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testing of an air-cycle refrigeration system for road transportAbstractThe environmental attractions of air-cycle refrigeration are considerable. Following a thermodynamic design analysis, an air-cycle demonstrator plant was constructed within the restricted physical envelope of an existing Thermo King SL200 trailer refrigeration unit. This unique plant operated satisfactorily, delivering sustainable cooling for refrigerated trailers using a completely natural and safe working fluid. The full load capacity of the air-cycle unit at −20 °C was 7,8 kW, 8% greater than the equivalent vapour-cycle unit, but the fuel consumption of the air-cycle plant was excessively high. However, at part load operation the disparity in fuel consumption dropped from approximately 200% to around 80%. The components used in the air-cycle demonstrator were not optimised and considerable potential exists for efficiency improvements, possibly to the point where the air-cycle system could rival the efficiency of the standard vapour-cycle system at part-load operation, which represents the biggest proportion of operating time for most : Air conditioner; Refrigerated transport; Thermodynamic cycle; Air; Centrifuge compressor; Turbine expander COP, NomenclaturePRCompressor or turbine pressure ratioTAHeat exchanger side A temperature (K)TBHeat exchanger side B temperature (K)TinletInlet temperature (K)ToutletOutlet temperature (K)ηcompCompressor isentropic efficiencyηturbTurbine isentropic efficiencyηheat exchangerHeat exchanger effectiveness1. IntroductionThe current legislative pressure on conventional refrigerants is well known. The reason why vapour-cycle refrigeration is preferred over air-cycle refrigeration is simply that in the great majority of cases vapour-cycle is the most energy efficient option. Consequently, as soon as alternative systems, such as non-HFC refrigerants or air-cycle systems are considered, the issue of increased energy consumption arises over legislation affecting HFC refrigerants and the desire to improve long-term system reliability led to the examination of the feasibility of an air-cycle system for refrigerated transport. With the support of Enterprise Ireland and Thermo King (Ireland), the authors undertook the design and construction of an air-cycle refrigeration demonstrator plant at LYIT and QUB. This was not the first time in recent years that air-cycle systems had been employed in transport. NormalAir Garrett developed and commercialised an air-cycle air conditioning pack that was fitted to high speed trains in Germany in the 90s. As part of an European funded programme, a range of applications for air-cycle refrigeration were investigated and several demonstrator plants were constructed. However, the authors are unaware of any other case where a self-contained air-cycle unit has been developed for the challenging application of trailer King decided that the demonstrator should be a trailer refrigeration unit, since those were the units with the largest refrigeration capacity but presented the greatest challenges with regard to physical packaging. Consequently, the main objective was to demonstrate that an air-cycle system could fit within the existing physical envelop and develop an equivalent level of cooling power to the existing vapour-cycle unit, but using only air as the working fluid. The salient performance specifications for the existing Thermo King SL200 vapour-cycle trailer refrigeration unit are listed .It was not the objective of the exercise to complete the design and development of a new refrigeration product that would be ready for manufacture. To limit the level of resources necessary, existing hardware was to be used where possible with the recognition that the efficiencies achieved would not be optimal. In practical terms, this meant using the chassis and panels for an existing SL200 unit along with the standard diesel engine and circulation fans. The turbomachinery used for compression and expansion was adapted from commercial . Thermodynamic modelling and design of the demonstrator plantThe thermodynamics of the air-cycle (or the reverse ‘Joule cycle’) are adequately presented in most thermodynamic textbooks and will not be repeated here. For anything other than the smallest flow rates, the most efficient machines available for the necessary compression and expansion processes are turbomachines. Considerations for the selection of turbomachinery for air-cycle refrigeration systems have been presented and discussed by Spence et al. [3]. a typical configuration of an air-cycle system, which is sometimes called the ‘boot-strap’ configuration. For mechanical convenience the compression process is divided into two stages, meaning that the turbine is not constrained to operate at the same speed as the primary compressor. Instead, the work recovered by the turbine during expansion is utilised in the secondary compressor. The two-stage compression also permits intercooling, which enhances the overall efficiency of the compression process. An ‘open system’ where the cold air is ejected directly into the cold space, removing the need for a heat exchanger in the cold space. In the interests of efficiency, the return air from the cold space is used to pre-cool the compressed air entering the turbine by means of a heat exchanger known as the ‘regenerator’ or the ‘recuperato ’. To support the design of the air-cycle demonstrator plant, and the selection of suitable components, a simple thermodynamic model of the air-cycle configuration shown in was developed. The compression and expansion processes were modelled using appropriate values of isentropic efficiency, as defined in heat exchange processes were modelled using values of heat exchanger effectiveness as defined in The model also made allowance for heat exchanger pressure drop. The system COP was determined from the ratio of the cooling power delivered to the power input to the primary compressor, as defined in illustrate air-cycle performance characteristics as determined from the thermodynamic model:illustrates the variation in air-cycle COP and expander outlet temperature over a range of cycle pressure ratios for a plant operating between −20 °C and +30 °C. The cycle pressure ratio is defined as the ratio of the maximum cycle pressure at secondary compressor outlet to the pressure at turbine outlet. For the ideal air-cycle, with no losses, the cycle COP increases with decreasing cycle pressure ratio and tends to infinity as the pressure ratio approaches unity. However, the introduction of real component efficiencies means that there is a definite peak value of COP that occurs at a certain pressure ratio for a particular cycle. However,illustrates, there is a broad range of pressure ratio and duty over which the system can be operated with only moderate variation of class of turbomachinery suitable for the demonstrator plant required speeds of around 50 000 rev/min. To simplify the mechanical arrangement and avoid the need for a high-speed electric