Printer Friendly

Dual Temperature Chilled Water Plant & Energy Savings.

In humid climates, the conventional paradigm is to determine the supply air temperature from the cooling coils that will result in the desired dew-point temperature in the space, and then determine the chilled water supply temperature that can deliver that temperature, based on the approach temperatures achievable with conventional equipment. While the conventional setpoints for supply air temperature at 55[degrees]F (12.8[degrees]C) and for chilled water at 44[degrees]F (6.7[degrees]C) may vary slightly with occupancy type and distribution system, they have become the standard for many types of air-conditioning systems and for rating air-conditioning equipment.

The reasons for the selection of 55[degrees]F (12.8[degrees]C) as the supply air temperature is that the dew-point temperature of air at 75[degrees]F (23.9[degrees]C) at 50% relative humidity is approximately 55[degrees]F (12.8[degrees]C). Assuming some fan heat is added to the stream and that the air off the coil is almost completely saturated, the moisture ratio of the supply air is sufficiently low to overcome indoor latent gains and maintain the conditions in the space at the required dew-point temperature.

Dehumidification, however, is only one of the processes provided by the building air-conditioning system. The system must also provide adequate ventilation air for the occupants and keep the temperature of internal spaces within allowable limits for comfort and functionality. In humid climates, outdoor ventilation air must be dehumidified if space humidity conditions are to be maintained, thus requiring either adequate cold water or desiccant devices to remove the humidity.

However, maintenance of space temperature by removing heat gain from the occupied space does not require chilled water at the same low temperature as the dehumidification task. Systems that generate chilled water at a calculated temperature for the sensible cooling task save energy because the lower compressor lift entailed by the higher chilled water temperature reduces the amount of work required by the compressor to generate a given amount of cooling. This column will discuss the design of a chilled water system for a project in New York City that can generate water at a low temperature for dehumidification and at a higher temperature for sensible cooling, while minimizing the number of machines required to provide the client-mandated redundancy.

Effect of Chilled Water Supply Temperature on Chiller Performance

Raising the temperature at which chilled water is generated reduces the energy required to generate a unit of cooling, while lowering it increases the energy requirement. Table 1 shows the performance requirements of Standard 90.1-2016 for large centrifugal chillers with variable speed drives and the adjusted required performance when supplying 55[degrees]F (12.8[degrees]C) chilled water or 42[degrees]F (5.6[degrees]C) chilled water. These two temperatures for chilled water were selected because they are the highest that are effective for the designed tasks of the air-conditioning system on this project, sensible cooling and dehumidification.

These performance numbers were generated using the Standard 90.1-2016 specification for condenser water flow, 3 gpm/ton (0.05 L/s x kW) with an entering condenser water temperature of 85[degrees]F (29.4[degrees]C), to illustrate the impact of changing the chilled water temperature. For reasons related to construction economy and simplification of the waterside economizer controls, the chillers were ultimately selected with a condenser water flow of approximately 1.85 gpm/ton (0.03 L/s x kW) and an entering condenser water temperature of 82.5[degrees]F (28.1[degrees]C). The impact of these changes will be described later in this article.

Chiller Plant Design to Generate Dual Chilled Water Stream

The project in New York City required N+1 redundancy, which if applied to both sources of chilled water could mean at least two additional chillers and significant additional first cost. The proposed solution uses double-ended chilled water return and chilled water supply headers for the chiller plant. Chillers are attached to the headers with the provision that one chiller on one end of the headers will be dedicated to making 42[degrees]F (5.6[degrees]C) chilled water, while several large chillers on the other end of the header will be dedicated to making 55[degrees]F (12.8[degrees]C) chilled water. One chiller will be located between these two sets, and the headers surrounding it will be partitioned by valves, so this chiller (the "swing" chiller) can be on either the low-temperature or tempered supply side. The return header is similarly configured and partitioned.

The swing chiller provides N+1 redundancy for both low or tempered chilled water supply. All chillers will be equipped with variable speed drives, and the chiller that is designated for "swing" duty was selected for variation in operating temperature range and chilled water flow. Chillers dedicated to either tempered or low-temperature chilled water are customized for that particular use. The low-temperature chiller is nominally 800 tons (2813 kW), while the tempered chillers are nominally 1,000 tons (3516 kW). The "swing" chiller is nominally 800 tons (2813 kW) at low temperature and 1,000 tons (3516 kW) at the higher temperature.

While the three chiller types are very similar, both the swing and tempered water chillers have larger evaporator shells and tube bundles to accommodate the higher chilled water flow rate required, not only by the larger capacity, but also by the inherently lower chilled water temperature rise available with the sensible only water-to-air heat transfer devices. The increased capacity of the chillers supplying tempered chilled water gives them a very low first cost per ton of refrigeration. The performance of the three chiller types with the actual condenser water temperature difference is shown in Table 2 (Page 71). Note that the performance of the "swing" chiller is slightly worse for a specific chilled water temperature than for the chiller customized to that temperature.

