HVAC总成设计指南
美国标准（American Standard）HVAC系统设计指南：2006年版说明书
•providing insights for today’s hvac system designer© 2006 American StandardAll rights reserved●1volume 35–4energy-saving control strategies forRooftop VAV SystemsRooftop variable-air-volume (VAV)systems are used to provide comfort in a wide range of building types and climates. This system consists of a packaged rooftop air conditioner that serves several individually-controlled zones. Each zone has a VAV terminal unit that is controlled by a temperature sensor in the zone.This EN discusses HVAC systemcontrol strategies that can be used to save energy in rooftop VAV systems.Optimal Start/Stop. In somebuildings, a simple time clock or time-of-day schedule is used to start and stop the HVAC system. During hours when the building is expected to be unoccupied, the system is shut off and the temperature is allowed to drift away from the occupied setpoint. The time at which the system starts again in the morning is typically set to ensure that the indoor temperature reaches the desired occupied setpoint prior to occupancy on either the coldest or warmest morning of the year. As a result, for most days, the system starts much earlier than needed. In turn, this increases the number of operating hours and system energy use.An alternative approach is a strategy called optimal start . This strategy utilizes a building automation system (BAS) to determine the length of time required to bring each zone from current temperature to the occupied setpoint temperature. The system waits as long as possible before starting, so that the temperature in each zone reaches occupied setpoint just in time for occupancy (Figure 1). This optimal starting time is determined using the differencebetween actual zone temperature and occupied setpoint. It compares this difference with the historicalperformance of how quickly the zone has been able to warm up or cool down.The optimal start strategy reduces the number of system operating hours and saves energy by avoiding the need to maintain the indoor temperature at occupied setpoint even though the building is unoccupied.A related strategy is optimal stop . As mentioned earlier, at the end of the occupied period, the system is shut off and the temperature is allowed to drift away from occupied setpoint.However, the building occupants may not mind if the indoor temperature drifts just a few degrees before they leave for the day.Optimal stop uses the BAS to determine how early heating and cooling can be shut off for each zone, so that the indoor temperature drifts only a few degrees from occupiedsetpoint (Figure 1). In this case, only cooling and heating are shut off; the supply fan continues to operate and the outdoor-air damper remains open to continue ventilating the building.The optimal stop strategy also reduces the number of system operating hours, saving energy by allowing indoor temperatures to drift early.Fan-Pressure Optimization. As cooling loads change, the VAV terminals modulate to vary airflow supplied to the zones. This causes the pressure inside the supply ductwork to change. In many systems, a pressure sensor is located approximately two-thirds of the distance down the main supply duct. The rooftop unit varies the capacity of the supply fan to maintain the static pressure in this location at a constant setpoint. With this approach, however, the system usually generates more static pressure at part load than necessary.When communicating controllers are used on the VAV terminals, it is possible to optimize this static-pressure control function to minimize duct pressure, and save fan energy. Each VAV unit controller knows theoccupied setpoint temperature mid 6 AM noon 6 PM midFigure 1. Optimal start and optimal stop2●Trane Engineers Newsletter volume 35–4providing insights for today’s HVAC system designercurrent position of its air-modulation damper. The BAS continually pollsthese individual controllers, looking for the VAV terminal with the most-open damper (Figure 2). The setpoint for the supply fan is then reset to provide just enough pressure so that at least one damper is nearly wide open. This results in the supply fan generating only enough static pressure to push the required quantity of air through this "critical" VAV terminal unit.This control strategy, sometimes called fan-pressure optimization, has several benefits:•Reduced supply fan energy use . At part-load conditions, the supply fan is able to operate at a lower static pressure and consume less energy (Figure 3).•Lower sound levels . The supply fan does not generate as much static pressure and will typically generate less noise. In addition, with lower pressures in the supply duct, the dampers in the VAV terminals will be more open, resulting in less regenerated noise.•Reduced risk of fan surge . Byallowing the fan to operate at a lower pressure when delivering reduced airflow, the fan operating point is kept further away from the surge region (Figure 3).•Flexibility of sensor location. Since this strategy uses the position of VAV dampers to reset the pressure setpoint, the static-pressure sensor can be located anywhere in the supply duct. It can even be located at the discharge of the fan, allowingit to be installed inside the rooftop unit and tested at the factory. In this location, it can also serve as the duct high-pressure sensor, protecting the ductwork from damage in the event of a fire damper closing.Supply-Air-T emperature Reset . In a VAV system, it is tempting to raise the supply-air (SA) temperature at part-load conditions to save compressor and/or reheat energy. Increasing the supply-air temperature reduces compressor energy because it allows thecompressor to operate at a warmer suction temperature. Thecorresponding higher suction pressure reduces the compressor lift, reducing the power required.In addition, supply-air-temperature reset makes an airside economizer more beneficial. When the outdoor air is cooler than the SA temperature setpoint, the compressors are shut off, and the outdoor- and return-air dampers modulate to