motor, the two-stage compression system shown was adopted. The existing Thermo King SL200 chassis incorporated a substantial system of belts and pulleys to power circulation fans, which severely restricted the useful space available for mounting heat exchangers. A simple thermodynamic model was used to assess the influence of heat exchanger performance on the efficiency of the plant so that the best compromise could be developed show the impact of intercooler and aftercooler effectiveness and pressure loss on the COP of the proposed two-stage system in incorporated an intercooler between the two compression stages. By dispensing with the intercooler and its associated duct work a larger aftercooler could be accommodated with improved effectiveness and reduced pressure loss. Analysis suggested that the improved performance from a larger aftercooler could compensate for the loss of the the impact of the recuperator effectiveness on the COP of the plant, which is clearly more significant than that of the other heat exchangers. As well as boosting cycle efficiency, increased recuperator effectiveness also moves the peak COP to a lower overall system pressure ratio. The impact of pressure loss in the recuperator is the same as for the intercooler and aftercooler shown in. The model did not distinguish between pressure losses in different locations; it was only the sum of the pressure losses that was significant. Any pressure loss in connecting duct work and headers was also lumped together with the heat exchanger pressure loss and analysed as a block pressure specific cooling capacity of the air-cycle increases with system pressure ratio. Consequently, if a higher system pressure ratio was used the required cooling duty could be achieved with a smaller flow rate of air. shows the mass flow rate of air required to deliver 7,5 kW of cooling power for varying system pressure the demonstrator system was to be based on commercially available turbomachinery, it became important to choose a pressure ratio and flow rate that could be accommodated efficiently by some existing compressor and turbine rotors. and were based on efficiencies of 81 and 85% for compression and expansion, respectively. While such efficiencies are attainable with optimised designs, they would not be realised using compromised turbocharger components. For the design of the demonstrator plant efficiencies of 78 and 80% were assumed to be realistically attainable for compression and turbomachinery efficiencies corresponded to higher cycle pressure ratios and flow rates in order to achieve the target cooling duty. The cycle design point was also compromised to help heat exchanger performance. The pressure losses in duct work and heat exchangers increased in proportion with the square of flow velocity. Selecting a higher cycle pressure ratio corresponded to a lower mass flow rate and also increased density at inlet to the aftercooler heat exchanger. The combined effect was a decrease in the mean velocity in the heat exchanger, a decrease in the expected pressure losses in the heat exchanger and duct work, and an increase in the effectiveness of the heat exchanger. Consequently, a system pressure ratio higher than the value corresponding to peak COP was chosen in order to achieve acceptable heat exchanger performance within the available physical space. The below optimum performance of turbomachinery and heat exchanger components, coupled with excessive bearing losses, meant that the predicted COP of the overall system dropped to around 0,41. The system pressure ratio at the design point was 2,14 and the corresponding mass flow rate of air was 0,278 kg/ moving the design point beyond the pressure ratio for peak COP, it was anticipated that the demonstrator plant would yield good part-load performance since the COP would not fall as the pressure ratio was reduced. Also, operating at part-load corresponded to lower flow velocities and anticipated improvements in heat exchanger performance. Part-load operation was achieved by reducing the speed of the primary compressor, resulting in a decrease in both pressure and mass flow rate throughout the . Prime mover and primary compressorThe existing diesel engine was judged adequate to power the demonstrator plant. The standard engine was a four cylinder, water cooled diesel engine fitted with a centrifugal clutch and all necessary ancillaries and was controlled by a microprocessor the thermodynamic model, the pressure ratio for the primary compressor was 1,70. The centrifugal compressor required a shaft speed of around 55 000 rev/min. Other alternatives were evaluated for primary compression with the aim of obtaining a suitable device that operated at a lower speed. Other commercially available devices such as Roots blowers and rotary piston blowers were all excluded on the basis of poor one-off gearbox was designed and manufactured as part of the project to step-up the engine shaft speed to around 55 000 rev/min. The gearbox was a two stage, three shaft unit which mounted directly on the end of the diesel engine and was driven through the existing centrifugal . Cold air unitThe secondary compressor and the expansion turbine were mounted on the same shaft in a free rotating unit. The combination of the secondary compressor and the turbine was designated as the ‘Cold Air Unit’ (CAU). While the CAU was mechanically equivalent to a turbocharger, a standard turbocharger would not satisfy the aerodynamic requirements efficiently since the pressure ratios and inlet densities for both the compressor and the turbine were significantly different from any turbocharger installation. Consequently, both the secondary compressor and the turbine stage were specially chosen and developed to deliver suitable turbochargers use plain oil fed journal bearings, which are low-cost, reliable and provide effective damping of shaft vibrations. However, plain bearings dissipate a substantial amount of shaft power through viscous losses in the oil films. A plain bearing arrangement for the CAU was expected to absorb 2–3 kW of mechanical power, which represented around 25% of the anticipated turbine power. Also, the clearances in plain bearings require larger blade tip clearances for both the compressor and the turbine with a consequential efficiency penalty. Given the pressurised inlet to the secondary compressor, the limited thrust capacity of the plain bearing arrangement was also a concern. A CAU utilising high-speed ball bearings, or air bearings, was identified as a preferable arrangement to plain bearings. Benefits would include greatly reduced bearing power losses, reduced turbomachinery tip clearance losses and increased thrust load capacity. However, adequate resources were not available to design a special one-off high speed ball bearing system. Consequently, a standard turbocharger plain bearing system was secondary compressor stage was a standard turbocharger compressor selected for a pressure ratio of 1,264. Secondary compressor and turbine selection were linked because of the requirement