hc T EGR in (19) ± f(t – τ) T c = -----------------------------------------m· c where the term f ( t – τ ) is incorporated into the model to account for the delay effects on temperature caused by the links and connections between the elements, i.e. ,hoses and tubes. Finally, the thermostat was simply modelled as a variable resistance valve, with a temperature dependent opening area, i.e., pressure loss and coolant flow. As a summary, Figure 3 shows a flowchart of the calculation procedure carried out in order to estimate coolant flow rates and temperatures.

3. ADJUSTMENT AND VALIDATION OF THE MODEL Once the thermo-hydrodynamic model has been constructed, experimental results are needed to characterize and adjust some of the constants used in the formulation and also to validate the model. In all the tests, the engine was equipped with magnetic type coolant flow meters, K type thermocouples, and pressure sensors. In addition, the main engine parameters, such as engine speed, brake torque, instantaneous fuel consumption, and intake and exhaust mass flows and temperatures were recorded. Finally, in-cylinder pressure was measured in the stationary points, whereas in the transient and driving cycle tests this was not possible as the glow plug inside the cylinder was needed in these tests and, therefore, a suitable location for the in-cylinder pressure sensor was not available. 3.1. Adjustment of the Model For this task, two kind of engine tests were performed:

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The effect of engine block sensible mass is shown in Figure 4, where the measured and modelled coolant temperatures at the engine outlet are compared for different masses, ranging from 1.9 kg and 14.9 kg. Model results follow the expected behaviour, i.e., the higher the sensible mass is, the longer the warming up time will be. From Figure 4, the mass that best fits the experimental results is close to 3.9 kg. With this mass, the temperature difference between the experiment and model is below 4%. A similar approach was followed for the adjustment of the thermal sensible mass for the other the elements in the cooling system.

Figure 3. Flowchart of the cooling system calculation. • Six different stationary points i.e., points of constant coolant temperature, engine speed and torque, in order to evaluate hydrodynamic and thermal resistances. From these tests, the pressure drop as a function of flow rate for all the components was obtained, as well as different heat exchanger data. • Warming up points, i.e., points at constant engine speed and torque, in order to adjust the heat capacitances of the elements. During warming up, cooling temperatures were raised from 20ºC to 80ºC. The main results after adjustment are related to the warming up process. First, the difference between the modelled and experimental coolant flows is acceptable, since there is a maximum 5% discrepancy in the warming up points. After matching the hydrodynamic response of the model to the real response, the thermal calibration was undertaken. This consisted of the thermal capacitance tuning of the model during the engine warm-up. The stationary heat model, based on a detailed and accurate model that predicts cycle averaged combustion chamber wall temperatures and total heat rejected through the boundaries, was validated in a previous work (Torregrosa et al., 2006a). However, a calibration procedure to assess the “lumped sensible thermal mass” is needed before using it in transient tests.

Figure 4. Coolant temperature behaviour for different thermal block sensible masses.

3.2. Validation of the Model Once the model was adjusted, a validation under different working conditions was performed. For this purpose, a NEDC (consisting of four consecutive Urban Driving Cycles, UDC, followed by one Extra Urban Driving Cycle, EUDC, as in Figure 5) was chosen under an engine-torque basis. This choice was made considering two different points: on one hand, in the future it will be necessary to evaluate modifications on the engine cooling system under this kind of tests, and on the other hand, a NEDC includes both transitory and stationary points. For these reasons, the model was run for this mode of operation and the results were compared with the real data obtained on the engine test bench. Figure 5 shows the comparison of the predicted and measured coolant temperatures during the NEDC driving cycle, together with the vehicle speed evolution, while Figure 6 shows a comparison of coolant flows in one of the cooling system branches. The model simulations reproduce the warm-up coolant temperature profile and the coolant flow until the thermostat opening temperature (83ºC) is reached. After this point, large deviations between the modelled and experimental results are found in both temperatures and flows, due to uncertainties in the thermostat and radiator modeling, caused by the lack of experimental data; tests were performed until thermostat opening. In order to overcome the deviation shown in Figure 5, additional tests should be performed to adjust the thermostat model, namely warm stationary tests and transitory tests above 80ºC. However, recalling that the purpose of the model was to simulate the warm-up process

Figure 5. Outlet engine coolant temperature profile over a driving cycle.