deliver the desired supply-air temperature. A warmer SA temperature setpointallows the compressors to be shut off sooner and increases the number of hours when the economizer is able to provide all the necessary cooling.For zones with very low cooling loads, when the supply airflow has been reduced to the minimum setting of the VAV terminal, raising the supply-air temperature also decreases the use of reheat at the zone level.However, because the supply air is warmer, zones that require cooling will need more air to satisfy the cooling load. This increases supply fan energy. Finally, in non-arid climates, warmer supply air means less dehumidification at the coil and higher humidity levels in the zones. If dehumidification is a concern, use caution when implementing this strategy.Supply-air-temperature reset should be implemented so that it minimizes overall system energy use. This requires considering the trade-off between compressor, reheat, and fanairflows t a t i c p r e s s u r eFigure 3. Benefits of fan-pressure optimizationFigure 2. Fan-pressure optimizationproviding insights for today’s HVAC system designerTrane Engineers Newsletter volume 35–4●3energy, as well as the impact on space humidity levels. Table 1 contains some general guidance to determine when this strategy might provide the most benefit.These competing issues are often best balanced by first reducing supply airflow, taking advantage of the significant energy savings fromunloading the fan. Once fan airflow has been reduced, raise the supply-airtemperature to minimize reheat energy and enhance the benefit of the airside economizer. While one could dream up numerous control schemes, the simplest approach is probably most common. Figure 4 shows an example of a supply-air-temperature reset strategy based on the changing outdoor dry-bulb temperature. When the outdoor temperature is warmer than 70°F , no reset takes place and the SA temperature setpoint remains at the design value of 55°F . When it is this warm outside, theoutdoor air provides little or no cooling benefit for economizing. The cooling load in most zones is likely highenough that reheat is not required to prevent overcooling. In addition, the colder (and drier) supply air allows the system to provide sufficiently dry air to the zones, improving part-load dehumidification.When the outdoor temperature is between 60°F and 70°F , the SAtemperature setpoint is reset at a 2-to-1 ratio. That is, for every 2°F change in outdoor temperature, the setpoint is reset 1°F . In this range, supply-air-temperature reset enhances the benefit provided by the economizerand it is likely that some zone-level reheat can be avoided.Finally, when the outdoor temperature is colder than 60°F , no further reset occurs, and the SA temperature setpoint remains at 60°F . Limiting the amount of reset to 60°F allows the system to satisfy the cooling loads in interior zones without needing to substantially oversize VAV terminals and ductwork.Alternatively, some systems reset the SA temperature setpoint based on the temperature in the "critical" zone. This is the zone that is most nearly at risk of overcooling, which would require activating local reheat. A buildingautomation system (BAS) monitors the temperature in all zones, finding the critical zone that is closest to heating setpoint. The rooftop unit then resets the SA temperature setpoint to prevent this critical zone from needing to activate reheat.When considering using supply-air-temperature reset in a rooftop VAV system:•First analyze the system to determine if the savings in compressor and reheat energy will outweigh the increase in fan energy.•If higher space humidity levels are a concern, consider either disabling reset when it is humid outside, or providing one or more humiditysensors to override the reset function whenever humidity in the zone exceeds some maximum limit.•For interior zones with near-constant cooling loads during occupiedperiods, calculate design airflows for those zones based on the warmer, reset supply-air temperature (60°F in the example from Figure 4). While this may require larger VAV terminals and ductwork, it allows the use ofsupply-air-temperature reset duringSource: California Energy Commission, Advanced Variable Air Volume System Design Guide, 2003.T able 1. Supply-air-temperature resetFigure 4. Supply-air-temperature reset based on OA temperature5560655080757061605958575655S A t e m p e r a t u r e s e t p o i n t , °Foutdoor dry-bulb temperature, °Fcooler weather, while still providing the necessary cooling to thoseweather-independent, interiorzones.•Design the air distribution system for low pressure losses and use the fan-pressure optimization strategyto minimize the fan energy penalty that accompanies a warmer SAtemperature.Ventilation Optimization . In a typical VAV system, the rooftop unit delivers fresh outdoor air to several, individually-controlled zones. Demand-controlled ventilation (DCV) involves resetting intake airflow in response to variations in zone population. Whilecommonly implemented using carbon dioxide (CO2) sensors, occupancy sensors, or time-of-day (TOD) schedules can also be used.Ventilation reset involves resetting intake airflow based on variations in system ventilation efficiency.One approach to optimizing ventilation in a multiple-zone VAV system is to combine the various DCV strategies at the zone level (using each where it best fits) with ventilation reset at the system level.With this strategy, CO2 sensors are installed only in those zones that are densely occupied and experience widely varying patterns of occupancy. For the example building in Figure 5, CO2 sensors are installed only in the conference room and the lounge. These zones are the best candidates for CO2 sensors, and provide "the biggest bang for the buck." These sensors reset the ventilation requirement for their respective zones based on measured CO2.However, zones that are less densely occupied or have a population that varies only a little (such as private offices, open plan office spaces, or many classrooms) are