R&T 2004 - VFD Condensers - Reindl

8/18/2019 R&T 2004 - VFD Condensers - Reindl

1/23

Douglas T. Reindl

Director, IRC

University of Wisconsin-Madison

University of Wisconsin-Madison

Emerging Technologies:

VFDs for Condensers

Emerging Technologies:Emerging Technologies:

VFDs VFDs for Condensersfor Condensers


2/23

We’ve looked at VFDs on

Evaporators and compressors, what is thepotential for application on condensers?


3/23

Head Pressure Control

Our heat rejection system controls headpressure

Evaporative condenser fan controls

on/off (single speed fans)

two-speed fans

variable speed fans


4/23

Floating Head Pressure Control

Consequences of lowering head pressure

increased evaporative condenser energy use

decreased compressor energy use

reduced high stage compression (on average)

Does the decrease in compressor energy

use outweigh the increase in condenserfan energy use???


5/23

Optimum Head Pressure

82 84 86 88 90 92 94 960

20

40

60

80

100

120

140

160

Saturated Condensing Temperature (F)

P o w e r ( k W )

Compressor

Compressor+Condenser

Condenser

Axial Fan

Tcond,opt = 87.1 F

Toa,wb=78°F


6/23


82 84 86 88 90 92 94 960

50

100

150

200

250

300

Saturated Condensing Temperature (F)

P o w e r ( k W )

Centrifugal Fan

Compressor

Compressor+Condenser

Condenser

Tcond,opt = 89.9 F

Toa,wb=78°F


7/23


Depends on:

Condenser fan type (axial vs. centrifugal)

Fan control strategy

Condenser sizing strategy

95°F saturated condensing (historic)

90°F saturated condensing (recommended)

85°F saturated condensing (possible not practical)


8/23

Case Study:Cold Storage Warehouse

Size34°F 39,000 (ft²)

0°F 52,000 (ft²)

600,000 (lbs/day, food)

Type

ammonia, single-stagecompression, liquidoverfeed evaporators

Operating Costs

9,000 ($/month)

4 Compressors Available

Instrumentation

Temp, Pressure,

Mass Flow!

Defrost Strategies

Head Pressure Control

The refrigeration system examined as part of this case study is a cold storage warehouse

facility located near Milwaukee, WI. The facility contains four types of refrigerated spaces – low temperature freezer (0°F), cooler (34°F), docks (45°F), and ripening rooms (45-64°F).

From a thermal mass perspective, the warehouse construction type can be considered

“lightweight” for all spaces. There is mostly insulation and very little mass in the walls and

roofs.

The freezer and cooler with its loading dock are separate buildings located adjacent to each

other. The banana and tomato ripening rooms are located in a heated space adjacent to the

cooler. The refrigerant used throughout this system is ammonia (R-717). Evaporators in the

freezer are top fed, pumped liquid overfeed. Cooler, and cooler dock evaporators are all

bottom feed pumped liquid overfeed where as the evaporators in the banana and tomato

ripening rooms are direct expansion controlled by thermal expansion valves and back

pressure regulators.

(NOTE: This case study was conducted by Manske, K. A. in partial fulfi llment of the

requirements for a MS degree in Mechanical Engineering under the direction of

Professor’s Reindl , D. T., and Klein, S.A. during 1998-1999. Portions of the thesisprepared by Mankse titled “ Performance Optimization of Industrial Refrigeration