One of the advantages of this separation of chilled water sources and uses is that it greatly facilitates meeting the waterside economizer requirement of Standard 90.1-2016, Section 6.5.1.2.1, requiring that 100% of the sensible load be met by the economizer at a 50[degrees]F dry bulb/45[degrees]F wet bulb (10[degrees]C/7.2[degrees]C) condition that has been identified as a very difficult requirement in most climate zones. (2)

The pumping design is configured as a primary only, variable flow design. Variable speed pump sets with multiple pumps are used for the tempered chilled water supply, the low-temperature chilled water supply, and the two condenser water systems to accommodate a very wide flow variation in these systems. The heat exchanger for the waterside economizer conditions only the return chilled water on the tempered side of the loop, because, when it is in operation, outdoor humidity is sufficiently low that no dehumidification is required of the supply air. A diagram for the system is shown in Figure 1.

Design of the Air System to Use Chilled Water Dual Stream

An air system that can take advantage of the dual temperature chilled water system should separate airflow into two streams: one that includes outdoor ventilation air and that will be conditioned using the low-temperature chilled water source, and one that operates only on recirculated room air using the tempered chilled water stream to provide sensible cooling only. Most commonly, these systems use a dedicated outdoor air system (DOAS) with the supply air chilled and/or dehumidified adequately to meet the internal latent load with the required outdoor air ventilation flow rate. With a sufficiently low coil face velocity, an appropriately robust cooling coil, and energy recovery on the inlet side to reduce the enthalpy of the air onto the coil, the DOAS should be able to achieve an approach of no more than 6[degrees]F (3.3[degrees]C) of the supply air to the entering chilled water temperature. Supply air with the resulting low dew point created by a DOAS, so configured, should be able to maintain the required space humidity condition, unless there is significant process latent gain in the space.

For this project, the selected system is an underfloor air-distribution system with DOAS and recirculating air towers with chilled water coils to distribute air to the access floor supply plenum. Supply air is supplied to the floor at 66[degrees]F (18.9[degrees]C), to minimize both the risk of overcooling occupants and the loss of cooling through "thermal degradation" due to heat gain through the structural floor slab. At the perimeter, a continuous line of in-floor hydronic convectors, controlled by air valves, can re-cool the supply air down to 60[degrees]F (15.6[degrees]C) to minimize airflow required to meet envelope (solar and conduction) loads. The two-pipe risers serving the convectors are connected to four-pipe changeover valves at the riser base, so the convectors can provide both recooling and heating at the perimeter. A diagram for the air system is shown in Figure 2.

Both the cooling coils in the air towers and the perimeter in-floor convectors use chilled water from the tempered loop, because condensation is not desired at either location. In fact, condensation at the perimeter would be a disaster. The chilled water use of this system is similar to that of a chilled beam system, either passive or active, that uses a DOAS to provide highly dehumidified ventilation air and requires a minimum 55[degrees]F (12.8[degrees]C) chilled water temperature, to avoid raining condensate down onto the work surfaces.

The dual temperature chilled water concept could also be used with variable flow fan coil systems. The lower pressure drop through a dry coil would enable the small fan coil motor to move air through a relatively more robust coil, allowing a close approach of supply air temperature to the entering chilled water temperature, and minimizing any additional fan energy entailed by a higher supply air temperature.

Effect of Higher Temperature Chilled Water on Cooling Transport Energy

Robust coils for those portions of the system providing sensible cooling can provide a close approach of the supply air temperature from the unit, with minimal or no additional air pressure drop through the system because the coils are dry. The resulting supply airflow rate often is no more than with a conventional system cooled by 44[degrees]F (6.7[degrees]C) chilled water. The most important transport energy disadvantage for the tempered chilled water system is pumping energy for the system because of the difficulty of achieving the same high chilled water temperature rise through a cooling coil that can be achieved with a low-temperature chilled water system.

High efficiency cooling systems today typically target at least 15[degrees]F (8.3[degrees]C) temperature rise in the system, and Standard 90.1-2016 now requires that rise as a minimum. (3) About the best that can be achieved with the air tower is approximately 11[degrees]F (6.1[degrees]C), while the highest temperature rise that can be achieved through the perimeter in-floor convectors is about 7[degrees]F (3.9[degrees]C). The resulting total temperature rise across the chilled water system is about 9.5[degrees]F (5.3[degrees]C). Thus, the anticipated chilled water pumping energy for the tempered loop is almost 60% greater than for the conventional system.

On the other hand, the low-temperature chilled water loop has potentially a greater temperature rise of 18[degrees]F (10[degrees]C), resulting in a 12% decrease in pumping energy. The expected net, given the greater annual run-hours of the tempered chilled water system is an increase in annual pumping energy for the dual temperature system.

