Thursday, December 16, 2010




Ring rolling is an advanced plastic forming technique to cause a ring to reduce its thickness, enlarge its diameter and shape a profile on its surface , which is used to manufacture precise seamless rings with various dimensions, shapes and materials. Ring rolling has many advantages such as high productivity, uniform quantity, smooth surface, material saving and it has been used in many industry fields, including aeronautics, astronautics, automobile and atomic energy etc .

At present, many research on ring rolling has been carried out by experimental method, analytical method and finite element method, which study various aspects like technology design ,deformation condition and material properties. A groove-section profiled ring is a typical kind of profiled rings, which is an important part of a ball bearing. Compared with the conventional manufacture methods of groove-section profiled rings like forging, flame cutting, ring rolling shows more economical advantages in productivity improving and material saving.

So, in recent years, many attempts have been made at the groove-section profile ring rolling. Hua et al. firstly gave the expression for the motion track of guide roll by analytic calculation. Subsequently, Hua et al. developed a complete 3D FE simulation model to simulate the ring rolling process and testified the model validity by experiment. Hua et al. proposed the plastic penetration condition by FE simulation. Qian et al. [20] investigated the ring gripping condition and its influence factors using statics theory. Qian [21] revealed the movement rules using theoretical calculation and FE simulation. These researches established a theoretical basis for the application and development of groove-section profile ring rolling, but they are still weak to instruct the actual production for lack of the research on technical design.




The working principle of groove-section profile ring rolling is illustrated by Fig. 1. The blank has a rectangular section shape. A closed rolling gap limits the axial spread of the ring and ensures the planeness of the ring end surface. The driver

roll takes active rotation. The mandrel is the pressure roll, which takes linear feed motion and passive rotation. The guide roll structure is shown in Fig. 2 [17]. The guide roll fixed on a frame and its center coincides with the center of the driver roll,

so it can rotate either around the driver roll center or around he axis of its own. At the end of the frame, a hydraulic cylinder produce a pressure P which makes the guide roll to keep contact with the ring, that help to maintain the stability of ring rolling process and improve the roundness of the rolled ring. In the rolling process, the blank is rolled into the rolling gap repeatedly, one rotation following another, while its thickness is reduced, diameter is enlarged and groove is formed.

Fig. 1. Schematic illustration of groove-section profile ring rolling.

Fig. 2. Schematic diagram of mechanism of guide roll.

A0 = B0H0


Refrigeration system.

2.2.1 Working Fluids

In the absorption refrigeration system two working fluids are used: a refrigerant and an absorbent. Among the most applied working fluids are the pair ammonia refrigerant–water absorbent and water refrigerant–lithium bromide absorbent. A limitation of the pair water–lithium bromide is the difficulty to operate at temperatures lower than . Besides, lithium bromide crystallizes at moderate concentration, and, at high concentration, the solution is corrosive to some metals and is of high cost.

From launch of absorption refrigeration system, the pair ammonia–water has largely been used. Both fluids are highly stable at a wide operating temperature and pressure range. Ammonia has a high enthalpy of vaporization, which is necessary for satisfactory system performance. The system can be used at low temperature, as the ammonia freezing point is. Besides, the pair ammonia–water is environment friend and of low cost.



Fig.3.1. Schematic of the Ammonia–water Absorption Refrigeration System.

Fig.3.1 shows a schematic of the basic ammonia–water absorption refrigeration cycle. High pressure ammonia vapor enters the condenser, where it transfers heat to the neighborhood. Liquid ammonia leaves the condenser and passes through an expansion valve, reaching the evaporator pressure. The refrigerant then enters the evaporator, where it receives heat from the cold source, turning into low pressure vapor. In the sequence, ammonia vapor enters the absorber, where a weak solution of water and low concentration ammonia absorbs the refrigerant and, at the same time, transfers heat to the neighborhood. The solution has now a high ammonia concentration, and is pumped to the vapor generator, where it receives heat from an external source. The ammonia in the solution then evaporates, separating from water and flowing to the condenser to start a new cycle. A weak water–ammonia solution leaves the vapor generator and enters the absorber to absorb ammonia vapor from the evaporator. A heat exchanger between the absorber and the vapor generator transfers heat from the weak solution leaving the vapor generator to the high ammonia concentration solution going into the vapor generator. That increases the cycle coefficient of performance. The absorption refrigeration system instantaneous coefficient of performance (COP) is given by:


Where is the absorption system instantaneous cooling capacity (W) and is the instantaneous heat transfer rate from the energy source (in this work engine exhaust gas is used) to the absorption refrigeration system (W).

From the First Law of Thermodynamics, the system cooling capacity is calculated as:


Where wall is the instantaneous heat transfer rate from the neighborhood to the cold source through the refrigerator walls (W), and are the initial and final system internal energy (J) at the time interval (s), respectively.

Was calculated considering radiation from the neighborhood () and natural convection from the environmental air () to the refrigerator external walls. This energy was conducted through the refrigerator walls and transferred by natural convection from refrigerator inner walls to the ambient air inside the refrigerator (). Thus,






where is the refrigerator wall thermal conductivity (W/m K), is the refrigerator wall area frontal to conduction heat flux (m), is the refrigerator wall thickness (m), is the external convection coefficient (W/m K), is the refrigerator total external surface area (m2), is the refrigerator external surface average temperature (K), is the environmental air temperature (K), is the refrigerator external surface emissivity, is the Stefan–Boltzmann constant (), is the neighborhood temperature (K), considered to be equal to the environmental air temperature, is the internal convection coefficient (W/m K), is the refrigerator total internal surface area (m), is the refrigerator internal surface average temperature (K), and is the ambient air temperature inside the refrigerator (K). Both the environmental air temperature and the ambient air temperature inside the refrigerator were measured, while the surface temperatures were determined by an iterative process using Eqns. (3)–(6).

The instantaneous internal energy inside the refrigerator was evaluated by:


Where is the dry air mass (kg), is the dry air specific internal energy (kJ/kg), is the water vapor mass (kg), is the water vapor specific internal energy (kJ/kg), is the condensed water mass (kg), and is the condensed water specific internal energy (kJ/kg).

Both dry air and water vapor were treated as ideal gases. Thus, the water vapor mass is calculated as follows:


where is the partial pressure of water vapor in the ambient air inside the refrigerator (Pa), is the measured volume inside the refrigerator (m), is the universal gas constant (8314J/kmol K), is the water vapor molecular weight (kg/kmol), and T is the average ambient temperature inside the refrigerator (K).

The water vapor partial pressure is calculated by:


Where is the measured relative humidity of the ambient air inside the refrigerator and is the water vapor saturation pressure at the measured temperature of the ambient air inside the refrigerator.

The dry air mass inside the refrigerator is evaluated in the following way:


Where is the dry air molecular weight (kg/kmol) and p is the absolute pressure inside the refrigerator (Pa).

The specific internal energies, , and were taken from thermodynamic tables as function of average ambient air temperature inside the refrigerator. The water vapor and condensed water specific internal energies were considered as approximately equal to that of the saturated vapor and saturated liquid at the same temperature, respectively.

The heat transfer rate from the engine exhaust gas to the absorption refrigeration system vapor generator is given by:


Where the engine exhaust gas is mass flow rate, is the engine exhaust gas specific enthalpy at the entrance of the vapor generator and is the engine exhaust gas specific enthalpy at the exit of the vapor generator. is the heat transfer from the engine exhaust gas to the environment though the insulated wall of the vapor generator, calculated in a similar procedure as described by Eqns.(3)–(8).

The engine exhaust gas mass flow rate was determined through the following relationship:

. (12)

where is the fuel density (kg/m), determined as a function of the temperature, is the measured fuel volume flow rate (m/s), and A/F is the air/fuel ratio, that is, the ratio of air mass flow rate and fuel flow rate.