to balance power and match the speed. Since most commercial turbines are sized for high temperature (and consequently low density) air at inlet, a special turbine stage was developed for the application. Cost considerations precluded the manufacture of a custom turbine rotor, so a commercially available rotor was used. The standard turbine rotor blade profile was substantially modified and vaned nozzles for turbine inlet were designed to match the modified rotor, in line with previous turbine investigations at QUB (Spence and Artt,). An exhaust diffuser was also incorporated into the turbine stage in order to improve turbine efficiency and to moderate the exhaust noise levels through reduced air velocity. The exhaust diffuser exited into a specially designed exhaust performance of the turbine stage was measured before the unit was incorporated into the complete demonstrator plant. The peak efficiency of the turbine was established at 81%.5. Heat exchangersDue to packaging constraints, the heat exchangers had to be specially designed with careful consideration being given to heat exchanger position and header geometry in an attempt to achieve the best performance from the heat exchangers. Tube and fin aluminium heat exchangers, similar to those used in automotive intercooler applications, were chosen primarily because they could be produced on a ‘one-off’ basis at a reasonable cost. There were other heat exchanger technologies available that would have yielded better performance from the available volume, but high one-off production costs precluded their use in the demonstrator different tube and fin heat exchangers were tested and used to validate a computational model. Once validated, the model was used to assess a wide range of possible heat exchanger configurations that could fit within the Thermo King SL200 chassis. Fitting the proposed heat exchangers within the existing chassis and around the mechanical drive system for the circulation fans, but while still achieving the necessary heat exchanger performance was very challenging. It was clear that potential heat exchanger performance was being sacrificed through the choice of tube and fin construction and by the constraints of the layout of the existing SL200 chassis. The final selection comprised two separate aftercooler units, while the single recuperator was a large, triple pass unit. Based on laboratory tests and the heat exchanger model, the anticipated effectiveness of both the recuperator and aftercooler units was 80%.6. InstrumentationA range of conventional pressure and temperature instrumentation was installed on the air-cycle demonstrator plant. Air temperature and pressure was logged at inlet and outlet from each heat exchanger, compressor and the turbine. The speed of the primary compressor was determined from the speed measurement on the diesel engine control unit, while the cold air unit was equipped with a magnetic speed counter. No air flow measurement was included on the demonstrator plant. Instead, the air flow rate was deduced from the previously obtained turbine performance map using the measurements of turbine pressure ratio and rotational . System testingDuring some preliminary tests a heat load was applied and the functionality of the demonstrator plant was established. Having assessed that it was capable of delivering approximately the required performance, the plant was transported to a Thermo King calorimeter test facility specifically for measuring the performance of transport refrigeration units. The calorimeter was ideally suited for accurately measuring the refrigeration capacity of the air-cycle demonstrator plant. The calorimeter was operated according to standard ARI 1100-2001; the absolute accuracy was better than 200W and all auxiliary instrumentation was calibrated against appropriate performance capacity of transport refrigeration units is generally rated at two operating conditions; 0 and −20 °C, and both at an ambient temperature of +30 °C. Along with the specified operating conditions of 0 and −20 °C, a further part-load condition at −20 °C was assessed. Considering that the air-cycle plant was only intended to demonstrate a concept and that there were concerns about the reliability of the gearbox and the cold air unit thrust bearing, it was decided to operate the plant only as long as was necessary to obtain stabilised measurements at each operating point. The demonstrator plant operated satisfactorily, allowing sufficient measurements to be obtained at each of the three operating conditions. The recorded performance is summarised .In total, the unit operated for approximately 3 h during the course of the various tests. While the demonstrator plant operated adequately to allow measurements, some smoke from the oil system breather suggested that the thrust bearing of the CAU was heavily overloaded and would fail, as had been anticipated at the design stage. Testing was concluded in case the bearing failed completely causing the destruction of the entire CAU. There was no evidence of any gearbox deterioration during . Discussion of measured performanceFrom the calorimeter performance measurements, the primary objective of the project had been achieved. A unique air-cycle refrigeration system had been developed within the same physical envelope as the existing Thermo King SL200 refrigeration unit, w
暖通专业的论文，最好是发国家级或者核心期刊了，不过审核也相当严的，
暖通空调就很好了
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你是学建筑环境也设备工程的不
testing of an air-cycle refrigeration system for road transportAbstractThe environmental attractions of air-cycle refrigeration are considerable. Following a thermodynamic design analysis, an air-cycle demonstrator plant was constructed within the restricted physical envelope of an existing Thermo King SL200 trailer refrigeration unit. This unique plant operated satisfactorily, delivering sustainable cooling for refrigerated trailers using a completely natural and safe working fluid. The full load capacity of the air-cycle unit at −20 °C was 7,8 kW, 8% greater than the equivalent vapour-cycle unit, but the fuel consumption of the air-cycle plant was excessively high. However, at part load operation the disparity in fuel consumption dropped from approximately 200% to around 80%. The components used in the air-cycle demonstrator were not optimised and considerable potential exists for efficiency improvements, possibly to the point where the air-cycle system could rival the efficiency of the standard vapour-cycle system at part-load operation, which represents the biggest proportion of operating time for most : Air conditioner; Refrigerated transport; Thermodynamic cycle; Air; Centrifuge compressor; Turbine expander COP, NomenclaturePRCompressor or turbine pressure ratioTAHeat exchanger side A temperature (K)TBHeat exchanger side B temperature (K)TinletInlet temperature (K)ToutletOutlet temperature (K)ηcompCompressor isentropic efficiencyηturbTurbine isentropic efficiencyηheat exchangerHeat exchanger effectiveness1. IntroductionThe current legislative pressure on