LAHU Heat Recovery System Optimal Operation and Control Schedules

Transcription

1 Yujie Cui Nexant, Inc., 44 South Broadway, 5th Floor, White Plains, NY Mingsheng Liu Architectural Engineering, University of Nebraska Lincoln, 206C, PKI, 1110 S. 67th St., Omaha, NE LAHU Heat Recovery System Optimal Operation and Control Schedules Optimal operation and control of heat recovery in an integrated Laboratory Air Handling Unit (LAHU) system differs substantially from that in conventional dedicated AHUs for laboratory buildings with a 100% outside air AHU for laboratory spaces, since the LAHU allows economizer operation for both offices and laboratories. Optimal operation and control schedules of the heat recovery systems in the LAHU have been developed to minimize the total thermal energy cost. This paper presents the procedure, methodology, and results of generic optimal heat recovery control schedules for the LAHU and investigates its impact on the LAHU potential thermal and pump energy savings. The optimal control schedule can potentially save 14% to 27% thermal energy and 17% to 100% pump energy during the winter under weather conditions that prevail in Omaha, Nebraska. The findings discussed in this paper also apply to any heat recovery system, where AHU has an economizer function. DOI: / Introduction Most modern laboratory buildings include both office and laboratory sections. A dedicated AHU for the laboratory section is conventionally designed with 100% outside air, and thus consumes a large amount of thermal energy to condition outside air. To reduce the thermal energy consumption utilizing a large amount of exhaust from laboratories, heat recovery system is often installed in laboratory buildings. With conventional optimal control, a heat recovery system is turned on and controlled at full operation when the absolute temperature difference of the outside air and the exhaust air is higher than a certain value. During the winter, the heat recovery system receives energy from the hot exhaust air and warms up the cold outside air. During the summer, the heat recovery system rejects energy to the cold exhaust air and cools down the hot outside air. Recently, the authors have developed the Laboratory Air Handling Unit LAHU, which provides for return air transfer from offices to laboratories, implementation of economizer in both office and laboratory sections, and different supply air temperatures to the office and the laboratory sections 1. Because the LAHU system allows economizer operation in both the office and the laboratory sections, the conventional heat recovery operation and control schedules do not apply to the LAHU system. In this paper we introduce the LAHU system, present the principle, methodology, and analytical results of the optimization of LAHU heat recovery system discharge air control, and investigate the impact of optimal schedules on potential thermal energy and heat recovery pump power savings. Lahu System Figure 1 shows the schematic diagram of a LAHU system. A central runaround coil heat recovery system is typically installed due to its high flexibility in laboratory buildings where the fresh air intake and the exhaust air are mostly remote. It is also free of cross contamination because heat is transferred between the two airstreams through the loop instead of via direct contact 2. The Contributed by the Solar Energy Division of ASME for publication in the JOUR- NAL OF SOLAR ENERGY ENGINEERING. Manuscript received March 2, 2005; final manuscript received August 22, Review conducted by Vince C. Mei. Paper presented at the 2004 International Solar Energy Conference ISEC2004, July 11 14, 2004, Portland, Oregon, USA. LAHU system serves both the office and laboratory sections with two supply air fans S.F.1 and S.F.2, one return air fan R.F., and fume hood exhaust fan s. Supply air fan 1 S.F.1 supplies air to the office section. Supply air fan 2 S.F.2 supplies air to the laboratory section. The return fan R.F. circulates air from the office section to either supply fan 1 or 2, or both. Return air can be sent either upstream or downstream of cooling coil 2, or in both locations. The return air distribution to each supply air fan is modulated using return air damper 1, 2, and 3 and the relief air damper. The outside air intake to the LAHU system is