ASSESSMENT OF THE INFLUENCE OF DIFFERENT COOLING SYSTEM CONFIGURATIONS

(i.e., before the thermostat is opened), it can be stated that for prediction purposes the developed cooling system model is able to predict the coolant temperature and flow behaviour during transient engine conditions, thus allowing for the accurate evaluation of the cooling system performance during the warm-up process.

4. MODIFICATIONS TO THE BASELINE COOLING SYSTEM During transient operation (warm-up, for instance), the storage of the heat rejected by the engine is controlled by the sensible thermal mass and the cooling system. Bearing in mind this fact, there are a few possibilities in order to reduce the warm-up period and, hence, to reduce both pollutant emissions and fuel consumption. The possibilities are either to reduce the sensible thermal mass of the elements or to modify the cooling system configuration. This work is centred on the second approach, i.e., changing the engine cooling system configuration. The cooling system parameters that affect the warm-up period are coolant temperature, total mass of coolant and coolant mass flow through the elements. The temperature is imposed by the homologation cycle (starting temperature of 20ºC), so that the real possibilities reduce to changing the total coolant mass involved in the heating period and/or to modify the coolant flow through the elements: • Assuming that the heat transferred to the coolant does not depend on the total mass, it is clear from equation (10) that the smaller the coolant mass is, the faster the temperature will grow. • A reduction in the flow rate would cause a rapid decrease of the heat transfer coefficient (h) and, therefore, a decrease of the heat transfer rate. Additionally, under the assumption that the heat rejected by the engine does not depend on the coolant flow rates, then lower flow rates lead to higher temperature differences, as shown in equation (10). Thus, if the coolant flow through the external circuit could be bypassed, the engine would warm up faster, and the temperature difference across the engine would be determined only by the delay associated with the circulating amount of coolant (similar to equation 19). However,

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the legislation requires a shorter time to clear ice or mist from the windscreen and to warm up the passenger compartment in winter, making it mandatory that the coolant flows through the heater core from the start of the engine. Hence, it is not possible to cut the flow to the heater core, yet it is possible to close the flow through the other external paths. Based on the previous analysis, the thermal response of the engine at the NEDC was evaluated for different strategies by using the model. The best strategy, i.e., the fastest warming up period, was then selected to be tested on the engine bench in order to corroborate the model results and to check the influence on both pollutant emissions and fuel consumption. The strategy consisted on incorporating two valves to the system. The first one, denoted by Vbp, was placed in the bypass branch with the aim of reducing the coolant flow rate, and the second, referred to as Vwt, was located in the water tank branch with the aim of reducing the total participating coolant volume. The flow through the engine is guaranteed due to the need to keep a coolant flow through the heater core in order to ensure passenger comfort. Five different configurations were modelled, with the assumption that heat rejection is the same for the baseline and the modified cooling systems. These configurations correspond to a 2×2 screening experimental design (two factors at two levels) plus a centre point. As the limit cases for the valves were fully opened or fully closed, the centre point was a simulation of a 50% reduction in the effective area. These points, and the adopted names, are presented in Table 2, where the position of each valve has been indicated with a number ranging from 0 to 1. The first number in the subscript corresponds to the bypass valve and the second to the water tank valve. The main reason to choose this kind of analysis is the possible presence of interactions in valve position, since when a valve is closed the entire hydrodynamic resistance is changed, and therefore the flow through the different branches (and therefore through the engine) changes when the configuration is changed. Since it would be dangerous to test the engine with all the valves fully closed during the whole cycle (due to the eventual formation of air bubbles and the low flow through the engine), it was decided to open both valves after the third urban cycle (i.e., after 585 s) in all the cases. The modelled results show that the lowest and the Table 2. New cooling system configurations.