conventional refrigerants is well known. The reason why vapour-cycle refrigeration is preferred over air-cycle refrigeration is simply that in the great majority of cases vapour-cycle is the most energy efficient option. Consequently, as soon as alternative systems, such as non-HFC refrigerants or air-cycle systems are considered, the issue of increased energy consumption arises over legislation affecting HFC refrigerants and the desire to improve long-term system reliability led to the examination of the feasibility of an air-cycle system for refrigerated transport. With the support of Enterprise Ireland and Thermo King (Ireland), the authors undertook the design and construction of an air-cycle refrigeration demonstrator plant at LYIT and QUB. This was not the first time in recent years that air-cycle systems had been employed in transport. NormalAir Garrett developed and commercialised an air-cycle air conditioning pack that was fitted to high speed trains in Germany in the 90s. As part of an European funded programme, a range of applications for air-cycle refrigeration were investigated and several demonstrator plants were constructed. However, the authors are unaware of any other case where a self-contained air-cycle unit has been developed for the challenging application of trailer King decided that the demonstrator should be a trailer refrigeration unit, since those were the units with the largest refrigeration capacity but presented the greatest challenges with regard to physical packaging. Consequently, the main objective was to demonstrate that an air-cycle system could fit within the existing physical envelop and develop an equivalent level of cooling power to the existing vapour-cycle unit, but using only air as the working fluid. The salient performance specifications for the existing Thermo King SL200 vapour-cycle trailer refrigeration unit are listed .It was not the objective of the exercise to complete the design and development of a new refrigeration product that would be ready for manufacture. To limit the level of resources necessary, existing hardware was to be used where possible with the recognition that the efficiencies achieved would not be optimal. In practical terms, this meant using the chassis and panels for an existing SL200 unit along with the standard diesel engine and circulation fans. The turbomachinery used for compression and expansion was adapted from commercial . Thermodynamic modelling and design of the demonstrator plantThe thermodynamics of the air-cycle (or the reverse ‘Joule cycle’) are adequately presented in most thermodynamic textbooks and will not be repeated here. For anything other than the smallest flow rates, the most efficient machines available for the necessary compression and expansion processes are turbomachines. Considerations for the selection of turbomachinery for air-cycle refrigeration systems have been presented and discussed by Spence et al. [3]. a typical configuration of an air-cycle system, which is sometimes called the ‘boot-strap’ configuration. For mechanical convenience the compression process is divided into two stages, meaning that the turbine is not constrained to operate at the same speed as the primary compressor. Instead, the work recovered by the turbine during expansion is utilised in the secondary compressor. The two-stage compression also permits intercooling, which enhances the overall efficiency of the compression process. An ‘open system’ where the cold air is ejected directly into the cold space, removing the need for a heat exchanger in the cold space. In the interests of efficiency, the return air from the cold space is used to pre-cool the compressed air entering the turbine by means of a heat exchanger known as the ‘regenerator’ or the ‘recuperato ’. To support the design of the air-cycle demonstrator plant, and the selection of suitable components, a simple thermodynamic model of the air-cycle configuration shown in was developed. The compression and expansion processes were modelled using appropriate values of isentropic efficiency, as defined in heat exchange processes were modelled using values of heat exchanger effectiveness as defined in The model also made allowance for heat exchanger pressure drop. The system COP was determined from the ratio of the cooling power delivered to the power input to the primary compressor, as defined in illustrate air-cycle performance characteristics as determined from the thermodynamic model:illustrates the variation in air-cycle COP and expander outlet temperature over a range of cycle pressure ratios for a plant operating between −20 °C and +30 °C. The cycle pressure ratio is defined as the ratio of the maximum cycle pressure at secondary compressor outlet to the pressure at turbine outlet. For the ideal air-cycle, with no losses, the cycle COP increases with decreasing cycle pressure ratio and tends to infinity as the pressure ratio approaches unity. However, the introduction of real component efficiencies means that there is a definite peak value of COP that occurs at a certain pressure ratio for a particular cycle. However,illustrates, there is a broad range of pressure ratio and duty over which the system can be operated with only moderate variation of class of turbomachinery suitable for the demonstrator plant required speeds of around 50 000 rev/min. To simplify the mechanical arrangement and avoid the need for a high-speed electric motor, the two-stage compression system shown was adopted. The existing Thermo King SL200 chassis incorporated a substantial system of belts and pulleys to power circulation fans, which severely restricted the useful space available for mounting heat exchangers. A simple thermodynamic model was used to assess the influence of heat exchanger performance on the efficiency of the plant so that the best compromise could be developed show the impact of intercooler and aftercooler effectiveness and pressure loss on the COP of the proposed two-stage system in incorporated an intercooler between the two compression stages. By dispensing with the intercooler and its associated duct work a larger aftercooler could be accommodated with improved effectiveness and reduced pressure loss. Analysis suggested that the improved performance from a larger aftercooler could compensate for the loss of the the impact of the recuperator effectiveness on the COP of the plant, which is clearly more significant than that of the other heat exchangers. As well as boosting cycle efficiency, increased recuperator effectiveness also moves the peak COP to a lower overall system pressure ratio. The impact of pressure loss in the recuperator is the same as for the intercooler and aftercooler shown in. The model did not distinguish between pressure losses in different locations; it was only the sum of the pressure losses that was significant. Any pressure loss in connecting duct work and headers was also lumped together with the heat exchanger pressure loss and analysed as a block pressure specific cooling capacity of the air-cycle increases with system