probably better suited for occupancy sensors. In Figure 5, each of the private offices has an occupancy sensor to indicate when the occupant is present. Whenunoccupied, the controller lowers theventilation requirement for the zone.Occupancy sensors are relativelyinexpensive, do not need to becalibrated, and are already used inmany zones to control the lights.Finally, zones that are sparselyoccupied or have predictableoccupancy patterns may be bestcontrolled using a time-of-dayschedule. This schedule can eitherindicate when the zone will normallybe occupied vs. unoccupied, or can beused to vary the zone ventilationrequirement based on anticipatedpopulation.These various zone-level DCVstrategies can be used to reset theventilation requirement for theirrespective zones for any given hour.This zone-level control is then tiedtogether using ventilation reset at thesystem level (Figure 6).In addition to resetting the zoneventilation requirement, the controlleron each VAV terminal continuouslymonitors primary airflow beingdelivered to the zone. The BASFigure 5. Demand-controlled ventilation at the zone levelUnoccupied Humidity ControlA VAV system typically dehumidifieseffectively over a wide range of operatingconditions because it continues to delivercold, dry air at part-load conditions. Aslong as supply-air-temperature reset isused with caution, and reheat is availablefor those VAV terminals that have highminimum airflow settings or experiencevery low cooling loads, a VAV system willtypically provide supply air at a dew pointthat's low enough to prevent elevatedindoor humidity levels during occupiedperiods.However, controlling humidity levels isn'tonly a priority when the building isoccupied. When indoor humidity rises toohigh during unoccupied times, one optionis to turn on the rooftop unit anddehumidify recirculated air to 55°F or so.However, there is typically very littlesensible load in the zones during theseperiods, so delivering this cold air willresult in overcooling. Reheat coils in theVAV terminals, and possibly a boiler andhot water pumps, will need to beactivated.An energy-saving alternative is to equipthe rooftop unit with hot gas reheat. Whenafter-hours dehumidification is needed,the rooftop unit turns on and diverts hotrefrigerant vapor leaving the compressorthrough a refrigerant-to-air heatexchanger that is located in the airstream,following the evaporator coil. Sensibleheat is transferred from the hot refrigerantto reheat the dehumidified air leaving theevaporator.This strategy uses heat recovered fromthe refrigeration circuit to reheat centrally,and saves energy by avoiding the use ofnew energy to reheat remotely at the VAVterminals.4●Trane Engineers Newsletter volume 35–4providing insights for today’s HVAC system designerproviding insights for today’s HVAC system designerTrane Engineers Newsletter volume 35–4●5periodically gathers this data from all VAV terminals and solves theventilation reset equations (prescribed by ASHRAE Standard 62) to determine how much outdoor air must be brought in at the rooftop unit to satisfy all zones served. Finally, the BAS sends this outdoor airflow setpoint to the rooftop unit which modulates a flow-measuring outdoor-air damper to maintain this new setpoint.In a DDC/VAV system, this strategy is fairly easy to implement because the necessary real-time information is already available digitally. Combining DCV at the zone level with ventilation reset at the system level has the following benefits:•Assures that each zone is properly ventilated without requiring a CO 2 sensor in every zone . CO 2 sensors are used only in those zones in which they will bring the most benefit. This minimizes installed cost and avoids the periodic calibration and cleaning required to assure proper sensor operation. For the other zones,occupancy sensors and time-of-day schedules are used to reduce ventilation.•Enables documentation of actual ventilation system performance . The VAV controllers communicate the ventilation airflow for every zone to the BAS, even for those zones that do not have a CO 2 sensor. The BAS can be used to generate reports showingventilation airflow (cfm) in every zone for every hour.•Uses system-level ventilation reset equations that are explicitly defined in an industry-wide standard . Using equations from ASHRAE 62 improves the "defend-ability" of the control strategy.Summary. The impact of any energy-saving strategy on the operating cost of a specific building depends on climate, building usage, and utility costs. Building analysis tools (like TRACE™ 700) can be used to analyze these strategies and convert energy savings to operating cost dollars that can be used to make financial decisions.Figure 7 shows the potential energy savings of using these variousstrategies in an office building that has a typical rooftop VAV system. The optimized system uses the optimal start, supply-air-temperature reset, and ventilation optimization strategies discussed in this EN. In addition, the supply fan is controlled based on fan-pressure optimization, rather than on a constant setpoint in the ductwork.The optimized rooftop VAV system reduced the HVAC energy use by about 30% for the building in both Atlanta and Los Angeles, and by 33% in Minneapolis.There is a real potential to save energy in rooftop VAV systems throughoptimized system control strategies. This savings reduces operating costs for the building owner and can help in achieving points toward LEED ® certification.Article by John Murphy, applicationsengineer,Trane. Y ou can find this and previous issues of the Engineers Newsletter at /engineersnewsletter. To comment,***********************.Figure 6. Ventilation reset at the system levelFigure 7. Energy-saving potential of optimized controlMinneapolisLos AngelesAtlanta20406080100H V A Ce n e r g y c o n s u m p t i o n , % of b a s eoptimized system controlsbase systemT raneA business of American Standard CompaniesFor more information, contact your local Trane***********************************Trane believes the facts and suggestions presented here to be accurate. However, final design andapplication decisions are your responsibility. Trane disclaims any responsibility for actions taken onthe material presented.6●Trane Engineers Newsletter volume 35–4ADM-APN022-EN (October 2006)。
