空调英文论文文献

空调英文论文文献

Central Air Conditioners for Consumers(Are you a partner? For Partners)
Heating and cooling costs the average homeowner about $1,000 a year — nearly half the home's total energy bill. If your central air conditioning unit is more than 12 years old, replacing it with an ENERGY STAR qualified model could cut your cooling costs by 30 percent.

Earning the ENERGY STAR means products meet strict energy efficiency guidelines set by the US Environmental Protection Agency and the Department of Energy.

ENERGY STAR qualified central air conditioners have a higher seasonal efficiency rating (SEER) than standard models, which makes them about 14% more efficient than standard models.
Remember, saving energy prevents pollution. By choosing ENERGY STAR and taking steps to optimize the performance of your cooling equipment, you are helping to prevent global warming and promoting cleaner air while enhancing the comfort of your home.

You may also be interested to know:

Though these products can be more expensive to purchase up front, the cost difference will be paid back over time through lower energy bills.
When buying new equipment, it is important to get a quality installation. Make sure you get a contractor who can do the job right.
You can get better performance out of your cooling equipment by sealing your home and making sure your ducts don't leak.
EPA offers additional suggestions for improving the performance of your cooling system as well as general home improvement advice.
Finding Qualified Equipment
The Consortium for Energy Efficiency (CEE) and the Air-Conditioning and Refrigeration Institute (ARI) have developed an online database which can be used to find qualifying ENERGY STAR equipment. All equipment listed in this online database meets the specification requirements for ENERGY STAR. This online database is solely maintained and operated by CEE and ARI.

谁能给几篇暖通空调类的英文文献，要附带汉语全文翻译的

testing of an air-cycle refrigeration system for road transport
Abstract
The 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 units.
Keywords: Air conditioner; Refrigerated transport; Thermodynamic cycle; Air; Centrifuge compressor; Turbine expander COP,
Nomenclature
PR
Compressor or turbine pressure ratio
TA
Heat exchanger side A temperature (K)
TB
Heat exchanger side B temperature (K)
Tinlet
Inlet temperature (K)
Toutlet
Outlet temperature (K)
ηcomp
Compressor isentropic efficiency
ηturb
Turbine isentropic efficiency
ηheat exchanger
Heat exchanger effectiveness
1. Introduction
The 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 immediately.
Concerns 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 refrigeration.
Thermo 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 turbochargers.
2. Thermodynamic modelling and design of the demonstrator plant
The 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 COP.
The 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 plant.
The 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 intercooler.
shows 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 loss.
The 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 ratios.
Since 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 expansion.
Lower 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/s.
By 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 cycle.
3. Prime mover and primary compressor
The 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 efficiency.
A 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 clutch.
4. Cold air unit
The 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 performance.
Most 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 used.
The 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 exchangers
Due 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 plant.
Several 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. Instrumentation
A 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 speed.
7. System testing
During 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 standards.
The 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 testing.
8. Discussion of measured performance
From 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

200分急求汽车空调英文文献及翻译！

Automotive Air Conditioning Systems

by Chris Bede

Today, as we drive our automobiles, a great many of us, can enjoy the same comfort levels that we are accustomed to at home and at work. With the push of a button or the slide of a lever, we make the seamless transition from heating to cooling and back again without ever wondering how this change occurs. That is, unless something goes awry.
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Since the advent of the automotive air conditioning system in the 1940's, many things have undergone extensive change. Improvements, such as computerized automatic temperature control (which allow you to set the desired temperature and have the system adjust automatically) and improvements to overall durability, have added complexity to today's modern air conditioning system. Unfortunately, the days of "do-it-yourself" repair to these systems, is almost a thing of the past.

To add to the complications, we now have tough environmental regulations that govern the very simplest of tasks, such as recharging the system with refrigerant R12 commonly referred to as Freon. (Freon is the trade name for the refrigerant R-12, that was manufactured by DuPont). Extensive scientific studies have proven the damaging effects of this refrigerant to our ozone layer, and its manufacture has been banned by the U.S. and many other countries that have joined together to sign the Montreal Protocol, a landmark agreement that was introduced in the 1980's to limit the production and use of chemicals known to deplete the ozone layer.