modulated by outside air damper 1 and 2. A heat recovery pump drives the water through the heat recovery coils to transfer heat between the laboratory exhaust and outside air to the LAHU system. Optimal Heat Recovery System Control The objective of the heat recovery optimization is to find the optimal heat recovery discharge air temperature of the LAHU that minimizes the total thermal energy including heating, Q h, cooling, Q C, and humidification energy, Q hu costs. The optimization parameters are the optimal heat recovery discharge air temperature T hr under given outside air conditions t oa,h oa, supply airflow, cold deck set points T c,1,t c,2, and energy prices d h,d c. Assuming heating and humidification have the same energy cost d h, the objective function is then defined as: Ḋ = d c Q c + d h Q h + Q hu. 1 Equation 2 shows the building thermal energy balances of the LAHU system control volume, as shown in Fig. 2: Q h + Q hu = Q c + Q re + Q e + Q eh Q i + Q env Q hg. When Eq. 2 is substituted into, Eq. 1 becomes: Ḋ = d h + d c Q c + d h Q re + Q e + Q eh Q i + d h Q env Q hg. 3 In Eq. 3, thermal energy that is carried by airstreams can be given as: Q re + Q e + Q eh Q i = Ṁ re + Ṁ e + Ṁ eh h r Ṁ i h hr. The airflow balance is / Vol. 128, AUGUST 2006 Copyright 2006 by ASME Transactions of the ASME

2 Fig. 1 Schematic of LAHU system with runaround coil heat recovery unit Fig. 2 LAHU system thermal energy flow and airflow Journal of Solar Energy Engineering AUGUST 2006, Vol. 128 / 361

3 Table 1 Outside air schedule constraints Outside air schedule constraints and objective functions in the dry coil mode Objective functions t hr T c,1 t hr T lec a A=d h 1 + h r W oa 1061 C p W oa t hr CL 2 1 t hr =T lec T f a A=d h 1 + h r W oa 1061 C p W oa t hr t hr T f b A=d h T c,2 T c,1 + T r T c,2 T r t hr h r W oa 1061 C p W oa t hr t hr T lec a A=d h 1 + h r W oa 1061 C p W oa t hr t hr =T lec T f a A=d h 1 + h r W oa 1061 C p W oa t hr CL 2 1 t hr =T f T t b A=d h T c,2 T c,1 + T r T c,2 T r t hr h r W oa 1061 C p W oa t hr t hr T t c A= d h +d c C p T r T c,1 t r T hr 1 T c,2 +d h 1 + h r W oa 1061 C p W oa t hr t hr =T c,1 T c,2 CL 2 1 t hr T lec d A= d h +d c C p t hr T c,1 +d h 1 + h r W oa 1061 C p W oa t hr t hr T lec e A= d h +d c C p t hr T c,1 +d h + ec,2 1 h r W oa 1061 C p W oa t hr CL 2 1 f A= d h +d c C p 1 t hr T r T c, t hr T r 1 T c,2,0 + +d h 1 + h r W oa 1061 C p W oa t hr t hr =T c,2 T r CL 2 1 g A= d h +d c C p t hr T c,1 + 1 t hr T c,2 +d h h r W oa 1061 C p W oa t hr CL 2 1 h A= d h +d c C p 1 t hr T r T c, t hr T r 1 T c,2,0 + +d h 1 + h r W oa 1061 C p W oa t hr t hr T r h A= d h +d c C p 1 t hr T r T c, t hr T r 1 T c,2,0 + +d h 1 + h r W oa 1061 C p W oa t hr where 1 = 1, +1 ; 2 = 0, ;CL 2 =d h h oa,v W r W oa d c C p T r t oa Ṁ i = Ṁ re + Ṁ e + Ṁ eh. 5 When Eq. 5 is introduced, Eq. 4 can be represented as Eq. 6 : Q re + Q e + Q eh Q i = Ṁ i h r h hr. Introducing Eq. 6 into Eq. 3, the thermal energy cost is or Ḋ = d h + d c Q c + d h Ṁ i h r h hr + d h Q env Q hg Ḋ = d h + d c Q c,1 + Q c,2 + d h Ṁ i,1 + Ṁ i,2 h r h hr + d h Q env Q hg. Since the term Q env Q hg is constant for a given building load, the objective function can be expressed as Eq. 9 : A = d h + d c Q c,1 + Q c,2 + d h Ṁ i,1 + Ṁ i,2 h r h hr. 9 When Eq. 9 is divided by the entire building supply airflow rate, the objective function Eq. 9 is rewritten as Eq. 10 : A = d h + d c q c,1 + q c,2 + d h h r h hr 10 In order to introduce discharge air temperature t hr into the objective function, Eq. 10, the relationship of heat recovery discharge air enthalpy h hr and its dry bulb air temperature t hr is given by Eq : h hr = C p t hr + W oa t hr. 11 Substituting Eq. 11 into Eq. 10, the objective function is rewritten as Eq. 12 : A = d h + d c q c,1 + q c,2 + d h h r W oa 1061 C p W oa t hr. 12 The objective function is constrained by heat recovery capacity and optimal outside air intake schedules. Heat Recovery Capacity Constraint. Discharge air condition is constrained by heat recovery system capacity under different outside air conditions. Since the runaround coil primarily has sensible heat transfer between the exhaust and outside air stream, the highest heat recovery discharge air temperature limit T L with the effectiveness is expressed as Eq : T L = t oa T r t oa. 13 The heat recovery