The exhaust gas enthalpy variation was calculated as follows:


Where , , and are the mass fraction, inlet enthalpy (kJ/kg) and outlet enthalpy (kJ/kg) of the exhaust gas components CO, CO, HC, HO, O, and N. The exhaust gas components concentration was measured in the experiments, while the enthalpy was taken from thermodynamic tables based on the exhaust gas temperature.



A commercial 215-l refrigerator built on an absorption refrigeration system was studied using the engine exhaust gas as energy source (Fig. 4). Temperature and humidity inside the refrigerator were monitored through two Pt100 thermometers and a thermo hygrometer. Measurement of the exhaust gas temperature was made through two K-type thermocouples installed in the refrigeration system heat exchanger inlet and outlet. A barometer and a liquid-in-bulb thermometer were used to measure ambient pressure and temperature. Ambient pressure was kept at 0.913 ± 0.007 bar, while ambient temperature was maintained at 300 ± 5 K.

Fig.4.1. Absorption Refrigerator Adapted to Engine Exhaust System.

A production 1.6-l, 8-valve, four-cylinder automotive engine with multipoint electronic fuel injection was used for the tests. The engine also featured compression ratio 9.5:1, 86.4 mm bore and 67.4 mm stroke. The engine was tested in a hydraulic dynamometer of maximum power 260 kW and maximum speed 6000 rev/min. The dynamometer was equipped with a load cell of measuring range up to 2224 N and uncertainty of ±0.4 N, and a magnetic speed sensor, which uncertainty was ±3 rev/min. Fuel consumption was measured by a turbine flow meter, of measuring range 0.038–100 l per minute and accuracy of 0.5% of the reading. To adapt the refrigerator to the engine the exhaust pipe configuration after the catalytic converter had to be modified, including removal of two plenum chambers and a flexible duct that directed the exhaust gas to an external chimney. These components were part of the standard laboratory setting, but could not be used when the refrigerator was installed.

The exhaust concentrations of carbon monoxide (CO), carbon dioxide (), hydrocarbons (HC) and oxygen () were measured by a non-dispersive infrared analyzer (NDIR). The equipment also allowed for determination of air/fuel ratio and mixture equivalence ratio (ratio between measured air/fuel ratio and stoichiometric air/ fuel ratio). The uncertainties associated to the measurements were ±0.16% for CO, ±0.14% for , ±6.8 ppm for HC, and ±0.01% for O2. A computer-based data acquisition system was used to monitor humidity and temperature inside the refrigerator, evaporator inlet and outlet temperatures, and fuel flow rate. Air humidity was measured with an uncertainty of ±0.06%. Cooling capacity and coefficient of performance were evaluated with uncertainties of ±1.84Wand ±0.0014, respectively. All uncertainties of measurements were evaluated according to ABNT/INMETRO [23] (similar to NIST TN 1297), for a confidence level of 95%. When powering the refrigerator with the engine exhaust gas it was verified that, for engine speeds over 2000 rev/min, the temperature in the refrigerator was increased. At such condition there was excessive energy being transferred from the high-temperature exhaust gas to the refrigerant, not allowing its condensation in the condenser due to the elevated sensible heat to be removed. As a consequence, the refrigerant temperature in the evaporator was above that inside the refrigerator. Thus, it was decided to perform the tests at a fixed engine speed of 1500 rev/min to avoid increasing temperature inside the refrigerator.



Figs. 5.1.–5.4. show the results obtained from the absorption refrigeration system using the engine exhaust gas as energy source for 25%, 50%, 75% and wide-open engine throttle valve. More than 800 data points were used to build each curve shown by the figures.

Fig.5.1. Time Variation of Refrigerator Average Temperature.