Now more than ever, your auto mechanic is at the mercy of this new environmental legislation. Not only is he required to be certified to purchase refrigerant and repair your air conditioner, his shop must also incur the cost of purchasing expensive dedicated equipment that insures the capture of these ozone depleting chemicals, should the system be opened up for repair. Simply put, if your mechanic has to spend more to repair your vehicle - he will have to charge you more. Basic knowledge of your air conditioning system is important, as this will allow you to make a more informed decision on your repair options.

Should a major problem arise from your air conditioner, you may encounter new terminology. Words like "retrofit" and "alternative refrigerant" are now in your mechanics glossary. You may be given an option of "retrofitting", as opposed to merely repairing and recharging with Freon. Retrofitting involves making the necessary changes to your system, which will allow it to use the new industry accepted, "environmentally friendly" refrigerant, R-134a. This new refrigerant has a higher operating pressure, therefore, your system, dependant on age, may require larger or more robust parts to counter its inherent high pressure characteristics. This, in some cases, will add significantly to the final cost of the repair. And if not performed properly, may reduce cooling efficiency which equates to higher operating costs and reduced comfort.

Vehicles are found to have primarily three different types of air conditioning systems. While each of the three types differ, the concept and design are very similar to one another. The most common components which make up these automotive systems are the following:

COMPRESSOR, CONDENSER, EVAPORATOR, ORIFICE TUBE, THERMAL EXPANSION VALVE , RECEIVER-DRIER, ACCUMULATOR. Note: if your car has an Orifice tube, it will not have a Thermal Expansion Valve as these two devices serve the same purpose. Also, you will either have a Receiver-Dryer or an Accumulator, but not both.

For more information on Air Conditioning, check out The Automotive Air Conditioning Information Server

COMPRESSOR
Commonly referred to as the heart of the system, the compressor is a belt driven pump that is fastened to the engine. It is responsible for compressing and transferring refrigerant gas.

The A/C system is split into two sides, a high pressure side and a low pressure side; defined as discharge and suction. Since the compressor is basically a pump, it must have an intake side and a discharge side. The intake, or suction side, draws in refrigerant gas from the outlet of the evaporator. In some cases it does this via the accumulator.

Once the refrigerant is drawn into the suction side, it is compressed and sent to the condenser, where it can then transfer the heat that is absorbed from the inside of the vehicle.

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CONDENSER
This is the area in which heat dissipation occurs. The condenser, in many cases, will have much the same appearance as the radiator in you car as the two have very similar functions. The condenser is designed to radiate heat. Its location is usually in front of the radiator, but in some cases, due to aerodynamic improvements to the body of a vehicle, its location may differ. Condensers must have good air flow anytime the system is in operation. On rear wheel drive vehicles, this is usually accomplished by taking advantage of your existing engine's cooling fan. On front wheel drive vehicles, condenser air flow is supplemented with one or more electric cooling fan(s).

As hot compressed gasses are introduced into the top of the condenser, they are cooled off. As the gas cools, it condenses and exits the bottom of the condenser as a high pressure liquid.

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EVAPORATOR
Located inside the vehicle, the evaporator serves as the heat absorption component. The evaporator provides several functions. Its primary duty is to remove heat from the inside of your vehicle. A secondary benefit is dehumidification. As warmer air travels through the aluminum fins of the cooler evaporator coil, the moisture contained in the air condenses on its surface. Dust and pollen passing through stick to its wet surfaces and drain off to the outside. On humid days you may have seen this as water dripping from the bottom of your vehicle. Rest assured this is perfectly normal.

The ideal temperature of the evaporator is 320 Fahrenheit or 00 Celsius. Refrigerant enters the bottom of the evaporator as a low pressure liquid. The warm air passing through the evaporator fins causes the refrigerant to boil (refrigerants have very low boiling points). As the refrigerant begins to boil, it can absorb large amounts of heat. This heat is then carried off with the refrigerant to the outside of the vehicle. Several other components work in conjunction with the evaporator. As mentioned above, the ideal temperature for an evaporator coil is 320 F. Temperature and pressure regulating devices must be used to control its temperature. While there are many variations of devices used, their main functions are the same; keeping pressure in the evaporator low and keeping the evaporator from freezing; A frozen evaporator coil will not absorb as much heat.