pressure ratio. Consequently, if a higher system pressure ratio was used the required cooling duty could be achieved with a smaller flow rate of air. shows the mass flow rate of air required to deliver 7,5 kW of cooling power for varying system pressure the demonstrator system was to be based on commercially available turbomachinery, it became important to choose a pressure ratio and flow rate that could be accommodated efficiently by some existing compressor and turbine rotors. and were based on efficiencies of 81 and 85% for compression and expansion, respectively. While such efficiencies are attainable with optimised designs, they would not be realised using compromised turbocharger components. For the design of the demonstrator plant efficiencies of 78 and 80% were assumed to be realistically attainable for compression and turbomachinery efficiencies corresponded to higher cycle pressure ratios and flow rates in order to achieve the target cooling duty. The cycle design point was also compromised to help heat exchanger performance. The pressure losses in duct work and heat exchangers increased in proportion with the square of flow velocity. Selecting a higher cycle pressure ratio corresponded to a lower mass flow rate and also increased density at inlet to the aftercooler heat exchanger. The combined effect was a decrease in the mean velocity in the heat exchanger, a decrease in the expected pressure losses in the heat exchanger and duct work, and an increase in the effectiveness of the heat exchanger. Consequently, a system pressure ratio higher than the value corresponding to peak COP was chosen in order to achieve acceptable heat exchanger performance within the available physical space. The below optimum performance of turbomachinery and heat exchanger components, coupled with excessive bearing losses, meant that the predicted COP of the overall system dropped to around 0,41. The system pressure ratio at the design point was 2,14 and the corresponding mass flow rate of air was 0,278 kg/ moving the design point beyond the pressure ratio for peak COP, it was anticipated that the demonstrator plant would yield good part-load performance since the COP would not fall as the pressure ratio was reduced. Also, operating at part-load corresponded to lower flow velocities and anticipated improvements in heat exchanger performance. Part-load operation was achieved by reducing the speed of the primary compressor, resulting in a decrease in both pressure and mass flow rate throughout the . Prime mover and primary compressorThe existing diesel engine was judged adequate to power the demonstrator plant. The standard engine was a four cylinder, water cooled diesel engine fitted with a centrifugal clutch and all necessary ancillaries and was controlled by a microprocessor the thermodynamic model, the pressure ratio for the primary compressor was 1,70. The centrifugal compressor required a shaft speed of around 55 000 rev/min. Other alternatives were evaluated for primary compression with the aim of obtaining a suitable device that operated at a lower speed. Other commercially available devices such as Roots blowers and rotary piston blowers were all excluded on the basis of poor one-off gearbox was designed and manufactured as part of the project to step-up the engine shaft speed to around 55 000 rev/min. The gearbox was a two stage, three shaft unit which mounted directly on the end of the diesel engine and was driven through the existing centrifugal . Cold air unitThe secondary compressor and the expansion turbine were mounted on the same shaft in a free rotating unit. The combination of the secondary compressor and the turbine was designated as the ‘Cold Air Unit’ (CAU). While the CAU was mechanically equivalent to a turbocharger, a standard turbocharger would not satisfy the aerodynamic requirements efficiently since the pressure ratios and inlet densities for both the compressor and the turbine were significantly different from any turbocharger installation. Consequently, both the secondary compressor and the turbine stage were specially chosen and developed to deliver suitable turbochargers use plain oil fed journal bearings, which are low-cost, reliable and provide effective damping of shaft vibrations. However, plain bearings dissipate a substantial amount of shaft power through viscous losses in the oil films. A plain bearing arrangement for the CAU was expected to absorb 2–3 kW of mechanical power, which represented around 25% of the anticipated turbine power. Also, the clearances in plain bearings require larger blade tip clearances for both the compressor and the turbine with a consequential efficiency penalty. Given the pressurised inlet to the secondary compressor, the limited thrust capacity of the plain bearing arrangement was also a concern. A CAU utilising high-speed ball bearings, or air bearings, was identified as a preferable arrangement to plain bearings. Benefits would include greatly reduced bearing power losses, reduced turbomachinery tip clearance losses and increased thrust load capacity. However, adequate resources were not available to design a special one-off high speed ball bearing system. Consequently, a standard turbocharger plain bearing system was secondary compressor stage was a standard turbocharger compressor selected for a pressure ratio of 1,264. Secondary compressor and turbine selection were linked because of the requirement to balance power and match the speed. Since most commercial turbines are sized for high temperature (and consequently low density) air at inlet, a special turbine stage was developed for the application. Cost considerations precluded the manufacture of a custom turbine rotor, so a commercially available rotor was used. The standard turbine rotor blade profile was substantially modified and vaned nozzles for turbine inlet were designed to match the modified rotor, in line with previous turbine investigations at QUB (Spence and Artt,). An exhaust diffuser was also incorporated into the turbine stage in order to improve turbine efficiency and to moderate the exhaust noise levels through reduced air velocity. The exhaust diffuser exited into a specially designed exhaust performance of the turbine stage was measured before the unit was incorporated into the complete demonstrator plant. The peak efficiency of the turbine was established at 81%.5. Heat exchangersDue to packaging constraints, the heat exchangers had to be specially designed with careful consideration being given to heat exchanger position and header geometry in an attempt to achieve the best performance from the heat exchangers. Tube and fin aluminium heat exchangers, similar to those used in automotive intercooler applications, were chosen primarily because they could be produced on a ‘one-off’ basis at a reasonable cost. There were other heat exchanger technologies available that would have yielded better performance from the available volume, but high one-off production costs precluded their use in the demonstrator different tube and fin heat exchangers