Figure 6. Coolant flows in the by-pass branch, during a driving cycle.

Configuration

Vbp

Vwt

C1,1

Opened

Opened

C0,0

Closed

Closed

C0,1

Closed

Opened

C1,0

Opened

Closed

C0.5,0.5s

50%

50%

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A. J. TORREGROSA, A. BROATCH, P. OLMEDA and C. ROMERO

Figure 7. Coolant temperature at the outlet of the engine block for the modelled cooling configurations. highest coolant mass flow rates are obtained with the C0,0 and C1,1 configurations, as expected, since the latter is the configuration in which the coolant can pass through the major number of paths and, thus, is subject to the minimum hydrodynamic resistance. Flows in configuration C0,1 are slightly higher than those in the C0,0 configuration, whereas flows in configuration C1,0 are slightly lower than those in C1,1. Figure 7 shows the modelled temperature response for the five different cases under a NEDC cycle, while Figure 8 shows the temperature difference at the engine water box between the proposed modifications and the baseline configuration. It is apparent that the configuration with the two valves fully closed is the fastest, as expected. In addition, the temperature difference in the C0,0 configuration has a particular behaviour: at the beginning, this difference increases until approximately 400 s (i.e., 2 UDC) and then decreases. This could be caused by the effect of the heat conduction through the coolant. The behaviour in the other proposed modifications is quite similar: a rapid increase in this difference at the beginning and then a constant difference behaviour. The observed trends can be explained by considering that the flow through the bypass increases the hydrodynamic impedance of the cooling system. As a result, the flow rate through the engine block decreases and the coolant leaving the engine reaches the desired 80oC temperature sooner when compared to the baseline configuration. For configurations C0,0 and C0,1, the same trend is observed in the change of the temperature difference, but the relative change is in favour of configuration C0,0. For this configuration, the coolant flowing from the block reaches a temperature of 80oC after a time of 539 seconds. Although configuration C0,1 is similar in flow rate magnitude to configuration C0,0, its associated higher coolant volume prevails, making it the slowest to reach the 80oC target temperature. The coolant warms up faster in configuration C1,0 than in configuration C1,1, because of the lower coolant volume involved, despite the fact that its coolant flow rates are higher.

Figure 8. Difference in modelled coolant temperature with respect to the baseline configuration. Nevertheless, to take benefit of the experimental design, the study was centred on the time required to reach 80oC at the engine water box. The inverse of the average temperature slope (ATS), which indicates how fast the engine warms up, is defined in equation (20), after the first three urban cycles, i.e., after 585 s of engine running. This definition was chosen due to the necessity of opening the valves after three urban cycles for configuration C0,0: 1 ∆t ------------------- = ------------------ATS ( T ) T 585 – T 0

(20)

where T585 is the temperature at the end of the third urban cycle and T0 is the temperature at the beginning of the tests (20ºC). Results from the experimental design are shown in Figure 9 for the time required to reach 80ºC, where the effect of each variable (Pareto chart) is shown. The R2 (which indicates the percentage of variation explained by a regression analysis; that is, how the variability of the studied parameter, T80, is explained by the variability of the independent variables, Vbp and Vwt) is 99.85%, and it is clear that, in this case, both the total coolant mass and the coolant flow have a significant effect on the warming up of the engine. Finally, the interaction between the two valves seems to be slightly significant. A similar analysis, in this case for the parameter ATS−1, is shown in Figure 10. In this case, although the R2 coefficient is high (almost 99%), there is only a significant effect due to the total coolant mass (water tank valve). Neither the effect of the by-pass valve (coolant flow) nor the interaction between the two valves was significant.

Figure 9. Results of the analysis design for T80.