Systems” , 1999 have been excerpted for this section. A complete copy of the Manske

thesis is available for download at: http://www.irc.wisc.edu/publications


9/23

Case Study:Cold Storage Warehouse

Condenser

Qreject

Evap

Qspace

Evap

Qspace

23°F

-10°F

Evap

Qspace

45-55°F

BPR

D

X

PLO

PLO

Design Loads Yearly Average Loads

Fruit Ripening = 90 tons

Cooler = 107 tons

Freezer = 106 tons

Fruit Ripening = 43 tons

Cooler = 58 tons

Freezer = 71 tons

HPR

There are three main vessels in the system as shown above. The first is the high pressure

receiver where liquid refrigerant draining from the condenser is stored. Liquid refrigerant from

the high pressure receiver is then throttled to either the intermediate pressure receiver or to

the direct expansion evaporators in the banana and tomato ripening rooms. The temperature

of the refrigerant in the banana/tomato room evaporators is regulated at a desired level by use

of a back-pressure regulator. The back-pressure regulator then throttles the refrigerant gas tothe intermediate pressure receiver which is at a lower temperature/pressure. Liquid in the

intermediate pressure receiver is then either pumped to the cooler and cooler dock

evaporators or throttled again to the low pressure receiver. Liquid refrigerant from the low

pressure receiver is pumped to freezer evaporators with a mechanical liquid recirculating

pump. Liquid levels in the intermediate and low pressure receivers are maintained at a near

constant level by a pilot operated, modulating expansion valve controlled by a float switch

located on the receiver tank.

A single-screw (Vilter model# VSS 451 connected to the low temperature vessel) and

reciprocating compressor (Vilter model# VMC 4412 connected to the high temperature vessel)

operate in parallel, each compressing to a common discharge header and a singleevaporative condenser. The suction line from the low pressure receiver leads to the screw

compressor. The suction line from the intermediate pressure receiver leads to the

reciprocating compressor. Additional compressors, in parallel piping arrangements to the

primary compressors, can be brought on-line if the load exceeds the capacity of the primary

compressors.


10/23

Control Options

Single speed fan with on/off control

most common head pressure control method

set cut-in (e.g. 150 psig) & cut-out pressures(e.g. 140 psig)

simple control method but results in

higher energy consumption vs. two-speed or VFD

higher maintenance (fan motors & belts)

liquid management problems w/multiple condensers


11/23

Control Options – Cont.

2-Speed fan control set high speed cut-in (e.g. 160 psig)

low-speed cut-in pressure (e.g. 150 psig), andlow-speed cut-out pressure (e.g. 140 psig)

relatively simple control method but results in

higher capital cost compared to single speed fan option

lower energy consumption vs. single-speed but slightly higherenergy consumption compared to variable speed

yields less system transients compared to single speed

sequencing speed controls requires attention


12/23

Control Options – Cont.

Variable speed fan (VFD) set a target head pressure modulate fan speed

to maintain head pressure

a very simple principle & method to implement

highest capital cost alternative

lowest energy consumption control alternative

modulate all condensers the same in systems withmultiple evaporative condensers

results in smoother system operation with minimaltransients


13/23

Condenser Fan Control Map

Strategy Mode 1 Mode 2 Mode 3 Mode 4 Mode 5

Small Motor off on off on

Large Motor off off on on

Small Motor off off on

Large Motor off on on

Small Motor off on on on

Large Motor off off half-speed on

Small Motor off half-speed half-speed on on

Large Motor off off half-speed half-speed on

Small Motor off

Large Motor off 5

variable speed

variable speed

1

2

3

4

The above map provides five different strategies that could be used for an evaporative

condenser that is equipped with twin motors, two-speed fans, or variable speed fans. The

“modes” are indicative of changes in head pressure (either increasing as one moves from left

to right or decreasing as one moves from right to left).

For example, strategy 3 would work as follows. In mode 1 all fans are off. As the head

pressure rises, the system responds by energizing a small fan motor in attempts to maintain

system head pressure. If the head pressure continues to rise and the setpoint is not satisfied,

mode three is initiated by the start of the larger fan motor to half-speed. As the head pressure

rises further, mode 4 dictates that the larger fan motor is tripped to run at high speed. The

exact opposite sequence occurs as the head pressure falls.


14/23

Condenser Fan Control Options

The above figure illustrates the required fan energy (expressed as a percentage of full-load

fan power) as a function of the evaporative condenser capacity for the five strategies listed

previously. The least efficient option is the on/off control (strategy 1) while the most efficient

option is the variable speed drive option. The two-speed fan option yields nearly all of the

part-load power and capacity benefits of the variable speed option but with much less costly

equipment.