capacity constraint of the objective function, Eq. 12, then requires that the heat recovery discharge air temperature not be lower than the outside air temperature t oa and not higher than the limit temperature T L if the outside air temperature is lower than room air temperature T r. It is the opposite if the outside air temperature is higher than the room air temperature. t oa t hr T L, t oa T r T L t hr t oa, t oa T r 14 Outside Air Intake Schedule Constraints. The objective function expressed in Eq. 12 relates directly with optimal outside air schedules 1, 2, which depend on the heat recovery discharge air condition t hr or h hr 4. Therefore, different objective functions under different heat recovery conditions are developed. In addition, the cooling energy q c,1,q c,2 and optimal outside air intakes 1, 2 have different formats in dry versus wet coil modes. The outside air intake schedule constraints are discussed next. Dry Coil Mode. In the dry coil mode, cooling energy q c,1,q c,2 is either zero or expressed by a temperature difference depending on the optimal outside air intake 1, 2. Objective function Eq. 12 is thus rewritten as: A = d h + d c 0, 1 t hr T r + T r T c, , 2 t hr T r + T r T c,2 + + d h h r W oa 1061 C p W oa t hr 15 When the optimal outside air intake schedule 1, 2 4 is substituted, the objective function, Eq. 15, is constrained by its corresponding heat recovery discharge air temperature condition related with the optimal schedule 1, 2. Following this approach, the detailed objective functions and corresponding outside air intake schedule constraints for the dry coil mode is developed and listed in Table 1. Parameter CL 2 is the ratio of humidification cost to mechanical cooling energy cost for conditioning the same amount of outside air 4. IfCL 2 is less than 1, the humidification cost is lower than the mechanical cool- 362 / Vol. 128, AUGUST 2006 Transactions of the ASME

4 Table 2 Outside air schedule constraints and objective functions in the wet coil mode Outside air schedule constraints Objective functions t hr T r and h hr h r A= d h +d c h hr h c, h hr h r h c,2 +d h h r C p t hr W oa t hr Where 2 = cr, t hr T r A= d h +d c 1 h hr h r h c, h hr h r h c,2 or +d h h r C p t hr W oa t hr t hr T r and h hr h r where 1 = 1, + 1 +, 2 = 0, ing energy cost. If CL 2 is higher than 1, the humidification cost is higher than mechanical cooling energy cost for handling the same amount of outside air. Where in Table 1, T lec = 1 T c,2 + 1 T r T f = T c,2 + T c,1 T c,2 + 1 T r 1 + T t = T r T r T c,1 T r Wet Coil Mode. In wet coil mode, if heat recovery discharge air enthalpy h hr is lower than room air enthalpy h r, the office uses 100% outside air and the laboratory optimizes the outside air intake to minimize reheat energy consumption. Otherwise, the building takes in the minimum outside air while the office accepts the maximum possible outside air. However, only sensible heat exchanges between the outside air and laboratory exhaust in the runaround coil heat recovery system. The optimal outside air intake schedules in the wet coil mode can then be categorized as Eq. 19 when heat recovery discharge air enthalpy is lower than room air enthalpy, and be categorized as Eq. 20 and Eq. 21 when heat recovery discharge air enthalpy is higher than room air enthalpy, as shown in Table 2. t hr T r and h hr h r 19 t hr T r and h hr h r 20 t hr T r 21 Taking the same approach as for the dry coil mode, the outside air schedule constraints and the corresponding objective functions can be determined and listed in Table 2. Optimization Methodology and Results. It has been proved that each objective function in Table 1 and Table 2 is either an increasing function a function with positive derivative or a decreasing function a function with negative derivative of heat recovery discharge air temperature t hr within its constraints. Therefore, each function has a local extreme point or a local optimal solution among the boundary points of its constraints due to continuity 5. All the local optimal solutions can be searched with this method using analytical derivatives. The global optimal solution can then be determined in terms of the continuity at all local optimal points solutions for all the objective functions 5. The optimal heat recovery discharge air controls