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Figure 10. Results of the analysis design for ATS−1. From both analyses, it is clear that the interaction between the two new valves has a minimal influence on the engine warming up. The search of possible interactions and the optimization of the working configuration was the main reason to use the experimental design analysis. Since the effects of interaction are negligible, it was decided to check the possible effects on fuel consumption and pollutant emissions only for the optimal configuration, i.e., C0,0. Finally, the advantage of the proposed cooling system modification is that it does not need to incorporate any sophisticated control strategy, since it only requires setting the coolant flow by means of the two mechanical valves added.

5. EXPERIMENTAL RESULTS An experimental facility conceived for the identification of performance characteristics of the cooling system was used. Two mechanical valves were installed: one with 40 mm diameter, placed between the water box and the engine entrance, and a second with 6 mm diameter, installed between the water box and the expansion tank. The first result is the engine thermal response under a NEDC cycle with the modifications performed on the cooling circuit. In this way, the model could be validated again by comparing the predicted temperature trace with the experimental one. Figure 11 shows these traces for configuration C0,0, and it can be observed that the model reasonably fits the experimental temperatures during the warm up process (both qualitatively and quantitatively). The deviations found at the end of the cycle have been

Figure 11. Profile of temperature response for modelled and experimental cooling systems corresponding to configuration C0,0 along a driving cycle.

Figure 12. Profile of temperature response for experimental cooling systems and ATS−1. already explained for the baseline case (poor modelling of the thermostat and the radiator behaviour). Therefore, the model reproduces the engine behaviour quite reasonably for a completely different engine cooling system configuration (totally opposite to the baseline). This result gave confidence on the results obtained in the previous analyses. In addition, the large drop in temperature after approximately 10 minutes is due to the valve opening, as mentioned earlier. Figure 12 shows the experimental temperature traces for both cooling configurations; the inverse thermal derivative after each UDC is also presented. It is clear that the modifications in the cooling system allow for a faster warm-up. This is clearer at the end of the third urban cycle, where the highest differences can be observed. The difference is not as clear at the beginning of the tests, since in the baseline system the coolant evacuates heat from the engine due to its convective effect, while in the new configuration the engine warming up is affected not only by the material sensible mass but also by the coolant sensible mass around the engine cylinders. The next step was to check the influence of system modifications on both fuel consumption and pollutant emissions. To measure this possible impact, additional instrumentation was installed. This consisted of a gravimetric type fuel flow meter, and a standard pollutant emissions equipment (HORIBA MEXA 7100DEGR) to measure instantaneous carbon monoxide (CO), unburned hydrocarbons (THC) and nitrogen oxides (NOx), sampled upstream of the catalyzer; instantaneous smoke measurements were performed by means of an opacimeter (AVL439) as proposed by different researchers (Arcoumanis and Megaritis, 1992; Jones et al., 1997; Cheung et al., 2000). Three tests were performed for each chosen configuration in order to check the repeatability of the results. Tests were performed only for the transient driving cycle operating conditions. Once the measurements were performed, special postprocessing of the signals was necessary, since it is well known that there is an important delay in pollutant measurement (Messer et al., 1995). This delay is neglected when stationary tests are performed (the engine is stable), but it could lead to important errors in transient tests.