Fig.5.1. presents the refrigerator average internal temperature at those engine operational conditions. In the first 30–45 min it is observed that the average refrigerator temperature increases with time. This phenomenon is characteristic of the system start-up working regime when there is a refrigerant hot flow inside the refrigerator evaporator that would heat the refrigerator interior while the refrigeration system cooling capacity increases (see Fig.5.3). After 30–45 min, depending on engine valve opening, the refrigeration system reaches the steady state condition and its interior starts cooling down, leading to the reduction of the refrigerator average temperature (Fig.5.1). The temperature dropped faster inside the refrigerator as the throttle valve opening was wider. Attainment of the refrigerator steady state temperature was faster as the engine throttle valve was widened. Overall, the refrigeration average internal temperature took around 3–3.5 h to reach the steady state condition, agreeing with the observations by Koehler et al. [15], whose system, developed for truck refrigeration, was indicated for long distance driving. The wider the valve opening, the lower the steady state average temperature attained inside the refrigerator. This internal temperature varied between 5 and , well below that obtained by Jiangzhou et al. [18] from a dedicated system. The absorption refrigeration system was shut down after approximately 3 h, when the steady state temperature was attained. After system shut down, the internal average refrigerator temperature started to increase. A longer period was required to attain a thermal equilibrium with the external environment temperature. This period was larger for widened engine valve openings.

Fig.5.2 Time Variation of Relative Humidity Inside the Refrigerator.

Fig.5.2. shows that, as the averaged temperature inside the refrigerator was risen from start up (see Fig.5.1), the relative humidity was also increased for all valve openings. A high slope reduction on the relative humidity inside the refrigerator was observed from the time the temperature started to fall (see Fig. 5.1), until reaching the refrigerator steady state condition. This trend is attributed to water vapor inside the refrigerator being condensed and, eventually, turning into ice with temperature decrease, thus reducing air humidity. The steady state relative humidity recorded was between 29%, for 25% throttle opening, and 35%, for wide open throttle. From the moment the refrigeration system was shut down the relative humidity showed a slight rise, and then, rose up at a high slope to values between 91% and 96%, for all valve openings. That is explained by water vapor being formed from ice melting with temperature increase after refrigeration system shut down, thus increasing air humidity.

Fig.5.3. Time Variation of Cooling Capacity.

In Fig.5.3. is shown the time variation of the absorption refrigeration system cooling capacity. Cooling capacity increases from the very beginning of the refrigeration system start up until reaching a maximum when the steady state condition is attained in the refrigerator interior, dropping from then on. Unlike the previous figures, here the results are shown until refrigeration system shut down. The maximum cooling capacity obtained for the tested refrigerator was between 14.9 W, for 25% throttle valve opening, and 18.4 W, for wide-open throttle. That is well below the value obtained by Jiangzhou et al. [18] from a dedicated system. However, the engine exhaust gas thermal power availability shown in Fig. 5.4, calculated by Eqns. but considering the exhaust gas would leave the heat exchanger at the ambient temperature, presents similar values to that obtained by those authors. For determination of the coefficient of performance, heat transfer from the exhaust gas was calculated with the outlet enthalpy evaluated at the measured exhaust gas temperature leaving the heat exchanger.

Fig.5.4. Engine Exhaust Gas Power Availability.

From Fig. 5.4. it can be inferred that the available exhaust gas power increases with wider valve opening. Power availability for 25% valve opening corresponds to approximately 18.8% that of 75% valve opening, and 18.4% that of wide-open valve power availability. These values characterize an exponential growth of the exhaust gas power availability as a function of engine valve opening. For 25% valve opening, the system shows the lowest cooling capacity (Fig. 5.3) and the highest COP (Fig.5.5). For wider valve openings the cooling capacity increases and the COP decreases, due to higher exhaust gas power availability. The cooling capacity and the COP obtained for 25% valve opening are close to the values reported in [4] for the refrigerator operation with liquefied petroleum gas as original fuel. The present results showed that there is no need to use all the available exhaust gas energy to operate the absorption refrigeration system for throttle valve openings wider than 25%. In such situations, a control system for the exhaust gas mass flow rate may be a useful tool to optimize the absorption refrigeration system operation for automotive application.

Fig.5.5. Time Variation of Refrigerator Coefficient of Performance.