are shown in Table 3. In the dry coil mode, the schedules are determined by the comparison of the office airflow ratio,, with two critical office airflow ratios, 1 and 2. The heat recovery system should be turned off whenever the optimal heat recovery discharge air temperature equals the outside air temperature, and turned on otherwise, in order to reduce the heat recovery system operation and maintenance costs. The heat recovery system typically operates when the outside air temperature is lower than the office cold deck set point in the winter and higher than a certain value in the summer. The optimal discharge air temperature schedule can be simplified as Eq. 22 for winter operation and Eq. 23 for summer operation. T hr = Max min T c,1,t f,t L,t oa 22 T hr = Max T c,1,t L 23 Since the building always accepts the least amount of total outside air with the optimal outside air intake, maximum heat recovery discharge air temperature Eq. 13 can be represented as Eq. 24 : 1 T L = t oa T r t oa 24 T f in Eq. 22 is defined as the LAHU Free-Cooling Temperature. When heat recovery discharge air temperature reaches this value, the LAHU uses economizer for both the office and laboratory sections with zero mechanical heating and cooling. LAHU uses the least total outside air with zero air relief. For winter operation, the optimal discharge air temperature is the higher value of the outside air temperature and the minimum value of the office cold deck T c,1, LAHU Free-Cooling Temperature T f and the limit temperature T L. For summer operation, the optimal heat recovery discharge air temperature is the maximum value of the cold deck set point of the office section T c,1 and the limit temperature T L. Actually, the LAHU optimal heat recovery control for the summer has the same schedule as in the conventional control. Coil mode Dry coil Wet coil Table 3 Optimal heat recovery control schedules Outside air condition Optimal heat recovery discharge air temperature Heat recovery system status 1 t oa T f min T f,t L On T r t oa T f t oa 1 and 2 t oa T lec min T lec,t L On T r t oa T lec t oa 1 and 2 t oa T c,1 min T c,1,t L On T r t oa T c,1 t oa t oa T r max T r,t L On t oa T r t oa t oa T r max T r,t L On 1 = T c,2 T c,1 T c,2 T c,1 + 1 T r T c,1 ; 2 =d h C p W oa d h +d c C p +d h 1 C p W oa Journal of Solar Energy Engineering AUGUST 2006, Vol. 128 / 363

5 Fig. 4 Heat recovery discharge air temperatures with the conventional and the optimal schedule of LAHU effectiveness =0.6 Fig. 3 Thermal energy savings of LAHU with heat recovery effectiveness of 0.6 Thermal Energy Savings The potential thermal energy savings of the LAHU with optimal heat recovery control schedules compared with LAHU with no heat recovery operation were calculated. Figure 3 presents the potential savings versus the outside air temperature for different office airflow ratios. The heat recovery has an effectiveness of 0.6 6,7. The cold deck set point is assumed to be 15.6 C 60 F for the office and 18.3 C 65 F for the laboratory in the winter. The cold deck set points are both 12.8 C 55 F in the summer. The laboratory maintains a supply air temperature of 18.3 C 65 F to minimize reheat. The room air condition is controlled at 23.9 C 75 F and 50%. The bin data of Omaha, Nebraska 8 are used. Figure 3 a shows the potential thermal energy savings in Btu/ cfm hr Btu per building airflow rate and per hour. During winter, the potential savings decrease as the outside air temperature increases, and increase as the office airflow decreases. The potential savings can be as high as 26.0 KJ/m Btu/cfm hr when the outside air temperature is 16.7 C 2.0 F and the office airflow ratio is 0.1. This thermal energy savings is about 43% of the total building heating energy consumption. The operation time of the heat recovery system decreases as the office airflow ratio increases. For example, when the office airflow is 0.7, the heat recovery system should be shut off due to zero thermal energy savings when the outside air temperature is 1.7 C 35 F or higher. This may eventually impact the heat recovery system design in certain climates. The heat recovery system should not be designed for mild winter climates or for LAHUs with relatively higher office airflow ratios. For example, if the office airflow ratio is around 0.7 and the typical