海上平台暖通空调系统(HVAC)设计手册
海上平台暖通空调系统（HVAC）设计手册（99版）中海石油生产研究中心机电部前言由于我国目前还没有出版一本关于海洋石油平台上采暖、通风和空调的设计手册或标准规范。
因此,我们在总结以往设计经验、参考国外和国内有关资料的基础上,编制了这本设计手册,以供我们在设计中参考。
由于我们的经验有限，文中难免有不完整或不妥之处，希望有关专家和使用者提供宝贵意见，以便我们进一步修改和完善。
中国海洋石油生产研究中心机电部编制王雅君校对赵虹审核王建丰一九九九年八月目录1 概述1.1 定义1.2 范围2 HVAC设计采用的标准和规范3 HVAC设计的条件3.1 室内外环境条件的确定3.2 其它有关资料的准备4空调负荷计算4.1夏季空调得热量计算4.2冬季围护结构热损失计算4.3空调送风量计算4.4空调新风量计算4.5排风量计算4.6空调热负荷计算4.7空调装置制冷量确定5 空调系统设计5.1 空调方式选择5.2 空调区域范围5.3 新风和回风系统设计5.4 排风系统设计5.5 空调设备与材料6 空调系统的控制和保护6.1 温湿度控制6.2 室内外压差控制6.3 安全保护措施7 平台的安全通风设计7.1 平台上通风系统的作用7.2 平台上需要通风的区域7.3 通风方式选择7.4 通风量计算7.5 风管截面选择7.6 气流组织7.7 风机的选择7.8 安全通风的保护措施7.9 风管设计注意事项7.10 控制与动力供应8 平台上典型房间的通风举例8.1 燃气轮机罩和燃气轮机间的通风8.2 柴油发电机房的通风8.3 蓄电池室的通风8.4 空调机房的通风8.5 消防泵房和泡沫站的通风8.6变压器间的通风8.7 配电室（开关间）的通风8.8 锅炉舱的通风8.9 厨房的通风9 小型冷库设计9.1 小型冷库的组成和主要参数9.2 冷库库容的确定9.3 冷库的结构9.4 冷库负荷计算9.5 制冷机组的选择和控制10 HVAC规格书编制10.1 HVAC规格书的范围10.2 HVAC规格书的内容简介11本手册编制所参考的资料12附图附图1 直接蒸发式空调系统（1）附图2 直接蒸发式空调系统（2）附图3 间接冷却式中央空气处理空调系统附图4 间接冷却式末端空气处理空调系统（1）附图5 间接冷却式末端空气处理空调系统（2）附图6 典型正压房间HVAC系统控制图附图7 危险区和非危险区的通风和门的布置图1.概述1.1定义HVAC—即Heating, Ventilation and Air-conditioning 的缩写，意为采暖、通风和空气调节。
美国标准 HVAC 制冷制热系统设计指南说明书
•providing insights for today’s hvac system designer© 2006 American Standard All rights reserved●1volume 35–3maintaining a comfortable environment inPlaces of AssemblyDesigning comfort systems for places of assembly (auditoriums,gymnasiums, arenas, houses of worship) presents some vexingchallenges. Such facilities often have acoustical requirements that place limits on equipment location and air distribution design. Many places of assembly experience extremely diverse loads and occupancyschedules, complicating part-load system control. But perhaps thebiggest challenge is occupancy itself, and its impact on ventilation and humidity control. Design guidelines that are commonly applied incommercial office space mayget us into trouble here.A simple example can illustrate some of these issues: a school gymnasium during a band concert. As this is a good high school band, both the bleachers and the floor are full. Occupancy is at the fire marshal’s rated seatingcapacity. The 18,000 ft 2 gymnasium is designed for 1200 people, including use of the gym floor. A load calculation reveals the following space loads:Roof 69,600 Btu/hr Wall 43,000 Btu/hr Glass 10,500 Btu/hr Lights 122,900 Btu/hrPeople 300,000 Btu/hr (sensible)240,000 Btu/hr (latent)Totals546,000 Btu/hr (sensible)240,000 Btu/hr (latent)Occupancy a Major FactorPeople constitute a significant portion of the space sensible cooling load, over 50 percent. However, it is the impact on humidity that makes occupancy a difficult load to manage. The space sensible heat ratio for this example is only 0.69 (Figure 1).If the target space comfort condition is 75°F and 50 percent relative humidity (RH), and the air distribution system is designed for 55°F supply air, the required supply airflow is over 4 cfmper square foot of floor area -- a huge amount! How did that happen?Humidity ratio tells the story. Humidity ratio is grains of water vapor per pound of air. The humidity ratio at 75°F dry bulb and 50 percent RH, the desired space condition, is 64.7 grains of water vapor per pound of air. The 55°F supply air has a humidity ratio of 60.4 gr/lb. At these conditions, each pound of supply air we introduce into the space can remove 4.3 grains of water vapor. If each occupant contributes 200 Btu/hr to the latent load, 1200 people add 230 pounds of water vapor (1,610,000 grains) to the air in the gymnasium. If each pound of supply air can remove only 4.3 grains of water vapor, it will take 374,000 pounds of supply air per hour. This equates to approximately 83,000 cfm, or 4.6 cfm/ft 2. That's a lot of air!Figure 1. Design sensible and latent loadsHumidity is the Driver.In this example, 83,000 cfm is required to handle humidity, but only 18,000 cfm of this must be outdoor air for ventilation (assuming 15 cfm of OA per person*). With a space sensible cooling load of 546,000 Btu/hr and a supply-air temperature of 55°F; approximately 25,000 cfm is required to maintain the space temperature at 75°F. In this case, 72% of the supply air must be outdoor air. While this is a high fraction of outdoor air, it is manageable. Ventilation air is not the culprit.This 25,000 cfm of supply air equates to 1.4 cfm/ft2. This is a large, but manageable supply air quantity. But we still need 83,000 cfm of supply air to control humidity. How do we better equip the supply air to handle the high latent load associated with this many people? Obviously the supply air needs to be drier. The drier the supply air (the lower the dew point), the more water vapor it will remove from the space. What supply air dew point is required to handle the space latent load?*While many local codes may still require 15 cfm/ person for ventilation, the most recent version of ASHRAE Standard 62.1-2004 has revised the minimum required ventilation rates for places of assembly.Calculating Specific Humidity. The key is another humidity measurement called specific humidity. Specific humidity is expressed as pounds of water vapor per pound of air. Suppose we choose to design the air distribution system for our example gymnasium for 25,000 cfm (114,000 pounds per hour). The 1200 people generate 227 pounds of water vapor each hour. Removing 227 pounds of water vapor with 114,000 pounds of air requires that the specific humidity of the supply air be 0.0020 lbw/lba drier than the space. The specific humidity at 75°F and 50 percent RH is 0.0092 lbw/lba. So the specific humidity of the supply air must be 0.0072 lbw/lba to offset the latent load of the people. This corresponds to a supply air dew point of about 48°F.So how do we create this 48°F dewpoint supply air? One common methodis to cool all the supply air to a dry-bulbtemperature of about 49°F to 50°F.Thisshould dehumidify