Growth and Refinement

By 1920, the car was a common fixture on both sides of the Atlantic, and automakers began to focus on improved performance for their vehicles. Cadillac had introduced the V-8 engine in 19 15 (Fig. 1.3), and by 19 16, eighteen companies were producing V-8s (Rinschler and Asmus, 1995). Packard introduced the straight eight in 1923, and by 1930, Cadillac introduced its 7.4L V-16. Engine performance was greatly improved with the development of the turbulent head by Harry Ricardo shortly after World War I (Fig. 1.4). The turbulent head aided combustion

This period also saw significant improvements in braking, lighting, tires, and windshields, and there was a dramatic shift in buyer preference toward closed cars. Until 1920, most cars were open-topped vehicles, which had obvious negative implications for driving in bad weather. The enclosed car isolated the occupants from rain, snow, and dust, but it also provided advan- tages in safety. Early closed vehicles had roofs made of fabric-covered wooden frames. In an accident, occupants sometimes were ejected through the roof (Yanik, 1996). Thus, work began on developing a steel roof. This was no small feat, as initial attempts with flat steel roofs produced a drumming sound when traveling. Harley Earl solved the problem by cum- ing the roof, and GM put his invention into production as the "Turret Top" in 1935 (Fig. 1 S). The enclosed, all-steel vehicles also prompted the first use of safety to market automobiles, and rollover tests such as those shown in Fig. 1.6 were used as advertisements to demonstrate the safety and sturdiness of such vehicles.

The 1930s also saw the advent of crash testing. General Motors used a test driver standing on the running board, who would direct the car down a hill toward a wall, jumping off at the last moment (Yanik, 1996). The only analysis that could be made at that time was to observe the resulting damage. Of course, the main event in the 1930s was the Great Depression, which brought a huge drop in demand for cars. Automakers were forced into a survival mode, and many automakers did not survive this period, notably Marmon, Peerless, Duesenberg, Cord, Auburn, Graham, Hupp, and Stutz.

Modem Development

World War I1 brought a halt to auto production in the United States as automakers switched to wartime materiel production. However, the halt was only temporary. At the conclusion of the war, not a single US. factory had been bombed. The same could not be said for Europe. Thus, American engineers in 1946 could immediately update their products, and the U.S. factories began churning out vehicles quickly. The four-year hiatus in auto production also created a pent-up demand for new vehicles, which spurred enormous growth in the U.S. economy.

The 1950s found the U.S. auto industry leading the world, and the cars reflected this general attitude. Cars were bedecked with ever more chrome trim, and tailfins rose in height until the 1959 Cadillac presented a practical limit to fin height. The Corvette was introduced in 1953 and has continued as "America's Sports Car" to this day. In 1955, Chevrolet introduced the now-famous "small-block" V-8. Initially, this was a 265-cubic-inch carbureted engine adver- tised at 180 hp (Autocar, 1996). The impact of this engine cannot be overstated. Until that point, the fastest cars also were the most expensive. Thus, a Cadillac could outrun a Buick, and so on down the cost ladder. The small-block V-8, under the workings of a skilled engine tuner, suddenly was enabling bargain-priced Chevys to outperform Cadillacs and Lincolns. Even today, 1950s-era cars with small-block Chevy engines are solid performers at the track (Fig. 1.7).

Transmission Matching and Vehicle Performance

To provide an appreciation of the issues involved in matching the transmission to a vehicle, use is made here of an extended worked example. This simple case study is restricted to steady-state operation and is concerned with sizing the engine for a given top-speed perfor- mance, determining the gear ratio required for a specified hill-starting performance, and cal- culating the fuel economy for constant-speed operation with different transmission options. This example demonstrates the need for computer modeling, which is introduced first in general terms. The chapter ends with an overview of the ADVISOR modeling software, a versatile package that is available free of charge.

Unfortunately, the power requirements of vehicles are characterized by part-load operation. Thus, it is necessary to consider not only the engine efficiency, but how it is matched to the vehicle through the transmission system. Because the principles in matching the gearbox and engine are essentially the same for any vehicle, it will be sufficient to discuss only one vehicle. The example used here is a vehicle with the specification as shown in Table 11.1.

Two-Degrees-of-Freedom Model (Quarter Car Model)

In a vehicle, the excitation of the vehicle spring-mass system is provided by the motion of the tirelunsprung mass and can be analyzed by the same techniques used to analyze support motion. The details of such an analysis are contained in vibration texts (Thomson, 1988), and only the highlights will be discussed here. Referring to Fig. 8.1, and isolating one spring-mass-damper system, the displacement of the vehicle body will be defined by x, whereas the displacement of the unsprung mass will be designated as y, as shown in Fig. 8.6.