winter outside air temperature is higher than 1.7 C 35 F, no heat recovery system is needed to reduce heating energy consumption in the LAHU. The initial cost can thus be reduced. In summer, when the outside air temperature is higher than the room air temperature, thermal energy savings increase as the outside air temperature increases for the same office airflow ratio. The savings in summer are comparatively less than in winter for each airflow ratio except for 0.9. Maximum thermal energy savings are achieved when the office airflow ratio is At this airflow ratio, 100% of the return air from office is directly mixed with the cold air from the cooling coil of the laboratory, downstream from the coil. The closer the office airflow ratio to 0.34, the higher is the percentage savings. The savings ranges from 5% to 25% for different airflow ratios in summer. Impacts of Optimal Heat Recovery Control Schedules The discharge air temperature set point of the heat recovery system is typically the cold deck set point for conventional schedules. In winter, the heat recovery system is on when the outside air temperature is lower than the cold deck temperature, if the circulation pump energy is ignored. The LAHU s optimal heat recovery system schedule is different than the conventional schedule. The discharge air temperatures of the heat recovery system under both optimal and conventional control schedules for the winter operation are simulated. The simulations are based on the same cold deck set points, the same design conditions and the same weather conditions as presented in the previous section. Figure 4 presents the results obtained for a heat recovery effectiveness of 0.6. The solid lines are the optimal LAHU schedules and the dashed lines are the conventional schedules. When the discharge air temperature is lower than the LAHU Free-Cooling Temperature as point A for office airflow ratio of 0.7, both schedules have the same discharge air temperature. When the discharge air temperature reaches the LAHU Free-Cooling Temperature, the conventional schedule may produce a higher discharge air temperature if there is no water flow modulation. The conventional control can also maintain the required discharge air temperature cold deck set point by modulating water flow through the coil. The circulation pump must be on when conventional heat recovery system control is used. When the optimal control schedule is used, the heat recovery system is modulated to maintain the heat recovery discharge air at the LAHU Free-Cooling Temperature. When the outside air temperature is higher than the LAHU 364 / Vol. 128, AUGUST 2006 Transactions of the ASME

6 outside air temperature than the typical control in the winter. Pump power and humidification energy is therefore reduced. Room relative humidity level is improved. Conclusions The optimal heat recovery discharge air temperature depends on the office airflow ratio and cold deck set points. During winter, the optimal discharge air temperature equals the outside air temperature when the outside air temperature equals or exceeds the LAHU Free-Cooling Temperature: under these conditions, heat recovery systems should be turned off. This general principle applies to any heat recovery system where the AHU has an economizer function. The optimal heat recovery control of LAHU, developed in this paper, can provide annual thermal energy savings of 14% to 27% when the office airflow ratio is 0.6 or lower. The optimal heat recovery system control schedule also significantly reduces the heat recovery circulation pump energy consumption. Pump energy savings range from 17% to 100% in winter when the office airflow ratio is between 0.4 and 0.9 under weather conditions that prevail in Omaha, Nebraska. If the LAHU is not equipped with a humidification system, the optimal operation and control schedules can also improve relative humidity when the outside air temperature is between the LAHU Free-Cooling coincident outside air temperature and the office cold deck set point. The heat recovery system should not be designed for mild winter climates or for LAHUs with relatively higher office airflow ratio. Economic analysis of heat recovery installation is needed. The findings discussed in this paper also apply to any heat recovery system where AHU has an economizer