the supply air to the48°F dew point required to offset thelatent load due to people.Supplying 50°F air to the gymnasiumprovides additional benefits. It reducesthe required airflow needed to offsetthe space sensible cooling load from25,000 cfm to only 20,000 cfm (1.1cfm/ft2). This concept is called cold airdistribution, and is a common designapproach when aggressive humiditycontrol is required or the design teamis seeking ways to reduce fan power orair handler footprint.1 All of thesebenefits may be attractive whendesigning for places of assembly. Coldair also requires careful diffuserselection, careful temperature control,and reliable control of buildingpressure. In addition, supply-airtemperatures below 50°F maypreclude the use of conventional,direct expansion (DX) equipment.The Desiccant Approach. But do weneed colder air, or do we need drier air?The truth is, we don't need air that iscolder; we only need air that is drier.Recent research in desiccants hasresulted in a Type III desiccant wheelthat is able to regenerate at lowtemperatures, often without the needto add heat. This allows the wheel tobe configured in series with a coolingcoil. This activated alumina desiccantwheel is available in a Trane systemcalled CDQ™ (Cool, Dry, Quiet).2 Theaddition of the CDQ wheel allows thesystem to deliver supply air at 48°Fdew point, while the cooling coil onlyneeds to cool the air to 54°F. With CDQthere is no need to design a cold airdistribution system. Since there is noneed to produce 50°F supply air, therequired capacity of the cooling load issubstantially reduced.With the cold air system, the requiredcooling coil capacity is about 150 tons(based on 1200 people and 18,000 cfmof outdoor air) and supply fan power isonly 10 kW. The CDQ system reducescooling coil capacity to about 140 tons,but increases fan power to 16 kWbecause of the higher airflow andadditional static pressure from thedesiccant wheel. Both are viableoptions. It is noteworthy that CDQ maybe an excellent means to achieve lowsupply air dew points withconventional DX equipment.Don't Forget Part Load Situations.Places of assembly often experiencevery diverse loads. It would be wise toevaluate the performance of these Figure 2. Air handling unit with a T ype III series desiccant wheel (T rane CDQ)2●Trane Engineers Newsletter volume 35–3providing insights for today’s HVAC system designersystems at part load. There are two part-load conditions we should evaluate. One is quite obvious, which is what happens when most of the people leave. Perhaps the remaining occupancy is only 40 people instead of 1200. This is an easy part-load condition to accommodate. The sensible loads drop to 256,000 Btu/hr and the latent load due to people drops to only 8000 Btu/hr. The resulting space sensible heat ratio increases to 0.97.If we supply air at 50°F with the cold air system, the required supply airflow is only 9400 cfm. This system is called "single zone VAV." Supply airflow is reduced to match the reduced sensible cooling load in the space. Single zone VAV is easy to control. The supply fan airflow is modulated based on space temperature. The 9400 cfm of 50°F air will remove the 256,000 Btu/hr of sensible heat and has the potential to remove 116 pounds of water vapor. However, at this reduced occupancy, the people add only 7.6 pounds of water vapor. The result is that space humidity is lowered to 40 percent RH. At this condition, the supply-air temperature could be reset upward to save some compressor energy. Problem with Constant Volume Systems. What happens if the cold air system is a constant volume design rather than VAV? The reduced sensible cooling load requires a warmer supply-air temperature, about 63°F for this example. At this supply-air temperature, the 20,000 cfm of supply air will remove 256,000 Btu/hr of sensible heat, but less than 7.6 pounds of water vapor. Space humidity rises to 65 percent RH, well above our target of 50 percent. Not only does a constant volume system use more fan energy at part load, but it is less adept at removing moisture. By comparison, a single zone VAV system reduces fan energy while adequately removing moisture. Single zone VAV with cold air provides humidity control at most load conditions, while simultaneouslysaving fan energy.How does CDQ fare with reducedoccupancy? If the supply fan delivers aconstant volume of air, the reducedsensible load requires the supply-airtemperature to increase to over 65°F.However, the CDQ desiccant wheelcan still deliver the supply air at 55°Fdew point (Figure 2). The resultingspace humidity rises to only 52 percentRH. Constant volume CDQ is certainlyadept at controlling space RH at loweroccupancy, but the benefits of VAV canbe applied to CDQ systems too.When Sensible Loads are Lighter.There is another part-load conditionthat can be even more sinister;reduced building-related sensible loadswhile the space is fully occupied. Whathappens with full occupancy (1200people) when there is envelope orglass loads? If the only loads in thespace are due to lighting and people,the sensible heat ratio drops to 0.63. Ifwe dim the lights, the situation getseven worse.This reduction in the space sensiblecooling load creates a sensible heatratio more severe than what thesystem was originally designed toaccommodate. Increasing the supply-air temperature or reducing supplyairflow in response to the reducedsensible load will hinder the ability toremove moisture. Reheat can helpwhen the sensible heat ratio is lowerthan design. Both cold air and CDQhave the ability to reduce sensiblecooling capacity while maintaining alower supply air dew point. Table 1compares these systems, with andwithout reheat, when all envelopeconduction and solar loads are absent.Both cold air and CDQ systemsperform well at this part-load condition.Single zone VAV results in a slightlyelevated space relative humidity, butstill well within the comfort zone. Thiscomfortable condition is achievedwithout reheat and uses less fanenergy. Some reheat and additional fanenergy may be needed if more precisehumidity control is