Notice that at zero fan power for all options, the capacity of the evaporative condenser is not

zero. This is due to the fact that natural convection will occur drawing air through the

condenser coils and rejecting heat yielding about 10% of the condenser’s heat rejection

capacity while the fans are idle. This assumes that the condenser coils are running wet i.e.

water continues to flow over the condenser coils.


15/23

Condenser Fan Controls

Source: Manske, K., 2000

May

Of course we do not want to just minimize the power of the evaporative condenser at the

expense of the system; consequently, we must look at the impacts or tradeoffs associated

with spending more energy on evaporative condenser fans vs. the reduction in compressor

power that accrues due to the lower head pressure.

The case study system had an oversized evaporative condenser. As a result, it was possible

to drive head pressures extremely low in the system. So low in fact that the incremental

expenditure of fan energy was not compensated for by an incremental reduction in

compressor energy demand.

The above plot shows the comparison between heat rejection system control strategies. The

point furthest to the left on the curves in the figure represents the system balance point head

pressure at which the condenser is operating at 100 percent capacity (for a given outdoor air

wet bulb and system load during a peak hour on a average day in May). Any further decrease

in condensing pressure would prevent the condenser from rejecting the required amount of

energy from the system. The figure shows that VFD fan control could save the system nearly

8% in combined compressor and condenser energy requirements if the head pressure were

raised to 125 psia. VFD fan control looses its advantages at low head pressures because thefans must run at near full speed most of the time anyway. At high head pressures the fans in

on/off control don’t stay on long because of the high rate of heat transfer that occurs.

However, at high head pressure an on/off control strategy would cycle the fans on and off

frequently which would cause excessive wear on the motors and fan belts. The figure also

shows that there is a different optimum head pressure for each type of condenser fan control.

It is also interesting to note that half-speed fan motors have energy requirements that are only

approximately one percent above the VFD motors at elevated head pressures. Since this

system has a minimum allowed head pressure of 130 psia, VFD and half-speed motors may

have very similar energy requirements for most of the year.


16/23

Optimum Head Pressure Control


This plot illustrates the preferred control head pressure control strategy for two different

evaporative condenser sizes. With an evaporative condenser sized for 95 F saturated

condensing temperature on a design day, the optimum head pressure is the lowest head

pressure achievable by running the evap condenser fans “full out”. If the condenser is

oversized (i.e. an oversized evap condenser is defined as one that yields a saturated

condensing temperature of 85 F on the design day), there is an optimum head pressure (i.e. ahead pressure greater than the minimum achievable that will minimize the combined power of

the compressor and condenser). In this case, the optimum head pressure is likely a function

of the outside air wetbulb temperature.

The dark set of lines is for the condenser that is currently installed in the system. The current

condenser is large enough to allow the system to balance out with a saturated condensingtemperature of 85°F on the design day. The compressor/condenser power with a smaller

condenser is given by the lighter colored line. The point furthest to the left on each line

represents the pressure at which the evaporative condenser has reached 100 percent

capacity. Given that the load is constant, it would be physically impossible to achieve a lower

head pressure without adding additional condensing capacity. Note, the above case assumesthat the refrigeration load is progressively decreasing during the winter months; however,

refrigeration load has little influence on the optimum head pressure.

Because of the presence of high temperature direct-expansion coils in the case study system,

the head pressure is not allowed to go below 130 psia. Therefore, the system cannot possibly

be operated at its ideal head pressure except for the months of June through September. It

must be operated above its optimum head pressure resulting in a slight excess of compressor

power.