were tested and used to validate a computational model. Once validated, the model was used to assess a wide range of possible heat exchanger configurations that could fit within the Thermo King SL200 chassis. Fitting the proposed heat exchangers within the existing chassis and around the mechanical drive system for the circulation fans, but while still achieving the necessary heat exchanger performance was very challenging. It was clear that potential heat exchanger performance was being sacrificed through the choice of tube and fin construction and by the constraints of the layout of the existing SL200 chassis. The final selection comprised two separate aftercooler units, while the single recuperator was a large, triple pass unit. Based on laboratory tests and the heat exchanger model, the anticipated effectiveness of both the recuperator and aftercooler units was 80%.6. InstrumentationA range of conventional pressure and temperature instrumentation was installed on the air-cycle demonstrator plant. Air temperature and pressure was logged at inlet and outlet from each heat exchanger, compressor and the turbine. The speed of the primary compressor was determined from the speed measurement on the diesel engine control unit, while the cold air unit was equipped with a magnetic speed counter. No air flow measurement was included on the demonstrator plant. Instead, the air flow rate was deduced from the previously obtained turbine performance map using the measurements of turbine pressure ratio and rotational . System testingDuring some preliminary tests a heat load was applied and the functionality of the demonstrator plant was established. Having assessed that it was capable of delivering approximately the required performance, the plant was transported to a Thermo King calorimeter test facility specifically for measuring the performance of transport refrigeration units. The calorimeter was ideally suited for accurately measuring the refrigeration capacity of the air-cycle demonstrator plant. The calorimeter was operated according to standard ARI 1100-2001; the absolute accuracy was better than 200W and all auxiliary instrumentation was calibrated against appropriate performance capacity of transport refrigeration units is generally rated at two operating conditions; 0 and −20 °C, and both at an ambient temperature of +30 °C. Along with the specified operating conditions of 0 and −20 °C, a further part-load condition at −20 °C was assessed. Considering that the air-cycle plant was only intended to demonstrate a concept and that there were concerns about the reliability of the gearbox and the cold air unit thrust bearing, it was decided to operate the plant only as long as was necessary to obtain stabilised measurements at each operating point. The demonstrator plant operated satisfactorily, allowing sufficient measurements to be obtained at each of the three operating conditions. The recorded performance is summarised .In total, the unit operated for approximately 3 h during the course of the various tests. While the demonstrator plant operated adequately to allow measurements, some smoke from the oil system breather suggested that the thrust bearing of the CAU was heavily overloaded and would fail, as had been anticipated at the design stage. Testing was concluded in case the bearing failed completely causing the destruction of the entire CAU. There was no evidence of any gearbox deterioration during . Discussion of measured performanceFrom the calorimeter performance measurements, the primary objective of the project had been achieved. A unique air-cycle refrigeration system had been developed within the same physical envelope as the existing Thermo King SL200 refrigeration unit, w
建筑热能通风空调、制冷与空调、城市建筑等等均可
暖通空调杂志就可以的
《暖通空调》创刊于 1971 年，是中国建筑科学类核心期刊， 国家期刊奖最高奖项获奖期刊， 中国暖通空调行业惟一的中央级科技期刊，由建设部主管， 亚太建设科技信息研究院、 中国建筑设计研究院、 中国建筑学会（暖通空调专业委员会）联合主办。 本刊以实用技术为主，兼具学术性和信息性，在行业中最具影响力，被誉为权威刊物，深受广大读者喜爱，发行量在国内同行业刊物中遥遥领先。 《暖通空调》始终以 “ 新颖、实用、准确、精练 ” 为办刊方针，以提高全行业素质、推动全行业技术交流与发展为宗旨，及时报道国家有关建筑节能和环境保护的重大技术政策，建筑环境与设备工程中供暖、通风、空调、制冷及洁净技术方面的研究成果、学术论文、先进技术、工程总结、设计经验、设备开发与运行管理以及行业学术活动与设备市场信息。 《暖通空调》是世界最著名的建筑专业数据库 —— 国际建筑文献数据库 ICONDA 收录期刊，中国科技论文与引文数据库统计分析数据源刊，中国科学引文数据库来源期刊，中国学术期刊综合评价数据库统计源期刊，中国核心期刊（遴选）数据库收录期刊，中国期刊全文数据库收录期刊。 《暖通空调》栏目设置：专题研讨、科技综述、标准规范、专业论坛、专题讲座、设备开发、设计参考、工程实例、技术交流、运行管理。 《暖通空调》发行对象：从事建筑环境与设备工程中供暖、通风、空调、制冷、洁净等相关领域的工程设计、科研教学、施工安装、设备制造、运行管理的专业技术人员、管理人员、院校师生、房地产开发商和业主，以及对暖通空调制冷技术感兴趣的各界朋友。 编辑单位：《暖通空调资讯》编辑部总编：王曙明执行总编：潘晓福执行主编：刘昊编辑部地址：常州市新北区黄山路99-5号4楼
暖通方面的论文在品学论文网很多的哦，你可以参考下，如果还有不清楚的地方，可以咨询下他们的在线辅导老师，我之前也是求助他们帮忙的，很快就给我了，当时还是品学论文的王老师帮忙的，态度不错，呵呵，相对于一些小机构和个人要靠谱的多
为了避免工程技术给人类社会与自然界可能带来的负面影响,必须对工程技术进行伦理控制。下面是我为大家整理的工程技术论文，供大家参考。
【摘要】光伏发电因其绿色环保、无污染、可再生等特点，在当前我国全面建成小康社会重要攻坚时期的社会经济形势下，大力发展光伏发电已经成为推进能源结构调整、促进各个地区经济健康可持续发展的重要改革 措施 。随着光伏发电的进一步推广和应用，电子信息工程技术将会起到越来越重要的作用，研究电子信息工程技术在光伏电场中的实践应用具有十分重要的现实意义。本文从相关概念切入话题，探讨光伏电场中应用电子信息工程技术的重要意义，并对其应用的基本原理和具体应用措施进行简要的分析。
【关键词】电子信息工程;光伏电场;实践应用
光伏发电是当前较为前沿和具有广阔发展前景的新型发电方式，其因为自身的绿色、无污染及可再生等特点受到社会各界的广泛关注。由于我国疆域辽阔，纬度跨越较大，光照资源极其丰富，所以在我国研究光伏发电相关问题具有十分重要的现实意义。据专家估计，到十三五结束时，我国的光伏发电将会占到全国总电力装机的6%左右，大量的光伏电场将会相继建成并且投入使用。在光伏电场中，电子信息工程技术也发挥着至关重要的作用，成为影响光伏发电技术不断向前进步的重要因素之一，研究电子信息工程技术在光伏电场中的应用不仅仅能够促进光伏发电技术的发展，对于电子信息工程技术本身也具有重要意义。
1相关概念综述
光伏发电中的“光伏”，实际上指的是光生伏特效应，即我们常说的光伏效应，它指的是半导体在受到光照射时能够产生电动势的现象。当前最为广泛的应用就是制作各种光电池等等，进一步发展为光伏发电。
光伏发电中的光主要指的是太阳光，光伏发电指的就是利用光生伏特效应基本原理，利用特制的太阳能电池，将太阳光能直接转化为电能的全部过程。由于太阳光是一种非常绿色环保，不会产生污染并且从某种程度上来说是取之不尽、用之不竭的能源，所以当前光伏发电已经成为受到广泛关注的一种新型能源利用方式。
电子信息工程则是依托于计算机技术发展的一门应用学科，它只要研究的对象是电子信息的处理和控制等等。基于电子信息业在当前已经成为全国五大支柱产业之一，电子信息工程专业在当前也成为非常热门的学科和专业。而光伏电场中的电子信息工程技术应用在当前仍然局限在电子信息工程技术专业本身的特点和范畴内，其主要发挥的作用仍然是信息的获取和处理。
2电子信息工程技术在光伏电场中应用的重要意义
电子信息工程技术在光伏电场中得以广泛应用，对于光伏发电的发展具有十分重要的现实意义，主要表现在以下两个方面：首先，它能够在获取数据、处理数据方面更加精确，为光伏电场作业提供更加准确的数据依据。要知道，光伏发电中基本上都是电子元器而很少有机械原件，相较起来更容易发生各种故障，需要做好更为精准的监控和控制。并且在光伏电场中，各项传感器测量的参数需要非常精确，参数的细微差别将会对整个发电系统的监控和处理都产生巨大的影响。其次，它大大解放了人力和物力资源，能够以充足的资源投入到更多的方面去确保光伏发电系统的正常运行。在计算机没有广泛应用之前，发电站的数据监测和处理只能够依靠人力，不仅给工作人员带来了巨大的工作压力，也容易出现各种细微的谬误。电子信息工程技术作为一项在当前非常成熟的技术，无论是数据监测还是数据采集又或者是数据统计都非常快捷和精确，解放了大量的人力物力。
3电子信息工程技术在光伏电场中应用的实际应用
电子信息工程技术在光伏电场中的实际应用主要表现在四个方面，分别是数据测量、数据采集、数据分析和数据统计。首先，数据测量中的实际应用。传感器是光伏发电中最重要的部分之一，其主要承担的是数据测量的重要任务。传感器测量的数据是否准确将会对整个发电系统产生巨大影响。电子信息工程技术的发展使得传感器测量的周期性误差、偶然性误差、量化性误差都进一步降低，测量数据更加精确。其次，数据采集中的实际应用。传感器可不仅仅是进行数据测量，其在测量出数据以后，会进一步进行数据采集并进行传送。在电子信息工程技术广泛应用之前，数据的采集和传输需要进行模拟转换，需要将数据先转化为模拟信号，再转化为数字信息，很容易出现失真情况。而电子信息工程技术可以将数据直接传输，最大可能地确保数据的精确性。再次，数据分析中的实际应用。这里的数据分析并不像字面上说的那样仅仅进行数据的分析，电子工程技术发展到今天甚至能够直接根据数据进行决策。举例来说，光能相较于水能来说，可控性更差，所以很容易出现孤岛现象，而利用电子信息工程技术，光伏并网的决策系统就能够在受到异常波形时及时作出分析和决策。最后，数据统计中的实际应用。传统的数据统计依赖于人力，容易出现错误。而数据统计在光伏发电中起到的作用是非常重要的，电场通过长期对数据的测量、收集和分析，能够据此作出进一步的决策和改善。电子信息工程技术的发展能够有效地统计电场运行以来的各项数据，对光伏发电过程不断改进，使其能够更加稳定、高效率地运行和发展。
4结语
当前的时代是计算机的时代和网络的时代，严格意义上来说电子信息工程技术已经不是一门前沿的学科，而成为在现实生活中应用非常广泛的成熟学科。但是由于电子信息工程技术本身无穷无尽的发展潜力，其可以与很多前沿的学科和实践活动相结合，形成创新性的实践应用，在光伏电场中发挥重要作用就是电子信息工程技术近些年来与实践领域相结合的最好例证。当前电子信息工程技术在光伏电场中的实际应用主要是在处理数据方面，最得到广泛应用的是在数据测量、数据采集、数据分析和数据统计中的应用，其仍然没有摆脱电子信息工程技术本身的特点。未来随着电子信息工程技术的不断发展和光伏发电的不断发展，相信二者会有更多的结合，为全面发展我国社会经济提供重要的基础性保障。
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[1]王本煜.电子信息工程技术在光伏电场中的应用[J].电子制作,2015,0(12):111~112.