desired.Each Situation Unique. Well, it was agreat concert, but this was a highschool concert band, not a rock band.Add smoke from a pyrotechnic display,or moisture from an Olympic sizeswimming pool, and designing acomfort system for "places ofassembly" can be even morechallenging. In addition to reheat, coldair distribution and the CDQ desiccantwheel, give us additional tools to dealwith high space latent loads. Simpleairside control schemes like singlezone VAV provide an easy means ofadapting to diverse part-load conditionswhile providing some energy savings.Article by Don Eppelheimer, applicationsengineer,Trane. Y ou can find this and previousissues of the Engineers Newsletter at/engineersnewsletter. Tocomment,***********************.1.A 2000 Engineers Newsletter (volume 29-2,"Cold Air makes Good $ense") provides moredetail on the benefits and design issues relatedcold air distribution systems.2A 2005 Engineers Newsletter (volume 34-4,"Advances in Desiccant-Based Dehumidification")provides more detail on the series configuration ofa Type III desiccant wheel (Trane CDQ).T able 1. System comparison at part load (no envelope conduction or solar loads)providing insights for today’s HVAC system designer Trane Engineers Newsletter volume 35–3●34●Trane Engineers Newsletter volume 35–3providing insights for today’s HVAC system designerT raneA business of American Standard CompaniesFor more information, contact your local Trane***********************************Tr a ne bel i eves the f ac ts a nd suggest i ons presented here to be acc ur a te. However, f i n a l des i gn a nda ppl ica t i on de ci s i ons a re your respons ib i l i ty. Tr a ne d i sc l ai ms a ny respons i b i l i ty for ac t i ons t a ken onthe m a ter ia l presented.5●Trane Engineers Newsletter volume 35–3ADM-APN021-EN (September 2006)6●Trane Engineers Newsletter volume 35–3providing insights for today’s HVAC system designer。
HVAC_Operation_Manual(使用手册)
工程机械用空调系统使用说明书KB Autotech Co., Ltd.Ⅰ. 产品的主要作用1. HVAC UNIT ASS'Y1) 蒸发器蒸发器使液态制冷剂经膨胀阀节流膨胀后吸收车厢内的空气热量，并利用蒸发风机将冷气打入车厢内的装置。
2) 温度传感器温度传感器是感知蒸发器室的出回风温度，控制压缩机离合器的吸合，使室内温度保持在设定的温度范围，并防止蒸发器结冰。
注）传感器应附着规定的位置。
（空调厂家设计的位置）3) 膨胀阀①膨胀阀是将高温高压制冷剂的液体节流降压，成为易蒸发的低温低压雾状制冷剂进入蒸发器（即：分开了制冷剂的高压侧与低压侧）②自动调节制冷剂流量③控制制冷剂流量，防止液击和异常过热发生4) 暖气暖气是将发热的发动机冷却水注入加热器芯体通过室内的冷空气使室内变热的热交换器5) 内外气滤网吸入内外气时起着阻挡灰尘等异物的作用。
注）应定期性清扫或更换滤网2. 压缩机压缩机是经过发动机皮带驱动获得动力，从蒸发器吸入被汽化的低温低压冷媒，经压缩转换成高温高压的气态冷媒并发送到冷凝器的作用。
3. 冷凝器冷凝器是利用发动机水箱风机散热将高温高压的制冷剂气体，转换成高温低压的制冷剂液体的热交换器。
4. 干燥瓶干燥瓶是将冷凝器中液化的冷媒通过干燥瓶吸收系统中的湿气，过滤冷媒中的杂质并贮存制冷剂的作用。
, Max Warm)时压缩机开始运作（因空调开关的启动压缩机有可能停止运作）Ⅲ. 空调装置及整备1. 安全注意事项1)R-134a 冷媒是具有强挥发性的化学物质，接触皮肤时会导致冻伤，因此操作冷媒时应带手套。
2)如果冷媒进入眼睛时应立即用清水清洗。
为了保护眼睛必须要带防护眼镜和手套（切记不要用手或手绢搓揉眼睛）3) R-134a 冷媒为高压物品，因此在操作时附近不能有明火、易燃品以及不得将冷媒容器放置在燥热的地方。
确认储藏地方是否为52℃以下（注：冷媒容器放置在燥热的环境下可能会导致破裂或者爆炸）。
汽车后排空调的HVAC总成的生产技术
图片简介:本技术公开一种汽车后排空调的HVAC总成，包括叠加设置并相互连接的分发器总成(1)和鼓风机总成(2)，分发器总成(1)的内部设有蒸发器总成(3)、加热器总成(4)和加热器温度传感器(5)，分发器总成(1)的内部还设有混合风门总成(6)和模式风门总成(7)，且混合风门总成(6)和模式风门总成(7)通过蒸发器总成(3)、加热器总成(4)实现汽车空调的冷暖控制。
具有上述结构的该种汽车后排空调的HVAC总成通过将蒸发器总成与加热器总成、加热器温度传感器设置在分发器总成内部，实现布局紧凑，减小外部尺寸的目的；同时，通过优化结构布局，增加后排出风，满足乘客的舒适性，解决整车中空调出风均匀的问题。
技术要求1.一种汽车后排空调的HVAC总成，其特征在于：所述的汽车后排空调的HVAC总成包括叠加设置并相互连接的分发器总成(1)和鼓风机总成(2)，所述的分发器总成(1)的内部设有蒸发器总成(3)、加热器总成(4)和加热器温度传感器(5)，所述的分发器总成(1)的内部还设有混合风门总成(6)和模式风门总成(7)，且混合风门总成(6)和模式风门总成(7)通过蒸发器总成(3)、加热器总成(4)实现汽车空调的冷暖控制。
2.根据权利要求1所述的一种汽车后排空调的HVAC总成，其特征在于：所述的汽车后排空调的HVAC总成还包括安装在分发器总成(1)上的后吹面风道，空调风经过后吹面风道的出风口引到后排整车出风口。
技术说明书一种汽车后排空调的HVAC总成技术领域本技术涉及汽车空调领域，尤其是涉及一种汽车后排空调的HVAC总成。
背景技术现有的汽车空调系统HVAC总成中，空调HVAC的出风结构布局大部分都是对前排进行出风，而对于七座汽车的后排出风则无法满足乘客要求。
而且，由于常规空调布局的局限性,气动方向上风阻较大；同时，目前汽车空调使用的的单驱或者双驱结构与驱动机构通用性较低，严重影响七座汽车的舒适性，造成用户的满意度降低。
HVAC总成
在同档次的汽车中进行性能评价后，该空调系统的性能必须被要求严格的客户接受。 3.7 鼓风机系统
手动空调系统内的鼓风机必须有三档或三档以上速度可供选择（不包括关闭状态），智能空调（通 常使用手动）的鼓风机有五种以上速度或者根据客户要求可供选择，以实现逐渐改变风量的目的。
TBD
热敏电阻
TBD
TBD
TBD
模式风门驱动器
TBD
TBD
TBD
混风风门驱动器
TBD
TBD
TBD
表 2 鼓风机速度
风机速度档位
风量百分比% 手动鼓风机
风量百分比% 智能鼓风机
0
TBD
TBD
1
TBD
TBD
2
TBD
TBD
3
TBD
TBD
4
TBD
TBD
5
TBD
在风机所有的速度下，给定风口处的风量占总风量的百分比必须满足表 3 的规定。风量分布率必须
水积聚在壳体表面。 c) 出风口处无水珠或微小的冰晶。
3.28 HVAC 系统总体可靠性 空调系统在各种耐久性测试前与测试后必须满足的功能要求见表 7。 表 7 功能测试要求
功能测试
耐久性测试
段落
风门性能 3.4.1
渗水现象 3.4.2
操作用力 3.4.3
线路连接 3.4.4
漏风现象 3.4.5
耐高温
45.0
25.0
100 km/h 下持续至 90min
50.0
25.0
怠速下持续 至 110min
40.0
25.0
3.24 车窗内表面除霜 按 GB11556-1994 汽车风窗玻璃除霜系统的性能要求及试验方法进行。
12860汽车空调HVAC结构设计指导书190402
汽车空调HV AC结构设计作者:蔡成龙单位:芜湖博耐尔汽车电气系统有限公司序:首先感激潘总给于我过开展Benchmark工作的机会，才有此HVAC结构设计的一点心得。
同时感谢博耐尔相关同事五年来的相互合作，给我人生一点美好回忆。
目录第1章HVAC介绍 (4)1.1 术语 (4)第2章HVAC 设计规范 (5)2.1．HVAC结构设计 (5)2.1.1各种HVAC一览 (5)2.1.2单区/多区HVAC (7)2.1.3多区HVAC的关键结构的设计因素 (8)2.1.4 HVAC 的噪音控制 (9)2.2加热功能 (10)2.2.1加热器芯的选择 (10)2.2.2 加热器芯在HVAC中的位置 (12)2.2.3 加热器芯与HVAC壳体的配合 (14)2.2.4 PTC 的选择 (14)2.2.5 PTC 在HVAC 中的装配位置 (15)2.3制冷功能 (17)2.3.1 蒸发器的选择 (17)2.3.2 蒸发器在HVAC中的位置 (17)2.4 进气与吹风功能 (20)2.4.1进气装置 (20)2.4.2鼓风装置 (22)2.5空气净化功能 (25)2.6空气混合与分发功能 (26)2.6.1混合功能 (26)2.6.2分发功能 (27)2.7运动机构 (29)2.7.1运动机构设计要素 (29)第3章HVAC主要功能评价试验及方法 (31)3.1试验准备和试验台架 (31)3.2风量试验 (32)3.3风量分配试验 (32)3.4温度平衡试验 (34)3.5蒸发器芯体换热性能试验 (35)3.6加热器芯体换热性能试验 (37)第4章其他设计要素 (38)4.1塑料壳体的装配 (38)4.2紧固件的选择与标准化 (41)4.3鼓风电机与微电机 (41)4.4温度传感器 (41)4.5 HVAC用非金属材料 (43)4.6材料的可回收与环保 (43)第1章HVAC介绍1.1 术语HVAC: 是英文Heating Ventilating Air Conditioning的缩写，即采暖，通风与空调；指安装在仪表板下具有加热、通风、空气调节功能的单元，包含鼓风机总成、加热器芯体、蒸发器芯体、混合风门、模式风门等主要部件。
HVAC总成设计指南
XXXX股份有限公司HV AC总成设计指南编制：审核：批准：设计指南编号：目录目录 (3)1 HV AC简要说明 (4)1.1 该零件综述 (4)1.2.适用范围 (4)1.3 HVAC基本组成 (4)1.4设计构想 (4)1.4.1 设计原则 (4)1.4.2 功能要求 (4)1.4.3 顾客要求 (4)1.4.4 性能要求 (5)1.4.5 设计步骤和参数 (5)2. HV AC的测试规范 (13)2.1 测试内容 (13)2.2 测试标准、方法 (13)3 一般注意事项 (14)4 图纸模式 (14)4.1 图纸主要内容和形式 (14)4.2 图纸的技术要求： (14)1 HV AC简要说明1.1 该零件综述HV AC为汽车提供制冷、取暖、除霜、空气过滤和湿度控制、车内出风大小控制的功能，使乘室内人员更加舒适，驾驶更加安全,由制冷装置、采暖装置、通风装置、净化装置、电控单元组成，这些装置组成了完整的HV AC功能。
HV AC的类型：HV AC的类型按功能分：冷热混合型、单制热型、单制冷型；HV AC的类型按控制方式分：自动、电动、手动；HV AC的类型按区域分：单区域、双区域、多区域。
1.2.适用范围设计指南本部分适用范围为乘用车HV AC总成。
1.3 HV AC基本组成HV AC总成的组成由：进风口、鼓风机、蜗壳、调速电阻、蒸发器芯体-膨胀阀、加热器芯体、风门、出风口、微电机、拉丝、运动机构、传感器、线束、紧固件等组成。
1.4设计构想1.4.1 设计原则1、满足整车提供的布置空间要求。
2、根据整车定位考虑HV AC的选型（自动控制、电动控制、手动拉丝控制、单区、双区）。
3、对HV AC单体的噪音值进行合理定义（最终满足整车的噪音要求）。
4、满足整车装配要求，方便安装和拆卸。
1.4.2 功能要求1、满足整车通风、制冷、制热、除霜除雾的功能要求。
2、满足相关功能指示、显示的功能。
1.4.3 顾客要求1、制冷、采暖效果良好。
Trane ReThink HVAC系统设计说明书