17/23

50 55 60 65 70 75 80120

130

140

150

160

170

180

190

200

210

220

230

1.5x106

1.7x106

1.9x106

2.1x106

2.3x106

2.5x106

2.6x106

2.8x106

3.0x106

3.2x106

3.4x106

Outside Air Wet Bulb Temperature [°F]

O p t i m u m H e a d P r e s s u r e [ p s i a ]

Calculated Ideal Head Pressure (Variable Evaporator Load)Calculated Ideal Head Pressure (Variable Evaporator Load)

Curve Fit (Variable Evaporator Load)Curve Fit (Variable Evaporator Load)

minimum head pressure T

o t a l S y s t e m H e a t R e j e c t i o n [ B t u / h r ]

Calculated Condenser Heat Rejection (Variable Evaporator Load)Calculated Condenser Heat Rejection (Variable Evaporator Load)

Calculated Ideal Head Pressure (Constant Evaporator Load)Calculated Ideal Head Pressure (Constant Evaporator Load)

Calculated Condenser Heat Rejection (Constant Evaporator Load)Calculated Condenser Heat Rejection (Constant Evaporator Load)

as required by dx txv



When performing the calculations to identify the optimum condensing pressure for the year,

we discovered that the optimum condensing pressure had a near linear relationship with the

outside air wet bulb temperature. The above curve illustrates the relationship between

optimum head pressure and outside air wetbulb temperature (lower curve) over a range of

evaporator load conditions (corresponding variability in heat rejection is shown by the points

above). In the case of this system, a very simple linear relationship was developed thatallows a supervisory reset on the system head pressure given the prevailing outside air wet

bulb temperature according to the following:

Phead,opt = -27.6 + 2.55 * Twb

where Phead,opt is the head pressure corresponding to minimum system power in psia and Twbis the outside air wet bulb temperature in F. This relationship assumes that the condensers

have variable speed drives. Keep in mind that the above relationship needs to have a lower

bound as dictated by the characteristics of each given system.


18/23

Optimizing Head Pressure

1. Measure the outdoor air wet bulb temperature2. Note the current condensing pressure and system electrical demand

3. Reset the condensing pressure down 5 psig & allow system to equilibrate

4. Note the new system electrical demand

5. Continue steps 3 and 4 until the lower condensing pressure limit setpoint isreached

6. Plot the system electrical demand vs. the condensing pressure and note thecondensing pressure corresponding to point of minimum system electricaldemand

7. Plot that single “optimum” condensing pressure point on a optimumcondensing pressure vs. outdoor air wet bulb temperature curve

8. Repeat the procedure from 1-7 to more fully develop a curve analogous to

the figure given on the previous page.

Procedure for Determining Optimum Relation Between Condensing Pressure and Outdoor

Wetbulb

The trajectory of optimum condensing pressures for corresponding outside air wet bulb

temperatures as shown on the previous page is specific to the existing ammonia system.

Each system will have its own unique trajectory. However, the following procedure can be

used to empirically develop the trajectory of optimum condensing pressures. Note, this

procedure needs to be executed during off-design periods of the year (during relatively lower

outside air wet bulb conditions). The procedure also requires the ability to continuously

monitor the outdoor air wet bulb temperature, condensing pressure, and the engine room total

electrical demand. We also recommend that other system state variables (such as suction

pressures, superheat – if applicable, etc.) be monitored to ensure reliable system operation

during the procedure.

1. Measure the outdoor air wet bulb temperature

2. Note the current condensing pressure and system electrical demand

3. Reset the condensing pressure down 5 psig (35 kPa) and allow the system to equilibrate4. Note the new system electrical demand

5. Continue steps 3 and 4 until the lower limit in condensing pressure setpoint is reached

6. Plot the system electrical demand vs. the condensing pressure and note the condensing

pressure that corresponds to the point of minimum system electrical demand

7. Plot that single “optimum” condensing pressure point on a optimum condensing pressure

vs. outdoor air wet bulb temperature curve

8. Repeat the procedure from 1-7 to more fully develop a curve analogous to the figure given

on the previous page.