Fig.5.5. shows the time variation of the refrigerator coefficient of performance until system shut down after the steady state condition was attained. The maximum coefficient of performance was 4.9%, for 25% open throttle. That is about five times lower than the COP obtained by Koehler et al. [15] from a system designed for truck refrigeration. For all other valve opening, the peak coefficient of performance attained was even lower, of about 1.2-1.4%. This result indicates that a dedicated system is necessary to take full advantage of the energy available in the exhaust gas (see Fig.5.4) and improve the coefficient of performance to reasonable values.

Figs 5.6--5.8 show the influence of the absorption refrigeration system installed in the engine exhaust on performance and emissions parameters. Results are presented for engine output power, specific fuel consumption, and carbon dioxide, carbon monoxide and hydrocarbon emissions. These results should be observed considering that, to introduce the refrigeration system the engine exhaust system was modified, as mentioned in the previous section.

Fig.5.6. Influence of Refrigerator on Engine Power Output.

Fig.5.7. Influence of Refrigerator on Specific Fuel Consumption.

Fig.5.6. shows that, with the refrigerator installed in the exhaust system the engine produced 20% more power in comparison to the original configuration, corresponding to 2.2kW at 25% throttle opening. On the other hand, the specific fuel consumption, that is, the fuel mass flow rate per unit power produced, was reduced for all valve openings, reaching up to 15% reduction at 25% valve opening (Fig.5.7). These results show that the pressure drop introduced in the exhaust system by the presence of the refrigeration system is low, in comparison to the other exhaust system components, which were removed. As the exhaust pressure was lower, the power required from the engine to expel the exhaust gas was reduced, thus increasing the available output power. As the fuel mass flow rate into the engine did not change substantially, the higher power output decreased the specific fuel consumption with the refrigerator installed.

Fig.5.8. Influence of Refrigerator on Carbon Dioxide Emissions.

Fig.5.9. Influence of Refrigerator on Carbon Monoxide Emissions.

Figs. 5.8 and 5.9 shows that carbon dioxide emissions increased and carbon monoxide emissions decreased in the presence of the refrigerator for all engine operating conditions tested, except by 25% valve opening, for which the trends were the opposite. For 50%, 75% and wide-open throttle the mass fuel consumption was slightly lower in the presence of the refrigerator, being the reason for the reduction of CO emissions. For 25% valve opening a higher fuel mass flow rate was recorded when the refrigerator was installed, being the reason for the increase of CO emissions at this condition. Fuel mass flow rate can be obtained through multiplying the engine output power (Fig.5.6) by the specific fuel consumption (Fig.5.7).

Fig.5.10. Influence of Refrigerator on Hydrocarbon Emissions.

Exhaust hydrocarbon emissions were consistently higher when the refrigerator was installed in the engine exhaust for all tested conditions (Fig.5.10). This is a consequence of lower pressure attained in the engine combustion chamber during the exhaust process, as the exhaust system pressure was lower. The lower exhaust pressure increased the unburned fuel mass flow rate out of the combustion chamber crevices and from the cylinder lubricating oil film into the burned gas, thus increasing unburned hydrocarbon formation (Sodré and Yates[24]).



v The engine exhaust gas was confirmed as a potential power source for absorption refrigeration systems.

v The domestic absorption refrigerator tested showed low coefficient of performance and did not provide the cooling capacity needed for automotive application. However, a dedicated absorption refrigeration system may be able to take advantage of the exhaust gas power availability and provide the cooling capacity required for automotive air conditioning.

v Introduction of the absorption refrigeration system in the engine exhaust system did not cause significant pressure drop in the exhaust flow, as the engine output power was increased and specific fuel consumption was decreased with removal of other exhaust system components.

v Overall, carbon monoxide emission was decreased when the absorption refrigerator was installed in the exhaust gas, while hydrocarbon emissions showed an increase.

v Changes in exhaust components concentration were a consequence of the major modifications in the exhaust system.


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