Heat rejection energy, on the other hand, is expected to be reduced for the dual temperature system, not only because of reduced compressor heat that must be rejected, but also because the cooling tower fans will run at lower speeds during many hours of the year while in waterside economizer mode. Less cooling tower airflow is required to produce 55[degrees]F (12.8[degrees]C) chilled water than lower temperature chilled water.

Conclusion

An annual energy model was prepared for one floor (36,000 net [ft.sup.2] [3,346 [m.sup.2]]) of a proposed, very densely occupied, 12-story "building-in-building" fit-out in a new office tower in Manhattan. The proposed design is as shown in Figure 2, providing 66[degrees]F (18.9[degrees]C) under-floor supply air that is recooled, at the perimeter, in summer to 60[degrees]F (15.6[degrees]C), and served by a central plant configured as shown in Figure 1. The baseline case is a conventional underfloor air system without perimeter recooling and providing 62[degrees]F (16.7[degrees]C) air to the plenum, using a chilled water plant generating 42[degrees]F (5.6[degrees]C) chilled water. Energy consumption for cooling production is reduced by about 28%, with large energy savings from both efficient chilled water production and increased hours of economizer operation, and a slight increase in chilled water pumping energy. Table 3 shows the output from the model for the different energy end-use components.

As a result of this design effort and analysis, the system has been approved for installation in a large office located in a new tower in New York City. The benefits of the project's energy savings were judged to far outweigh the small additional first cost required to implement the system.

BY DANIEL H. NALL, P.E., BEMP, HBDP, FAIA, FELLOW/LIFE MEMBER ASHRAE

Daniel H. Nall, P.E., FAIA, is vice president at Syska Hennessy Group, New York.

References

(1.) ANSI/ASHRAE/IES Standard 90.1-2016, Energy Standard for Buildings Except Low-Rise Residential Buildings, Table 6.8.1-1, p. 65.

(2.) Nall, D. 2014. "Waterside economizers and 90.1." ASHRAE Journal 56(8).

(3.) ANSI/ASHRAE/IES Standard 90.1-2016, Energy Standard for Buildings Except Low-Rise Residential Buildings, Paragraph 6.5.4.7, p. 100.

Caption: FIGURE 1 Flow diagram of dual temperature chilled water chiller plant.

Caption: FIGURE 2 Underfloor air distribution system with separate sensible cooling and dehumidification components.
TABLE 1 Effect of chilled water temperature on chiller
performance.

STANDARD        STANDARD 90.1-2016     STANDARD 90.1-2016
90.1-2016       55[degrees]F           42[degrees]F
STANDARD        CHILLED WATER          CHILLED WATER
CONDITIONS

FL       IPLV   [FL.sub.   [PLV.sub.   [FL.sub.   [PLV.sub.
                    adj]        adj]       adj]        adj]

KW/TON

0.585   0.390     0.463       0.309      0.611       0.407

TABLE 2 Chiller performance for different services.

TEMPERED CHILLED       "SWING" CHILLED             LOW-TEMPERATURE
WATER (55[degrees]F)   WATER (55[degrees]F/        CHILLED WATER
                       42[degrees]F)               (42[degrees]F)

[FL.sub.   [PLV.sub.      [FL.sub.     [PLV.sub.      FL    IPLV
adj]           adj]           adj]          adj]

KW/TON

0.451         0.242    0.467/0.590   0.250/0.387   0.575   0.377

COP

7.80          14.53      7.53/5.96    14.06/9.09    6.11    9.33

TABLE 3 Results of energy analysis of dual temperature
chiller plant.

                          42[degrees]F    DUAL TEMPERATURE
                         CHILLED WATER      CHILLED WATER
                        PLANT (MBTU/YR)    PLANT (MBTU/YR)

CHILLER ENERGY                   199.4              125.2
HEAT REJECTION ENERGY             51.0               41.2
CHILLED WATER PUMPS               14.9               23.9
                                 265.3              190.3
SAVINGS                             na              28.3%
COPYRIGHT 2017 American Society of Heating, Refrigerating, and Air-Conditioning Engineers, Inc. (ASHRAE)
No portion of this article can be reproduced without the express written permission from the copyright holder.
Copyright 2017 Gale, Cengage Learning. All rights reserved.

Article Details
Printer friendly Cite/link Email Feedback
Title Annotation:COLUMN: ENGINEER'S NOTEBOOK
Author:Nall, Daniel H.
Publication:ASHRAE Journal
Article Type:Column
Geographic Code:1USA
Date:Jun 1, 2017
Words:2555
Previous Article:Overview of Fluids For AC Applications.
Next Article:Inward Drive-Outward Drying: Walter Payton Does Permeance.
Topics:

Terms of use | Privacy policy | Copyright © 2021 Farlex, Inc. | Feedback | For webmasters |