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[3]秦志龙.计及相关性的含风电场和光伏电站电力系统可靠性评估[D].重庆:重庆大学,2013,08(11):101~102.
【 文章 摘要】人才培养方案的制定关乎学校的生存和发展。本文根据陕西国防工业职业技术学院在国家级骨干示范院校建设对供热通风与空调工程技术专业人才培养方案的制定中，对有关人才培养模式和教学模式制定的改革探索。
【关键词】人才培养方案;供热通风与空调工程技术专业;人才培养模式;教学模式
0引言
人才培养方案的制定关乎一个学校的生存和发展。本文根据陕西国防工业职业技术学院在国家级骨干示范院校建设中对供热通风与空调工程技术专业人才培养方案的制定中，对有关人才培养模式和教学模式的改革探索，从而促进 教育 教学的发展。
1我院供热通风与空调工程技术专业人才培养模式的构建
我院在供热通风与空调工程技术专业人才培养模式构建中，依托西安大金空调有限公司、海尔空调工程有限公司等校企合作工作站，以就业为导向，以空调工程施工为载体，以供热通风与空调工程技术企业岗位职业能力培养为主线，引入制冷行业职业技能鉴定标准，参照职业岗位任职要求，由行业企业的专家与学校共同构建工作过程系统化课程体系，共同设计、制订、实施人才培养方案.
理论学习阶段的构建
理论学习是指公共基础学习领域、专业基础学习领域、专业核心学习领域及拓展学习领域相关理论课程的学习。在此阶段，一部分课程采用理论学习与技能训练交替进行，一部分课程采用“教、学、做”于一体的教学模式，遵循学生认知规律，灵活应用讲授法、任务驱动法、项目导向法、案例分析法、角色扮演法、现场教学法等 教学 方法 循序渐进、由浅入深地安排课程内容，使学生在“做中学”，从而实现知识及能力的逐级提升。
岗位实操阶段的构建
在校内理论学习、技能训练及模拟训练的基础上，在订单培养企业岗位进行生产实习及顶岗实习，进行和企业产品生高职供热通风与空调工程技术专业人才培养方案制定的探索曹振华陕西国防工业职业技术学院建筑与热能工程学院西安710302产相适应的专业核心课程学习，形成“边工作边学习，为工作而学习”的教学模式。顶岗实习时，学生在实习基地以职业人的身份参与企业生产活动，承担工作岗位规定的责任和义务，增加了学生对生产过程包括设计原理、生产设备、工艺流程、 规章制度 等的切身认识，使学生及时掌握最新工艺和技能，强化学生的专业能力、协作精神和责任意识，使学生的课堂知识真正转化成工作能力。并引入供热通风与空调工程技术专业相关的国家职业资格考试，要求学生获得相应的职业技能资格证书(如：制冷工、钣金工等)，实现人才培养规格与社会用人单位岗位需求的最大限度接轨。
2我院供热通风与空调工程技术专业教学模式的构建
我院针对供热通风与空调工程技术的专业特点和相关企业对高职人才能力的要求，以校内、外实训基地为载体，共同实施“6学期3阶段”的多学期、分段式教学组织模式。具体如下：第一阶段：第1、2学期，本阶段完成专业通用能力的培养。在学校进行公共基础领域、专业基本学习领域课程的理论学习及专业通用能力训练。
让学生学习相关的 公共基础知识 和专业基础知识，在校内、外实训基地及国防教育基地完成制冷基本技能操作训练和国防 拓展训练 ，在企业进行专业认知实习，了解专业具体产品生产组织、生产工艺，加强学生间的交流、合作与自我学习等能力的培养，将职业素质教育渗透到教学过程中，将校园 文化 与军工文化相融合，实现学生达到制冷行业通用能力的培养目标。第二阶段：第3、4、5学期，本阶段完成专业核心能力培养。第3、4学期，完成专业核心领域课程的理论学习，在校内实训基地完成专业核心能力技能训练、课程仿真训练及综合仿真训练。
充分利用校内实训资源，选择典型工程施工或设备做为教学载体，开展教学活动。[3]获取专业技能证书，实行“双证书”制。第5学期，利用3周在生产现场进行实习，利用12周完成专业拓展课程学习，拓展专业视野，为就职可能面临的转岗、转业做好准备。后7周进行 毕业 设计，也可在企业边进行生产实习边完成，在校外实训基地根据岗位实际生产进行选题，通过实操，进一步掌握工程管理、设备维护等相关知识，获取企业上岗证书。
或利用前13周在订单企业结合企业产品生产工艺完成专业校企合作开发课程的学习。第三阶段：第6学期，本阶段完成专业综合能力培养。学生到校外实训基地或订单企业顶岗实习，校企共同制订顶岗实习标准，将就业与实习有机结合，在真实的职业情境中，培养学生的专业综合能力。
学生与企业签订顶岗实习协议，以企业员工的身份参与企业生产，企业技术人员现场指导，专职教师负责实习辅导和学生管理。在实习过程中企业与学校联合对学生进行质量教育、成本教育、保密教育和 安全教育 ，培养学生的职业道德、职业技能及国防精神。
3 总结
随着高职高专教育教学改革步伐的不断加快，我们对高职供热通风与空调工程技术专业课程改革的认识也在逐渐加深，我们将随着社会和企业的需求不断及时修正人才培养方案，不断探索科学的教学评价和考核方式，培养出合乎社会要求的和一批批理论扎实、实践能力过硬的供热通风与空调工程技术专业高技能应用型人才。
【参考文献】
[1]戴路玲;涂中强.高职制冷专业校企合作、工学结合人才培养模式建设[J].供热通风与空调工程技术(四川),2009(05):89-92.
[2]吕君;宋永军.供热通风与空调工程技术专业办学模式的探索——以黑龙江建筑职业技术学院为例[J].中国科技信息,2012(23):160.