Once the optimum condensing pressure trajectory curve is developed, it can be programmed


19/23

Economic Benefits of Drive

Annual savings for warehouse

13 kW (peak) reduction

97,140 kWh reduction (~5%)

$3,856 per year in electrical operatingcosts (~5%)

Drive cost = $6,900

Simple payback of 1.8 yrs


20/232

Final Thoughts

VFDs on condensers can provideeconomic and operating cost benefitson the high-side

Take advantage of lowering headpressure

Consider barriers to lowering headpressure


21/232


Head pressure limits dictated by: hot gas defrost requirements

setting of defrost relief valves

sizing of hot gas main

condensate management in hot gas main

DX evaporators

most thermostatic expansion valves need at least75 psig differential pressure to function properly

liquid injection oil cooling check manufacturer’s requirements for TXV

pressure differential

As with most things, there are limits to lowering system head pressure. We do not want to

create problems by trying to improve the efficiency of our systems. The above items are

some of the more common factors constraining or limiting our ability to lower system head

pressure. Keep in mind that these items may not necessarily be unmovable barriers;

however, changes in components or system arrangements may be required to overcome their

limiting effects on the system.

Hot Gas Defrost:

Many industrial refrigeration systems utilize hot gaseous refrigerant to defrost evaporators. In

cases where defrost relief valves are installed, a sufficient pressure differential (e.g. 75 psig)

across the valve must be created to open the valve. Sizing of the hot gas main may also

impose constraints. If a hot gas main is undersized, hot gas (at a sufficient rate) will not be

delivered to the evaporator without a high differential pressure. For larger size hot gas mains,

a much lower differential pressure will allow adequate flow of hot gas to defrosting

evaporator(s). Finally, if condensate is not properly managed in hot gas mains, hydraulic

shock can cause catastrophic failures of hot gas piping on a call for defrost. Also, the

condensate effectively decreases the pipe size causing similar symptoms as an undersizedline with regard to head pressure requirements. All of these deficiencies can be overcome in

the long run; however, they do create real barriers to lowering head pressure in the short run.

DX Evaporators:

In systems that utilize direct-expansion evaporators, a minimum differential pressure is

required across the thermostatic expansion valves (TXV). The minimum pressure differential

is dependent on the specific valve selection but is routinely on the order of 75 psig. When we

drop the pressure differential across the TXV below the minimum, we lose controllability of the

valve (control engineers call this “control authority”). What results is an inability to properly

modulate refrigerant to the evaporator. Since our evaporator pressure i.e. downstream of the

TXV the pressure is fixed (to satisfy our temperature requirements for meeting load), the head


22/232


Head pressure limits dictated by: evaporative condenser selection

oversized evaporative condensers results in anoptimum head pressure that depends on outdoorair wet bulb temperature

evaporative condenser fan controls

VFD fans are preferred but 2-speed fans yieldconsiderable benefits

thermosiphon oil coolers

Evaporative Condenser Selection:

If an evaporative condenser is too small, the system head pressure will rise until its heat

rejection capacity is sufficient to reject the needed heat from the system.

Fan Controls:

Although fan controls themselves do not necessarily limit head pressure, there are methods of

fan controls that lead to more stable and efficient system operation. Two speed condenser

fans or variable frequency drive (VFD) fans have better capacity modulating capability and

result in more stable head pressures – leading to more stable system operation. In addition to

their stability, two-speed and VFD controlled fans will result in improved energy due to their

better part-load performance as compared to single speed fans.

Thermosiphon Oil Cooling (TSOC):

TSOC improves compressor efficiency by using a thermosiphon effect coupled with thesystem’s evaporative condenser to reject heat from the compressor’s oil.


23/23


Head pressure limits dictated by: hand expansion valve settings

significantly lowering head pressure will likely requireseasonal HEV adjustments

this constraint can be overcome by the use of motorizedvalves or pulse width valves

oil separator sizing

gas driven systems (transfer systems & gas pumpers)

controlled-pressure receiver setpoints

heat recovery engineering and operations (knowledge & willingness)

Documents

R&T 2004 - VFD Condensers - Reindl