Rental ServicesReThink HVAC Transforming your business is Ea a SyWhy ReThink?Time for a ReThink.Businesses require cooling or heating for process or comfort. This is the task of your HVAC system, asignificant piece of equipment and traditionally a fixed asset/fixed cost. Businesses today scale up anddown rapidly, with cooling or heating needs following these trends. The dynamics of the demand are often inconflict with the static nature of the installation.Equipment repairs, replacement and upgrades represent substantial capital investments, not without risk.Just as cloud computing services have replaced in-house systems, Trane Ea a sy is a short-term or long-termrental program giving you E quipment a s a S ervice (EaaS). So you pay for using the equipment to deliver thecooling and heating you need, without the risk and burden of ownership.Trane Ea a Sy makes HVAC simple and gives you what you want, when you need it:• A flexible solution, not a fixed asset.• Paying for the capacity you actually need.• Easy switching to more powerful and efficient equipment.• An end to maintenance headaches.Now you can focus on what you do best: running your operation.How it worksHaving assessed your requirements in detail, Trane Engineers will propose the right equipment to meet yourneeds in a flexible and customizable formula, with an agreed and all-inclusive monthly rate.Now you can enjoy full control of your operating costs, lower risk and maximum flexibility.We call it ReThinking. You’ll call it Ea a Sy.• Ea a Sy COOLING• Ea a Sy FREE COOLING• Ea a Sy FREE HEATING• Ea a Sy STANDBY• Ea a Sy PARTIAL UTILIZATIONTrane ServicesThe real expertise of a manufacturerAt Trane, we are committed to providing a comprehensive portfolio of HVAC solutions throughout your system lifecycle.Breakdown resolutionNo one plans for breakdowns, but when they happen you need the right partner. Our expert Service Engineers use the latest diagnostic tools to guide you through your options to Repair, Renew, Replace or ReThink.Secure operationsAt every point during the lifetime of your equipment - installation, commissioning, maintenance or breakdown - Trane can offer an effective solution with commissioning, first-aid kits and service agreements.System upgradeTrane Building AdvantageTrane is committed to bringing the latest technological advantages to our customers through a wide portfolio of solutions which increase the Efficiency, Reliability and Sustainability of their HVAC plants. Our Service Engineers use their expertise together with the latest diagnostic tools to future-proof your system and make it “better than before”.Equipment rentalFor special events, exceptional needs or when you want to ReThink HVAC management, Trane Rental Services have the right solution. With our extensive fleet of equipment, we can perfectly match your temporary heating and cooling requirements.Contact usWith over 1000 of the best trained sales engineers and service technicians in the industry, Trane is in the best position to serve your needs. Just call us and we will help you select the ideal Trane EaaSy solution.• Systems approach • Dependable installations • Energy saving solutions • Operating cost optimization • Chiller plant management solutions •Chilled water production solutions.SRV-SLB245-GB October 2020© 2020 Trane. All Rights Reserved.Trane – by Trane Technologies (NYSE: TT), a global climate innovator – creates comfortable, energy efficient indoor environments through a broad portfolio of heating, ventilating and air conditioning systems and controls, services, parts and supply. For more information, please visit trane.eu or .。
HVAC风门设计指南
2.1 HVAC风门简介及其功能类型空调箱要实现进风切换车内外循环进气通道、调节冷暖风量混合比例、以及各种模式（吹面、吹脚、除霜）出风功能，需靠各种相关风门的运动来实现密封、分流其功能。
按照上述功能HVAC一般有三种功能类型风门，分别是内外循环风门、冷暖风门、模式风门。
内外循环风门：通过旋转内外循环风门的位置，实现车外空气和车内空气的流通通道，从而使HVAC进气装置实现外循环或内循环。
另外通过电气控制或机构结构限制也可实现带一定比例内外循环风门开度增加舒适度和改善空气质量。
图2.1.1 一种内外循环机构装置示意冷暖风门：通过调整冷暖风门的不同开度、位置，从而改变冷空气和热空气的不同混合比例实现不同的冷暖温度。
图2 冷暖风门处于全冷模式示意图2.1.3 冷暖风门处于全热模式示意图2.1.4 冷暖风门处于冷暖混合模式示意模式风门：为了更好的提升车内空调的舒适性，需要使用模式风门实现空调箱不同模式的风量分发功能。
其中模式风门主要有吹面风门、吹脚风门、除霜风门。
根据HVAC的结构特征也可以实现不同的模式风门共用一个风门（如吹面、吹脚共用一个吹面吹脚风门，吹面、除霜共用一个吹面除霜风门），同时根据HVAC功能及结构特征一个模式风门也可有多个风门（如吹面风门有前吹面和后吹面风门、吹脚风门有前吹脚风门和后吹脚风门等）。
通过吹面(Vent.),吹脚(Feet),除霜(Def.)模式风门，根据不同需要组成组合成多种的气流分发模式，如吹面模式、吹面吹脚模式、吹脚模式、吹脚除霜模式、除霜模式、吹面吹脚除霜模式，其中吹面模式、吹面吹脚模式、吹脚模式、吹脚除霜模式、除霜模式这5种模式是目前最为常见的气流分发模式。
图2.1.5 两种不同的风门位置实现相应的气流分发模式示意2.2 HVAC风门的结构类型和材料选型根据风门的形状结构不同，一般有平板风门、蝶形风门、弧形风门、薄膜风门、齿轮风门等，如下图：平板风门结构案例蝶形风门结构案例图2.2.3 弧形风门结构案例图2.2.4 薄膜风门结构案例图2.2.5 齿轮风门结构案例缺点结构三（7202等项目）需要风轴门内加转轴的建议使用结构四（6014等项目）与结构一相似，可考虑优化成结构一结构五（6101等项目）由于受力面积小，此结构不推荐以上是一些常见的成熟门轴和曲柄对接结构及尺寸，仅供选型参考。
PDMS软件工程设计指南-Design模块-暖通建模（Hvac）
HVAC工程设计指南一、HVAC简介 1、HVAC是指“H eating, V entilation and A ir C onditioning ducting networks”即采暖、通风及配送管路系统；2、在PDMS软件提供的“HVAC”设计系统中，用户通过简单的选择管路元件，完成管路系统的设计，最后通过确省的系统设置得到“HVAC”系统的平面布置图；3、“HVAC”设计系统中提供了“HVAC”系统设计所需要的大部分设计元件模型（超过100多种）、扶助元件（“stiffening frames, access panels, splitter plates ”等）及在线的设备模型；4、通过系统提供的元件规范库的功能，用户可以建立用户及公司标准的“HVAC”系统元件库及设计等级；5、根据详细的“HVAC”系统元件的设计，系统进行在线的或后期的模型碰撞检查；二、“HVAC”系统的建立1．进入PDMS，DESIGN设计环境* Project –设置项目名称；* Username –设置用户名称；* Password –设置用户密码；* MDB –设置用户数据库；* Module –设置用户要进入的模块名称；（进行“HVAC”建模，选择“Design”模块）* Read Only –设置数据库的读写属性，即用户可以选择以读写或只读的方式进入相应的设计模块；* Load from –设置用户进入设计模块的环境设置文件；** User’s Binary –用户的环境设置文件；** Default Binary –系统确省的环境设置文件；** Select Binary –选择特定的环境设置文件；** Macro File –系统初始的环境设置文件；* 单击“OK”按钮，进入PDMS软件设计模块；2、DESIGN设计环境简介* 标题栏–显示当前模块名称；主工具栏图形显示窗体三维视图工具条数据导航窗体* 主菜单栏–显示程序的主要功能；* 主工具栏–程序主要便捷工具；* 图形显示窗体–三维设计模型显示窗体；* 三维视图工具条–控制图形显示窗体三维图形显示的状态及工具；（注：“DESIGN”设计模块的菜单、窗体及工具栏的详细说明及设置见《DESIGN设计模块菜单》中的详细描述。
HVAC总成标准
Q/XZ 南京协众汽车空调集团有限公司企业标准Q/XZ 04－2017代替Q/XZ 04－2013HVAC总成标准2017－05－08发布2017－05－20实施目次前言 (2)1 范围 (3)2规范性引用文件 (3)3 术语和定义 (3)4 使用环境 (3)5 试验总体要求 (4)6 材料 (4)7 测试过程和要求 (4)7.1 形式与功能 (4)7.2 气流特性 (6)7.3 热交换性能 (8)7.4 NVH (10)7.5 总成水管理 (11)7.6 耐久性能 (12)7.7 电磁兼容性要求 (13)8产品供货要求 (14)9检验规则 (14)10包装、运输和储存 (17)1前言本标准是根据GB/T1.1-2009的要求，作为公司企业标准文件编写时规定的纲领性文件，并为公司企业标准编写提供了依据。
本标准代替Q/XZ 04-2013 《HVAC总成标准》（试行），与Q/XZ 04-2013 《HVAC总成标准》（试行）相比，主要技术变化如下：标准修改了总成水管理中的测试方法及评价指标本标准由南京协众汽车空调集团有限公司提出。
本标准由南京协众汽车空调集团有限公司研究院归口管理。