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工程技术论文范文三：化学生产中化学工程技术的应用
摘要：随着我国科学技术的不断发展，化学工程技术在化学生产中的应用越来越广泛。化学工程技术作为化学生产中重要的一项技术，不仅能够有效的节约在化学生产中所需要的时间，而且还能够提高化学工程的生产效率。因此，本文通过对化学工程技术的技术概念进行了阐述后，又详细的介绍了超临界流体技术、传热技术以及绿色化学反应技术在化学生产中的应用，并且分析了现如今的化学工程技术存在的问题，同时提出了相应的对策，从而使得化学工程技术在化学生产中能够有更好的发展。
关键词：化学工程技术;化学生产;应用;分析
在我国，科学技术一直是我们的一项重要的生产技术，随着科技的快速发展，在化学生产过程中也开始广泛的采用化工技术。化学工程技术主要是一项研究化学生产过程中需要采用的相关技术，其主要目的是对化学工程产品进行开发、设计、制造和管理。由于化学工程技术能够有效的提高产品的质量，同时也能够提升化学生产中的工作效率，因此我们对化学工程技术有了更广泛的关注，并不断的将其拓展到化学生产中的各个领域，使得化学工程技术能够发展的更好，进而不断的推进我国的经济发展和科技发展，使我们的生活条件更加优越。
1化学工程技术的技术概念阐述
现如今，化学产品已经成为了人们生活中非常常见的物品，例如药物、食品和日用品，还有农业药物和工厂生产所需的原料等等。因此化学工程技术变成为了一项炙手可热的技术，不断的受到人们的关注。化学工程技术是根据化学理论基础与相关的技术相结合的一项应用于化学生产中的技术，利用化学设备，通过一系列的化学反应进行产品的大量生产。在化学生产的过程中，化学的反应物和设备对于工程的技术要求是非常高的，而化学工程技术的优势就在于能够满足化学反应的要求，进而提高了化学产品的质量。除此之外，化学工程技术还有一项更大的优势就是对废物的处理，这项技术能够尽可能不对环境造成很大的影响，正符合我国当前对生产的要求。
2化学工程技术在化学生产中的应用
超临界流体技术在化学生产中的应用
超临界流体技术主要的内容是，控制一定的温度和压力，使得需要的流体处于液体与气体中间的状态。这种流体的特点集合了气液的优点，它的粘度低与气体相似，它的密度很高与液体相似，这就导致它的扩散能力很强，介于气体和液体之间。同时它还拥有很强的溶解能力和压缩能力。将这种技术应用于化学生产中，通过控制温度与压力，得到超临界流体，利用其拥有的优势来达到节省能耗的目的。现如今，我们将这种技术应用于更过多领域，比如，高分子材料、复合材料、有机物材料和无机物材料。
传热技术在化学生产中的应用
化学工程之中的传热技术主要是分为两方面，一方面是微细尺度传热技术，另一方面是强化传热过程。首先微细尺度传热，是以热对流、热传导、热辐射为主要的内容，从空间尺度和时间尺度微细进行讨论和研究的一项传热技术。这项技术在微米、纳米科学中得到了广泛的应用，并取得了不错的成绩，因此人们更加关注它在化学生产中的应用。强化传热过程，主要的重点是通过调试换热器设备，不断改进生产过程中的传热系数，使其能够有能力不断的对外放热。为了强化传热过程，就要增加冷热流体间的温差，这就必须通过改变换热的面积来提高传热系数，从而来提高传热的效率，使得在化学生产的过程节能减耗。
绿色化学反应技术在化学生产中的应用
通常化学生产的产品一般对我们生活有一些影响的，因此我们就需要采用绿色化学反应来防止化学生产的过程中对环境造成污染，这是从源头来解决污染问题的技术方法。绿色化学只得就是通过使用化学的技术与方法，结合相关的知识来解决化学对人们和环境造成的危害。主要要求就是，化学生产过程中用到的试剂、催化剂、反应原料，和反应完成后的产物与副产物都必须对人类和环境无危害，同时也要保证绿色环保。
例如，采用绿色无毒的原料方面，可以将石油原料装换成生物原料。像是在化学产品尼龙的生产过程中，原先采用的是含苯的石油化工原料，我们将可以其原料改换成生物原料，一样也可以制成尼龙，不仅保护了环境，而且也保护了人体收到伤害。除此之外，这项技术在绿色食品生产中也起到了很大的作用，绿色食物是对人体很有益的，在其生产过程中一般禁止使用化学药剂，这样不仅减少了对人体的伤害，同时也减少了对环境的影响。
然而生产绿色食品的代价就是成本高，为了可以降低成本又能够有质量，我们可以将化学技术与生物技术相结合，开发基因技术，提高并促进农作物的产量和质量，生物技术与化学反应技术相结合可以在以下过程中充分的利用。
3现今化学工程技术存在的问题
化学工程技术需要进一步的提高
现如今，我国的化学工程技术应用的领域非常更广泛，但是仍存在一些不足。滴状冷凝在工业上的应用仍然不能有很好的表现，因为在获得滴状冷凝后，冷凝的液滴不能够被长久的保存，所以，我们应该在这问题上有进一步的研究，从而来解决这个问题。使得我国的化学工程技术能够有更好的发展，人们能够有更好的生活条件。
化学工程技术的人才匮乏
在化学工程中存在的另一个严重的问题就是技术人才问题，只有用化学专业技术强的人才，才能够更好的提高化学生产的质量。而我国现在就存在这样的问题，化学领域的工作人员的普遍的技术能力和专业能力不强，主要是由于我国的教育体制问题，当代的大学生理论要点掌握很好，但实际操作方面却严重的匮乏，这就导致技术型人才的缺乏，从而影响了化学工程技术的进步。
4对化学工程技术的发展提出对策
不断提升化学工程技术
随着我国的科技不断的发展，化学工程技术也会越来越进步，我们应该不断的更新技术，以此来适应社会科技的发展。应该在巩固传统的化学技术的同时不断的添加新型技术，并抛弃不利的部分，从而实现化学工程技术有更好的发展。
培养化学技术人才
人才的重要性是我们有目共睹的，化学技术人才对于化学工程的发展有着至关重要的作用。因此为了化学工程技术能够有更好的发展，我们重点培养化学技术人才，化学生产企业可以通过与相关专业的院校进行合作，让专业对口的大学生能够有机会到生产工厂进行相关的实习操作，从而来培养理论知识牢固并且有一定的操作能力的技术人才来工作。
5结语
化学工程技术在化学生产过程中的应用广泛，它不仅促进了社会经济的发展，更是提高了人们的生活水平，通过技术和人才的不断涌进，我国的化学工程技术会有更好的发展。
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