function, like that in hospitals and schools. Fig. 5 Thermal energy savings of LAHU with optimal heat recovery control effectiveness=0.6 Free-Cooling Temperature as point B for office airflow ratio of 0.7, the optimal discharge air temperature is the outside air temperature, the circulation pump is turned off. The required AHU supply air temperature is maintained by modulating return and outside air flows. For example, the optimal control schedule for an airflow ratio of 0.7 turns off the heat recovery system when the outside air temperature is higher than 1.7 C 35 F for climate conditions prevailing in Omaha, Nebraska. With 2600 fewer hours of operation than the conventional schedule, the optimal LAHU heat recovery control may reduce annual pump energy consumption by kwh assume 25 horse power pump. When the outside air temperature is between the LAHU Free- Cooling coincident outside air temperature and the office cold deck set point, optimal heat recovery control also reduces the amount of outside air intake. Since the outside air humidity ratio is lower than the required room air humidity ratio, less outside air intake also improves the room relative humidity level if the LAHU has no humidification system. If the LAHU is equipped with a humidification system, the optimal heat recovery control schedules reduce humidification energy due to the reduced total outside air intake when the outside air temperature is between the LAHU Free-Cooling coincident outside air temperature and the office cold deck set point. Within this temperature range, the optimal control schedule significantly increases the potential energy savings. For example, the maximum potential energy savings can be as high as 20% when the heat recovery has an effectiveness of 0.6, as shown in Fig. 5. The findings discussed in this section also apply to any heat recovery system where AHU has an economizer function. For example, heat recovery systems serving AHUs with an economizer in hospitals and schools can be turned on at a much lower Acknowledgment The authors would like to express gratitude to Ms. Deborah Derrick for her editorial assistance. Nomenclature C p air specific heat, Btu/lbm F kj/kg K CL 2 ratio of humidification cost to mechanical cooling energy cost for conditioning the same amount of outside air d energy cost, $/Btu $/kj h air enthalpy, Btu/lbm kj/kg Ṁ airflow rate, lbm/h kg/s Q air thermal energy flow rate, Btu/h W q air thermal energy flow per unit mass, Btu/lbm kj/kg T optimal air temperature or temperature set point, F C t air temperature, F C W air moisture, lbm/lbm kg/kg outside air intake ratio heat recovery effectiveness office section relief airflow ratio ratio of office supply airflow rate to the total building supply airflow rate office section common exhaust airflow ratio Subscripts c air cooling cr critical e common exhaust eh the laboratory section fume hood exhaust env building envelope f free cooling h heating Journal of Solar Energy Engineering AUGUST 2006, Vol. 128 / 365

7 hg internal heat gain hr heat recovery hu humidification i entering air L capacity limit lec economizer in laboratory section m mixed air oa outside air ph preheat Rh reheat r room air re relief air s supply air,1 the office section,2 the laboratory section Superscripts + maximum minimum References 1 Cui, Y., and Liu, M., 2005, Improving Laboratory Buildings Energy Performance and Indoor Air Quality Using a Laboratory Air Handling Unit System LAHU, ASHRAE Trans., 111, pp ASHRAE handbook, 2000, 2000 ASHRAE Handbooks: HVAC Systems and Equipment, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc., Atlanta. 3 ASHRAE, 2001, 2001 ASHRAE Handbooks: Fundamentals, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc., Atlanta. 4 Cui, Y., and Liu, M., 2004, Optimal Airflow Control of Laboratory Air Handling Unit LAHU Systems, J. Sol. Energy Eng., 126, pp Thomas, G., and Finney, R., 1996, Calculus and Analytical Geometry, 9th ed., Addison-Wesley, Reading, MA. 6 Carnes, L., 1988, Air-to-Air Heat Recovery Systems For Research Laboratories, ASHRAE Trans., 90, pp Moyer, R. C., 1978, Energy Recovery Performance in the Research Laboratory, ASHRAE Trans., 84, pp Degelman, L., 1986, Bin Data Weather, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc., Atlanta. 366 / Vol. 128, AUGUST 2006 Transactions of the ASME