本标准起草单位：协众研究院本标准修订人：王迎本标准所代替标准的历次版本发布情况为：Q/XZ 04-2013。
2HVAC总成标准1 范围本标准规定了空调箱的总体性能、风量分配和可靠性等要求。
本标准适用于协众空调箱产品的开发和验证。
2 规范性引用文件下列文件中的条款通过本标准的引用而成为本标准的条款。
凡是注日期的引用文件，其随后所有的修改单（不包括勘误的内容）或修订版均不适用于本标准。
然而，鼓励根据本标准达成协议的各单位研究是否可使用这些文件的最新版本。
凡是不注日期的引用文件，其最新版本适用于本标准。
QC/T 413 汽车电器设备基本技术条件Q/FT B102 车辆产品零部件可追溯性标识规定XZCX7.3.1(1)-ZY01 南京协众集团产品包装规范BAS-362 汽车电气及电气设备气候环境要求及其试验方法BAS-363 汽车电气及电气设备电气性能要求及其试验方法BAS-364 汽车电气及电气设备机械性能要求及其试验方法BAS-365 汽车电气及电气设备EMC性能要求及其试验方法BAS-338 汽车用ABS类材料QC/T625 汽车用涂镀层和化学处理层GB 11555 汽车风窗玻璃除霜和除雾系统的性能和试验方法GB 18655 用于保护车载接收机的无线电骚扰特性的限值和测量方法QC/T198 汽车用开关通用技术条件QC/T266 汽车零件未注公差尺寸的极限偏差一般要求QC/T413 汽车电气设备基本技术条件3 术语和定义下列术语和定义适用于本标准：TXV –热力膨胀阀；HVAC Module –空调箱总成，含蒸发器芯体、暖风芯体、膨胀阀、鼓风机组件、风门执行电机、调速电阻、调速模块和蒸发器芯体热敏电阻等；REC - 内循环模式。
HVAC基础知识
1.1.7 HVAC 总成的壳体设计 1.1.7.1 HVAC 壳体设计要考虑的问题
①、 首先要考虑 HVAC 总成本身的最大尺寸（长、宽、高），以及整车可 以提供的最大的布置空间（在副驾驶的仪表台下）
②、 根据实际的布置空间及 HVAC 总成的大概的重量来确定 HVAC 的固定 方式及固定结构的设计，一般固定在防火墙上和 IP 支架上
根据整车内空间的大小，乘员的多少来计算需要多大的热换能力，具体计 算过程参考蒸发器芯体的计算过程及方法，来计算加热器芯体的尺寸，并选择适 合的加热器芯体。 1.1.6 鼓风机的设计与选配 1.1.6.1 鼓风机的作用
鼓风机是把加热器芯体和蒸发器芯体产生的热量和冷度吹到室内为驾驶 员和乘客提供热风和冷风来提高整车的舒适性；鼓风机为除霜除雾提供强有力的 风源；鼓风机为驾驶室内的通风、换气提供动力。 1.1.6.2 鼓风机的类型：有刷电机、无刷电机； 1.1.6.3 鼓风机的功率：鼓风机在轿车上的额定电压为 12V，功率为 100W～500W 鼓风机根据实际的需要相适应的功率，具体如下：
⑤、 无论温度和分配器的调节装置处于何位置，也无论鼓风机处于何档 位，都不允许发出如哨声、嘘声、呜声和振颤抖声一类的干扰性噪声。
⑥、 左、右吹脚出风口之间的温差最大为 3℃，右侧温度高一些。 ⑦、 在最大“COOL”位置，侧出风口和中左、中右出风口之间的温差最
大为 3℃，在整个温度调整过程中，所有出风口之间的总温差不得超 过 8℃。 ⑧、 在最大“COOL”位置，侧出风口和中左、中右出风口之间的风速偏 差最大为 0.6M/S，在整个温度调整过程中，所有出风口之间的总温差 不得超过 2M/S。 1.1.7.2 HVAC 壳体的材料要求：HVAC 壳体 PP-T20，风门 PP-T40，传动机构 POM， 原材料最好是从以下供应商选择：PC：德国拜尔、GE；POM：宝理、杜邦； PP：南京金杉、南京聚隆； 1.1.7.3 HVAC 总成的布置及相关设计注意事项 ①、HVAC 总成的布置及相关设计;在正式模具开模前首先要用手工样件
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XXXX股份有限公司HV AC总成设计指南编制：审核：批准：设计指南编号：目录目录 (3)1 HV AC简要说明 (4)1.1 该零件综述 (4)1.2.适用范围 (4)1.3 HVAC基本组成 (4)1.4设计构想 (4)1.4.1 设计原则 (4)1.4.2 功能要求 (4)1.4.3 顾客要求 (4)1.4.4 性能要求 (5)1.4.5 设计步骤和参数 (5)2. HV AC的测试规范 (13)2.1 测试内容 (13)2.2 测试标准、方法 (13)3 一般注意事项 (14)4 图纸模式 (14)4.1 图纸主要内容和形式 (14)4.2 图纸的技术要求： (14)1 HV AC简要说明1.1 该零件综述HV AC为汽车提供制冷、取暖、除霜、空气过滤和湿度控制、车内出风大小控制的功能，使乘室内人员更加舒适，驾驶更加安全,由制冷装置、采暖装置、通风装置、净化装置、电控单元组成，这些装置组成了完整的HV AC功能。
HV AC的类型：HV AC的类型按功能分：冷热混合型、单制热型、单制冷型；HV AC的类型按控制方式分：自动、电动、手动；HV AC的类型按区域分：单区域、双区域、多区域。
1.2.适用范围设计指南本部分适用范围为乘用车HV AC总成。
1.3 HV AC基本组成HV AC总成的组成由：进风口、鼓风机、蜗壳、调速电阻、蒸发器芯体-膨胀阀、加热器芯体、风门、出风口、微电机、拉丝、运动机构、传感器、线束、紧固件等组成。
1.4设计构想1.4.1 设计原则1、满足整车提供的布置空间要求。
2、根据整车定位考虑HV AC的选型（自动控制、电动控制、手动拉丝控制、单区、双区）。
3、对HV AC单体的噪音值进行合理定义（最终满足整车的噪音要求）。
4、满足整车装配要求，方便安装和拆卸。
1.4.2 功能要求1、满足整车通风、制冷、制热、除霜除雾的功能要求。
2、满足相关功能指示、显示的功能。
1.4.3 顾客要求1、制冷、采暖效果良好。
2、除霜除雾效率高。
3、空调开启时噪音小。
4、安装拆卸方便，便于维修，使用寿命长。
1.4.4 性能要求DVP试验主要分为：可靠性试验、性能实验、法规试验、舒适性试验。
可靠性实验：道路试验、抗振强度、耐温度实验、耐压力实验、耐久实验、腐蚀实验、过电压、过电流实验、绝缘性能、阻燃性能等。
性能实验：制冷能力、采暖能力、空气体积流量、滤清器过滤性能。
法规实验：鼓风机、微电机EMC实验，电源线射频传导发射、射频辐射发射、瞬态电压发射等。
舒适性实验：噪音、操纵性能、线性实验、挥发性、气味性实验等。
1.4.5 设计步骤和参数1.4.5.1 HV AC设计的步骤1.4.5.1.1 设计输入1、整车空调系统配置表的确定根据空调系统的VTS，由采购部门、销售公司、项目组、设计部门共同讨论确定空调的配置，明确空调系统的档次，是否采用功能或结构上的新技术等。
进而确定空调的类型：手动空调、全自动空调或是电动空调。
2、根据整车空间的大小、销售市场、顾客需求及竞争车型的空调配置来确认是不是需要开发多个HV AC总成（如果需要开发两个：一般情况下一个布置在仪表台下，一个布置在顶部或是布置在车后部）；3、根据整车的配置来确定是采用单区域还是采用多区域空调，然后对HV AC的结构进行设计4、根据销售区域的不同来确定是冷暖分开型、或是冷暖合一型或是全功能的HV AC。
1.4.5.2 HV AC的设计1.蒸发器芯体的设计⑴.蒸发器芯体的类型：管片式、管带式、层叠式。
⑵.蒸发器的作用蒸发器是汽车空调制冷系统中另一个热交换器，结构形式与冷凝器基本相同，其作用是将经过节流降压后的液态制冷剂在蒸发器内沸腾汽化，吸收蒸发器表面周围的热量而之降温，电动机驱动的鼓风机再将冷风吹到车室内，使进入其中的制冷剂与其外部空气完成热交换，蒸发器外部的空气放热冷却，达到降温的目的；蒸发器是由铝制芯管和散热片组成。
要求蒸发器具有效率高、尺寸小、质量轻等特点。
⑶.蒸发器芯体的设计主要是制冷能力的计算 （下面以某一车型实例为主）：根据现有的条件，计算汽车空调蒸发器总传热外表面积，对比分析现有蒸发器的制冷能力。
①.工况要求制冷剂为R134a,换热量Q=5809w 。
制冷剂循环量mr q =0.046kg/s,蒸发温度ft =2℃,蒸发器进风干球温度1a t =24℃．湿球温度‘1a t =17℃,出风口干球温度2a t =7.25℃．湿球温度‘2a t =6.5℃,取迎风面风速为a υ=3m/s.②.计算1. 按热平衡关系计算制冷剂进出口参数根据制冷量和制冷剂循环量,可求出制冷剂进出口比焓差r h ∆=12r r h h -=Q/mr q =5809/0.046j/kg=126.28kj/kg制冷剂入口干度通常在20%~30%之间,取制冷剂进口干度为0.3,根据蒸发温度查HFC134a 和空气热力性质表,可计算得制冷剂出口比焓值2r h =1r h +r h ∆=(261.62+126.2⑻kj/kg=387.9kj/kg蒸发器出口制冷剂温度2r t =8℃,过热度为6℃ 2.干工况下空气侧表面传热系数aα根据已知条件,按空气进口的平均温度的平均值a t =20℃,查取空气的热力性能表和热物理性质图计算空气侧表面传热系数a α:32108.11057.2055.10--⨯⨯⨯==l a p Nu λαw/(㎡·k)=143.56w/(㎡·k)3.湿工况下空气侧表面传热系数aeq ，α根据已知条件,查空气进出口风温下的热力性质表和热物理性质图,可以求得:a eq ，α=ξaα=1.855×143.56 w/(㎡·k)=266.3 w/(㎡·k)4.计算制冷剂侧表面传热系数r α 查ft =2℃下的HFC134a 饱和状态下的热力性质表和热物理性质图,计算出r α=3965 w/（㎡·k ）5.计算总传热系数及传热面积如果忽略管壁热阻及接触热阻，忽略制冷剂侧污垢热阻，取空气侧污垢热阻ar =0.0003㎡·k/w ，则传热系数K 为K =3.26610003.00283.04248.0396511111++⨯=++aeq a r a r r A A ，ααw/（㎡·k ）=127.5w/（㎡·k ） 对数平均温差mt ∆为m t ∆=⎪⎪⎭⎫⎝⎛---f a f a a a t t t t t t 2121ln =⎪⎭⎫ ⎝⎛---210224ln 1024℃=13.84℃ 由于管片式蒸发器的流程较少,另外湿工况在增大空气侧表面传热系数的同时也增加了液膜阻力,因此空气侧的实际表面传热系数低于计算结果,所以要加上一个修正因子,ϕ=1.2,故所需传热面积(以外表面为基准)Ao 为ϕQ/(K·m t∆)=1.2×5809/(127.5×13.8⑷㎡=3.95㎡Ao=所需圆管长度L为L=Ao/45Aa=3.95/(45×0.424⑻m=0.2066m取L=210mm另实际传热面积与计算传热面积的对比可以通过比较管长来反映出制冷能力, XX项目引用XXXX车型的蒸发器管长为263mm,计算长度仅为:210mm。
通过以上计算可以推算出空调蒸发器的外型尺寸，根据推算出来的外型尺寸的来选择合适的蒸发器芯体。
2、加热器芯体的设计⑴.加热器芯体的作用加热器芯体是为整车提供热量转化的，一般都是铝质材料，因为铝质的材料导热系数比较大，重量轻，适合在汽车有效的空间上布置。
加热器芯体的热源来源于发动机防冻液，故加热器芯体的换热能力和换热速率很大程度上取决于发动机的防冻液的温度变化。
⑵.加热器芯体的结构形式：管片式、管带式、层叠式、平行流。
目前大多数汽车空调的加热器芯体采用层叠式，层叠式热交换能力比较高，一般可以达到80％以上，比较先进的平行流式热交换能力更好一点可达到90％以上。
⑶.加热器芯体的选型根据整车内空间的大小，乘员的多少来计算需要多大的热换能力，具体计算过程参考蒸发器芯体的计算过程及方法，来计算加热器芯体的尺寸，并选择适合的加热器芯体。
3、鼓风机的设计与选配⑴.鼓风机的作用鼓风机是把加热器芯体和蒸发器芯体产生的热量和冷度吹到室内为驾驶员和乘客提供热风和冷风来提高整车的舒适性；鼓风机为除霜除雾提供强有力的风源；鼓风机为驾驶室内的通风、换气提供动力。
⑵.鼓风机的类型：有刷电机、无刷电机。
⑶.鼓风机的功率：鼓风机在轿车上的额定电压为12V，功率为100W～500W鼓风机根据实际的需要相适应的功率，具体如下：①.根据整车的室内空间的大小，采用CFD分析，使空气在室内的流动在规定的时间内温度场的分析来确定鼓风机的风量及风速；②.进行除霜除雾的CFD的分析，根据企业标准和国标以及美标fmvss103/saej902v003的规定时间，在有效的时间内把玻璃上的霜或是雾除掉，根据除霜的效果来推算鼓风机的风量及风速是否满足要求，风量小，增加鼓风机的风量（也就是功率）；③.鼓风机的选择还要考虑到噪声的问题，一般噪音有严格的规定，详见国标，鼓风机的噪音和功率是成正比的，就是说功率越大噪音越大；所以在选择鼓风机时要考虑鼓风机的功率和噪音的平衡点；④.鼓风机在高档车上一般采用无刷电机，这样即可以降低噪音，也不会损失相应的功率。
4、HV AC总成的壳体设计⑴.HV AC壳体设计要考虑的问题①.首先要考虑HV AC总成本身的最大尺寸（长、宽、高），以及整车可以提供的最大的布置空间（在副驾驶的仪表台下）；②.根据实际的布置空间及HV AC总成的大概的重量来确定HV AC的固定方式及固定结构的设计，一般固定在防火墙上和IP支架上；③.根据整车的布置舒适性来考虑设计后排脚部出风口的布置及HV AC本身的出风口设计；还要考虑是否向B柱处通风的配置（出于整车舒适性配置的要求）；④.根据整车舒适性要求考虑是采用自动控制、电动控制、手动拉丝控制，如果采用拉丝控制，就要考虑拉丝的布置，需要给除霜除雾及吹面的风道留有足够的布置空间，布置空间是否合适采用CFD分析验证；⑤.无论温度和分配器的调节装置处于何位置，也无论鼓风机处于何档位，都不允许发出如哨声、嘘声、呜声和振颤抖声一类的干扰性噪声；⑥.在蒸发器上出现的冷凝水必须通过排水管正常排出壳体，而且在路试试验中不允许出现水流不畅的现象；⑦.HV AC与相关联的对接风道间均应增加密封措施（如加贴密封海绵等）：1、以防止漏风量过大，导致HV AC功效下降，不能满足整车性能要求；2、密封不严风噪加大，导致整车噪音过大；⑧.与HV AC相关联的对接风道走向应避免过渡不自然（出现直角或锐角转弯）：1、造成风道内风阻过大或扰流，导致HV AC功效下降，不能满足整车性能要求；2、风管截面积小风阻大，导致噪音增大。