At this point, it is most illustrative to assume that the motion of the unsprung mass is har- monic, which is not a bad assumption given that the tire and unsprung mass constitute a damped vibratory system. Before proceeding, the concept of transmissibility must be intro- duced. Transmissibility is defined as the ratio of the transmitted force to the ratio of the exciting force. Because in this case the exciting force is provided by the unsprung mass and tire, and as such is proportional to the displacement of the unsprung mass, the transmissibility is given by (Thomson, 1988)

Vehicle Cooling Systems

Gruden and Kuper (1987) conducted a systematic study of the energy balance in a spark ignition engine and presented a series of contour plots for the different energy flows (i.e., fuel in, brake power, coolant, oil, and exhaust) as functions of bmep and engine speed for a 2.5-liter engine. They also presented contour plots of the brake, mechanical, and indicated efficien- cies. The brake efficiency results imply that the engine has been tuned for maximum economy at part load, whereas at full throttle, the mixture has been richened to give the maximum power. The mechanical efficiency is directly affected by the load (with zero mechanical efficiency by definition at no load). Also, the mechanical efficiency at full load falls from approximately 90% at 1000 rpm to 70% at 6000 rpm. At 6000 rpm, the frictional losses represent approximately 34 kW. Friction dissipates useful work as heat, some of which appears in the coolant and some in the oil. The heat loss recorded to the oil is almost solely a function of speed, with approximately 5 kW dissipated at 3000 rpm, and 15 kW dissipated at 6000 rpm.

Figure 5.12 shows the contours of the energy flow to the coolant as a function of the load and speed. For convenience, the brake power output hyperbolas (calculated from the bmep and speed) also have been added. At a bmep of approximately 1 bar, the energy flow to the coolant is approximately twice the brake power output, whereas at a load of 3 bar bmep, the energy flow to the coolant is comparable to the brake power output. In the load range 8-10 bar bmep, the energy flow to the coolant is approximately half the brake power output. However, of greater importance to the vehicle cooling system are the absolute values of the heat rejection. Figure 5.12 shows that heat rejected to the coolant is a stronger function of speed than load. Ancillaries The heat rejection to a direct injection diesel engine is approximately a third lower, and this can lead to the need for a supplementary heater for passenger compartment heating. In both diesel engines and spark ignition engines, approximately half the heat flow to the coolant

Ancillaries

a film of lubricant of sufficient thickness to prevent metal-to-metal contact. The flow of oil and its pressure between the bearing surfaces are governed by their motion and the laws of fluid mechanics. The oil film pressure is produced by the moving surface drawing oil into a wedge-shaped zone, at a velocity high enough to create a film pressure that is sufficient to separate the surfaces. In the case of a journal and a bearing, the wedge shape is provided by the journal running with a slight eccentricity in the bearing. Hydrodynamic lubrication does not require a supply of lubricant under pressure to separate the surfaces (unlike hydrostatic lubrication), but it does require an adequate supply of oil. It is convenient to use a pressurized oil supply, but because the film pressures are much greater, the oil must be introduced in a way that does not disturb the film pressure. Ancillaries As the bearing pressure is increased and either the viscosity or the sliding velocity is reduced, the separation between the bearing surfaces reduces until contact occurs between the asperi- ties of the two surfaces-point A on Fig. 5.9. As the bearing separation reduces, the solid-to- solid contact increases and the coefficient of friction rises rapidly, leading ultimately to the boundary lubrication mode shown in Fig. 5.10. The transition to boundary lubrication is controlled by the surface finish of the bearing surfaces, and the chemical composition of the lubricant becomes more important than its viscosity. The real area of contact is governed by the geometry of the asperities and the strength of the contacting surfaces. In choosing bearing materials that have boundary lubrication, it is essential to choose combinations of material that will not cold weld or "pick up" when the solid-to-solid contact occurs. The lubricant also convects heat from the bearing surfaces, and there will be additives to neutralize the effect of acidic combustion products.

In reality, the shaft does not remain concentric in the bearing, nor is the pressure uniform around the bearing. The eccentricity and the pressure distribution are shown schematically in Fig. 5.1 1. An early and reliable solution to this problem was published by Raimondi and Boyd (1958). This technique uses tabulated or graphical data to calculate the coefficient of friction, film pressure, point of maximum film pressure, side flow, and bearing temperature rise on the basis of the Sommerfeld number for the bearing under analysis. A thorough treat- ment of this method is found in Shigley and Mischke (2001).

Aurther

Richard Stone

RJeffrey K. Ball

Downlaod Book