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Besides, this delay depends on the length of the measurement line, the type of pollutant and the temperature at which the sample is taken. Therefore a procedure must be applied in order to reduce potential errors in pollutant measurements. In this work the procedure proposed in (Arregle et al., 2006) was applied. It consists of applying a cross correlation method to align the desired pollutant measurement with the instantaneous exhaust mass flow rate, obtained as the sum between the instantaneous air and fuel mass flow rates. Additionally, the filter smoke number, FSN, is calculated from opacity measurements using the same correlation as in (Arregle et al., 2006). From this FSN, the particle mass of dry soot may be obtained by using one of the different correlations available. In this work the correlation proposed in (Alkidas, 1984) was used. Finally, the total particle mass (PM) was obtained from the estimated mass of dry soot and the HC measurements. The results after the post-processing were fuel consumption, in l/km, and pollutants, in g/km. These results were analyzed as follows. The accumulated values for each test and each configuration were extracted for the five different intervals in the NEDC: one for each urban driving cycle (UDC) cycle and one for the extra urban driving cycle (EUDC). With these values, a mean and a confidence interval for the three tests on both configurations (proposed and baseline) was calculated. Then the loss or gain (as a percentage) with the new configuration with respect to the baseline was calculated for all the studied parameters at the different intervals Figure 13 shows the comparison between the two configurations for pollutant emissions. It is apparent than the faster warming up of the engine allows a reduction on HC and CO emissions during the urban cycles; the experimental reduction obtained was much larger than the measurement dispersion, while there is no significant difference on the extra-urban cycle. This result can be explained considering that there is an inverse relationship between these emissions and the temperature, i.e., the lower the temperature, the higher are these emissions. The largest reduction in both pollutants takes place in the third urban cycle, due to the highest temperature difference between configurations, whereas in the case of the extra-urban cycle, the engine temperature has already reached its nominal value in both configurations, and thus the small differences found are mainly due to measurement dispersion. In the case of NOx, the differences are smaller than the measurement dispersion, but the tendency seems to follow a trend inverse to that found for HC and CO, i.e., a slight increase of this pollutant during the warming up (urban cycles) in the proposed configuration due to the higher temperatures. However, the total effect on this pollutant is negligible, since the emissions in the extra-urban cycle are close to 70% of the total emission of NOx, and therefore adverse effects on the urban cycles is compensated. For PM emissions, the dispersion is higher than the

Figure 13. Comparison between pollutant emissions for C1,1 against C0,0. (a) HC, (b) CO, (c) NOx and (d) PM. measured effects. Nevertheless, the tendency in the urban cycles seems to be a slight reduction, probably due to a higher oxidation of the particles at higher temperatures, with almost the same results in the extra urban cycle. Due to the high dispersion in this pollutant, extra tests were performed on the baseline configuration but the results did not show any improvement on the confidence level. The reduction in fuel consumption during the urban cycles, shown in Figure 14, can be explained considering that higher temperatures imply a lower oil viscosity, and thus lower friction losses. The reason for the negligible difference in fuel consumption in the extra urban cycle is

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reductions of 12% and 16% in the total CO and HC emissions, respectively, whereas a negligible impact on NOx and PM emissions was found.

REFERENCES

Figure 14. Comparison between fuel consumption for C1,1 against C0,0. the same given in the case of HC and CO; i.e., the engine has reached its nominal temperature in both cases.

6. CONCLUSIONS A thermo-hydrodynamic model was developed for the analysis of the thermal performance of a Diesel engine during the warm-up phase. The model is based on the thermal and hydrodynamic characteristics of a real engine instrumented and tested on an engine test bench. Pressure drop as a function of mass flow rate, and heat flow rates were obtained for all the components in the system under different engine operating conditions. With the aid of the model developed, the improvements in thermal performance associated with achievable modifications in the cooling system design were studied. From the simulations, the potential of reducing the warm-up period under constant load transient driving operations and also under NEDC road cycle operation was observed. The modifications proposed were implemented in the actual engine. In general, the model predictions were in reasonable agreement with the experimental data both in trend and in magnitude, indicating that the model is suitable to study the engine’s thermal performance. For complex conditions, such as NEDC conditions, the model simulations reasonably follow the warm-up coolant temperature profile and the target temperatures observed in tests. Concerning the influence of the cooling system architecture on engine performance, it can be concluded that significant improvements can be achieved by implementing cooling systems sufficiently flexible so as to control the total coolant volume as well as the coolant flow rates through the cooling circuits, as a strategy to speed up the engine warm-up, which leads to a reduction of fuel consumption and pollutant emissions. For the modelled engine, it was demonstrated that changing the baseline cooling system C1,1 by the C0,0 configuration produced a significant reduction of 22.3% in the time needed for the engine to reach the target temperature of 80ºC. This reduction in the warm-up time resulted in a reduction of the total fuel consumption of 1.62%, and

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