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PRESSURE REGULATING DEVICES
Controlling the evaporator temperature can be accomplished by controlling refrigerant pressure and flow into the evaporator. Many variations of pressure regulators have been introduced since the 1940's. Listed below, are the most commonly found.

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ORIFICE TUBE
The orifice tube, probably the most commonly used, can be found in most GM and Ford models. It is located in the inlet tube of the evaporator, or in the liquid line, somewhere between the outlet of the condenser and the inlet of the evaporator. This point can be found in a properly functioning system by locating the area between the outlet of the condenser and the inlet of the evaporator that suddenly makes the change from hot to cold. You should then see small dimples placed in the line that keep the orifice tube from moving. Most of the orifice tubes in use today measure approximately three inches in length and consist of a small brass tube, surrounded by plastic, and covered with a filter screen at each end. It is not uncommon for these tubes to become clogged with small debris. While inexpensive, usually between three to five dollars, the labor to replace one involves recovering the refrigerant, opening the system up, replacing the orifice tube, evacuating and then recharging. With this in mind, it might make sense to install a larger pre filter in front of the orifice tube to minimize the risk of of this problem reoccurring. Some Ford models have a permanently affixed orifice tube in the liquid line. These can be cut out and replaced with a combination filter/orifice assembly.

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THERMAL EXPANSION VALVE
Another common refrigerant regulator is the thermal expansion valve, or TXV. Commonly used on import and aftermarket systems. This type of valve can sense both temperature and pressure, and is very efficient at regulating refrigerant flow to the evaporator. Several variations of this valve are commonly found. Another example of a thermal expansion valve is Chrysler's "H block" type. This type of valve is usually located at the firewall, between the evaporator inlet and outlet tubes and the liquid and suction lines. These types of valves, although efficient, have some disadvantages over orifice tube systems. Like orifice tubes these valves can become clogged with debris, but also have small moving parts that may stick and malfunction due to corrosion.

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RECEIVER-DRIER
The receiver-drier is used on the high side of systems that use a thermal expansion valve. This type of metering valve requires liquid refrigerant. To ensure that the valve gets liquid refrigerant, a receiver is used. The primary function of the receiver-drier is to separate gas and liquid. The secondary purpose is to remove moisture and filter out dirt. The receiver-drier usually has a sight glass in the top. This sight glass is often used to charge the system. Under normal operating conditions, vapor bubbles should not be visible in the sight glass. The use of the sight glass to charge the system is not recommended in R-134a systems as cloudiness and oil that has separated from the refrigerant can be mistaken for bubbles. This type of mistake can lead to a dangerous overcharged condition. There are variations of receiver-driers and several different desiccant materials are in use. Some of the moisture removing desiccants found within are not compatible with R-134a. The desiccant type is usually identified on a sticker that is affixed to the receiver-drier. Newer receiver-driers use desiccant type XH-7 and are compatible with both R-12 and R-134a refrigerants.

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ACCUMULATOR
Accumulators are used on systems that accommodate an orifice tube to meter refrigerants into the evaporator. It is connected directly to the evaporator outlet and stores excess liquid refrigerant. Introduction of liquid refrigerant into a compressor can do serious damage. Compressors are designed to compress gas not liquid. The chief role of the accumulator is to isolate the compressor from any damaging liquid refrigerant. Accumulators, like receiver-driers, also remove debris and moisture from a system. It is a good idea to replace the accumulator each time the system is opened up for major repair and anytime moisture and/or debris is of concern. Moisture is enemy number one for your A/C system. Moisture in a system mixes with refrigerant and forms a corrosive acid. When in doubt, it may be to your advantage to change the Accumulator or receiver in your system. While this may be a temporary discomfort for your wallet, it is of long term benefit to your air conditioning system.

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